\"\ FAT
     WATER POLLUTION CONTROL HESEARCH SERIES • 16130EES11/70
   RESEARCH ON
   DRY - TYPE COOLING TOWERS
   FOR THERMAL ELECTRIC
   GENERATION
   Part I
ENVIRONMENTAL PROTECTION AGENCY • WATER QUALITY OFFICE

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          WATER POLLUTION CONTROL RESEARCH SERIES

The Water Pollution Control Research Series describes the
results and progress in the control and abatement of pollu-
tion of our Nation's waters.  They provide a central source
of information on the research, development, and demon-
stration activities of the Water Quality Office, Environ-
mental Protection Agency, through inhouse research and grants
and contracts with Federal, State, and local agencies, re-
search institutions, and industrial organizations.

A triplicate abstract card sheet is included in the report
to facilitate information retrieval.  Space is provided on
the card for the user's accession number and for additional
uniterms.
Inquiries pertaining to the Water Pollution Control Research
Reports should be directed to the Head, Project Reports
System, Office of Research and Development, Water Quality
Office, Environmental Protection Agency, Washington, D.C. 20242.

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            RESEARCH ON  DRY-TYPE COOLING TOWERS FOR

              THERMAL ELECTRIC GENERATION: PART I
                                by

                         John P. Rossie
                               and
                       Edward A. Cecil

                   R. W.  Beck and Associates
               600 Western  Federal Savings Bldg.
                    Denver,  Colorado  80202
                             for the

                      WATER QUALITY OFFICE

                 ENVIRONMENTAL PROTECTION AGENCY
                      Project # 16130  EES
                      Contract # 14-12-823
                          November 1970
For sale by the Superintendent ol Documents, U.S. Government Printing Office, Washington, D.C., 20402 - Price $2.50

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                 EPA Review Notice
This report has been reviewed by the Water Quality Office,
EPA, and approved for publication.  Approval does not signi-
fy that the contents necessarily reflect the views and poli-
cies of the Environmental Protection Agency, nor does mention
of trade names or commercial products constitute endorsement
or recommendation for use.

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                                  FOREWORD
       The production of electrical power requires that large amounts of waste  heat
from the generating process be rejected  to a heat sink. The usual method of accom-
plishing heat rejection has been to use circulating water, either from a natural body
of water (once-through system) or from an evaporative-type cooling tower or cooling
lake, to carry away the waste heat.   The use of a once-through system results in the
addition of heat to the natural body of water.  The use of an  evaporative-type
cooling tower or cooling lake results in  the consumption of water to replace that
lost by evaporation  in the cooling process.

       A method of waste heat rejection by means of air-cooled heat exchangers,
which transfer heat  directly to the atmosphere without addition of heat to natural
bodies of water or evaporation loss of water, is available to the utility industry.

       In this report, information is presented on the theory of dry cooling as it
would apply to steam-electric generating plants; operating results are summarized
for several  existing  dry cooling tower installations; the comments of various equip-
ment manufacturers  are summarized; and the results of economic analyses made for
dry cooling systems are presented for  800-mw fossil-fueled and nuclear-fueled
generating units for 27 representative sites in the United States reflecting a range
of fixed-charge rates, fuel costs, and weather conditions.

       Following is a summary of certain of the more important conclusions reached
as a result  of the study:

       1 .     There is need for a method of disposing of waste heat from
              steam-electric generating plants which does not add heat
              to natural  bodies  of water or require large quantities of
              make-up water for evaporative-type cooling towers.

       2.     Steam-electric generating plants equipped  with dry-type
              cooling systems which discharge waste heat directly to the
              atmosphere are in successful operation in  Europe.  Two
              small  generating  units  in the United States are also
              equipped with dry-type cooling systems.

              In a number of such plants,  dry-type cooling systems were
              selected either as a  result of better economic evaluation
              as compared to evaporative-type cooling, or because of
              an insufficient make-up water supply for an evaporative-
              type cooling tower.

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3.    As a result of technology and experience gained with air-
      cooled heat exchangers in industry,  United States  manu-
      facturers can design and produce such dry-type cooling
      towers for generating plants. Air-cooled heat exchangers
      are now commonly used in the petroleum refining industry,
      and in petro-chemical  and chemical process plants.  Such
      air-cooled plants have  been built and are in operation with
      heat rejection  loads equivalent to  large steam-electric
      generating plants.   A  number  of plants dissipate up to
      2 billion Btu per hour, a heat rejection load equivalent to
      a 425-mw generating plant.

4.    The performance of a dry-type  cooling system is measured
      by  the temperature  difference   between  the condensing
      steam of the turbine exhaust and the ambient air entering
      the cooling coils (called "initial temperature difference",
      or ITD) required to reject the design heat load.

      The  capital cost of a  dry-type  cooling system increases
      with decreasing ITD; i.e., the capital cost of a 40°F ITD
      system will be higher than  the  capital cost of a 60°F ITD
      system for the same heat rejection load.  Conversely, more
      efficient turbine operation will be obtained with the lower
      ITD  (more expensive system).

5.    A generating  plant  equipped with a dry-type cooling
      system of optimum economic size will experience some
      loss of generating capability as a result of increased tur-
      bine back pressure during hot weather.

      In this report,  it was assumed that such lost capacity was
      replaced by means of  peaking  plants for a capital cost of
      $100perkw.  Other methods  of restoring capacity  of
      fossil-fueled plants  are  available including: removing
      feedwater heaters from service;  use of a second  steam
      admission point on the  turbine with increased boiler capa-
      city; and use of over-pressure throttle steam.  Because of
      reactor licensing limitations, such methods would not apply
      to nuclear plants.

6.    Turbine manufacturers  are currently performing research
      on a new line  of utility turbines especially designed for
      high back-pressure operation and are also studying the
      feasibility of modifying present designs to operate at the
                                n

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      high tack pressures that will  be  encountered with dry-
      type coolit,g tower operation.

7.    For a typical 800-mw  generating plant  in the Chicago
      area, the capital cost (based on 1970 price and wage
      levels), ITD,  loss of capability during hot weather  and
      cost of replacing such lost capability Tor economically
      optimized dry-type cooling systems are esT'iwiated to be
      as  follows:

                        Fossil-Fueled Plant

                                       Mechanical     Natural
                   Type of tower:          Draft          Draft

      Capital cost,  $/kw	       $17           $20
      Initial temperature difference ..       60°F         56°F
      %  loss of capability during
         hot weather	       7.6%        6.4%
      Penalty for loss of capability
         at $1 OOAw replacement* ...       $  8           $ 6
      Capital cost of dry tower
         system plus replacement
         peaking capacity	       $25           $26


                      Nuclear-Fueled Plant

                                       Mechanical     Natural
                   Type of tower:          Draft          Draft

      Capital cost,  $/kw	       $23           $27
      Initial temperature difference ..       65°F         62°F
      % loss of capability during
         hot weather	      13.6%        12.5%
      Penalty for loss of capability
         at $100Aw replacement* ...       $14           $13
      Capital cost of dry tower
         system plus replacement
         peaking capacity	       $37           $40
      *On  the basis of 800-mw capacity (capital cost of peaking
       capacity required  to restore  lost capability during hot
       weather, $ divided by 800,000 kw).
                               in

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       The foregoing analysis is on the basis of 5,250,000
       annual production (75 percent plant factor).

 8.    The use of dry-type cooling system." *"'fh steam-electric
       generating plants will elimina** rhe need for a large
       supply of water as a bas'« site requirement and will re-
       sult in greater freedom of plant siting than has been
       possible.

 9.    Ther*  are large  deposits of coal and lignite in the United
       Sfates  which are not yet fully developed-notably in
      Arizona, Montana, North Dakota, Utah and Wyoming-
      which  lack sufficient local water supplies for the make-up
      requirements of evaporative cooling means.  Except for
      the use of dry-type cooling  systems, the alternatives avail-
      able for development of these coal and  lignite supplies for
      large generating plants are to bring water to the mine-
      mouth  plant sites or to transport the fuel to a plant site
      where water is available.

      The use of dry-type cooling  systems with mine-mouth
      generating plants in these areas opens up new possibilities
      for use  of the important fuel reserves.

10.   The results of the economic studies made in this report
      indicate that the total bus-bar power costs of a typical
      large fossil-fueled generating plant equipped with a dry
      tower cooling system will be approximately 0.48 mills
      per kwh higher than the total bus-bar cost, including
      fixed charges, of a similar plant equipped with an evap-
      orative-type cooling tower,  a difference of approximately
      7 to 10 percent.   When considered at retail level for
      residential service with all costs of generation, transmis-
      sion and distribution reflected, the increase in cost for
      the dry-type production will be about 2  to  5 percent,
      depending  upon rates.  A 2 to 5 percent increase in a
      $20 residential monthly electric bill is equivalent to 40$
      to $1.00, and an increase of even  this amount would not
     occur unless all generating plants in a utility system are
     cooled  by a dry-type system.   For industrial power service,
     the increase would be approximately from 2 to 6 percent.

     There are a number of possible savings available to a
     utility with dry-type cooling systems which would tend
                               IV

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to reduce or possibly offset the increased production costs
from a dry-type cooling system:

a.    Possible fuel cost savings as a result  of the greater
      flexibility of plant  location with a dry-type cool-
      ing system.  A savings of approximately 5<£ per
      million Btu in fuel would entirely offset the cost
      difference of approximately 0.48 mills per kwh
      estimated above.

b.    Possible transmission cost savings as a result of
      greater flexibility of plant location.

c.    Possible savings as a result of the economies of an
      additional unit at an existing facility where in-
      adequate water supply would otherwise rule out
      the addition.

d.    Possible savings in cooling water make-up when
      compared to an evaporative-type cooling tower
      plant.  For a cooling water make-up cost of $100
      per acre  foot, the water savings for the dry tower
      installation would approximate 0.2 mills per kwh.

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                            TABLE OF CONTENTS

Section     	Description	     Page

           FOREWORD	      ?

           LIST OF FIGURES	      vi

           LIST OF TABLES 	      vii

   I        INTRODUCTION 	       1

           Purpose of Report	       1

           Heat Rejection in Power Production	       1

           Existing and Estimated Power Generating Capacity and
           Requirements in the United States	       2

               Increased size of generating units and plants  	       3

           Water Requirements	       °

           Presently Used Methods of Rejecting  Heat from Gener-
           ating Stations  	

               Once-through circulating water systems  	       6
               Cooling lakes	       6
               Wet-type cooling towers  	       6
               Spray ponds, or spray canals	       8

           Consumptive Use of Water by Generating Stations	       8

           Recent Legislation Governing Thermal Discharges  to
           Natural  Waters 	       9

           List of Generating Plants Equipped with  Dry-Type  Cooling
           Towers in Operation and Currently Under Construction	      11

           Description of Dry-Type Cooling Towers	      13

           Conventional Evaporative-Type System	      13
                                     VI

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                            TABLE OF CONTENTS

Section     	Description	     Page

           Dry-Type Systems	      15

               Indirect system  	      15
               Direct system	      18
               Comparison of indirect and direct systems	      20

           Use of Air Cooling by Industry 	      20

               Extent of air cooling in industry	      21

           Use of Air Cooler with Refuse Incinerators  	      22

   II        FUNDAMENTALS	      24

           Design and Construction Considerations  	      24

           Codes and Testing	      24

           Fin Types  	      24

               General  	      24
               Tension-wound, footed fin	      27
               Embedded fin	      27
               Extruded fin	      27
               Wrapped-on overlapped, footed  fin	      27
               Plate-type fin  	      27

           Types of Air-Cooled Exchange Systems  	      28

           Theory of Heat Transfer from Air-Cooled Coils	      28

               Basic theory	      30
               Indirect system	      31
               Direct system	      33
               Design of ai r coolers	      37
               Initial temperature difference	      37
               Dry cooling tower heat balance  	      38

           Effectiveness—N.  Approach	      42
                                     VII

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                            TABLE OF CONTENTS

Section     	Description	     Page

           Theory of Thermodynamic Cycles 	      43

               The Carnot cycle	      43
               The Rankine cycle	      45
               Improvements to the Rankine cycle	      50

  III        PERFORMANCE	      52

           Performance of Dry-Type Cooling Towers 	      52

               Natural-draft towers	      53
               Mechanical-draft towers  	      57
               Tower performance for varying load and ambient
                  air temperatures	      57
               Design ITD	      64

           Performance of Turbine Used with Dry-Type Cooling Towers  .      67

               Effect of back pressure on heat rejection of turbine	      69

           Combining Performance of Cooling Tower and Turbine 	      71

           Comparison of Performance of Dry Tower and Conventional
           Cool ing Systems	      71

           Application  of Present Large-Turbine Design to
           Dry-Type Cool ing Towers	      76

               Available designs  	      76
               Possible future designs  	      78

           Use of Recovery Turbine with  Main Circulating Pumps 	      81

           Use of Multi-Pressure (Series-Connected) Direct-Contact
           Condensers with Dry-Type Cooling Towers 	      82

           Effect of Air Temperature at Site	      87

               Turbine  performance	      87
               Freezing	      89
                                    VIII

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                            TABLE OF CONTENTS

Section     	Description	    Page

               Auxi I iary power	      89
               Natural-draft cooling tower  	      90
               Mechanical-draft cooling tower	      90
               Cooling water for auxiliary purposes 	      90

           Effect of Precipitation and Humidity  	      90

               Rain	      90
               Hail	      91
               Sleet or snow	      91
               Humidity	      91

           Effect of Wind Velocity and Direction	      91

               Natural-draft cooling towers	      91
               Mechanical-draft cooling towers 	      93

           Effect of Dust	      93

           Effect of Radiation and Cloud Cover  	      94

           Effect of Topography	      94

           Effect of Elevation  	      94

  IV       STRUCTURES AND MATERIALS	      97

           General  	      97

           Reinforced Concrete Structure, Natural-Draft Tower  	      97

           Structural Steel Natural-Draft Towers	      97

           Design Loadings	      98

           Cost Comparison  	• • •      98

           Corrosion of Coils and Fins  	      99

               Marley Company —  Summary on Corrosion
               and Fouling  	      99
                                     IX

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                            T/^BLE OF CONTENTS

Section     	Description	       Page

               Hudson Products  	     101

                   Process Industry Air Cooled Heat Exchanger
                      Experience Record	     101
                   Extended Surface Materials and Corrosion
                      Resistance Properties	     102
                   Aluminum Fin Corrosion and Its Prevention	     103
                   Protective Coatings	     105
                   Simulated Corrosion Tests	     106
                   Fin Surface Foul ing	     106
                   Power Plant Operation	     106

           Effect of Corrosion on Performance  of Coi Is  	     107

  V       AUXILIARY EQUIPMENT	     108

           General 	     108

           Condensers	     108

           Air Removal Equipment	     HO

           Pumps  	     112

           Recovery Turbi nes	     H2

           Auxiliary Cooling	     113

  VI       DRY-TYPE COOLING TOWER USE WITH BINARY CYCLES .     117

           General 	     117

           Description of Steam-Ammonia Binary Cycle  	     117

           Conclusions 	     117

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                            TABLE OF CONTENTS

Section	      Description       	     Page

  VII       METEOROLOGICAL  CONSIDERATIONS  	     120

           Possible Effects of Dry-Type Cooling Towers on Local
           Meteorological Conditions	     120

               Air temperature	     120
               Cloudiness  	     120
               Fog  	     120
               Precipitation 	     121
               Air currents  	     121

           Dry-Type Cooling Towers and Air Pollution 	     121

           Comparative Effects of Various Cooling Methods	     122

               Once-through cooling  	     122
               Cooling ponds 	     122
               Wet (evaporative) type cooling towers	     122
               Natural-draft versus mechanical-draft towers	     122

           Conclusions 	      123

  VIII      DISCUSSION WITH MANUFACTURERS	      124

           Introduction	      124

           Dr. Ldszlo Heller and Hoterv 	      124

           M.A.N. (Maschinenfabrik Augsburg-Nurnberg) 	      126

           GEA - Gesel Ischaft  Fur Luftkondensation	      126

           English Electric Company	      129

           Brown  Boveri Corporati on	      130

           United States Turbine Manufacturers 	      130

           Hudson Products Corporation	      131
                                     XI

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                             TABLE OF CONTENTS

Section     	Description	    Page

           The Marl ey Company   	    131

           I ngersol I-Rand Company   	    133

           GKN Birwelco Limited   	    133

  IX        OPTIMIZATION PROGRAMS	    134

           I ntroduction  	    134

           Method of Analysis and Description of Tower
           Optimization Program	     135

           Factors Affecting the Economic Optimization of
           Dry-Type Cool ing Towers  	    139

                Performance related to ITD  	     1 39
                Capital cost of the dry cooling system 	     139
                Elevation  	     141
                Fixed-charge rate 	    141
                Ambient air temperatures   	    141
                Fuel costs  	    141
                Turbine performance  	    '41
                Auxiliary power requirements  	     '42
                Replacement of capacity losses  	    '42

           Method of Analysis and Description of the Economic
           Optimization Program  	   142

  X        RESULTS OF. THE ECONOMIC OPTIMIZATION 	   149

  XI        DISCUSSION OF RESULTS 	   191

           General	    191

           Effect of Fixed-Charge Rate . *	   193

           Effect of Fuel Cost	    193

           Effect of Air Temperatures 	    194

           Effect of Assumptions as to Lost Capacity   	     194
                                       XII

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                           TABLE OF CONTENTS

Section                         Description                           Page

  XII       ECONOMIC COMPARISON OF THE DRY-TYPE AND  THE
           EVAPORATIVE-TYPE COOLING SYSTEMS	       204

 XIII       REFERENCES 	       207

           APPENDIX FOREWORD	       211

           APPENDIX FIGURES 	       212

           APPENDIX TABLES	       214

 XIV       APPENDICES	       215

           Appendix A — Field Trips to Dry Cooling Tower Installations       215

              RUGELEY STATION	       215

                  Introduction	       215

                  Description of Station  	       215

                  Water Circuit	       216

                  Design Parameters	       220

                  Capital Costs of the Dry Tower	       220

                  Manpower Requirements of the Tower	       222

                  Winter Operation 	       222

                  Description of System Components	       223

                      Cooling coils	       223
                      Tower shel I	       223
                      Condenser	       223
                      Sector valves	       225

                  Auxiliary Power Requirements	       225
                                  XIII

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                            TABLE OF CONTENTS



Section     	         Description	      Page



                   Turbine Cycle Performance	       225



                   Corrosion Problems	       228



                   Effect of Wind on Tower Performance  	       228



                   Water-Side Chemistry 	       229



                   Maintenance	       229



                   Concl usion  	       231



               IBBENBUREN PLANT	       232



                   Introduction	       232



                   Description of Plant	       232



                   Water Circuit 	       234



                   Design Parameters	       236


                                                                       9^0
                   Capital Costs	       zjy



                   Manpower Requirements of the Tower 	       239



                   Winter Operation	       242



                   Auxi liary Power Requirements  	       242



                   Turbine Cycle Performance	       243



                   Corrosion Problems	       243



                   Effect of Wind on Performance	       243



                   Water-Side  Chemistry  	       245



                   Maintenance	      248



                   Concl usion   	      248
                                      XIV

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                           TABLE OF CONTENTS




Section                          Description                           Page




               VOLKSWAGEN PLANT	      249




                   Introduction	      249




                   Description of Station	      249




                   Condensation Circuit 	      251




                   Design Parameters	      253




                   Manpower Requirements of the Tower	      257




                   Freezing Problems	      258




                   Auxiliary Power Requirements	      261




                   Turbine Cycle  Performance 	      261




                   Corrosion Problems 	      261




                   Effect of Wind on Performance 	      262




                   Water-Side Chemistry	      262




                   Maintenance  	      262




                   Conclusion	      263




              GYONGYOS STATION  	      264




                   Introduction	      264




                   Description of Station	      264




                   Water Circuit	     265




                   Design Parameters	     269




                  Capital Costs of the Dry Tower	     270




                  Manpower Requirements of the Tower	     270
                                   xv

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                            TABLE OF CONTENTS

Section     	    Description	      Page

                   Winter Operation 	      270

                   Turbine Cycle Performance  	      272

                   Corrosion Problems 	      272

                   Concl usion	      273

               NEIL SIMPSON STATION 	      274

                   Introduction	      274

                   Description of Station	      274

                   Design Parameters	      276

                   Capital Cost	      276

                                                                       ")7f\
                   Manpower Requirements	      *'°

                   Winter Operation 	      276

                   Description of System Components 	      278

                        Air-cooled condensation system 	      278
                        Cooling coils	      278
                        Auxiliary power requirements 	      278

                   Turbine Cycle Performance  	      280

                   Corrosion Problems 	      280

                   Effect of Wind on Cooling Tower  Performance  ....      280

                   Maintenance 	      280

                   Conclusion	      280

           Appendix B — Engineering Weather Data 	      283
                                      XVI

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                           TABLE OF CONTENTS

Section     	Description	      Page

           Appendix C — General Specifications for Dry-Type
           Cooling System Applications	      312

           Appendix D — Testing Upon Completion of Project	      315

           Appendix E — Cooling System Cost Structure	      317

           APPENDIX REFERENCES	      321

  XV       ACKNOWLEDGMENTS	      322
                                   XVII

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                               LIST OF FIGURES

                                                                        Page

 1     National Power Survey Regions	        4

 2     Water Consumption Versus Wet-Bulb Temperature	        9

 3     Evaporative Cooling Tower Condensing System	       14

 4     Indirect Dry-Type Cooling Tower Condensing System With
       Natural-Draft Tower	       16

 5     Indirect Dry-Type Cooling Tower Condensing System With
       Mechanical-Draft Tower	       17

 6     Direct-Type  Cooling Tower Condensing System	       19

 7     Horizontal Air-Cooled Heat Exchanger 	       25

 8     Heat Exchanger — Fin Tube Types 	       26

 9     Air-Evaporative Cooled Heat Exchanger Systems  	       29

10     Temperature  Diagrams of Direct and Indirect Dry  Cooling Tower
       Heat Transfer Systems  	       32

11      Heat Transfer Effectiveness as a Function of Number of Transfer
       Units (Ntu) Crossflow Exchanger With Air Mixed	       44

12     Carnot Cycle Plotted on Temperature-Entropy Diagram	       45

13     Diagram of Rankine Cycle 	       47

14     Typical Flow  Diagram for  Regenerative Reheat Cycle  	       51

15     Coil Performance Versus Air and Water Flow	       54

16     Coil Performance Versus Water Flow, Tower  Height and Initial
       Temperature  Difference (ITD)	       56

17     Cooling Units Required for Mechanical-Draft Dry-Type
       Cooling System  	       58
                                     XVIII

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                               LIST OF FIGURES

                                                                         Page

18     Curve of Full Load Auxiliary Power Requirements Versus ITD —
       Mechanical-Draft Dry-Type Cooling System	       59

19     Natural-Draft, Dry-Type Tower Performance Capability With
       Variation of Initial Temperature Difference  	       61

20     Graph of Calculated Operating Characteristics (Predicted
       Performance) for Direct Air-Cooled Condensing System, Neil
       Simpson Plant, Wyodak, Wyoming (from GEA)	       63

21     Natural-Draft Dry-Type Cooling Tower Operating Characteristics       65

22     Dry-Type Cooling Tower System: Turbine Back Pressure Variation
       With Initial Temperature Difference (ITD) for Given Ambient Air
       Temperatures 	       66

23     Diagram  of Steam Expansion Line	       68

24     Dry-Type Cooling Tower and Turbine  Curves 	       70

25     Typical Average Monthly Temperatures, Dry and Wet Bulb  	       73

26     Comparison of Dry Tower and  Evaporative Tower Performance ....       74

27     Approximate Mean Monthly Temperature of Water  from Surface
       Sources for July and August  	       77

28     Estimated Turbine Generator,  Full Load, Heat Rate Variation
       With Elevated Exhaust Pressures 	       79

29     Pressure Head Diagram for Circulating Water System of Indirect
       Dry Tower Equipped With Water Turbine	       83

30     Circulating Water for 4-Flow  Exhaust Turbines With Surface
       Condensers	       85

31     Temperature-Pressure Diagram of Parallel and Series-Connected,
       Direct-Contact Condensers and Dry Cooling Towers 	       86
                                     XIX

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                               LIST OF FIGURES

                                                                        Page

32     Relation of Natural-Draft Dry-Type Cooling Tower Height at
       Various Elevations to Height at Sea Level	      96

33     Cross Section of Direct-Contact Condenser Used at Rugeley
       Station	       109

34     M.A.M. Direct-Contact Condenser	      Ill

35     Auxiliary Cooling by Steam-Jet Refrigeration  	      115

36     Flow Diagram of Binary Cycle With Dry Cooling Tower	      118

37     Direct Condensing System, Utrillas Power Station, Spain	      128

38     Cooling Tower  Dimensions as a  Function of Initial Temperature
       Difference  and  Elevation for Natural-Draft Cooling  Towers —
       Steel and Aluminum Construction — 800-Mw Generating Capacity..      137

39     Ground Area Requirement as a Function of Initial Temperature
       Difference  for Mechanical-Draft Dry Cooling Towers —
       800-Mw Unit	      138

40     Relationship of Dry Cooling System Capital Cost to ITD and
       Elevation  	      140

41     Typical Curves of Total Annual Cost (Cooling System, Peaking
       Capacity Loss Penalty and Total Plant Fuel)  Variation With ITD
       for Summer and  Winter Peaking Assumptions	      151

42     Economically Optimum Values of Initial Temperature Difference
       (°F) — Fossil-Fueled Generating Unit— Natural-Draft Tower  ....      153

43     Economically Optimum Values of Initial Temperature Difference
       (°F) — Fossil-Fueled Generating Unit— Mechanical-Draft Tower .      154

44     Economically Optimum Values of Initial Temperature Difference
       (°F)-Nuclear-Fueled Generating Unit-Natural-Draft Tower...      155

45     Economically Optimum Values of Initial Temperature Difference
       (°F) —Nuclear-Fueled Generating Unit —Mechanical-Draft
       Tower  	      156
                                      xx

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                               LIST OF FIGURES

                                                                        Page

46     Generating Capacity Losses as Percent of Rated Load for the Range
       of Economically Optimum Values of ITD Shown on Figure 42 —
       Fossil-Fueled Generating Unit — Natural-Draft Tower	     158

47     Generating Capacity Losses as Percent of Rated Load for the Range
       of Economically Optimum Values of ITD Shown on Figure 43 —
       Fossil-Fueled Generating Unit— Mechanical-Draft Tower  	     159

48     Generating Capacity Losses as Percent of Rated Load for the Range
       of Economically Optimum Values of ITD Shown on Figure 44 —
       Nuclear-Fueled Generating Unit— Natural-Draft Tower	     160

49     Generating Capacity Losses as Percent of Rated Load for the Range
       of Economically Optimum Values of ITD Shown on Figure 45 —
       Nuclear-Fueled Generating Unit — Mechanical-Draft Tower	     161

50     Capital  Cost of the Dry Cooling System ($/Kw) for the Range of
       Economically Optimum Values of ITD  Shown on Figure 42 —
       Fossil-Fueled Generating Unit— Natural-Draft Tower	     162

51     Capital  Cost of the Dry Cooling System ($/Kw) for the Range of
       Economically Optimum Values of ITD  Shown on Figure 43 —
       Fossil-Fueled Generating Unit— Mechanical-Draft Tower             163

52     Capital  Cost of the Dry Cooling System ($/Kw) for the Range of
       Economically Optimum Values of ITD  Shown on Figure 44 —
       Nuclear-Fueled Generating Unit —Natural-Draft Tower  	     164

53     Capital  Cost of the Dry Cooling System ($/Kw) for the Range of
       Economically Optimum Values of ITD  Shown on Figure 45 —
       Nuclear-Fueled Generating Unit —Mechanical-Draft Tower	     165

54     Capital  Cost of the Dry Cooling System ($/Kw) Plus Capital Cost
       of Peaking Capacity ($/Kw) for the Range of Economically Opti-
       mum Values of ITD Shown on Figure 42 — Fossil-Fueled Generat-
       ing Unit — Natural-Draft Tower	     166

55     Capital  Cost of the Dry Cooling System ($/Kw) Plus Capital Cost
       of Peaking Capacity ($/Kw) for the Range of Economically Opti-
       mum values of ITD Shown on Figure 43— Fossil-Fueled Generat-
       ing Unit — Mechanical-Draft Tower	     167
                                     XXI

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                               LI STOP FIGURES

                                                                        Page

56     Capital Cost of the Dry Cooling System ($/Kw) Plus Capital Cost
       of Peaking Capacity ($/Kw)  for the Range of Economically
       Optimum Values of ITD Shown on  Figure 44 — Nuclear-Fueled
       Generating Unit — Natural-Draft Tower	     168

57     Capital Cost of the Dry Cooling System ($/Kw) Plus Capital Cost
       of Peaking Capacity ($/Kw)  for the Range of Economically
       Optimum Values of ITD Shown on  Figure 45 —Nuclear-Fueled
       Generating Unit — Mechanical-Draft Tower	     169

58     Relationship of Economically Optimum Initial  Temperature
       Difference to Ambient Air Temperatures for the Sites Studied —
       Natural-Draft Dry Cooling System fora Fossil-Fueled 800-Mw
       Generating Unit  	     195

59     Relationship of Economically Optimum Initial  Temperature
       Difference to Ambient Air Temperatures for the Sites Studied —
       Mechanical-Draft Dry Cooling System for a Fossil-Fueled 800-Mw
       Generating Unit  	     1 96

60     Relationship of Economically Optimum Initial  Temperature
       Difference to Ambient Air Temperatures for the Sites Studied —
       Natural-Draft Dry Cooling System for a Nuclear-Fueled 800-Mw
       Generating Unit  	     197

61     Relationship of Economically Optimum Initial  Temperature
       Difference to Ambient Air Temperatures for the Sites Studied —
       Mechanical-Draft Dry Cooling System for a Nuclear-Fueled
       800-Mw Generating Unit	     198

62     Relationship of Economically Optimum Initial  Temperature
       Difference to Ambient Air Temperatures at Sea-Level Elevation —
       Natural-Draft Dry Cooling System for a Fossil-Fueled 800-Mw
       Generating Unit  	     199

63     Relationship of Economically Optimum Ini.tial  Temperature
       Difference to Ambient Air Temperatures at Sea-Level Elevation —
       Mechanical-Draft Dry Cooling System for a Fossil-Fueled  800-Mw
       Generating Unit	     200
                                     XXII

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                               LIST OF FIGURES

                                                                        Page

64     Relationship of Economically Optimum Initial Temperature
       Difference to Ambient Air Temperatures at Sea-Level Elevation —
       Natural-Draft Dry Cooling System for a Nuclear-Fueled 800-Mw
       Generating Unit  	   201

65     Relationship of Economically Optimum Initial Temperature
       Difference to Ambient Air Temperatures at Sea-Level Elevation —
       Mechanical-Draft Dry Cooling System for a Nuclear-Fueled
       800-Mw Generating Unit	   202
                                     XXIII

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                                LIST OF TABLES

                                                                         Page

 1     Predicted Increase in Future  Electrical Requirements	      3

 2     Estimated Number of Thermal Generating Plant Sites, 500-Mw
       Capacity and Above for Year 1990	      5

 3     Generating Plants With Dry-Type  Cooling Towers	     11

 4     Refuse Incinerators With Air-Cooled Condensing Systems 	     22

 5     Possible Variations in Back Pressure and Ambient Air for a
       Given ITD	     64

 6     Variations in 800-Mw Turbine-Generator Capability Due to
       Changes in Back Pressure  With a 60°F ITD Dry-Type Tower	     88

 7     Computer Printout — Natural-Draft Cooling Tower  System —
       Sizing and Costing Program	     136

 8     Heat Rejection Versus Back Pressure for an 800-Mw Generat-
       ing Unit (Full Throttle Flow Performance)	     144

 9     Computer Printout, Economic Optimization, 800-Mw Fossil-
       Fueled Generating Unit,  Natural-Draft Tower, Burlington,
       Vermont  	     148

10     Economic Optimization Analysis,  Summary of  Sites, Site Data
       and Study Assumptions  	     150

11     Economically Optimum Values of Initial  Temperature Difference
       (°F),  Fossil-Fueled Generating Unit, Natural-Draft Tower	     170

12     Economically Optimum Values of Initial  Temperature Difference
       (°F),  Fossil-Fueled Generating Unit, Mechanical-Draft Tower ....     171

13     Economically Optimum Values of Initial  Temperature Difference
       (°F),  Nuclear-Fueled Generating Unit,  Natural-Draft Tower	     172

14     Economically Optimum Values of Initial  Temperature Difference
       (°F), Nuclear-Fueled Generating  Unit, Mechanical-Draft Tower .  .     173
                                     XXIV

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                                   LIST OF TABLES

                                                                        Page

15     Capital Cost- of the Dry Cooling System  ($/Kw) for the Eco-
       nomically Optimum Values of Initial Temperature Difference
       Shown in Table 11, Fossil-Fueled Generating Unit, Natural-
       Draft Tower	     174

16     Capital Cost of the Dry Cooling System  ($/Kw) for the Eco-
       nomically Optimum Values of Initial Temperature Difference
       Shown in Table 12, Fossil-Fueled Generating Unit, Mechan-
       ical-Draft Tower	     174

17     Capital Cost of the Dry Cooling System  ($/Kw) for the Eco-
       nomically Optimum Values of Initial Temperature Difference
       Shown in Table 13, Nuclear-Fueled Generating Unit,
       Natural-Draft Tower	     176

18     Capital Cost of the Dry Cooling System  ($/Kw) for the Eco-
       nomically Optimum Values of Initial Temperature Difference
       Shown in Table 14, Nuclear-Fueled Generating Unit,
       Mechanical-Draft Tower  	     177

19     Capital Cost of the Dry Cooling System  ($/Kw) Plus Capital
       Cost of Peaking Capacity ($/Kw) for the Economically
       Optimum Values of Initial Temperature Difference Shown in
       Table 11, Fossil-Fueled Generating Unit, Natural-Draft
       Tower 	     178

20     Capital Cost of the Dry Cooling System  ($/Kw) Plus Capital
       Cost of Peaking Capacity ($/Kw) for the Economically
       Optimum Values of Initial Temperature Difference Shown in
       Table 12, Fossil-Fueled Generating Unit, Mechanical-
       Draft Tower 	     179

21     Capital Cost of the Dry Cooling System  ($/Kw) Plus Capital
       Cost of Peaking Capacity ($/Kw) for the Economically
       Optimum Values of Initial Temperature Difference Shown in
       Table 13, Nuclear-Fueled Generating Unit, Natural-
       Draft Tower	      180
                                    xxv

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                                LIST OF TABLES

                                                                        Page

 22     Capital Cost of the Dry Cooling System ($/Kw) Plus Capital
        Cost of Peaking Capacity ($/Kw) for the Economically
        Optimum Values of Initial Temperature Difference Shown in
        Table 14, Nuclear-Fueled Generating Unit,  Mechanical-
        Draft Tower	     181

 23     Optimized Total Annual Costs  (in Mills per Kwh)  Influenced
        by the Cooling System — 800-Mw, Fossil-Fueled Generating
        Unit, Natural-Draft, Dry-Type Cooling Tower System 	     183

 24     Optimized Total Annual Costs  (in Mills per Kwh)  Influenced
        by the Cooling System — 800-Mw, Fossil-Fueled Generating
        Unit, Mechanical-Draft, Dry-Type Cooling Tower System	     184

 25     Optimized Total Annual Costs  (in Mills per Kwh)  Influenced
        by the Cooling System — 800-Mw, Nuclear-Fueled Generat-
        ing Unit,  Natural-Draft, Dry-Type Cooling Tower System	     185

 26     Optimized Total  Annual Costs (in Mills per Kwh)  Influenced
        by the Cooling System — 800-Mw, Nuclear-Fueled Generat-
        ing Unit,  Mechanical-Draft, Dry-Type Cooling Tower System	     186
                    »
 27     Auxiliary Capacity Required (in Mw) for Cooling  System
        Pumps at the Optimum ITD for an 800-Mw, Fossil-Fueled Unit,
        Natural-Draft, Dry-Type Cooling Tower System	     187

28      Auxiliary  Capacity Required (in Mw) for Cooling System Pumps
        and Fans at the Optimum ITD for an 800-Mw,  Fossil-Fueled
        Generating Unit,  Mechanical-Draft, Dry-Type Cooling  Tower
        System	     188

29     Auxiliary  Capacity Required  (in Mw) for Cooling System Pumps
       at the Optimum ITD for an 800-Mw,  Nuclear-Fueled Generat-
       ing Unit,  Natural-Draft,  Dry-Type Cooling Tower System	     189

30     Auxiliary  Capacity Required (in Mw) for Cooling System Pumps
       and Fans at the Optimum ITD for an 800-Mw,  Nuclear-Fueled
       Generating Unit,  Mechanical-Draft, Dry-Type Cooling Tower
       System	     190
                                    XXVI

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                                   LIST OF TABLES

                                                                         Page

31     Initial Temperature Differences of Dry Cooling Systems,
       Existing Installations Visited  	     192

32     Effect of Fuel Cost on Optimum ITD (Chicago Fossil-Fueled
       Plant, 15% Fixed-Charge Rate)	     194

33     Effect of Peaking Capacity  Cost on Optimum ITD (Fossil-
       Fueled Plant, Chicago, Fuel Cost = 35$ per Million Btu,
       Fixed-Charge Rate - 15%)	     203

34     Monetary Considerations —Dry-Type and Evaporative-Type
       Cooling Tower Systems — Mechanical-Draft,  800-Mw —
       Northern United States	     205
                                     XXVII

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                                  SECTION
                                INTRODUCTION
Purpose of Report

        The purpose of this report is to present the results of research conducted by
R. W.  Beck and Associates in connection with the use of dry-type cooling towers
with steam-electric generating plants.  Dry-type cooling towers transfer the heat
of condensation of the turbine exhaust steam to the atmosphere by means of air-
cooled heat exchangers with no evaporation loss of circulating water to the atmos-
phere .

        Because of the growing shortage of large volumes of water for industrial and
power generation cooling services, the concern with the effects of adding heat to
natural  bodies of water and the consumptive use of water with  evaporative-type
cooling towers, it is important to have available in one publication a source of tech-
nical  information covering the  present state of the art of dry-type cooling towers.
The dry-type  tower has no consumptive use  of water by evaporation, nor does  it re-
quire that water of high salinity content be drained off from the cooling water cycle
and wasted, as is the  case with the conventional evaporative cooling tower.

        Nearly all of  the technology associated with the dry-type tower for steam-
electric generating plants has been developed in Europe.   However, United States
manufacturers have adequate  know-how and experience in the  design and construc-
tion of liquid-to-gas heat exchangers in industry, especially in chemical and
refinery processes, to design and produce dry cooling towers.

        There are a number of steam-electric generating plants in successful opera-
tion in Europe with dry-type  towers, but to date only two small dry tower gener-
ating units have been constructed in the United States. The largest is a 20.18-mw,
nameplate capacity, generating unit of the Neil Simpson Plant of the  Black Hills
Power and Light Company at Wyodak,  Wyoming, placed into service in  1969. This
was preceded by a 3-mw unit installed  in 1962 at the same location.

Heat Rejection in Power  Production

        The production of electrical power requires that enormous amounts of waste
heat be rejected.  In  the case of the conventional fossil-fired steam-electric  gen-
erating unit,  the waste heat is rejected partly to the atmosphere  in the form of
products of combustion from the steam-generating equipment, but the larger part is
rejected to the cooling-water circuit.

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        By far the greatest heat rejection is from the main steam condenser.  Other
 minor heat rejections are from the generator I  R losses and the mechanical losses
 from  the turbine and auxiliary rotating equipment.  For a modern fossil-fired plant,
 approximately 4,800 Btu are rejected to the circulating water for each kwh  of
 energy  produced.

        With a pressurized-water or boiling-water nuclear generating  plant,  the
 heat'rejection to the  circulating water is approximately 50 percent greater than for
 a fossil-fueled plant.  However, the use of the high-temperature gas-cooled  reactor
 nuclear plants will result in waste  heat rejections to  the circulating water which
 are comparable to those  experienced by fossil-fueled generating plants.

        When it is realized that the  heat rejection to the circulating water from a
 modern  fossil-fueled plant is equivalent to approximately half of the fuel burned in
 the boiler, and the heat rejection from a typical nuclear  plant amounts  to approxi-
 mately two-thirds of the nuclear heat generated, one can appreciate the enormity
 of the thermal problem.

 Existing and  Estimated Future Power Generating
 Capacity and Requirements in the United States

        The  National  Power Survey (1), a report written by the Federal Power Com-
 mission, has  projected that the electrical-energy requirements of the United States
will increase from 1 .6 trillion kwh in 1970 to 2.8 trillion kwh in 1980—a 75 per-
 cent increase in 10 years.  The  report also predicts that by the year 1980, 87 per-
cent of  the energy will be generated by either fossil- or nuclear-fueled  plants in
the ratio of  68 percent fossil  fuel to 19 percent nuclear fuel. At the present,
either once-through condenser cooling  (using natural bodies  of water) or evapora-
 tive-type cooling  towers are  used for generating station heat dissipation. However,
presently unused water, formerly available for evaporative-type cooling purposes,
is rapidly decreasing because of other higher priority uses.  In addition, large
blocks of power generation are creating increasingly undesirable thermal pollution
problems in once-through condenser  cooling  installations.  Future increases in
power generation will  place a great  strain upon our available water supply.  There-
fore,  waste heat removal from the projected  increase of generation will  require
that new methods be investigated for its disposal.  Table 1 (2) illustrates  the increase
in peak  demands and energy requirements from  1970 to 1990.

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                                    TABLE 1

                Predicted Increase in Future Electrical Requirements

                            Ratio of 1980 to 1970       Ratio of 1990 to 1970
                               Peak                       Peak"
       Region                 Demand  Energy             Demand   Energy

       Northeast                1.8      1.8                 3.2      3.2
       East Central              1.9      1.8                 3.4      3.4
       Southeast                2.1      2.2                 4.1      4.1
       West Central             2.0      2.0                 3.8      3.8
       South Central            2.3      2.3                 4.5      4.7
       West                    2.0      2.0                 4.0      4.0

       The areas comprising the six  National Power Survey regions  are shown in
Figure 1  (2).

       Increased size of generating  units and plants. Traditionally, the economics
of capital construction costs and operating  efficiency have resulted  in  a trend to-
wards larger generating units.  Construction costs per kw of unit capacity decrease
with unit size,  and plant labor requirements are more nearly  proportional  to
machine units than to plant kw capacity.  Although a small unit is not inherently
less efficient than a  large unit, the costs required  for adding  specific features re-
sulting in  higher net plant efficiencies are  prohibitive for small units.  Examples
are high pressures, high temperatures, reheat, superheat and  automated features.
From a maximum size of approximately 200 mw  at the end of  World  War II, the size
of generating units ordered has increased to 1,300 mw, and utility industry leaders
predict that, by 1990, units of 2,000-mw size will be in use.

       In recent years, a number of electrical  utilities have  formed power pools  in
which the utilities join in the construction  of generating units larger than any which
the individual utilities could accommodate alone.  Since large blocks of power can
be transmitted long distances to load centers of widely separated pool  members,
the construction of extra-high-voltage transmission systems has contributed greatly
to the feasibility of such large generating units.  Table 2 illustrates the trend to-
wards larger sized units and generating plants.

       From the foregoing,  we can only conclude that the problem of disposal of
waste heat from steam-electric generating  plants-will become more  acute  in the
future.

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FEDERAL POWER COMMISSION
POWER SUPPLY AREA

REGIONS SELECTED FOR UPDATING
THE NATIONAL POWER SURVEY
         FIGURE  I- NATIONAL  POWER  SURVEY REGIONS  (2)

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                                                                                    TABLE 2

                                                                 Estimated Number of Thermal Generating Plant Sites
                                                                    500-Mw Capacity and Above for Year 1990
                                                                    Fossil-Fueled Plants by Mw Capacities
Oi
 Northeast:
      Total Sites
      New Site	
      Cooling Towers .

 Southeast:
      Total Sites	
      New Sites
      Cooling Towers .

 East Central:
      Total Sites	
      New Sites
      Cooling Towers .

 South Central:
      Total Sites	
      New Sites
      Cooling Towers .

 West Central:
      Total Sites
      New Sites
      Cooling Towers .

 West:
      Total Sites	
      New Sites
      Cooling Towers .

Total U.S.:
      Total Sites
      New Sites
      Cooling Towers .
                                                                500
                                                                 to
                                                               1,000
                                                                 22
                                                                 15
                                                                  3
                                                                 30
                                                                  5
                                                                  7
                                                                 37
                                                                 17
                                                                 17
                                                                 11
                                                                  1
                                                                  1
                                                                 U
                                                                 4
                                                                 10
                                                                129
                                                                30
                                                                38
1,000
  to
2,000
  15
   1
   5
  12
  3
  1
 26
  9
  8
 31
 12
  7
 12
  5
  4
 19
  4
 10
115
 34
 35
                                                                      2,000
                                                                       to
                                                                      4,000
23
19
 3
                   Over
                   4,000
45
25
 9
                  Total
                                                                                                        41
                                                                                                         5
                                                                                                         8
                    34
                     6
                     4
                    62
                    16
                    17
 92
 49
 27
                    25
                     6
                     6
                    38
                     9
                    21
292
 91
 83
                    Nuclear-Fueled Plants by Mw Capacities
                 "500T700057000
                  to        to        to       Over
                 1,000    2,000     4,000     4,000     Total
                                      7
                                      6
                                      2
                  10
                   8
                   7
                                               28
                                               23
                                               15
                            19
                            18
                             4
                                                         22
                                                         18
                                                         14
                            10
                             8
                             4
                             6
                             4
                             4
73
63
33
                                                                   17
                                                                   14
                                                                    3
          21
          14
           7
                                                                    9
                                                                    5
                                                                    3
                                      9
                                      9
                                      4
73
58
19
                                                                            13
                                                                            12
                                                                             2
26
20
                              45
                              38
                               9
                    60
                    45
                    32
                                                                                      21
                                                                                      17
                                                                                       6
                                                                                      22
                                                                                      22
                                                                                       5
                             19
                             11
                              8
                             33
                             31
                             15
200
164
 75

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 Water Requirements

         Almost1 one-half of all water utilized in the United States is used for indus-
 trial cooling, including cooling water for condensing steam in power plants (3).
 Of the estimated 50  trillion gallons of water used by industry  in 1964, approximately
 81 percent was for electrical power production.

         Depending upon the turbine cycle heat rate and the temperature rise of the
 circulating water selected in the plant design, approximately 300 to 900 gpm permw
 must be pumped  through the condensing system of a steam-electric generating plant.

        The heat of condensation of the turbine exhaust steam  is transferred  to the
 circulating water and ultimately to the atmosphere through one of several methods.

 Presently Used Methods of Rejecting
 Heat from Generating Stations

        Once-through circulating water systems.  Where sufficient volume of circu-
 lating water is available, as on a large river such as the Ohio or Missouri, circulat-
 ing water is often taken directly from the river by means of an intake system,
 pumped  through the condensers and then discharged back to the river at a location
 selected to prevent recirculation of the heated water back to the intake.

        Generally, the once-through circulating system is the  least expensive of
 the several  types used, and utilities have used this method of providing circulating
 water wherever conditions permit.

        Once-through systems are also used on natural lakes, ocean estuaries,
 rivers, including tidal rivers, flood-control reservoirs, water-storage reservoirs,
 navigational reservoirs and  hydroelectric lakes.

        Cooling lakes.  Another method of providing circulating water for the
 steam-electric generating plant is to construct a pond or lake for the specific pur-
 pose of providing a source of circulating water and for dissipating the heat of con-
 densation from the surface of the water.  Generally, the surface area of the lake is
 sized for approximately one acre per mw of generating capacity, with variations
 above and below this figure in specific instances.

        Wet-type cooling towers. Where insufficient water is  available for a once-
 through circulating water system, an evaporative (wet-type) cooling tower is often
 used to dissipate  the heat of condensation which has been transferred to the water
 in the condenser.  Until the  present concern  about thermal pollution became prev-
 alent, the use of evaporative cooling towers  with steam-electric generating  plants
 was generally with smaller plants and, primarily, in locations where insufficient
water was available.

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       In the evaporative-type cooling tower,  the circulating water is  pumped
through the surface condenser where it picks up the heat of condensation of  the
exhaust steam and is then pumped through the cooling tower where it is broken up
into small drops by splashing down through the "packing" or "fill" of the tower.

       More than 75 percent of the heat is removed by evaporation and the re-
mainder is  transferred to air by convection.  After passing down through the tower
fill, the waterfalls into a storage basin beneath  the tower and is recirculated through
the condenser.

       Evaporative-type cooling  towers are generally either of the crossflow design
in which the  air flows horizontally through the falling water or the counterflow
design in which the air flows upward through the water.

       Although  the majority of cooling towers constructed in the United States
have been  of the  mechanical-draft type which use motor-driven fans to move the
air through the tower, natural-draft type towers are also constructed which utilize
the stack effect of the tall  tower structure for the movement of air. Approximately
20 natural-draft towers are either in operation or  under construction in the United
States  (3).  The natural-draft towers,  usually hyperbolic in shape, are from 300 to
500 feet in height and over 300 feet in diameter at the base,  with a height-to-
diameter ratio of more than one but less than 1.5.  Natural-draft towers are com-
monly  used in Europe for steam-electric generating plants, where economics favor
their selection over the mechanical-draft type.

       During 1965 in the  North  Atlantic region  as defined by the Federal Water
Quality Administration (FWQA),  only one plant (100 mw) out of 101 utilized cool-
ing towers; in the Southeast region, two out of 61; in the Great Lakes region,  none
out of  54; and for the country as a whole, 116 out of 514 plants utilized cooling
towers (Federal Water Quality Administration, 1968).

       At  the present time, approximately 30 states  have  adopted legislation con-
cerning water temperature standards which have been approved by the Division of
Standards of the  FWQA  (4).

       The standards recommended by the National Technical Advisory  Committee
on Water Quality Criteria and the passage of state legislation controlling tempera-
ture rise can  be expected to result in a reduction  in the percentage of generating
plants  constructed for  use  with  the once-through circulating water system and an
increase in the use of evaporative water towers,  where make-up water for such
towers is available, and in dry-type cooling towers where make-up water is scarce
or expensive.

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        Spray ponds, or spray canals. One method of rejecting heat from the con-
denser is to use a spray pond in conjunction with a recirculating supply of condens-
ing water.  In this method of rejecting heat to the atmosphere, the circulating
water is sprayed into the air through discharge pipes and spray nozzles located above
a pond which serves as a reservoir for receiving and  holding the water  for recircula-
tion.

        Except for certain small installations, spray ponds are not generally used for
steam-electric generating stations.

Consumptive Use of Water by Generating Stations

        All of the above methods  of removing  heat from the condensing exhaust
steam of a turbine result in ultimate  rejection of the heat to the atmosphere.  Since
a certain amount of cooling in each  of these different methods is accomplished by
the process of evaporating water, all of the methods result in a certain consumptive
use of circulating water.  In addition to  losses by evaporation,  the cooling towers,
cooling ponds and spray ponds must waste a certain percentage  of the water circu-
lated in order to maintain the concentration of dissolved solids  to limits compatible
with operation without objectionable deposit of scale on the plant equipment.

        The amount of evaporation experienced is dependent upon a number of vari-
ables, including:

        1.     Wet-bulb temperature  of air.

       2.     Relative  humidity.

       3.     Cloud cover.

       4.     Wind speed.

       5.     Range of cooling.

        For a site in the northeastern United States, the evaporation-loss  curve
shown in Figure 2, assuming a relative humidity of 60 percent,  cloud cover of 70
percent, wind speed at 8 mph, and a 20°F range with wet-bulb temperature vary-
ing from 40 to 80 F, is reproduced from (5).  This curve shows the estimated water
consumption,  in  gallons per kwh, for the following cooling methods:

        1 .     Cooling pond with a surface area of 2 acres per mw.

       2.     Cooling pond with a surface area of 1 acre per mw.
                                       8

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UJ
z
o
       3.

       4.

       5.

       6.
Mechanical-draft evaporative cooling tower.

Spray pond.

Natural-draft evaporative cooling tower.

Natural lake or river.
                                                          2  ACRES/MW
     1.3 -i
                                                          I  ACRE/MW
                                           NATURAL LAKE
                                           OR RIVER
                                              MECHANICAL DRAFT C.T.
                                              SPRAY PONDS
                                              NATURAL  DRAFT C.T
                   50          60          70           80
                      WET BULB  TEMPERATURE -  °F

          FIGURE  2—WATER  CONSUMPTION  VERSUS
                 WET  BULB TEMPERATURE (5)
       At the present time, the water consumption from evaporation as a result of
power generation is estimated to be 10 gallons per day per person and  is expected
to increase at a faster rate than the growth rate of power production because sup-
plementary cooling systems, such as  cooling towers and cooling lakes, will be
utilized on a larger proportion of new generation  capacity in the future.

Recent Legislation Governing Thermal
Discharges to Natural Waters
       Although there is much controversy as to the effects of temperature upon
aquatic life, one fact is undisputed:

             Present and contemplated legislation sets definite
             temperature limits upon circulating water dis-
             charged from steam-electric generating plants.

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For an excellent discussion of the effects of temperature on aquatic organisms, the
reader is referred to (6).

       The following is a brief review of existing and proposed laws governing en-
       sntal questions associated with the utility industry  as reported  in (7).
vironmental questions
             A new law (H.R. 4148) would require the Atomic  Energy
             Commission to obtain assurances that a nuclear plant will
             operate in  conformity with applicable water quality
             standards.  This  is aimed primarily at the thermal pollu-
             tion problem.  Similar assurances would have to be ob-
             tained for other  electric power plants which require
             federal permits or licenses such as the many power plants
             which require  federal permits from the Corps of Engineers
             if their construction plans include structures on navigable
             waters.

             The National Environmental Policy Act passed in 1969
             could have a bearing on federal activities in the power
             field, for it requires  all federal agencies to consider en-
             vironmental factors in carrying out their programs.

             The Federal Power Commission licenses only nonfederal
             hydroelectric plants and major electrical transmission
             lines and has authority to weigh the recreation, wilder-
             ness, fish and wildlife, and esthetic values of  these
             projects.

             Only 20 states require licenses for new generating plants
             and most of them consider reliability and safety alone.
             However, there have been recent law enactments in some
             states  for control over power plant siting.  A 1968
             Maryland law requires public hearings to consider the
             effects of the plant on the environment,  including ther-
             mal effects. Washington, Vermont and Maine  have re-
             cently passed similar  legislation.

             The Federal Water  Pollution Control Act, as amended by
             the Water Quality Control Act of 1965,  authorizes the
             states and the Federal Government to establish water
             quality standards for interstate (including coastal) waters (8).
                                     10

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             The water quality standards submitted by the states are
             subject to review by the Department of Interior and, if
             found to be consistent with Paragraph 3  of Section 10
             of the Act, will be approved as Federal  Standards by the
             Secretary  of the Interior.

             Paragraph 3,  Section 10 reads as follows:

             "Standards of quality established pursuant to this sub-
             section shall be such as to protect the public health or
             welfare,  enhance the quality of water and serve the
             purposes of this Act.  In establishing such standards, the
             Secretary, the  Hearing Board, or the appropriate state
             authority shall  take into consideration their use and
             value for public water  supplies, propagation of fish and
             wildlife, recreational purposes, and agricultural, in-
             dustrial,  and other  legitimate uses."

             If a state does not adopt water quality standards con-
             sistent with the above paragraph, the Act provides the
             Secretary  with  the opportunity to set the standards.  In
             April, 1970,  as reported in  (4), 20 states had not yet
             received full  approval  of their water-temperature stand-
             ards.

List of Generating Plants Equipped with Dry-Type  Cooling
Towers in Operation and Currently Under Construction

       Table 3 shows a  listing of the major steam-electric generating plants, either
in operation or currently under construction, which are equipped with dry-type
cooling towers.

                                    TABLE 3

                 Generating Plants with Dry-Type Cooling Towers

                                                Type of              Year
        Location               Rating           Dry Tower        Commissioned

Rugeley, England*               120 mw       Heller                  1962

Ibbenburen, Germany*           150 mw       Heller                  1967

Wolfsburg, Germany*          3-50 mw       GEA  Direct           1961-67
                                       11

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                              TABLE 3 (continued)
Location
Grootvlei, South Africa
Gyongybs, Hungary*
Razdan, USSR
Wyodak, Wyoming, USA*
Utrillas, Spain
Quetta, West Pakistan
Bavaria
Windhok, South Africa
Switzerland
Luxemburg
Rome, Italy
Cologne, Germany
Sindelfingen, Germany
Worms, Germany
Chile
Ludwigshafen, Germany
Eilenburg, Germany
Dunaujvarus, Hungary
Rating
200 mw
2-100 mw
2-200 mw
3-220 mw
22 mw
3 mw
160 mw
7.5 mw
40 mw
3-30 mw
4.3 mw
13 mw
2-30 mw
28 mw
ll&15mw
5 mw
3.6 mw
38 mw
5.3 mw
16 mw
Type of
Dry Tower
MAN/Birwelco
(Indirect)
Heller
Heller
Heller
GEA Direct
Direct
GEA Direct
Baldwin-Lima-
Hamilton (Direct)
GEA Direct
GEA Direct
GEA Direct
GEA Direct
GEA Direct
, GEA Direct
GEA Direct
GEA Direct
GEA Direct
GEA Direct
Heller
Heller
Year
Commissioned
1971
1969
Under Constr.
1970-72
1969
1962
1970
1964
1960
1971
1969
1956
1957
1958
1 960-61
1962
1963
1966
NA
1961
*Visited during study.
                                     12

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Description of Dry-Type Cooling Towers

       General.  The use of air for condensing turbine  exhaust  steam  is not a new
concept since it is reported that condensation by air cooling has been used in small
industrial power plants for over 50 years  (9).  However, the application of air con-
densing to relatively large generating units has been limited, since the use  of air
condensation will generally  result  in an increase in construction costs.

       There are  two basic  types  of air-cooled condensing systems—the indirect
system and the direct system.  The indirect system utilizes a direct-contact con-
denser at the turbine to condense  the exhaust steam.   Water from the condenser is
pumped to the dry-type tower for  cooling and recirculation to the spray jets  in the
condenser.  The indirect system is often referred to as the Heller system since the
concept of the use of the indirect system of condensation by air cooling for use with
a steam turbine-generator was presented by Dr. Laszlo Heller at the World Power
Conference in Vienna in  1956 (10).  Dr. Heller, who is Head of the Department of
Energetics of  the Technical  University of Budapest, Hungary and also serves as
Director of the Hungarian engineering  firm called Hoterv (charged with the de-
velopment of dry towers), along with his assistant, Dr. L. Forgo, developed  and
perfected much of the special equipment required for use with the  air-condensing
systems—notably the heat exchanger coil,  the automatic  controls and  the direct-
contact condenser.   However, a strict  interpretation of the use of the term "Heller
system" would limit it to  an  indirect system using the  Heller-Forgo coil,  since at
least one indirect system  using other coil designs is under construction.

       In the direct system,  steam is condensed in the coils without the use of a
direct-contact condenser or  circulating water.

Conventional Evaporative-Type System

       An understanding of  the conventional evaporative (wet-type) tower cycle
is useful in considering the two types of air-cooled condensing systems.  Figure 3
shows the  schematic arrangement of an evaporative-type cooling tower serving a
condensing turbine.

       Condensing water is circulated  through the tubes of a surface  condenser
and carries away the  heat of condensation of the turbine exhaust steam.  The ex-
haust steam comes into contact with the exterior surfaces of the tubes, and con-
denses as it gives up heat to the water.

       The warm circulating water is piped to the evaporative cooling tower where
it flows over the packing or fill, which may be closely spaced strips of asbestos-
cement or wood, to break up the circulating water into small drops through which
air is pulled by the tower fan. By a combination of evaporation and convection,
                                       13

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   V  AIR  FLOW  /
                  FAN
                     -COOLING  TOWER
    A   &   As
          »    »   i \
/—I I	1  I	I  I	I L_l\
 y  '  *   *   6  VL
'  .—'  '	' '—' '	'  V^FILL OR PACKING
 X o   4   &    4   4 .
/I—I I	II	II	I L_J\
       AfR FLOW \\
 / ,               \
   <   44*446
                                    SURFACE
                                    CONDENSER
    CIRCULATING  WATER
          PUMP
CONDENSATE PUMP
                                                                   TO  BOILER
                                                               FEEDWATER CIRCUIT
                    FIGURE 3 —EVAPORATIVE COOLING TOWER
                               CONDENSING  SYSTEM

-------
the temperature of the circulating water is reduced and the water is again pumped
through the condenser in a continuous cycle.  The condensed steam (condensate) is
removed from the condenser by the condensate pump and returned  to  the  boiler
feedwater circuit.

Dry-Type  Systems

        An explanation of the two basic air-cooled condensing systems follows:

        Indirect system.  For the indirect dry-type cooling tower,  the principal
components are:

        1 .    A direct-contact steam condenser.

        2 .    Circulating water pumps.

        3.    Water-recovery turbine (optional).

        4.    Cooling coils.

        5.    Means for moving air across the coils; either a
             natural-draft tower or  a mechanical-draft fan.

        Figure 4  shows a diagram of the  indirect-type system with  a  natural-draft
tower.

        Either a mechanical-draft or  a natural-draft tower is used  with the indirect
system. The  choice is dependent upon the economics of each particular case,  and
such factors  as fuel cost,  comparative costs of construction of the two types, cost
of money, and other pertinent factors are considered.   Figure 5 shows the diagram-
matic arrangement of an indirect dry-type cooling tower with a mechanical-draft
tower.

        Water from the cooling coils  is sprayed into the direct-contact steam con-
denser and mixes directly with the exhaust steam from the  turbine. The water from
the tower and the  condensed steam falls to the bottom where it is removed by cir-
culating and  condensate pumps.  The greater part of the water flows through the
pipes to the cooling coils, and an amount equal to the exhaust steam from the tur-
bine is directed  back to the boiler feedwater circuit for re-evaporation in the cycle,
Since the cooling tower circulating water and the boiler feedwater are intimately
mixed, the circulating water must be of condensate purity.
                                       15

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                                                     STEAM
                                                        BINE
                       NATURAL-
                       DRAFT  TOWER
       , AIR
U U   / FLOW
	  COOLING COILS
                        EXHAUST
                        STEAM
                                                             DIRECT-CONTACT
                                                             CONDENSER
                                                          CIRCULATING
                                                          MOTOR
                                                     PUMP
                                                             O
                                       WATER RECOVERY
                                       TURBINE
                                         CIRCULATING
                                         WATER PUMP
                                                              TO BOILER
                                                              FEEDWATER
                                                              CIRCUIT
          FIGURE  4 —  INDIRECT, DRY-TYPE  COOLING  TOWER
          CONDENSING SYSTEM  WITH  NATURAL-DRAFT  TOWER

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                     MECHANICAL-
                     DRAFT TOWER
      AIR
-H-  /  FLOW
.... .   COOLING COILS
                                                   STEAM
                                                   TURBINE
                        EXHAUST
                        STEAM
                                                           DIRECT  CONTACT
                                                        CIRCULATING  PUMP
                                                        MOTOR
                      WATER RECOVERY
                      TURBINE
                                                        CIRCULATING
                                                        WATER PUMP
                                                             TO  BOILER
                                                             FEEDWATER
                                                             CIRCUIT
       FIGURE 5 —  INDIRECT, DRY-TYPE COOLING TOWER
      CONDENSING SYSTEM  WITH MECHANICAL-DRAFT TOWER

-------
        In the Heller system, the cooling coils are mounted vertically, and the
warm circulating water enters the bottom of the coils, flows upward in the inner
rows of coils to the top water boxes, and then is directed downward through the
outer rows of coils.  The outer rows of coils come into contact with the entering
air,  thereby providing the greatest cooling range in water temperature.

        To prevent drawing air into the system in case of leaks in the cooling coils,
a positive pressure head of approximately 3 feet is imposed at the top of the coils.
This  is accomplished by means of either a throttling valve in  the circulating water
discharge from the tower, or, if a water-recovery turbine is used, by varying the
position of the adjustable turbine vanes.  In order to recover some of the pressure
head between the  cooling coils  and the condenser, in some installations water-
recovery turbines are coupled to the drive shaft of the circulating water pump to
recover the available energy.

        After passing through the recovery turbine, the circulating water is again
sprayed into the direct-contact condenser and recycled  through the cooling  system.

        Note that the circulating water does not come into direct contact with  the
cooling air; therefore, there is no evaporation loss of water as with the wet-type
tower.

        Direct system.  The principal components of the  direct air-cooled condens-
ing system are:

        1 .    Exhaust steam trunk.

        2.    Cooling coils.

        3.    Motor-driven fans.

        4.    Condensate pumps.

        Figure 6 shows a djagram of a typical direct air-condensing system.  Turbine
exhaust steam is conveyed through the exhaust steam trunk, which is large in dia-
meter to minimize  the pressure drop, to the air-cooled coils where cooling air pass-
ing over the finned-coil surfaces condenses the steam.   Shown here in  the simplest
form, the steam enters the top of the coil section and condenses as it travels down-
ward with the steam and condensate flowing in the same direction.  In actual
installations, provisions are made for  removal of noncondensable gases and air and
for prevention of freezing during cold weather.  The most common system in the
United States is to use horizontal tube bundles with 80 to 90  percent of the tubes
as the main condenser and 10 to 20 percent as an after-condenser to condense the
remaining steam that is not condensed in the main condenser. The steam and con-
                                       18

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               STEAM
               HEADER
                                  COOLING
                                    COILS
CONDENSATE
HEADER
TO BOILER
FEEDWATER
CIRCUIT
                                      EXHAUST  STEAM
                             CONDENSATE
                             HEADER
    EXHAUST-
    STEAM
    TRUNK
                                                  STEAM
                                                  TURBINE
CONDENSATE
RECEIVER
          CONDENSATE
          PUMP
EXHAUST
STEAM
             FIGURE 6—DIRECT-TYPE COOLING TOWER  CONDENSING  SYSTEM

-------
 densate flow concurrently, minimizing pressure loss and increasing the heat transfer
 coefficient.  The purpose of  the foregoing coil arrangement is to minimize noncon-
 densable gas blanketing of the main condenser as the residual noncondensable gases
 are swept out of the main condenser with  the residual steam. The presence of ex-
 cessive buildup of noncondensable gas in  the main condenser would be deleterious
 to effective condensation. Freeze  protection is usually accomplished by recircula-
 tion of warm air combined with the use of fan control.

        GEA of Bochum, Germany uses a method of direct condensation in which a
 certain percentage of the cooling coils are constructed so that the remaining steam,
 after passing down through a  condensing unit, enters the bottom headers of the
 aftercooling  coils,  and the condensate and steam flow in opposite directions in
 order to obtain better control of condensate temperature during cold-weather oper-
 ation. Only noncondensables remain in these latter coils near  the upper ends after
 all the steam has been condensed, thus preventing freeze-up in that region of the
 heat exchangers.

       The condensed steam  from the cooling coils fjows by gravity to condensate
 receivers from which it is pumped back to the boiler circuit by a condensate pump.

       Comparison of indirect  and direct  systems.  The principal difference be-
 tween the two systems is the large volume of exhaust steam which must be handled
 in the direct  system as compared to the smaller volume of circulating water in the
 indirect system.

       Although discussions with users of  the direct systems did not indicate that
 any adverse experiences as a result of condenser air leaks have  been encountered
 with the direct systems, the fact that all the cooling coils are under a high vacuum
 during operation is sometimes considered a disadvantage when compared  to the
 indirect systems with positive water pressures in the cooling coils (9).

 Use of Air Cooling by Industry

       The use of finned-tube  heat exchangers to dissipate waste heat to the at-
mosphere has been accepted by industry for well over 75 years.

       Common applications of the finned-tube heat exchanger are the automobile
radiator and steam or hot-water heating systems.  Also,  radiator-type heat ex-
changers have been used on stationary, internal-combustion, engine-driven gener-
ators up to 3,000 kw in size.

       In recent years, especially since the late 1940's, industry has turned more
to the  use of air cooling for discharging large amounts of heat to the atmosphere  in
                                     20

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processes where heretofore water-type heat exchangers or evaporative-type cool-
ing towers have been used.

       In 1969, it was estimated that the chemical process industry spent $50 mil-
lion on air coolers (11).

       The experience gained by the chemical and process industries in air cooling
can be of value to the electrical generating utilities as they consider air-cooled
condensing systems.

       Extent of air cooling in industry (13).  Industrial air cooling in natural-gas
processes was first used approximately 30 years ago in the arid south-central and
southwestern parts of the United States where water is not readily available. Later,
the petroleum  refining industry, petro-chemical  and chemical process industries
began to use air cooling on a wide scale.

       Plants  originally equipped with water-cooled systems and new process
plants often utilize air to remove from 60 to 100 percent of the waste  heat. Process
plants with indirect air-cooling systems as large as 2 billion Btu per hour—the heat
rejection equivalent of a fossil-fueled electrical  generating plant of over 400 mw—
have  been built  by the Hudson  Products Corporation of Houston, Texas. Direct
condensing systems have been supplied, on a smaller scale, to these same process
industries.

       The selection of  air cooling is being done on the basis of economic  justifi-
cation, taking into account first cost, operating and maintenance costs, and plant
capability.  Important factors in the evaluation of air cooling are the increasing
costs  of securing,  treating and  disposing of cooling water, and environmental limi-
tations.

       The problems peculiar to the various processes have resulted in the use of
many exotic materials for the air-cooled coils, including carbon steel, alloy steel,
stainless steel, nickel,  copper  admiralty, aluminum, cupro-nickel, hastelloy,
titanium and karbate. Process  coil design pressures and temperatures often  far ex-
ceed  those required  for  steam-electric plant air-condensing systems,  ranging up to
15,000psi and 1,000°F.

        The most common tube size is 1 inch O.D., although diameters from 5/8
inch  to 1-1/2 inch are used, with tube lengths up to 40 feet.  Except for special
high-temperature  process coolers, or for economizers where combustion products
are in contact with the  fins, fin material is predominately aluminum.
                                       21

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 Use of Air Cooler with Refuse Incinerators

        An important use of air-cooled heat exchangers in recent years in Europe
 has been that of steam condensing for municipal-type refuse incinerators. Modern
 refuse incineration makes use of a water-cooled furnace in place of the refractory-
 lined furnaces previously used.  In order to eliminate odors and to insure that all
 putrescible material is consumed,  the furnace temperature of the incinerator  should
 be in the range  of 1,500°F to 1 /850°F.  These high temperatures, when used with
 the older type incinerators, resulted in high maintenance costs for the refractory-
 lined furnaces; whereas, water-cooled furnaces can withstand high furnace temper-
 atures without deterioration.

        In order to absorb the heat from the combustion,  steam is generated in the
 process. In certain instances where there is a ready market for the steam, it is
 used for heating, process or power generation. However, there are installations
 where it is not feasible or practical to sell the steam generated in  the incineration
 process and,  in  order to condense  the steam and reuse the condensate, air-cooled
 condensers have been used.  A number of incinerators have been constructed where
 all steam is condensed in the air condenser and the condensate recycled through
 the boiler.  There are others where the air-cooled condenser is used during seasons
 where there is no market for the steam.

        Refuse incinerators which utilize air-cooled condensing systems have been
 installed at the  plants  listed in Table 4.

                                    TABLE  4
              Refuse Incinerators with Air-Cooled Condensing Systems

                               Condensing
                                Capacity          Condensing      Construction
	Location	     (Ibs. of steam/hr.)      Pressure           Year

Darmstadt, Germany              67,000           610   pjsi          1967

Vienna, Austria                  97,000           280   psi          1970

Biel,  Switzerland                 18,250           298   psi          1967

Bremen, Germany               220,000           212   psi          1968
                                 66,000            18.5 psi          1968
                                      22

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                             TABLE 4 (continued)
Location
Hagen, Germany
Bad Godesberg, Germany
Toulouse, France
Condensing
Capacity
(Ibs. of steam/hr.)
35,000
35,000
22,000
45,000
Condensing
Pressure
214 psi
214 psi
156 psi
44 psi
Construction
Year
1966
1966
1966
1969
From:  GEA Air-Cooled Steam Condenser Summary of Installations.
                                    23

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                                  SECTION II
                                FUNDAMENTALS
Design and Construction Considerations

       The  trend in industrial air-cooling design is to standardization of basic unit
components  including the air-handling unit, supports and structures, tube bundles,
and rotating mechanical parts, with maximum shop assembly and testing.

       V-belts as well  as direct-coupled, in-line, gear-reducer motors are used
to drive the fans.  Fans with a minimum of 4 to 8 blades with blade widths up to
18 inches are used.  Fans with diameters of 8 to 14 feet are common in industrial
air coolers.

       Figure 7 shows a cross section of a typical industrial air-cooled heat ex-
changer.

Codes and Testing (13)

       Except for proprietary header designs based on actual strain-gauge tests,
Section VIII for Unfired Pressure Vessels of the ASME Code is generally used for
header construction.

       Field test codes are in the process of being developed for air-cooled equip-
ment by committees of the ASME and AICHE, and there is currently an American
Petroleum Institute (API) specification.

       Fan ratings are ordinarily based on wind-tunnel tests  in accordance with
Bulletin 210, April 1962 Edition of Standard Test Code for Air Moving  Services,
adopted by the Air Moving and Conditioning Association.  Noise tests are con-
ducted in accordance with the United States Acoustical Society standards with
current maximum  limitations as required by the Walsh-Healy Act, effective July,
1969.  The trend  is towards even more stringent noise limitations which usually
results in an increase in first cost and in operating costs.

Fin Types

       General.  In general, there are five types of fins used with industrial air
coolers.   Figure 8 shows a cross-section view of these and  also the triangular pitch
used with circular tubes.  The fin tube cross sections are taken through Section AA
of Figure 8(a).
                                      24

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      SEAL  DISC
             TIP SEAL
PLENUM-
                                              'AN  RING
        ooo
        OO
TUBE    OOO
BUNDLE 0000o
 ooo         oo
OOO  TUBE    OOO
o°o°   BUNDLE n0o0n°o
MOTOR
                                      MACHINERY
                                      MOUNT
          FIGURE  7—HORIZONTAL  AIR COOLED
                 HEAT  EXCHANGER (12)
                         25

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   (a)  FIN TUBES IN TRIANGULAR
       PITCH ARRANGEMENT
(b) L-SHAPE
   FOOTED FIN
  (c) EMBEDDED
     FIN
(d) EXTRUDED
   FIN
(c) OVERLAPPED
   FOOTED FIN
         FROM (14)
(f ) HELLER-FORGO
   SLOTTED PLATE FINS
                        FROM (15)
    FIGURE  8—HEAT EXCHANGER- FIN TUBE TYPES
                            26

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       Tension-wound, footed, Figure 8(b).  This type of finned tube is generally
used for temperatures up to approximately 225°F.  The fins are predominately alu-
minum and the tubes of commercially available material suitable for the service.
Specially designed machinery wraps the fin around the tube.  The  L-shaped foot
provides  heat transfer surface  between the  tube and the fin.  This is the least expen-
sive type of finned tube, but has the poorest  bond and the  least thermal capability.

       Embedded fin, Figure  8(c).  This fin  type is used for temperatures up to
750°F.  The fins, usually of aluminum or steel, are wrapped around the tube and
fitted into grooves which have been cut or rolled into the tube. The embedded  fin
is locked into place  in the  wrapping process by rollers which press tube metal
against the base of the fin  to  form the necessary bond.  This fin tube is the most
versatile and physically rugged tube.

       Extruded fin, Figure 8(d).  This fin is used for medium-temperature service
up to 550°F. The finned section of the  tube  is extruded from a heavy-wall alumi-
num outer tube which has been fitted over  the outside of an inside tube of suitable
material  for the particular service requirement.  The extrusion process provides the
necessary bond between the inner and outer tubes. This type of tube has been used
for  chemical process applications where corrosion  has been a problem and in cases
where supplementary water sprays are used.

       United States manufacturers use this type of fin extensively for the process
industries, and offer slotted plate-type fins as well as extruded fins for the utility
industry.

       Wrapped-on  overlapped, footed fin,  Figure 8(e).  This fin is used  for tem-
peratures up to 450°F.  The tube is constructed by wrapping the fin around the
tube and is the same as the tension-wound, footed fin,  except for the double foot
which affords better protection against corrosion.

       Plate-type fin, Figure 8(f). In addition  to the above-described finned
tubes, the plate-type fin is also used in industrial work.   In the plate-type con-
struction, flat plates are drilled or punched for the tube and the plates are placed
over the  tubes.  In the United States, plate-type  finned coils are  always furnished
with  collars  integral with the fins,  but in  Europe  separate collars are often used.
The purpose of the collar is to increase  the contact surface between the tube and
the plate fin.  As shown in Rgure 8(f), the Forgo coil, developed for use with  the
Heller system, is of the plate-fin design.  Note the slots or louvers in the aluminum
fins.   The purpose  of the  slots is  to improve heat transfer rate by preventing  the
thickening of the boundary  layer which forms with uninterrupted  air flow over a
flat surface.  Since  each slot results in  another boundary  layer, the heat  transfer
is improved.
                                       27

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 Types of Air-Cooled Exchange Systems

        There are a  number of different types of systems used in air-cooling practice
 in industry.  Figure 9, from (14), illustrates the most commonly used systems.

        Direct air cooling, as illustrated in Figure 9(a), is commonly used and is
 the simplest system.

        Direct air cooling with warm air recirculation, Figure 9(b),  is used to pre-
 vent freezing of process fluid during cold weather, or where it is desired to keep
 the cooling air temperature high. Air coolers have been used where air tempera-
 tures reach 50°F below zero with this design.

        In  the indirect cooling system with a liquid-cooled exchanger,  Figure 9(c),
 the process fluid is cooled in a shell-and-tube heat exchanger by a  cooling fluid
 which is recirculated in the secondary air-cooled heat exchanger.

        The indirect cooling system with spray condenser,  Figure 9(d),  is similar to
 the indirect cooling system with a liquid-cooled heat exchanger, Figure 9(c), with
 the substitution of a spray condenser for the closed heat exchanger,  and closely
 resembles the Heller indirect air-cooling system for steam-electric generating plants.

        The combination air cooler-cooling tower unit, Figure 9(e), performs a
 double function of cooling air by humidification spray to supply the  air cooler with
 inlet air of lower temperature than the available ambient dry-bulb temperature and
 also supplies cooling water which is used to remove process heat from miscellaenous
 cooling services.

        The direct, V-type, air-cooled exchanger with sprays, Figure 9(f), re-
 quires less  space than the conventional  horizontal air cooler, but is  more expensive
 for a given duty and requires more horsepower.  By spraying water of condensate or
 demineralized quality onto the fin tubes during a few hours of extremely high dry-
 bulb air temperature, the process outlet temperatures from  the cooling coils can be
 reduced.

Theory of Heat Transfer from
Air-Cooled Coils

        Although it is recognized that there are differences of opinions among
designers of air-cooled heat exchangers as to methods of determining heat rejec-
tion performance and, consequently, the material  used herein  from various
references  is not always in agreement, it is felt that all pertinent information
available should be  included in order to present as complete a picture as possible
of the state of the art of dry-type cooling towers.
                                      28

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AIR COOLER-y
PROCESS
           )   I   t
                   FINNEO
                   COIL
     (a) DIRECT COOLING
         AIR

AIR COOLER t  f LOUVERS
        rtHltffflll	
 FINNED
 COIL —

               imt'
          K — ^ __ ./
LOUVERS
      vtMt/tttt-
                                   PROCESS
                                   EXCHANGER
                                   (d)  INDIRECT COOLING WITH
                                       SPRAY CONDENSER
                                                     AIR COOLER
                 LOUVERS


                 WARM AIR
                 RECIRCULATION
                       PROCESS
                         ^M

                     COOLING
f   !
                                     AIR r
                                                            PROCESS
                                                        WATER
                                                        COOLED
                                                        EQUIPMENT
  (b) DIRECT COOLING
      WARM AIR RECIRCULATION
                                 (e) COMBINATION AIR COOLER-
                                     COOLING TOWER UNIT
      PROCESS
      EXCHANGER
 PROCESS
                           AIR COOLER
 (c) INDIRECT COOLING WITH  LIQUID
     COOLED EXCHANGER
                                    TUBE
                                    BUNDLE
                                                           SPRAYS
                                                          LOUVERS
                                     (f ) DIRECT COOLING-
                                         "V"TYPE AIR COOLED
                                         EXCHANGER WITH SPRAYS
               FIGURE  9 — EVAPORATIVE-TYPE
               HEAT  EXCHANGER  SYSTEMS  (14)
                                29

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        Basic theory  (12).  The resistance to the flow of heat from a hot fluid inside
a finned tube to cooler air flowing across the outer surfaces  of a clean, corrosion-
free tube can be expressed as six separate components of resistance:

        1 .    The internal film  resistance to the flow of heat between
              the  hot fluid in the tube and the internal  surfaces of the
              metal tube.

        2.    The resistance  to conduction of heat through a fouling
              resistance deposited on  the inside wall of the tube.

        3.    The resistance  to conduction of heat through the metal
             wall of the tube.

        4.   The resistance  to flow  of heat across the bond or gap
              between the inner tube  metal and the fin  muff or collar.

        5.    The resistance  to flow of heat  through the fin from the
              inner periphery of the  fin to  the outer  periphery of
             the  fin.

        6.   The  air-film  resistance to the flow of heat from the
             surface of the fin to the air passing over it.

        Of the  six resistances, the air-film resistance is  the most significant.  The
other resistances are, in general,  relatively low compared to the air-film resistance
which impedes the flow of air from the fin surface to the air.  Because of the
greater  resistance  to transfer of heat from the metal fins to the air,  it is necessary
to increase the heat transfer surface in contact with the air, which  accounts for
the use  of fins in air-cooled heat transfer surface.  Typical ratios of fin surface to
tube surface are from 10 to 30.

        The transfer of heat from the inside fluid to the air is influenced by  a num-
ber of variables:

        1 .    The  temperature difference between the fluid and the air.

        2.    The  design and surface arrangement of the coil.

        3.    The  velocity and character of air flow across the tubes.

        4.    The  velocity and physical properties of the fluid inside
             the tubes.
                                      30

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       The driving force for the transfer of heat between the fluid inside the tubes
and the air flowing across the tubes is a function of the logarithmic mean tempera-
ture difference (LMTD) between the fluid and the air.  The LMTD is expressed by
the following formula:

             LMTD  =  GTTD-LTTD                                      [1]
             where:   LMTD  =  logarithmic mean temperature
                                difference, °F

                      GTTD  =  greater terminal temperature
                                difference between the hot
                                fluid and the cold fluid, °F

                      LTTD   =  lesser terminal temperature
                                difference between the hot
                                fluid and the cold fluid, °F

        Figure 10 illustrates the basic temperature diagram as it applies to dry-type
cooling tower coils.

        Indirect system.  Figure 10 (a) shows the temperature relationship that exists
in the indirect system.  The left side of Figure 10 (a) represents the temperatures
that exist in  the direct-contact condenser where the cool circulating water from
the tower mixes with the turbine exhaust steam.  The upper line represents the tem-
perature level of the condensing steam. Since condensation takes place at a con-
stant temperature corresponding to the saturated steam temperature of the turbine
back pressure, this  line is horizontal and at temperature Tsj .  The lower curve on
the left side of the diagram represents the  temperature condition of the circulating
water heated from TW2 to  Twi  as the exhaust steam transfers the  heat of conden-
sation to the  water.

        The difference in temperature between Ts] and Twi  represents subcooling
of the condensate and circulating water below the saturated steam temperature of
the exhaust pressure and is a thermal loss to the turbine cycle.  In the typical
direct-contact condenser,  the subcooling is approximately 3°F, but it is possible
to have a condenser in which no subcooling exists,  in which case Tsi  and Twi are
the same.

        The diagram on the right-hand side of Figure 10(a) illustrates the tempera-
ture condition of the circulating water and the cooling air as the  water flows
through the coils.  The air at ambient temperature Tai  comes into contact first
                                      31

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I
UJ
(E
ID
or
UJ
o.
DIRECT-CONTACT
CONDENSER

TURBINE EXHAUST
STEAM
TS|
   COOLING COILS

  /TRANSFER OF HOT CIRCULATING
 /WATER FROM CONDENSER
/  TO TOWER	
i.
     TRANSFER OF COLD CIRCULATING WATER
     FROM TOWER TO CONDENSER
     (I)
      I
   NOTE:
                 (2)   (3)
    THE ABOVE SKETCH DOES NOT IMPLY  FLOW RELATIONSHIPS
       (a)  INDIRECT  SYSTEM
                  WATER AND STEAM ENTERING CONDENSER
                  WATER LEAVING CONDENSER
                  WATER ENTERING TOWER AND AIR LEAVING TOWER
                  AIR ENTERING TOWER AND WATER LEAVING TOWER
                       CONDENSING
                 TS,    COILS
                   (I)
                   (2)
                   (3)
                   (4)
     TURBINE EXHAUST
     STEAM
    TRANSFER  OF
    EXHAUST STEAM
    FROM TURBINE TO
    CONDENSING COILS
     (5)
             (b)
                                           AMBIENT AIR
                      (6)             (7)

             DIRECT  SYSTEM
              (5)  STEAM  LEAVING TURBINE
              (6)  STEAM  ENTERING CONDENSING COILS AND AIR
                  LEAVING CONDENSING COILS
              (7)  AIR ENTERING CONDENSING COILS AND HOT
                  WATER(CONDENSATE) LEAVING COILS
        FIGURE   10  — TEMPERATURE  DIAGRAMS OF
        DIRECT  AND INDIRECT DRY COOLING TOWER
                HEAT-TRANSFER  SYSTEMS
                             32

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with the cooled water at TW2 and is heated to Ta2 as the water cools from Tw"|  to
TW2- The diagrams shown are for counterflow of air and water.   In actual practice,
crossflow correction factor is used to compensate for the deviation in heat exchanger
performance because of the crossflow condition.

        Direct system.   Figure 10 (b) shows the temperature relationship between the
turbine  exhaust steam and the cooling air as they exist in the direct air-condensing
system.   No circulating water is used in the direct condensing system and the ex-
haust steam with temperature at Ts] is conveyed through the exhaust steam trunk to
the condensing coils.  The difference in temperature between Tsi  and TS2  is the
result of pressure drop in the  exhaust steam trunk and also is a loss to the cycle; TS2
represents the temperature level in the coils at which condensation takes place and
is at a constant level since the steam condensation takes place at a saturated tem-
perature corresponding to the steam pressure in the coils.  The lower line of this
diagram shows the temperature rise of the air as it flows past the  coils and picks up
the heat of condensation .

        Further subcooling below  temperature TS2 can result from improper design or
operation of the condenser.

        The heat transfer from the coils to the air is expressed by the general for-
mula:

             Q  =  U LMTD A  F_                                         [2]
                                  y

             where:     Q   =   total  heat transfer of the coil,
                                 Btu/hr.

                        U   =   over-all coefficient of heat
                                 transfer, Btu/(hr.  ft.2°F)

                    LMTD   =    logarithmic mean temperature
                                 difference between fluid in-
                                 side the coil and the air, °F

                        F_  =    dimension less crossflow cor-
                                 rection factor — usually
                                 around 1 .0
                                                           2
                        A   =    area of the coil surface, ft.

        The over-all heat transfer coefficient (12)  must be applied to the proper
area.  It is sometimes applied to the inside surface of the tubes,  sometimes to the
outside  surface of the bare tubes and sometimes to the total outside extended
                                      33

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surface.  In any case, the area chosen must be consistent with the properly applied
over-all U.

        The over-all resistance to heat flow is the sum of the six individual resist-
ances, which are set forth on page 30 of this report.

        Since the air-film resistance is a function of the velocity of the air and the
geometry of the fin surface only, and since the efficiency of the fin is a function
of the air-film coefficient, the geometry of the fin  surface, and the conductivity
of the metal of the fin, these resistances are usually combined into one resistance
for a particular geometry and fin metal.  This one resistance is then only a function
of air velocity and is usually determined from wind-tunnel  tests for any particular
surface.

       Thus, the over-all U and R can be expressed by the following  equations:
             ^  ~   ra + "77
                              or

             i     i+M    +   +A+II
             u     ha   AJ  L rg    rf   kt    h. J
                                                            2
             where:   A   =  area of the extended surface, ft.
                                                             2
                      A.   =  area of the inside tube surface, ft.

                      h   =  apparent coefficient of heat trans-
                             fer of a  finned surface,  Btu/(hr. ft.*   F)

                      h.   =  coefficient of heat transfer on inside
                       '      of tube, Btu/(hr.  ft.2 °F)

                      r    =  resistance  to air (°F hr.)/Btu

                      r.   =  resistance  to flow from fin surface to
                             air(°Fhr.)/Btu

                      r    =  resistance  to flow through metal

                             (°Fhr.)/Btu

                      R    =  thermal resistance ^ F hr.)/Btu
                                      34

-------
                      t    = tube thickness, ft.

                      k.   = thermal conductivity of the tube
                             metal,  Btu/(hr. ft.2 °F)

                      r    = gap or bond resistance  (°F hr.)/Btu
                       y
                      r,   = fouling resistance (°F hr.)/Btu

       The apparent coefficient of heat transfer of the external surface, hQ, is, as
noted above, a combination of the heat transfer from the collar of the fin which
has 100 percent efficiency and the heat transfer from the fin which has incorporated
in it the average  resistance to flow of heat through the fin metal to every part of
the fin surface.  The apparent coefficient of heat transfer of the external surface
can be expressed  as follows:
-  hr
                                                                             [5]
             where:   Ar   =  surface area of fins,  ft.^

                      AQ   =  total area of fin  surface
                              and collar, ft.*
                           =  mean surface coefficient of
                              heat transfer of a finned
                              surface, Btu/(hr. ft.2 °F)
                       ^r   =  fin efficiency

        Since the ratio  Ar/AQ is usually .91  to  .97,  this is not a significant cor-
rection.  Fin efficiency can be calculated by methods of Gardner (16)  or others,
and correlation exists for approximate calculations of hr for many geometries [see
Kays and  London (1 7) J .  However, the only truly accurate method of determining
h  for any particular geometry is by wind-tunnel tests where hQ is plotted against
face velocity of the air.

        Since fin efficiency is a function of  fin height and fin thickness, and since
surface  ratio is also a function if fin  height, the design  of a fin surface is an eco-
nomic balance  between increasing fin height,   and consequently surface ratio, and
decreasing fin efficiency.

        Mechanical fabrication technology also imposes limits on increasing sur-
face ratios.
                                       35

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        The fin efficiency, ir, is given by the following formula:
                                                                             „
where:
                           =  g
                       t    =  fin thickness,  ft.

                       r.   =  radius of curvature of fin
                       f      tip, ft.

                       r    =  radius of curvature of fin
                       f      root, ft.

                       hf   =  mean surface coefficient of
                              heat transfer of a finned
                              surface,  Btu/(hr. ft.2  °F)

                       k    =  thermal  conductivity of fin,
                              Bru/(hr.  ft.2 °F)

                       g    =  denotes  "a function of"

        Good design of a fin surface dictates  obtaining the maximum air side film
coefficient for a minimum expenditure of pressure loss of air passing through the
surface. This can be accomplished by louvering or serrating the fin, thus inter-
rupting the boundary  layer of air which  is the resistance to heat transfer.  This
interruption of the boundary layer considerably increases the air-film coefficient
at a modest  increase in pressure loss.

        Good heat transfer from finned coils is also dependent  upon a good mechan-
ical bond between the  tube and the fin.  The effectiveness of  the bond depends on
the as-manufactured compression pressure between the fin  collar and the tube.  For
low compression pressures it is possible for only a fraction  of the two surfaces to be
in contact.  This results in an air gap over part of the contacting surfaces  and con-
sequent loss of heat transfer.  As the metal  temperature of the  coil rises from the
as-manufactured temperature, if the fin is aluminum and the tube  is steel, the alu-
                                      36

-------
minum will  expand more than the steel, thereby relaxing the initial contact pres-
sure.

        For  finned coils,  the heat transfer to the air stream is dependent upon so
many  factors that reliable rating and performance information for any specific coil
design must  be verified by actual tests (18) .

        Design of air coolers (12).  The design of an air cooler for any process con-
dition involves several  trial-and-error procedures.  The size of the air cooler is
not initially known and, therefore, the exit air temperature will  not be known and
neither will the transfer rate in the tubes, since velocity  in the  tubes cannot be
calculated.

        The  initial  step is to assume an air temperature rise in the air cooler com-
patible with the process fluid inlet and outlet temperatures and to calculate the
LMTD based on this assumption.  A "U" is assumed, based on experience with sim-
ilar equipment.   From the calculated LMTD and assumed "U",  a required surface
can be calculated. A trial arrangement of this surface is made in the most econom-
ical way as to length of tube, width and number of bundles, and depth of tube rows.
With  this arrangement and assuming an air face velocity compatible with desired
pressure loss, the temperature rise of the  air is calculated.  If this rise  does not
match the assumed rise, the cooler must be rearranged until it does match by vary-
ing the surface, or face velocity  of the air, or both. Then the tube side  passes can
be arranged to suit the required pressure drop,  and the tube side film coefficient
can be calculated.  If summation of the individual resistances does not equal the
assumed "U", the process must be repeated until balance is achieved.  Knowing now
the number  of crossflow passes, the LMTD correction factor must  be incorporated.

        Initial temperature difference.  Rather than to use the logarithmic mean
temperature  difference  between  the  fluid in the  coil and the air-cooling coil,
designers of air-cooled  heat transfer  surfaces  have found  it more convenient to
express coil performance as a function of the initial temperature  difference (ITD)
between the fluid  entering  the  coil and  the air entering the coil (ambient air).
Ignoring the subcooling effect, the ITD is identical to GTTD expressed in Equation
[l] for direct systems and is equal to GTTD plus the cooling range for indirect
systems.

        Cheshire  and Daltry (19) have developed an expression for the frontal area
of the cooling coils which  utilizes the  initial air temperature difference between
the fluid in the coil and the ambient air, and also takes into account the variations
in height and depth of coolers, number of water passes, air flow, and water flow.

             A   =  Q/    '      +  PH    .     1    \                    f71
               f                    **
                                       37

-------
                                                            2
             where:  Ac   =  frontal area of cooling coils,  ft.

                      Q   =  heat to be dissipated, Btu/hr.

                      At   =  maximum temperature difference
                              between the water and the air, °F
                              (which is the same as ITD)

                      Va   =  air velocity at cooler, ft. /sec.

                      ^   =  air density, Ibs./ft.3

                      p    =  number of water passes

                      H    =  height of cooler, ft.

                      n    =  number of tube rows
                      Vw =  water velocity in cooler, ft ./sec.

                      Uc  =  over-all crossflow heat transfer
                             coefficient, Btu/(hr. °Fft.2)
                       , ,  «,  g  =   constants


       Although calculations of heat transfer surfaces usually develop a logarith-
mic function,  the linear function as derived by Cheshire and Daltry may well be
accurate within the design limits of the dry-type cooling towers.

       Dry cooling tower heat balance (20)  .  In the system heat balance for a dry
tower of  the indirect type with a steam-electric generating  plant, the heat trans-
ferred to the circulating water, the heat rejected to the air, and  the  heat rejected
by the coil are all equal .  The following are  the basic formulas for these heat
quantities:

       Heat rejected to air:

             GHRa   =  WaCa
-------
Heat rejected by finned heat exchanger:

      GHRC  =  UALMTDFg                                       [10]

      where:  CQ      =  specific heat of air at constant
                         pressure, Btu/(lb. °F)

              C       =  specific heat of water,
                w
                         Btu/(lb.

              GHRQ  =  gross heat rejected to air,
                         Btu/hr.

              GHRC  =  gross heat rejected by coil,
                         Btu/hr.

              GHR   =  gross heat rejected to circulat-
                         ing water, Btu/hr.

              T_i     =  temperature of air entering
                           .1  Or
                         COll,  F

              To     -  temperature of air leaving
                           • I  Or-
                         coil,  F

              T  i     =  temperature of water entering
                           • i  O i-
                         coil,  F

              To     =  temperature of water leaving
                         coil,°F

              WQ     =  weight of air,  Ibs./hr.

              W      =  weight of circulating water,
                         Ibs./hr.

              F       =  crossflow correction factor, a
                         function of air and water tem-
                         peratures and  pass arrange-
                         ments, typicajly varies from
                         0.9 to 1 .0 for  large  heat
                         exchangers (21)
                              39

-------
        For a thermal-electric system in balance, it is obvious that:

             GHR   =  GHR    =   GHR                                   [ll]
                  awe                                  u  J

        Solving simultaneous equations yields the following expression for gross heat
rejected at the balance point, GHRi :

             GHR
                  b
                            ca     ww cw
             where:   z     =  F^ U A
                                g     V wa ca     wwcw
)                 D3]
                      ITD   =  the initial temperature difference
                               between the water entering the
                               coil and the ambient air entering
                               the coil,   F

        Note the definition of ITD as used by Gates is different from that shown in
Figure 10, which shows ITD as the initial temperature difference between the
saturated  temperature corresponding to the  turbine back pressure, rather than the
difference between the circulating water entering the coil and the air surrounding
the coil.  Smith and Larinoff (13) have used the  definition of ITD for coil perform-
ance as the difference between temperature of saturated steam at turbine back
pressure and ambient air, and this definition is generally used in European practice.

       The numerical difference between  the two methods of defining ITD is the
subcooling of the circulating water below  the saturated temperature of the exhaust
steam at the turbine and can  amount to approximately 3°F in practice.  Either
method of handling ITD is satisfactory as long as the subcooling effect is taken into
account.  The method used by Gates permits direct use of ITD without correction
for subcooling effect on coil  performance,  and perhaps is the most logical method
when  viewed from  a coil performance standpoint, whereas the other method would
seem to be the better selection when considering  over-all system performance be-
cause it relates tower performance directly to turbine back  pressure.  The foregoing
example of different use of terms points up  the need for standards of terms and defi-
nitions for the dry-type cooling  tower industry; undoubtedly, such standardization
will be forthcoming if dry-type cooling towers come into general use.
                                     40

-------
       Cotes expresses U as:



            U  =  -L+^u'..    ^                               D4
                   h1 o    h.            hw

            where:   h'o  =  a collection of all conductances
                             other than the inside coefficient,
                             based on actual test data,
                             Btu/(hr. ft.2°F)

                     A    =  ratio of area of fin surface to
                             inside area of tube

                     hw   =  tube wall conductance,
                             Btu/(hr. ft.2°F)

       The equation for inside film heat transfer coefficient for water at ordinary
temperatures is (18):
             h.  =
                _   150(1  + .011 t)V°'8
            where:   V     = water velocity, ft./sec.

                     d     = inside diameter of tube, inches

                     t     = temperature of water,  F

       Equation [15] is a simplification of the more general equation for liquids
in fully developed turbulent flow given by McAdams  (22) as:

                       •  /^\-8  /MfC  \-33 Kr
            h=  '  -023 (f)    (nf)    ^

            where:   d     = inside diameter, ft.
                                                       2
                     G     = mass velocity, lbs./(hr. ft.  )

                     Mf     = absolute viscosity, lbs./(hr. ft.)
                                    41

-------
                      Kf    =  thermal conductivity of fluid,
                               Btu/(hr. ft.2°F/ft.)

                      C    =  specific heat at constant pres-
                        P       sure,  Btu/(lb. °F)

        The inside tube transfer rate for isothermal condensing fluids such as steam
 is still controversial .   The  Heat Transfer Research Institute  of Alhambra, California
 is embarked on an extensive experimental and correlation program for predicting
 isothermal  and nonisothermal condensing rates.  The methods now in common usage
 for isothermal condensation are the method of Dukler (23) , the method of Kirkbride
 (24)— which are modified Nusselt correlations — and the method of Akers (25) which
 takes into account the shear effect on  the condensate caused by the velocity of  the
 uncondensed vapor (12) .

 Effectiveness— N^  Approach

        The technique of arriving at an optimum heat exchanger design is a complex
 one due to the mathematics involved.  Even more significant, however,  are the
 many qualitatave judgements that must be introduced into the analysis.   The vari-
 ables previously described are complex functions that do not lend themselves to
 ease of  evaluation. Consequently, except for simple configurations, model tests
generally are used to establish their effect in a given cooling element.

        Kays and London  (1 7)  describe another approach to heat exchanger design
 in terms that allow a better visualization of the interaction of various major para-
meters on the efficiency of a heat  exchanger.  They describe exchangers i-n terms
of "effectiveness" and  "number of  heat transfer units".

        The "effectiveness" term defines the heat transfer performance.  This term
compares the actual heat transfer rate to the maximum possible heat transfer rate
and is a measure of the heat transfer effectiveness of the cooling element.
             E  =
                            Ow- ~Ta.
                              wm    um
             where:   E    = exchanger heat transfer effective-
                              ness, nondimensional

                      W    = water flow rate, Ibs./hr.

                      WQ   = air flow rate, Ibs./hr.
                                     42

-------
                      C    =  specific heat of water, Ibs./hr.

                      C_a  =  specific heat of air,  Btu/(lb.°F)

                      TW   =  temperature of water, °F

                      TQ   =  temperature of air, °F

       The number of heat transfer units, N^,  is a nondimensional expression of
the heat transfer size of the exchanger.

             N.u  =
       When the N^ is small,  the exchanger effectiveness is also small .   It is
apparent from an examination of Equation [18] that the cost of attaining a large
Nj.y and, consequently, a high  degree of effectiveness is tied closely to the capital
investment required to provide a large heat transfer area or an improvement in the
conductance, U .

       The relationship between E and Nj^ for a crossflow cooling situation with
the air assumed to be mixed is shown in Figure 1 1 , taken from (17) .  The effective-
ness of a heat transfer element increases sharply with a greater number of heat
transfer units until the curve levels out and becomes almost asymtotic. Considerable
expense is required  to obtain  the last 20 to 30 percent of effectiveness.  Therefore,
the optimum cost-effective cooling element may not be the most efficient one.

Theory of Thermodynamic Cycles

       The Carnot cycle. The  understanding of the basic Carnot cycle is useful in
studying the improvement of the Rankine cycle, which is described in subsequent
paragraphs.  The  Carnot cycle is shown in Figure 12,  plotted on a temperature-
entropy diagram.

       The simplest statement of the second law of thermodynamics is that heat
will not flow of its own accord  from a cold body to a hot body.  The second law
may be rephrased to state that not all of a given quantity of heat can be converted
into useful work.

       The Carnot cycle, comprising two constant-entropy processes and two
constant-pressure processes, all of which are reversible,  is the most efficient
power plant cycle conceivable. Temperature 2 - 3 is the maximum temperature
available to the cycle and temperature 1  - 4 is the lowest temperature available .
                                      43

-------
     100
  UJ


  CO
  CO
  UJ
  o
  UJ
  u.
  u.
  UJ
       012345

    NO. OF TRANSFER UNITS, NTU max = AU/ WA C
FIGURE  II—HEAT TRANSFER EFFECTIVENESS AS A

FUNCTION OF  NUMBER OF TRANSFER UNITS (NTU)

   CROSSFLOW  EXCHANGER WITH  AIR MIXED(I7)
                      44

-------
 UJ
 a:
 ac
 UJ
 a.
 2
 UJ
HEAT AVAILABLE
      FOR
     WORK
                        HEAT UNAVAILABLE
                              FOR
                             WORK
                                          3 ( MAXIMUM TEMPERATURE )
                  4  ( MINIMUM  TEMPERATURE)
                             ENTROPY

              FIGURE 12 —CARNOT CYCLE PLOTTED ON
                TEMPERATURE— ENTROPY DIAGRAM

In power plant practice, temperature 1 - 4 is the temperature of the circulating
water, or, in the case of an air-condensing system, the ambient air temperature.

       The thermal  efficiency of the Carnot cycle is expressed as follows:

            ~     .    i   rr. .         heat available for work
            Carnot cycle efficiency =
               total heat supplied

       (T2-T1)(S4-S1)      T2-l
                                                                      [19]
       The Carnot cycle represents the highest possible efficiency of a cycle.
Although such efficiency is unobtainable on a practicable basis, the cycle provides
a basis for measuring the efficiencies of power systems.

       The Rankine cycle. The general energy equation expresses the first law of
thermodynamics as it applies to steady-flow processes, such as apply to the steam
power plant cycle, as (26):
                                    45

-------
             PE] + KE,  + WH1 +  Q   =  PE2 +  KE2 + WH2 + Wk          [20]

             where:   PE  =  potential energy,  ft-lb. or Btu

                      KE  =  kinetic energy, ft-lb. or Btu

                      H   =  total  enthalpy,  Btu/lb.

                      Q   =  heat  transferred to or from the
                             system,  Btu

                      Wk  =  work done on or by the system,
                             Btu

                      W  =  weight of fluid, Ibs.

       The heat balance for the power plant assumes the cycle to be a closed sys-
tem.  Changes in potential  energy are not significant and by treating the power
plant components as integral units, changes in kinetic energy do  not have to  be
considered.  The general energy equation for a steam power plant, then, is written
as:

             Wk  =  W(Hj -H2) + Q                                     [21]

             where:   H.  =  enthalpy of entering steam or
                             water, Btu/lb.

                      hL  =  enthalpy of leaving steam or
                             water after expansion, Btu/lb.

                      Q   =  heat  added to the  system be-
                             tween conditions 1 and 2,
                             Btu/hr.

                      Wk  =  work done on or by the system
                             between conditions 1 and 2,
                             Btu/hr.

                      W  =  flow  of steam or water, Ibs./hr.

       Figure 13 shows the diagram of an elementary steam plant cycle known as
the Rankine cycle.
                                     46

-------
  Qs HEAT
(HEAT SUPPLIED
  TO STEAM)
               BOILER
                                     t
WKbfp
(PUMP WORK)
                                                                       Qr  HEAT

                                                                     (HEAT REJECTED
                                                                     FROM  CYCLE)
                     FIGURE  13—DIAGRAM OF RANKINE CYCLE

-------
 Referring to Figure 13,

 a.    Heat added to the cycle by the boiler:

                    Q\AA /LJ    LJ \                                    Pool
              s   =  W(H] — r\4)                                    [22J

 b.    Heat rejected from the cycle by the condenser:

            Qr   =  W(H2-H3)                                    £3]

 c.    Work done by turbine:

            Wk.  =  W(H,-H9) = W^hK-rM                   [24]
               i          I    ^        r   i    £.                        J

            where:         'L   =   over-all turbine efficiency

                    Hi—H«i   =   total isentropic available
                                  energy, Btu/lb.

d.    Generator output:

            Wkg  =  W»7g (Hi - H2)  =  >7g Wkt                      [25]

            where:         »?  =   over-all generator
                                  efficiency

e.    Boiler feed pump work:

                    =   W(H4-H3)                                 [26]
f.     Thermal efficiency based on gross generator output:

                                        Wk
            Gross thermal efficiency   =     9                       [27]
                                           s

g.    Thermal efficiency based on net generator output:

                                       Wk  -Wk
            Net thermal efficiency  =     9	bfp
                              48

-------
      h.    Gross turbine cycle heat rate:

                  Gross heat rate  =  	.	3/f13rr.  .	
                                     gross thermal  efficiency

                  3,413 Qs  _    Qs

                     Wkg     "  KWg~

      i.    Net turbine cycle  heat rate:

                  M .  ,   .    .              3,413
                  Net  heat rate  =  ——r	'-—^r-.	
                                    net thermal  efficiency

                  3,413 Qs             Qs
                                    KVKWbfP

             Plant net heat rate, taking into account boiler
             efficiency and auxiliary power requirements:

                  Net plant heat rate  =
                                            KWg - K Wbfp - KWq)

                  where:   \       =  boiler efficiency

                           KW     =  generator output at generator
                                      terminals, kw

                           KWL e   =  boiler feed pump power,  kw
                              bfp
                           KW     =  auxiliary power (excluding
                              a
                                      boiler feed pump), kw
[29]
[30]
                           Wk      = generator work, Btu/hr.
                              57
                           Wki r    = boiler feed pump work, Btu/hr.
                              bfp

       In modern power plant design, the regenerative reheat cycle is used rather
than the Rankine cycle. The modern regenerative reheat cycle, however, is but
an improvement of the basic Rankine cycle and, therefore, an understanding of the
simple Rankine cycle is essential.
                                     49

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       Improvements to the Rankine cycle.  The basic efficiency of the Rankine
cycle can be improved by increasing the temperature of the steam from  the boiler
by superheating, by reheating the steam to its maximum temperature after it has
performed a certain amount of work in the turbine, and by means of regenerative
feedwater heating,

       The regenerative reheat cycle is used with all  large, modern steam-
generating plants.   Figure 14 shows the basic diagram  of a regenerative reheat
turbine cycle.

       The equation for the heat rate of the regenerative reheat turbine cycle,
more commonly referred to as the reheat turbine cycle, is:

                               W.(H.-h, ) + W,t  (H,  -H   )
             ^    ,   ,   ,       tv  t   fw7     rhtrv hr   cr'             p,^
             Gross neat rate  =  	             [321
                                      generator output

             where:   W.     = throttle flow, Ibs./hr.

                      WrL(.r   = reheater flow, Ibs./hr.

                      Hf     = throttle enthalpy, Btu/lb.

                      h,      = final feedwater enthalpy, Btu/lb.

                      H,      - enthalpy leaving reheater, Btu/lb.

                      Hcr    = enthalpy entering  reheater, Btu/lb.
                                     50

-------
W, LBS/HR   H, ENTHALPY
                                                                                          CONDENSATE
                                                                                             PUMP
                                BOILER FEED PUMP
         FIGURE 14 —TYPICAL FLOW DIAGRAM FOR REGENERATIVE REHEAT CYCLE

-------
                                  SECTION
                                 PERFORMANCE
Performance of Dry-Type Cooling Towers

       The concept of initial temperature difference (ITD), discussed in Section II
under "Theory  of  Heat Transfer from  Air-Cooled Coils" and  as  illustrated in
Figure 10, is essential in understanding the performance of a dry-type coo I ing system
under varying ambient air temperatures and turbine loads.  In this report, the ITD is
considered to be the difference between the saturated steam temperature correspond-
ing to the turbine back pressure at the exhaust flange and the ambient air tempera-
ture,  since this method  directly indicates  the  effect of variations in ambient air
temperature upon  turbine performance.   In adopting  this definition of ITD,  there
must be compensation for condensate subcooling in the indirect-type cooling system
and for steam-pressure loss in the exhaust  steam  trunk for the direct-type condensing
system when considering cooling coil  performance.   Depending upon the design and
performance of the system, there may be a difference of from one-half degree to
several degrees between the ITD as defined above and the ITD defined as the differ-
ence between the temperature of either warm circulating water entering the coils in
the indirect system or the temperature  of condensing steam in the direct system  and
the ambient air temperature.

       As discussed in the section on theory, the actual heat transfer coefficient,
U, for a cooling element in a  dry-type tower  is  designed  on a trial basis  with a
testing program to establish  or verify  the design factors.   From data obtained on
existing and proposed natural-draft installations, the U factor varied  from 184 to
238 Btu/(hr. °F ft.2) per row  and  averaged 202 Btu/(hr. °F ft.2) per row  on  a
frontal-area basis.  The value  would be less if calculated on a total cooling surface
area.  No indirect,  mechanical-draft  installations of any size for generating units
have been constructed, so their U  factor must be estimated on a theoretical  basis.
It would appear that the pressure drop  across the cooling element in a mechanical-
draft tower would be of less concern than  in a natural-draft tower where a slight
increase in the  pressure drop would increase the tower height by many feet with a
consequent increase  in construction cost.  In a mechanical-draft tower, an  increase
in  pressure drop would be offset by  increasing fan horsepower.  Therefore, it is
likely that the cooling elements could be  closer together in a mechanical-draft
tower and the U factor on a frontal-area basis would be higher than for a  natural-
draft tower-say in the range of 300 to 350 Btu/(hr. °F ft.') per row.

       An analysis  of the  heat transfer effectiveness "E" versus heat transfer size
"Nt(J" indicates that existing units  are designed for far less  than the  ideal maximum
heat transfer effectiveness due to cost  considerations in developing a cooling system
for the lowest expenditure of capital and operating expense.
                                     52

-------
       The heat rejection performance of the tower and the thermodynamic perform-
ance of the turbine are the two most significant factors in the operation of  a dry-
type condensing system.  The complex relationships which exist between the tower
and the turbine must be determined in order to predict the performance of a combin-
ation of turbine-generator and dry-type cooling tower.

       Since  the  performance  of a dry-type cooling tower system and the turbine
which it serves are so closely related,  the  complete condensing system (cooling
coils, method  of moying air, pumps, piping, condenser) and turbine  can best be
considered as one  integral unit  in studies of economic comparisons of various systems
of a dry tower for  a specific turbine.

       The  performance data and  prices for dry-type coils used in the economic
evaluation of this  report were furnished by the Hudson Products Corporation  for
mechanical-draft  towers and by Dr. Heller for natural-draft towers,  and represent
heat exchangers actually being  offered to the  utility  industry  by established
manufacturers.

       Certain of the heat exchange data supplied  by the manufacturers who co-
operated in this research are of a proprietary nature.  This proprietary information
was incorporated into our analyses and is reflected in the results shown herein; how-
ever, these data  are not included in our report in their original form, but may  be
available by direct inquiry to the original sources.

       Natural-draft towers.   During  conferences with Dr. Heller in Budapest,
basic information  was obtained regarding the performance of Heller-type  towers
which shows the relationship between heat transfer, flow of air and water,  natural-
draft tower height, and other factors.   This information  is shown on  the following
generalized curves.

       Figure 15 shows the relative coil performance in  heat rejection along with
the relative water- and air-pressure losses through the heat exchanger  coils.  The
range of values reflected by the curves  of Figure  15 are as follows:

            Air flow - 400,000 to 1,000,000 pounds per hour per column

            Air-pressure drop through coils - 0 to 0.4 inches head loss
                (water gauge)

            Water flow - 130,000 to 260,000 pounds per hour per column

            Water head loss through coils - 0 to  50 feet pressure head

            Heat transfer per 20-meter  heat exchanger column — 0 to
                100,000 Btu per hour per °F.
                                      53

-------
                RELATIVE  WATER  FLOW

              1.5      1.0      0.5       0
                  1.2



                  1.0
                          NOTE:
UNITS FOR ABOVE CURVES

ARE SHOWN  IN RELATIVE

MAGNITUDE ONLY TO

PROTECT PROPRIETARY
INFORMATION
                                                      Q.

                                                    8 2
                                                      UJ
                                                      tr
                                                      to o
                                                    5 UJ O
                                                      or
                                      3 ^
                                        UJ

                                      *£
                                                    I
                                                      UJ
                                                      Q:
1.5      2.0      2.5

     RELATIVE AIR  FLOW
                                       3.0
                3.5
FIGURE  15—COIL PERFORMANCE VERSUS AIR AND WATER FLOW
                               54

-------
       Figure 16 illustrates the relative  coil performance  as  related to the rela-
tive  height  of the  natural-draft  tower, water flow and ITD.  The range of values
reflected by the  curves of Figure 16 are as follows:

            Air flow - 400,000 to 1,000,000 pounds per hour per column

            Water flow - 90,000 to 300,000 pounds per hour per column

            Tower height - 0 to 400 feet

            Heat transfer per 15-meter heat exchanger column —  0 to
                90,000  Btu per hour per °F.

       These curves are  based on a combination of basic theory and proprietary data
developed in Hungary, so  they were analyzed  to determine their applicability to
conditions in the  United  States.

       When air  inside a natural-draft cooling  tower is heated by the coils, a draft
is  created,  causing an  upward flow of air.  At some flow of air, tower conditions
reach an equilibrium where the draft created by the heated air matches draft losses
caused by the flow of air.  The equation  for theoretical stack effect of heated air,
as taken  from (27),  is:

             D   =   .256  h p1 (j-  - y-J


             where:   D  =  draft, inches ^O

                      h   =  effective stack height, ft.

                      p1  =  atmospheric pressure, inches  Hg

                      TQ  =  ambient air temperature, °R

                      T   -  inside tower air  temperature, °R
                       &

       The  effective tower height must be corrected for a portion of the coil height,
elevation of the tower, ambient air temperature, and variation in atmospheric pres-
sure  due  to tower height.

       Draft losses may  be grouped into three  categories:  1)  draft loss across the
coil, 2)  exit loss (27), and 3) draft losses within the tower.  The heat transfer coil
used is a key factor in tower design.  The heat  transfer rate, water and air  flow
rates, and water and air  pressure drops through  and across the coil are interrelated.
For purposes of this study,  information on  coils  developed by Dr. Heller,  as shown
on Figure 15, was used.
                                      55

-------
                                               \Y
                              NOTE: UNITS FOR ABOVE CURVES
                                   ARE SHOWN IN RELATIVE
                                   MAGNITUDE ONLY TO
                                   PROTECT PROPRIETARY
                                   INFORMATION
             10.0     20.0      30.0     40.0     50.0
RELATIVE  TOWER  HEIGHT  TIMES  INITIAL  TEMPERATURE  DIFFERENCE
   FIGURE  16-COIL  PERFORMANCE  VERSUS  WATER FLOW,
 TOWER HEIGHT AND  INITIAL TEMPERATURE DIFFERENCE (ITD)
                             56

-------
       The exit loss is a function of the exit velocity of the heated air as it leaves
the tower.  The exit velocity is determined by the volume of discharged air and the
upper tower diameter.  There are several alternatives for estimating exit loss.  In
this study,  the exit  velocity was considered a function of the air flow per cooling
element.  Empirical data supplied by Dr.  Heller was used.

       The draft losses within the tower consist of losses due to frictional resistance
and changes in tower cross section.  These losses are small in comparison  with the
coil and exit losses.

       Mechanical-draft towers.  The sizing of mechanical draft, either forced or
induced, is much simpler than for natural  draft.  Since the required  air flow is
assured by the fans, the principal concern is the characteristics of the cooling coils
and the pressure loss across them.

       For purposes of this study, mechanical-draft data supplied by  Hudson
Products Corporation was utilized.  Some  of the information received was proprie-
tary in nature and, consequently, only a portion of it is illustrated.   Figure 17
shows the effect of cooling tower size upon  turbine back pressure for a range of am-
bient air temperatures.  The number of cooling units ranges from approximately 15
to 45 for a cooling range of 45°F to 15^F  for an 800-mw fossil-fueled unit.   From
these curves,  it is readily recognized that the use of 15 cooling units would resultin
extremely high turbine back pressure in a  location subject to high ambient air tem-
peratures.  On the other hand, 45 cooling units would provide extremely low back
pressures but at a considerably  higher cooling tower investment.

       Figure  18  illustrates the power requirements necessary to drive the fans and
pumps over a range of  ITD1 s from 30°F to  80°F.

       Variations in altitude will not affect tower configuration, but fans and drive
mechanisms must be varied in size to move the appropriate mass of air.

       Tower performance for varying load and ambient air temperatures.  The  op-
timum size and cost of natural- and mechanical-draft towers was established with
the analyses described above.  For the economic optimization of towers at a speci-
fic site,  it was necessary to consider their operation at part loads and with the full
range of temperatures that they would experience.  Therefore, it was necessary to
develop a means of evaluating their performance under these conditions.

        From an analysis of performance data of natural-draft,  dry-type cooling
towers that  have been designed in Europe, the relationship between ITD and heat
rejection for typical natural-draft towers  can be determined.
                                       57

-------
  100--
u.
e

 I
UJ
tr

-------
Oi
          CO
          t-
          z
          UJ
          s
          UJ
30
          O 25
          or
          UJ
         Q.


         Q.
            20
             30°
                                                    800  MW, FOSSIL FUEL PLANT
40°         50°          60°         70°         80°


     INITIAL TEMPERATURE DIFFERENCE ( ITD), ° F
                                                                    90°
                FIGURE 18—CURVE OF FULL LOAD AUXILIARY POWER REQUIREMENTS

                VERSUS ITD — MECHANICAL- DRAFT, DRY-TYPE COOLING SYSTEM

-------
        The performance of natural-draft towers was found to be reasonably ex-
pressed by the equation:

             1TD  =   AQb                                               [34]

             where:   ITD =  initial temperature difference, °F

                      Q   =  heat rejection,  106 Btu/hr.

                      A   =  tower constant

                      b    =  constant, depending on natural-
                             or mechanical-draft towers

        The performance of four existing or designed natural-draft towers was ana-
lyzed by computer with the results shown below.

        For the Ibbenbiiren and Rugeley plants, which have been built and are  in
operation:

             Ibbenburen (150mw)     ITD   -  0.501Q'717

             Rugeley  (120 mw)        ITD   =  0.247Q'793

        From information received from  M.A.N.  (Maschinenfabrik Augsburg-
Nurnberg) for a 200-mw, indirect, dry-type  cooling system,  designed but  not built,
using two different types of cooling coils:

             "A" coils               ITD   =  0.410 Q-762

             "B" coils               ITD   =  0.544Q'730

        The four exponents of the equations as found above are similar, ranging from
0.72 to 0.79 with an average of 0.75.

        Figure  19 shows the plot of the equation ITD  =  A Q  for the four natural-
draft towers analyzed.  Note that the approximate ITD for the Rugeley Station  is
35°F; for Ibbenburen,  50.5°F; and  for the 200-mw station with the "A" coil and "B"
coil,  80°F.

        The performance curves published by  Smith and Larinoff (13)  indicate that
the relationship between ITD and heat rejection for a mechanical-draft dry tower  is
approximately  linear.   The operating curves of the  Neil Simpson plant at Wyodak,
Wyoming, furnished by GEA, Gesellschaft fur Luftkondensation, also indicate a
near linear relationship for heat rejection versus ambient air temperature for a  fixed
                                     60

-------
                             SIGN POINT
                                RMANCE CURV
                       ON THIS CHART ARE IN NO WAY INTENDED
                       TO REFLECT SUPERIORITY OF ONE TYPE
                       OF COIL OVER ANOTHER
10
         200    400    600    800   1000    1200

             HEAT  REJECTION  ( I06 BTU / HOUR )
1400
     FIGURE  19—NATURAL-DRAFT, DRY-TYPE TOWER
       PERFORMANCE CAPABILITY  WITH VARIATION
       OF INITIAL TEMPERATURE  DIFFERENCE
                          61

-------
 turbine back pressure (saturated steam temperature), Figure 20.  Since ambient air
 temperature, used as the abscissa in Figure 20, is equal to the saturated steam tem-
 perature minus ITD, the ambient air temperature  varies inversely as the ITD for  a
 fixed saturated steam temperature  (turbine back pressure as used in the figure).
 Therefore, the slope of the curves on Figure 20 can also be considered to represent
 ITD versus heat rejection.  Appendix A, in the description of the Volkswagen plant,
 describes how Figure 20 is used as a guide for control of the cooling system.

        Based on the above information, the exponent  "b"  in equation [19] was
 established at 0.75 for natural-draff and 0.91 for mechanical-draft towers  to deter-
 mine part-load operation and to evaluate the effects of the variation in temperature
 during the year.

        The above findings as to the 0.75 exponent for natural-draft towers corres-
 ponds with the statement in (28), the only publication  found that had any reference
 to natural-draft,  dry-type cooling tower performance.  Some of the approximating
 rules it contains are:

             The  ITD is a function of heat rejection to somewhat more
             than the 2/3 power.

             The  air mass flow is a  function of heat rejection to the
             1/3  power.

             The  rise in air temperature is a  function of heat rejection
             to the 2/3 power.

             The  air flow becomes less  at partial heat rejection load
             because the chimney action of the tower  is decreased.

        The somewhat more than 2/3 exponential  curve for  ITD seems to be  reason-
ably close to our 0.75.  It is also stated in (28) that, with  mechanical-draft towers,
the ITD is approximately proportional to the heat  rejection of the tower.  This in-
formation is in accordance with the curves published in (13) and as shown in
Figure 20.

        The design point for the dry-type cooling  tower system for the Rugeley Sta-
tion is reported to  be 1 .3 inches Hg turbine back  pressure at 52°F ambient air tem-
perature.  Since the saturated  steam temperature  corresponding  to 1 .3 inches Hg  is
87°F, the ITD is 35°F (87°F -52°F).  The design  point could have been taken at
57°F ambient air temperature and 1 .5 inches  Hg turbine back pressure, since this
condition also represents an ITD of 35  F.
                                     62

-------
              calculated operating characteristic of the air-cooled steam condensing plant
              Black Hilts Power & Light Company -Wyodak
X»«W


190 see
Ethovti Steam Rot*
    of E*rtoust Steam
Exftousl Prrssm*
inifl Tfmprraluft of Cooling Air
Baromrltr
Tol
at Itw fan Shafts
FIGURE 20-GRAPH OF CALCULATED OPERATING CHARACTERISTICS
(PREDICTED PERFORMANCE)FOR DIRECT, AIR-COOLED CONDENSING
SYSTEM.NEIL SIMPSON PLANT, WYODAK .WYOMING (FROM GEA)

-------
        Design ITD.  The fact that a dry-type cooling tower system can have a num-
 ber of combinations of turbine back pressure design points and air temperatures for
 the same size system is illustrated in (13).  A system designed for 10 inches Hg tur-
 bine back pressure and an ambient air temperature of 100°F is identical in size and
 performance to a 6.9-inch Hg back pressure and 85°F ambient air design, or 3.6
 inches Hg back pressure and 60  F ambient air design since the ITD is 61  5°F in  each
 case, as shown in Table 5 below.

                                    TABLE 5

                      Possible Variations in Back Pressure and
                           Ambient Air for a  Given ITD

           Turbine     Saturated Steam         Air
        Back Pressure     Temperature      Temperature     Design ITD
         (Inches Hg)          (°F)               (°F)            (°F)
            10.0            161.5              100           61.5
             6.9            146.5               85           61.5
             3.6            121.5               60           61.5

       Undoubtedly, in order to forestall  confusion as to tower performance, stand-
ardization of back pressure and air temperature design will be accomplished when
air cooling systems are more prevalent.

       The performance of the dry tower for various ambient air temperatures and
heat rejection loads can be illustrated by performance curves in which the ambient
air temperature is plotted against turbine back pressure and heat rejection.
Figure 21  shows typical curves of dry tower performance plotted on the foregoing
basis for a natural-draft tower.

       Note that this tower has a dual rating 4.0 x 109 Btu heat rejection at 50°F
ITD and 6.Ox 109 Btu  at 67.8°F, illustrating the relationship of ITD to heat rejec-
tion capability.

       Figure 22 compares turbine back pressure as a function of ITD for various
ambient air temperatures.  This set of curves shows clearly how the dry cooling
tower design (ITD for heat rejection from a turbine operating at  full load) influences
turbine back pressure at various ambient air temperatures.  As an example, with an
ambient air temperature of 100°F/ an ITD  of 80°F will  result in  a  turbine back
pressure of 15.3 inches Hg; an ITD of 40 F will result in a turbine back pressure of
5.89 inches  Hg.
                                     64

-------
    16
    14
a   *
Z
o
    10
III
e
3
CO
to
teJ
oe
a.
u
<
a

UJ
—   4
a
oc
                  COOLING TOWER SYSTEM OPERATING CHARACTERISTIC CURVE- ITD=AQb

                      __J	I	I	I	1
4-
+
                               4567

                             HEAT REJECTION ( I09 BTU/HOUR)


                 FIGURE 21 —NATURAL-DRAFT, DRY-TYPE COOLING TOWER

                             OPERATING CHARACTERISTICS


             (ITD = 50°F AT 4.0x I09 BTU/HR   a ITD=67.8°F AT 6.0 x I09 BTU/ HR )

-------
                                                       70e
80*
                   INITIAL TEMPERATURF DIFFERENCE (ITD), °F
 FIGURE 22-DRY-TYPE COOLING TOWER SYSTEM: TURBINE BACK PRESSURE VARIATION
WITH INITIAL TEMPERATURE  DIFFERENCE (ITD)  FOR GIVEN AMBIENT AIR TEMPERATURES

-------
Performance of Turbine Used With
Dry-Type Cooling Towers

       An explanation of the turbine expansion  line and the available energy of
the steam as it flows through the turbine is useful in understanding the effect of back-
pressure variation on turbine efficiency.  The Mollier Chart is  a plot of steam
properties made from steam tables where enthalpy is plotted against entropy, with
other parameters such as temperature, degrees of superheat and percent moisture
also shown.

       Figure 23 shows how the expansion line  for a reheat turbine would appear
plotted on a Mollier Chart.

       Turbine throttle steam at enthalpy Ht expands through the high-pressure tur-
bine from point 1  to point 2, returns to the boiler at reheater pressure where the
steam temperature is raised to the hot reheat temperature (generally the same as the
initial throttle temperature), flows to the  intermediate section of the turbine at
enthalpy H^,. at point 3, and then expands through  the intermediate- and low-
pressure turbines to the exhaust pressure.

       Figure 23 shows the steam expanding to two exhaust pressures,  "A" and "B",
which will serve to illustrate the change in efficiency and turbine capability with
an increase  in back pressure.

       Under the first assumed operating  condition, the steam expands to point "A",
2 inches Hg back pressure.  In passing through the intermediate-pressure and low-
pressure turbines,  each pound  of  steam  does work  equivalent to Hg - H
-------
o >•
II
    UJ

-------
       The turbine cycle heat rate may  be expressed  as a  function of heat input
and generation.


             Heat rate   =  \>ea* inPut                                      [35]
                           kw  output

Therefore, the change in power output is inversely proportional to the change in
heat rate.

       Effect of back pressure on heat rejection of turbine.  The amount  of heat
rejected per pound of steam flowing to the condenser is expressed by the equation:

             Qr  =   Wc(Hx-hc)

             where:  Qf  =  heat rejection by exhaust steam,
                             Btu/hr.

                     W  =  flow of exhaust steam  to con-
                             denser, Ibs./hr.

                      HX  =  enthalpy of exhaust steam at
                             expansion  line end point,
                             Btu/lb.

                      h   =  enthalpy of condensate, Btu/lb.

       Assuming no subcooling of the condensate, h   is the enthalpy of saturated
liquid corresponding  to  the exhaust pressure of the  turbine.  The effect of  in-
creased turbine exhaust  pressure is increased heat  rejection per kwh.

       The capability of a turbine manufactured under current standards of design
and construction will be reduced  for turbine back  pressures above 3.5 inches Hg.
Figure 24(a) shows the effect of increased turbine back pressure upon the  heat re-
jection and capability of a nominal 800-mw turbine-generator unit operating  at
2,400 psi, 1,000°F/1,000°F throttle conditions.  Curve 1, Figure 24(a), shows the
operation of a turbine at full  throttle turbine output.  Below 3.5 inches Hg back
pressure, the capability is greater than 800 mw; at 3.5 inches Hg the capability is
800 mw. At 6.4 inches Hg back  pressure, the capability is 770 mw; at 9.3 inches
Hg,  the capability is 735 mw; and at 14 inches Hg it is 692 mw.

       Curve 2 shows the total heat rejection for the same unit when operating at
600 mw, and Curve 3 shows the total heat rejection when operating at  400 mw.
Additional throttle flow is required to maintain a constant load when there is an
                                      69

-------
 —  15.0
 o
 I

 CO
 UJ
 ~  10.0
 UJ
 IT

 CO
 CO
 UJ
 ce
 a.

 *   5.0
 a:
     1.0
	 1 —
TURBINE- GENERATOR
LIMITED BY BACK PF
3.5" Hg.






DESIGN OUT
JESSURE AB

f/1"
f 1
PUT
OVE

, /
1
i/*- FULL
/ / TUR
i 	 	 	
f. — 735 MW
i
^--770 MW
x 800 MW
! MW
. THROTTLE
IINE OUTPUT







                  1.0
                      2.0
                                      3.0
                                                 4.0
                                                     5.0
                                                                                7.0
                            HEAT  REJECTION  K I09  BTU / HOUR

       (a) EFFECT OF INCREASED  TURBINE BACK  PRESSURE  ON  TOTAL HEAT REJECTION  TO

          CONDENSER  AT  VARIOUS TURBINE  LOADS FOR  AN 8OO MW FOSSIL-FIRED PLANT
              50° ITD AT 4 X I09   BTU / HOUR
                                                                     6.0
                     2.0       3.0        4.0       5 0
                     HEAT REJECTION X I09 BTU  /HOUR

(b) DRY  COOLING  TOWER  PERFORMANCE  CURVES  REPLOTTED  FROM  FIGURE 21
                                                                                7.O
~  15.0
 o
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—  10.0
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-------
increase in turbine back pressure, assuming the turbine  is operating in a load range
which will  permit the throttle to pass the required amount of steam.

Combining  Performance of Cooling
Tower and Turbine

       Figure 24 (b)  is a plot of the heat rejection versus turbine back pressure of
the dry-type cooling tower, the performance of which is plotted on Figure 21,  re-
vised to show only the performance below 5 billion Btu per hour since this is the
range applicable to an 800-mw fossil-fueled generating unit.

       If the turbine heat rejection curves from Figure 24(a) are plotted  on
Figure  24 (b), as shown on Figure 24 (c),  the intersection of the turbine performance
curves  and  the tower performance curves represent the turbine back pressures which
will prevail with varying turbine-generator loads and air temperatures.

       From the combined curve, it is seen that at 800-mw load and 70  F air  tem-
perature, the back pressure will be approximately 3.5 inches Hg.  For full throttle
flow and 100°F air,  the back pressure will be 7.9 inches Hg at approximately 750-
mw maximum capability.

       Similarly, at  600 mw and 90°F, the back  pressure will be 4.7 inches Hg and
at40°F,  1  .Oinch Hg.

Comparison of Performance of Dry Tower
and Conventional Cooling Systems

       Since the variation in annual dry-bulb temperature  is greater than that of
natural water temperature or water temperature from an evaporative-type cooling
tower, the change in turbine back pressure for a generating unit equipped with a
dry-type cooling tower will cover a wider range  than that of a unit with  a surface-
type condenser and conventional cooling system.

       The wet-bulb temperature of the air is an important parameter in  the design
and performance of evaporative-type cooling towers since the wet-bulb temperature
of the  air is the lowest temperature to which the water circulating through the tower
can be cooled.  The term  "approach" is used in evaporative tower terminology to
designate the difference between the temperature of the cooled water leaving  the
cooling tower and the wet-bulb temperature of the ambient air.  The design wet-
bulb temperature of the air for a specific site is generally selected as that wet-bulb
temperature which is exceeded for no more than a small percentage of the time on
the average.
                                      71

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        The proper selection of the design conditions for an evaporative-type
cooling tower for use with a steam-electric generating unit is a complex process and
takes into account the capital costs of the tower for various approaches,  turbine
back pressure variation,  pumping and fan  power costs and, in general, requires an
analysis comparable to the economic selection of a dry-type cooling system.

        A wet-type  cooling tower  with  a  15 F approach will cool  the circulating
water to within 15°F of the ambient air wet-bulb temperature at design heat rejec-
tion load.  Carrying the design heat rejection load from the condenser, such a tower
would cool the water to 100°F when the wet-bulb temperature is 85°F.

        Figure 25 shows the variation  in average monthly dry-bulb and wet-bulb tem-
peratures for four locations in  the United States.

        Figure 26 from  (9) shows a diagrammatic  comparison of the turbine exhaust
pressures obtained with typical systems using dry- and evaporative-type cooling
towers.  Figure 26(a) shows the variation in back pressure and saturated steam tem-
perature corresponding to turbine back pressure under full-load conditions as func-
tions of ambient air temperature for a typical location .  Note that for the dry tower
(Curve  1), the variation  in turbine back  pressure has a greater range than  for the
evaporative tower.  This same trend is shown  in  Figure 26(b) where  the variation in
turbine back pressure for the two types of cooling systems is shown plotted against
varying turbine load with constant ambient air temperature. Curve  2 of Figure  26(a)
shows the operating characteristics of a dry tower with a smaller ITD than  the dry-
type cooling tower shown by Curve 1 . Since the turbine cycle efficiency is adversely
affected by rise in back pressure, the greater range in turbine back  pressure exper-
ienced with the dry tower results in a wider range of turbine heat rates as  compared
to wet tower operation.

        Also, greater loss of capability will generally be experienced with units
equipped with dry-type cooling towers.  Economic studies undertaken in this report
indicate that ITD of dry-type towers will be from 55 F to 60°F in areas where aver-
age conditions prevail.

        In a location typical of the eastern part of the United States, design para-
meters might be as follows:

             Dry-bulb temperature at 1 percent level:      90 F
                                                           o.
             Wet-bulb temperature at 1 percent level:      76 F

                 (Design temperatures at the 1 percent level  are
                 the temperatures which are equalled or ex-
                 ceeded by 1 percent of the 2,928 hours of June,
                 July, August and September in an average year.)
                                      72

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                             OMAHA, NEBRASKA
        BOSTON,  MASSACHUSETTS
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                                                          1         1    1
                                                          MIAMI, FLORIDA
CASPER, WYOMING
                        JAN  FEB MAR  APR  MAY JUNE JULY  AUG SEPT  OCT  NOV  DEC
                                                                                       \

                                                                                                DRY BULB TEMPERATURE
                                                                                                WET BULB TEMPERATURE
                                                                                       Source = U.S. Weother  Bureau
                                                                             JAN  FEB  MAR  APR MAY  JUNE JULY  AUG  SEPT OCT  NOV DEC
                                                  FIGURE 25—TYPICAL AVERAGE MONTHLY TEMPERATURES,

                                                                    DRY AND WET BULB

-------
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  AMBIENT AIR  TEMPERATURE — °F


   (a)  COMPARISON OF TURBINE
BACK  PRESSURE  AND  EXHAUST
TEMPERATURE AT CONSTANT LOAD
AS A FUNCTION  OF AMBIENT AIR
TEMPERATURE
                                          50%

                                     TURBINE OUTPUT
                                                         5.89  co

                                                              o
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                                                              co
                                                         1.93  co

                                                              £
                                                         1.03  %
                                                              CD

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                                                         0.52 I
                                                         100%
                                      (b) COMPARISON OF TURBINE
                                   BACK PRESSURE  AND  EXHAUST
                                   TEMPERATURE  AT  CONSTANT  AIR
                                   TEMPERATURE  AS  A FUNCTION OF
                                   TURBINE  EXHAUST  STEAM  LOAD
          FIGURE  26—COMPARISON  OF DRY TOWER AND
             EVAPORATIVE  TOWER  PERFORMANCE  (9)
                                74

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       For an evaporative-type cooling tower to serve an 800-mw generating unit,
a typical tower  might have the following characteristics:

             24°F water cooling range

             18 F approach to the wet bulb

              6°F condenser terminal temperature difference (TTD)

                 (TTD is the difference between the saturated
                 steam temperature of the turbine  exhaust and
                 the temperature  of  the  circulating water
                 leaving the condenser.)

       When 76 F wet-bulb  temperature is experienced and the turbine-generator
is at full throttle steam flow, the condensing temperature for the unit equipped with
the evaporative tower is determined in the following manner:

             Wet-bulb temperature

             Water cooling range

             Approach to wet-bulb temperature

             Terminal temperture difference

                 Condensing temperature:

        For the  same generating unit equipped with  a dry-type cooling tower of 60°F
ITD, the condensing temperature for 1 percent of the summer hours could be expected
to be 150°F, determined by  adding 60°F to the  90°F dry-bulb design temperature.

        The 124 F condensing temperature with  the wet-type tower corresponds  to
3.8 inches Hg back pressure, as compared to 7.6 inches Hg for 150°F with the dry-
type tower.

        Thus, for approximately 29  hours  per year the unit with  the evaporative
tower would suffer a 3-mw loss of capability (0.4 percent) while the  unit with the
dry-type tower would lose 47 mw (5.9 percent).

        Since the highest wet-bulb temperatures experienced in the United  States at
the 1 percent level are approximately 82°F, it  can reasonably be expected that loss
of capability because of high turbine  back pressure during hot weather will not be a
major factor with turbine-generator units equipped with evaporative-type cooling
towers.
                                      75

-------
        However, there are a number of specific locations in the United States
where the dry-bulb temperatures at the 1 percent level exceed 95°F.  With 50°F to
60°F ITD, the design range which appears to be a typical economic selection for the
United States, the condensing temperatures will be 145 F to 155 F, corresponding
to 6.7 inches Hg to 8.6 inches Hg back pressure, with a loss of rated generating
capability of approximately 5 to 7 percent (see Table 6,  page
        The turbine back pressures to be expected with once-through systems are
comparable to the back pressures  experienced with a typical evaporative-type cool-
ing tower, and, generally, the loss of capability during the summer with either the
evaporative-type tower or the once-through system would not be a major factor.

        As reported in (29),  the highest expected sea-water temperature in Miami is
approximately 86°F; in  Boston Harbor, 76°F; and in New York City, 78°F.  Also,
there are  few large rivers or lakes which would be considered for a once-through
condensing system where the maximum summer water temperature exceeds 85°F.

        Figure 27, reproduced from  USGS Water Supply Paper 520, shows the ap-
proximate mean monthly temperature of water from surface sources during the months
of July and August for the United States.

        It must  be emphasized  that the 60°F ITD used in the foregoing example for
a dry tower, the 24°F cooling water range with 6°F terminal temperature difference,
and the 18°F approach for the evaporative tower were selected only for  the purpose
of illustrating that the increase in turbine back pressure above 3.5 inches Hg and
resulting loss of turbine capability are more  significant factors with a  dry tower than
with an evaporative tower.  Either type of tower can be selected  to have more or
less loss in capability if economic considerations justify different  design parameters.

Application of  Present Large-Turbine
Design to Dry-Type Cooling Towers

        Available designs.  The only design  of large turbine-generators presently
available  from either United States or European manufacturers limits operation to
turbine back pressures below 5 inches Hg.  Historically,  the economics of large
utility turbine-generator operations  have been such that with conventional cooling
systems of the once-through or the evaporative cooling tower type, turbine back
pressures have been limited to an  upper  range of approximately 2.5 to 3 inches Hg.
The turbine ratings of presently available  turbine-generator units  are on the  basis of
maximum guaranteed kilowatt output at  3.5 inches Hg, with reduced capability for
back pressures above 3.5 inches Hg.

        According  to one leading  turbine-generator manufacturer, the experience
with high-back-pressure operation is limited  to small 3,600-rpm units with short tur-
bine buckets and small exhaust hoods.  If  the present design of large turbine-
                                      76

-------
25
             115°
                                                                                79°
              FIGURE 27 —APPROXIMATE MEAN MONTHLY TEMPERATURE OF WATER
                       FROM  SURFACE SOURCES FOR JULY AND AUGUST

-------
generators were to be used for operation at back pressures above 5 inches Hg,  prob-
lems would be anticipated in the following areas, unless certain modifications were
made.

        1 .    Bucket heating and vibration.

       2.    Thermal distortion of the exhaust hood and diaphragms which
             would cause misalignment and rubbing.

       3.    Abnormal  stress caused by thermal  cycling.

       Possible future designs. There are at least three possible approaches that
turbine-generator manufacturers might take to provide a turbine which will operate
satisfactorily at back  pressures above  5 inches Hg.

       1 .    Eliminate  the Last Row of Blades in the
             Tow-Pressure Turbine of Present Design

             This  method has been used by at least one European manufacturer on
             a 200-mw turbine for use with an indirect-type air-cooling system.
             The standard 200-mw turbine designed for 2 inches Hg back pressure
             is modified by removing the last row of blades, 28 inches long, and
             leaving the next row of 22-inch blades as the last stage to make it
             possible to operate the unit at higher back pressures.  The 200-mw
             turbine air-cooling system combination is designed for 6.6 inches
             Hg with 60°F ambient air.  With 90°F ambient air, the  turbine back
             pressure will rise to 15.6 inches Hg.

             In addition to the loss of capability which occurs at high back pres-
             sure, the capability of the turbine at all  back pressures  will be less
             than the capability with the row of 28-inch blades intact as a re-
             sult of the shortening of the steam  path.  For this reason, a further
             modification was made by enlarging the steam flow area of the high-
             pressure and intermediate-pressure turbines.  Also, provision was
             made to introduce steam into the turbine downstream of  the initial
             stage during times of high ambient air temperature to  compensate for
             loss of capability as a result of high turbine back pressure.

       2.    Design of  a Large Turbine to Operate
             at High Back Pressure              ~

             One approach to the problem of turbine operation with dry towers is
             to design a new line of  turbines, Curve No.  2 on Figure 28, for
             operation  at back pressures from 2  inches Hg to approximately 15
                                      78

-------
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                                          EXHAUST PRESSURE (INCHES  Hg)
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                        FIGURE 28—ESTIMATED TURBINE-GENERATOR, FULL LOAD, HEAT  RATE

                                  VARIATION WITH  ELEVATED  EXHAUST  PRESSURES

-------
             inches Hg.  The last-stage blades would be on the order of from 15
             inches to 20 inches in length, rather than the 26-inch to 33.5-
             inch blades used with large utility turbines currently designed for
             operation  up to 3.5 inches Hg back pressure and 3,600 rpm.

             The exhaust structure would be considerably redesigned and all
             stages of the turbine would have to be stronger and pass more steam
             flow than  present designs in order to compensate for the thermody-
             namic loss associated with  high exhaust pressures.

             Such a turbine is not now available, nor would we expect any de-
             velopment to be started by manufacturers until there was a demand
             for a large number of high-back-pressure turbines.

       3.    Modification  of Turbine of Present Design
             (Curve 3, Figure 28)

             Another method  available for operation at high back pressure would
             be the modification of a turbine of present design standards so that
             it would be suitable for operation at back pressures up to approxi-
             mately 15 inches Hg.

             The high-pressure and intermediate-pressure sections would essen-
             tially be the same as for the conventional turbines, except for the
             changes in steam flow to achieve over-all performance and rating
             differences, and the low-pressure turbine would have the same total
             exhaust annulus  as for the 3.5-inch Hg design.  However, the last
             several rows of blades would be redesigned for additional structural
             strength and would be  limited  to lengths between 25 and 30 inches.
             The smaller hood structure and shorter  bearing span that go with
             this length of  last-stage blade would help to solve the mechanical
             problems associated with high  exhaust  pressure.

        Figure 28 shows several curves representing the relative heat rates of the
different types of turbines described above.   The ordinate of the chart shows the
ratio of the heat rate of the  particular turbine under consideration to the heat rate
of the basic turbine at 3.5 inches Hg back pressure operation, represented  by
Curve No. 1 .  Note that this curve stops  at 5 inches Hg,  since 5 inches Hg is the
limit of back pressure operation recommended by the manufacturer.

        Curve No. 2 represents the  relative  heat rates for a  turbine especially de-
signed for high-back-pressure operation as described in paragraph 2 above. This
turbine  would have its best performance between 2 inches and 8 inches Hg  back
pressure, with increasing heat rate above  8 inches.  Note that the  turbine  designed
                                      80

-------
for high back pressure has poorer heat rate performance below 8 inches than the
modified turbine of present design, but has better performance above 8 inches. The
dashed-line curve represents the  heat rate performance of a turbine of high back
pressure design with different characteristics than the  turbine of Curve  No. 2.  By
tailoring the turbine design to the specific economic considerations for any particu-
lar application, it is theoretically possible to have a number  of such designs repre-
sented with performance between Curves 2 and 3.

       Curve  No.  3 shows the heat rate performance  expected from a conventional
turbine modified as described in paragraph 3 above.

       Correspondence with another major United States turbine manufacturer indi-
cates that this manufacturer is in the initial stages of a study  for high-back-pressure
application with dry-type cooling towers and believes that,  although it is theoreti-
cally possible  to modify present turbine designs for high-back-pressure operation,
such modification may not be economically or technically feasible with the present
state of the art.

       The economic evaluation studies in this report were performed on the basis of
information furnished by turbine-generator manufacturers for  turbine cycle heat rates
obtainable with presently designed large turbine-generators modified to operate at
back pressures higher than 5 inches  Hg, the present limit of back-pressure operation.
Although manufacturers are currently studying designs of large, high-back-pressure
turbines  for operation with dry-type cooling towers, no information as  to price or
performance is yet available.  The economic  results obtained in  this study may be
modified somewhat when turbine-generators designed  especially  for operation at high
back pressure are considered.  Both  the loss of capacity and the heat rate  character-
istics of  such turbines will be different from the characteristics of  conventional  tur-
bines.  However,  cursory studies indicate that the changed characteristics may not
significantly alter the production costs as found herein.

Use  of Recovery Turbine With
Main Circulating Pumps

       The circulating water system of the indirect, dry-type cooling  tower system
is usually designed so that a positive water pressure head of approximately 3 feet at
the highest elevation of the cooling coils exists at all times during operation.  The
purpose of this positive pressure is to prevent air leakage into the coils in case of
leaks.  Also,  with  positive water  pressure in the coils, any leaks will be apparent
to the operators.

        In  order to maintain positive water pressure in the coils, a restriction of flow
must be  imposed in the circulating water piping between the cooling tower and the
condenser. This restriction could be accomplished by the use of a throttling valve
which would be adjusted for varying circulating water flows  to maintain the desired
                                       81

-------
pressure at the high point of the coils.  However,  in order to recover the head that
would be lost across a control  valve,  a water turbine  (usually of the Francis design)
may be installed in place of a throttling valve.  Such a turbine is able to convert
approximately 85 to 90 percent of the head drop across the turbine into useful energy
and provide from 20 to 40 percent of  the power required for the main circulating
pumps.

       Generally, the recovery turbines are directly connected to the circulating
pumps on the same shaft as the pump-driving motor, but the hydraulic turbine could
also be used to drive a small generator.

       Figure 29 shows a diagram of  the pressures in the circulating water system of
an indirect dry tower system equipped with a recovery turbine.

Use of Multi-Pressure (Series Connected)
Direct-Contact Condensers With
Dry-Type Cooling Towers

       The large volumes of the steam flow to the low-pressure end of turbine-
generators in sizes above approximately 300 mw require that multiple low-pressure
turbines and condensers be used.  The circulating water flow through the condensers
may be either in parallel, with the flow divided for equal volume of flow through
the multiple condensers, or the flow may be in series, with the total  volume of cir-
culating water flowing through each condenser.  Either arrangement of flow can be
used with conventional surface condensers or with direct-contact condensers and
dry towers.

       Parallel circulating water flow through the condensers results in the same
pressure in  all condensers and  also the same final temperature of the circulating water
leaving each condenser.  With the  flow of circulating water through  the condensers
in series, the pressure in the first condenser will be lower than  the pressure in the
following condensers as a result of the increase in circulating water temperature
entering the following condensers.  The total rise  in the circulating water tempera-
ture will be the same for either parallel flow or series flow.

       In  the case of series connection of circulating water flow through the  con-
densers, the average pressure  in the condensers will be lower than the back pressure
obtained with the  circulating water flow in parallel through the condensers, assum-
ing the same quantity of exhaust steam and circulating water.  Because of the  lower
average back pressure, there is a slight  thermodynamic advantage for series connec-
tion of circulating water (also called  multi-pressure condensers), and, for this rea-
son, a number of  large turbine-generator units in the  United States with surface
condensers have been constructed with series connection of circulating water.
                                      82

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    DRY TOWER EQUIPPED WITH WATER TURBINE
                       83

-------
        Figure 30 shows a diagrammatic arrangement of the two types of circulating
water connections with surface flow condensers.

        With surface-type condensers, one circulating pump can  handle the  total
flow of circulating water for series-connected condensers since the circulating water
flows through the tubes of the first condenser and, subsequently, through the  tubes
of the following condensers in an integral hydraulic circuit without coming into
contact with the steam.  However, in the case of the direct-contact type steam con-
denser used with dry-type cooling towers,  the circulating water and steam are inti-
mately mixed in each condenser shell so that it is necessary to convey the mixture of
condensed steam and circulating water from the first condenser to each subsequent
condenser, either  by pumping or by gravity flow.  A design has been developed  by
Dr. Heller which takes advantage of the low pressure  drop of the spray nozzles in the
direct-contact condenser to permit the transfer of circulating water from one  conden-
ser to the next by  means of gravity. The downstream condensers are located at a suf-
ficiently lower elevation than the upstream condensers to permit gravity flow  (30).

        Multi-pressure operation of condensers and series connection of circulating
water flow has a distinct advantage with a dry tower installation because multi-
pressure operation  results in a greater ITD with the same average turbine back pres-
sure and circulating water  flow, as compared to the ITD obtained with single-pressure
condenser operation and parallel  circulating water flow.

        Figure 31 shows the temperature and pressure relations which exist in single-
pressure  and muIti-pressure condenser installations for the same heat rejection and
circulating water flow for direct-contact type condensers.   In  Figure 31 (a),  for
single-pressure operation TWJ is the temperature of circulating water to the conden-
sers, R is the rise in circulating water temperature in  the condensers, and Tp is the
temperature of saturated steam in the condensers  (which is the same as TW2 ,  the cir-
culating water temperature leaving the condenser, assuming no subcooling of  con-
densate).   The initial temperature difference for the mu I ti-pressure condensers is
ITDp,  and numerically is the difference  in degrees  Farenheit between Tp and  the
ambient air temperature.

        Figure 31 (b) shows the temperature-pressure relationship which exists in  a
multi-pressure condensing system  with the same heat rejection, the same quantity of
circulating water flow, and designed for the  same average turbine back pressure as
the single-pressure condenser system in Figure 31  (a).

        Half of the steam is condensed in Shell  No. 1 and half in Shell  No. 2.
Consequently, one-half of the rise in circulating water temperature occurs in Shell
No. 1  and one-half in Shell No. 2.  The condenser pressure in Shell No. 1  is  the
saturated steam pressure corresponding to the temperature of the circulating  water
leaving Shell No.  1, TW4, which is also the temperature of the circulating water
                                       84

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STEAM FROM REHEATER
STEAM  FROM
SUPERHEATER
                             LOW PRESSURE
                               TURBINES
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REHEATER
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                TURBINE
 CIRCULATING WATER IN
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SURFACE
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                                  CIRCULATING WATER OUT
      (a)  DIAGRAM OF PARALLEL CONNECTION OF
          CIRCULATING WATER FOR SURFACE CONDENSERS

STEAM FROM REHEATER
STEAM FROM
SUPERHEATER
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      (b)  DIAGRAM OF  SERIES CONNECTION OF
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    FIGURE  30 — CIRCULATING WATER FOR  4 FLOW
    EXHAUST TURBINES WITH SURFACE CONDENSERS
                           85

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00
o
                     TEMP.  OF AMBIENT AIR
                   PARALLEL FLOW OF
                   CIRCULATING WATER
                   (SINGLE-PRESSURE)
                          (a)
CONDENSER  CONDENSER
  SHELL  I  SHELL
   NO. I   I  NO. 2
   TEMP OF AMBIENT AIR
   SERIES FLOW  OF
   CIRCULATING WATER
   (MULTI-PRESSURE)
          (b)
                            FIGURE 31 —TEMPERATURE-PRESSURE DIAGRAM
                        OF PARALLEL-AND SERIES-CONNECTED, DIRECT-CONTACT
                              CONDENSERS AND DRY COOLING TOWERS (30)

-------
entering Shell No. 2.  The condenser pressure in Shell No.  2 is the saturated steam
pressure corresponding to the temperature of the  circulating water leaving Shell
No. 2, TW5 .  TW5 is also the temperature of the water entering the cooling coils
of'the dry tower, assuming no subcooling.

       The average of the condenser pressures of Shell No.  1  and Shell No. 2 for
the series connection is equal to the condenser pressure in the parallel  flow conden-
ser, but since TW5 is greater than TW2 by the amount R/4, the ITD of the multi-
pressure condensing system is greater than the ITD of the single-pressure condensing
system by the quantity R/4.

       Since the capital cost of a dry tower is inversely proportional to the ITD, it
can be expected that a  less  expensive dry tower  can be constructed for the same  tur-
bine back pressure design and circulating water flow rate if  the circulating water is
connected in series through  the condensers.

       This conclusion  has been verified by actual  studies made by Dr. Heller1 s
group for specific plant installations with the result that the estimated  capital cost
of the  dry tower system  can  be reduced by as much as 10 to 12 percent if the circu-
lating  water is connected in series  (30).

Effect  of Air Temperature at Site

       Turbine performance.  The variation in ambient dry-bulb air temperature at
the site has an effect on dry-type cooling tower  performance.  The expected air
temperatures at any particular location must be taken into account in  selection of
the ITD of the tower design. Generally, at locations with lower average air tem-
peratures, dry-type cooling towers with greater  ITD will be  selected than for sites
with high average air temperatures.

       With any particular  tower design, an increase in air temperature will result
in higher turbine back pressure and a consequent increase in plant heat rate.  When
the back pressure exceeds the maximum point at which rated turbine capability can
be  achieved  (3.5 inches Hg for turbines of present design standards), the capability
of the  turbine-gen era tor is reduced.

       Table 6 illustrates the estimated loss of capability at high ambient air tem-
peratures for an 800-mw turbine-generator unit with steam conditions of 2,400
pounds per square inch, 1,000 F/1,000°F, and  guaranteed  to deliver  rated capa-
bility at 3.5 Inches Hg back pressure when equipped with a dry-type cooling tower
of  60°F ITD.
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       Table 6 shows the reduction in turbine-generator output for back pressures
from 1 .0 inches Hg to 14 inches Hg back pressure with the corresponding ambient
air temperatures for the 60°F ITD tower.

                                   TABLE 6

               Variation in 800-Mw Turbine-Generator Capability
                     Due to Changes in Back Pressure With a
                           60°F ITD Dry-Type Tower


       Ambient Air          Turbine                         Percent
       Temperature       Back Pressure        Output           Rated
           (°F)              (In. Hg)           (mw)         Capability

             19                1.0           809.98           101.2
             32                 1.5           809.48           101.2
             41                 2.0           808.98           101.1
             49                2.5           806.89           100.9
             55                3.0           803.92           100.5
             60                3.5           800.00           100.0
             65                 4.0           795.54            99.4
             70                 4.5           790.27            98.8
             74                 5.0           784.03            98.0
             77                5.5           777.44            97.2
             81                 6.0           771.04            96.4
             87                 7.0           759.53            94.9
             92                 8.0           748.52            93.6
             97                9.0           738.16            92.3
           101                10.0           728.25            91.0
           106                11.0           718.60            89.8
           109                12.0           709.12            88.6
           113                13.0           700.34            87.5
           116                14.0           692.14            86.5

       Table 6 was prepared from performance data furnished by General Electric
Company for a tandem-compound, 6-flow turbine-generator modified for operation
at high back pressure and is on the basis of full throttle flow performance.

       Table 6 does not  reflect any possible recovery of capability which might be
obtained  by taking feedwater heaters out of service, use of over-pressure  throttle
steam, or by providing a second  steam admission point on  the turbine with increased
boiler capacity.
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       The 60  F ITD was arbitrarily selected for the table in order to indicate per-
formance for back pressure ranges up to 14 inches Hg with air temperatures which
represent a typical site in the United States.  Other ITD selections would  result in
different air temperature-back pressure combinations at full throttle conditions.
For example, with a 50 F ITD tower, 3.5 inches Hg back pressure would be  obtained
with 70°F air rather than at 60 F as shown in Table 6.

        Freezing. Air temperatures below 32  F cause potential problems of coil
freezing.  Provisions must be  made in the design of the system to prevent coil freez-
ing during cold weather. The problem of freezing is especially prevalent  during
periods of light load and during start-up.

       The freezing problems which the operators  of the existing dry tower plants
have experienced and the measures taken to remedy freezing  are reported  in some
detail  in Appendix A.   It is likely that with a proper automatic control system for
start-up operation and  shutdown and with adequate alarms, freezing of dry-type
cooling towers will not be a problem.  However, unless a completely automatic  con-
trol system for tower operation is provided, much of the success in preventing freez-
ing lies with the plant  operators.  Thorough training must be given to the  operators
before a plant is placed into service.

        Historically, the freezing of coils which has occurred generally has been
during the early period of initial service before the operators were thoroughly famil-
iar with procedures to  prevent freezing.  A complete automatic control system to
manage as many operations as possible  is desirable with a dry tower system.  Such an
automatic control system should  initiate and automatically accomplish all functions
of taking cooling coils out of service when such action is required because of cold
weather and should automatically return the cooling sections to service later on.
The control  system should also generally perform all tower operations which are  nec-
essary to prevent freezing or  to operate the tower  during freezing weather to the
extent that reliance upon the judgement of operators is minimized.

        Auxiliary power.  With a mechanical-draft tower, more fans will  be re-
quired during hot weather than during  cold weather; consequently, the fan auxiliary
power requirements of  the tower will be greater during hot weather.

        Since the volumetric  capacity  of the fans to move the required cubic feet per
minute of air must be based upon the highest air temperature expected, air tempera-
ture variations at the site must be taken into account in selecting the fans and motor
drives.  The horsepower of the motor driving the fan,  however, must  be based upon
the coldest air temperature expected since the density of the air increases with the
lower temperature, while the volume delivered by the fan remains constant for fixed-
pitch fans.  Variable-pitch fans  can be used to reduce fan power requirements at
either part loads or during  cold weather.
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        With a tower design which has multiple circulating water pumps, it may be
 possible to take some of the pumps out of service during cold weather thereby reduc-
 ing auxiliary power requirements.

        Natural-draft cooling tower.  The performance of a natural-draft, dry-type
 cooling tower, with a given ITD and heat rejection load, will be affected by the
 ambient air temperature in two ways.  First, the available draft for moving air
 through the coils is less at elevated  ambient air  temperatures than at lower ambient
 air temperatures.  The available draft  is reduced, for example, by 14 percent when
 the ambient air temperature increases from 50 F to 86  F.  The second effect  is  an
 increase in draft  loss through the  tower as the ambient air temperature is increased.
 This effect is caused by the increased volume and corresponding increased velocity
 of air required to move  the same air  mass across  the coils as compared to operation
 at the  lower temperatures.

        As a result of the two effects,  the design height of the natural-draft tower
 must be increased to maintain  the required heat rejection performance at higher am-
 bient air temperature locations.

        Mechanical-draft cooling tower. Increased ambient air temperatures result
 in greater air volume requirements for  the same mass flow of air (cooling capability
 requirements).  This greater air volume, in turn,  results in  increased air pressure
 drop across the heat exchangers.  The  combination of increased air volume and  pres-
 sure loss requires increased fan horsepower  for mechanical-draft cooling systems
 operating under higher ambient air temperatures.

        Cooling water for auxiliary purposes.  The cooling surfaces supplied  with
 standard design of generator cooling, turbine oil cooling, and other auxiliary plant
 services generally require cooling water of a maximum temperature of 95°F.  Because
 the ambient air temperature will be above the temperature at which 95°F water can
 be obtained from  the dry-type cooling  tower during part of the average  year, it is
necessary to install  means of cooling sufficient water for auxiliaries to a maximum
temperature of 95°F at all times.  The  description of the equipment available for
this service is found in Section  V  of  this report.

 Effect of Precipitation and Humidity

       Rain.  Based upon the  reports of the operations of dry tower plants which
 have been constructed,  rain has an effect upon the performance of dry towers.  At
the Rugeley  Station in England and at the Ibbenburen Station in Germany, rain  re-
sults in poorer tower performance. Both of these towers are of the natural-draft type.
Rain reduces the draft through  the tower because it cools the air inside the tower,
consequently reducing the thermal lift. The reduction  in draft diminishes the air
flow through the cooling coils which causes a higher turbine back pressure.  Con-
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versely,  the effect of a wetted coil surface is to increase heat rejection performance
because of the evaporation of the water on the coil surface.  In the case of the large
natural-draft, dry-type cooling towers, the small gain in performance from wetted
coil surface during rain is nullified by the  loss of draft because of the rain cooling
the heated air inside the tower shell.  A third possible minor effect of rain on the
performance of cooling coils is an increase in air-pressure drop through the coils be-
cause of the reduction in air passage area caused by water on the coil surface.

        Since the flow of air through the cooling coils of a  mechanical-draft, dry-
type cooling tower is not dependent upon thermal lift, rain does  not have an adverse
effect upon the performance of mechanical-draft towers. No adverse effects from
rain have been reported for the  Volkswagen plant or the Neil Simpson plant.

        Hail.   In areas where hail storms occur, some protection in the form of hail
screens should be considered for cooling coils, especially if the coils are installed in
a horizontal position.   The  degree of protection will  be influenced by the structural
strength  of the coil fins and the ability of the fins to withstand distortion or damage
from hailstones.

        In process industries located in areas prone to hailstorms, it has been cus-
tomary to use hail screens with forced-draft fans but often not with induced-draft
fans, where the fans themselves provide protection for the coil fins.

        Sleet or snow.  No adverse effects resulting from sleet or snow plugging air
passages of the cooling coils of the dry-type tower have been  reported.

        The Ibbenburen plant, located in an area where freezing rain occurs regu-
larly during the winter, has not experienced trouble. No problems were encountered
at the Neil Simpson plant at Wyodak, Wyoming during heavy  snowstorms.

        Dr. Heller has advised that plants installed in northern Russia have had no
plugging problems caused by snow or sleet.  However, louvers on the coil face, or
across the area of air  inlet, which could afford protection against sleet or snow by
closing off the coil sections exposed to the wind, should be considered for natural-
draft, dry-type cooling towers to be located in severe-weather zones.

        Humidity.  Since the temperature of the cooling-coil  surfaces is above the
dew-point temperature of the air passing through them,  the humidity of the air has
no noticeable effect  upon  coil performance.  However, fog improves performance
of the Rugeley tower, according to published reports (31).

 Effect of Wind Velocity and Direction

        Natural-draft cooling towers.   In general, wind causes  poorer performance
 of a nature I-draft, dry-type cooling tower than the performance obtained under con-
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ditions of no wind.  This adverse effect is generally considered to be due to the low-
pressure area which develops on the lee side of the tower as a result of air  current
eddies caused by the high air flow around the tower.  The low pressure on the down-
wind, or lee side of the tower produces a reduction in static pressure available for
producing air flow through that part of the cooling sectors and a higher turbine back
pressure is experienced under these conditions.

       At Rugeley, no increase in  turbine back pressure is experienced until the
wind speed reaches 10 mph, but at  Ibbenburen the effect of the wind is felt at lower
speeds. The increase in back pressure at Rugeley  increases approximately 0.1 inch
Hg as compared to approximately 0.3 inch Hg for  Ibbenburen (32) (33).

       Dr.  Heller has advised that his nature I-draft,  dry-type towers are designed
to maintain  guaranteed heat rejection at 4 meters  per second wind velocity (9 mph),
which Is the German standard for the industry for both wet- and dry-type cooling
towers. One  interesting aspect of the effect of wind  upon cooling tower performance
in raising turbine back pressure is that wind has an adverse effect upon a natural-
draft, wet-type cooling tower as well as upon a natural-draft, dry-type cooling
tower.  The cooling air flow through the natural-draft, wet-type  tower is subject to
the same influences of eddies on the lee side of the tower as is the natural-draft,
dry-type tower.  However, the large mass of cooling  water in the wet-type tower
storage basin with which  the cooled water from the tower  mixes before  returning to
the condenser has a dampening effect upon any immediate influence of the  wind on
turbine back pressure. With the dry-type tower, the  effect of wind is felt  imme-
diately since there is no large storage of circulating water.

       According to Dr.  Heller, a  wind velocity of 4 meters per  second, for which
natural-draft towers are designed, causes a reduction  in heat dissipation of approxi-
mately 5 percent as compared to calm conditions.

       Reports of tests on the Rugeley tower  (32) indicate that under high-wind con-
ditions the static pressure on the lee side of the tower was actually higher than the
pressure inside  the tower and not lower as had been predicted from wind-tunnel tests.
These tests also indicate that tower  performance  is adversely affected by the tangen-
tial wind components which tend to reduce air flow through the downstream coolers
exposed to the  tangential winds.  The air flow reduction due to the combination of
the above factors more than offsets the beneficial effects of the increased air flow
through the coolers  on the upwind side of the tower.

       M.A.N., a  European supplier of natural-draft, dry-type  cooling tower sys-
tems, uses a design with the cooling sections  in a horizontal position inside the base
of the tower shell with air flow upward through the coils with a clear space beneath
the tube bundles for air passage.  M.A.N. has indicated that wind-tunnel  tests
have shown reduced wind influence on tower  performance  with this coil  arrangement.
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A 200-mw installation of such a design is now under construction in South Africa at
Grootvlei Power Station and is scheduled for 1972 operation.

       Mechanical-draft cooling towers.  The influence of average wind velocities
upon the performance of  dry-type  towers equipped with motor-driven fans to move
the air through the  cooling coils is almost negligible.  Operating results of  the
Volkswagen plant in Germany and the Neil Simpson plant in Vfyoming have not in-
dicated any significant influence on tower performance as a result of wind. Both of
these plants  are equipped with GEA direct-type,  air-cooled condensing systems
which utilize mechanical draft to move air across the condensing coils.

Effect of Dust
        The deposit of dust on the outside surface of cooling-coil fins and tubes has
not caused significant difficulties in cooling tower operations. By cleaning the coils
periodically with either water or air pressure, the operators of the existing dry-type
tower installations have been able to keep the exterior cooling surface sufficiently
clean so that performance has not been affected.

        The experience with  the Rugeley Station tower, however, indicates that
local coal dust, which at that station is reported to contain a percentage of chloride
compounds, may have been a factor in  the severe  corrosion which occurred in its
cooling coils.  The problem was determined to have been corrosion cells set up in
the minute cracks between fins and spacer collars  of the Forgo coils,  likely due to
high humidity and atmospheric pollution which resulted in deposits of moisture and
chlorides (32).  The source of the chlorides has been variously attributed to carry-
over from adjacent wet-type cooling towers, the salt-bearing coal dust, and salt-
laden fog from the sea coast  approximately 150 miles away, but no definite conclu-
sion has been announced.

        Although there is a possibility that coal  dust was a significant factor in cor-
rosion of the coils at Rugeley, the fact that other dry tower stations have not
experienced  such corrosion,  although coal dust, soot, and dirt have built up on the
exterior cooling-coil surfaces, would lead to the  conclusion that the Rugeley ex-
perience is unique and that utilities considering the use of dry towers  could expect
little trouble from exterior dirt on the cooling coils.

        The plugging of air-cooled coils, as a result of vegetation and debris in the
air stream, presents a more serious problem than deterioration of performance from
dust and soot, especially at  power plant sites in rural areas where material  such as
cottonwood seed may be present in the air during  certain seasons of the year.   How-
ever, even this problem can be readily overcome  by seasonal application of screens
and/or vacuum cleaning.
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Effect of Radiation and Cloud Cover

        Since the cooling surfaces of the dry cooling coils are of such a structural
configuration that only a negligible portion of the fin and tube areas are exposed,
the effect of radiation from the sun is negligible.

        Actual operating experience at Rugeley, as reported by  Christopher (31),
substantiates the above conclusion; intermittent sun produces only a flicker on the
turbine  vacuum gauge.  However, sunshine and cloudiness are reported to have an
influence on the air-cooled, direct-condensing system at the Volkswagen plant.

Effect of Topography

        The  topography of the plant site utilizing dry-type cooling towers is gener-
ally of no great concern in influencing tower performance.  The same considerations
which govern plant site selection for generating plants with  evaporative-type cool-
ing towers or with once-through cooling  systems will hold true for dry-type tower
sites.

        Flat, level terrain is  to be preferred.  Differences in site elevation may
affect the pumping head and  auxiliary power requirements of the circulating water
system,  depending upon the individual design. In general,  the  site  location prob-
lems associated with  dry-type cooling towers would seem to  be of less magnitude
than the site problems of wet-type cooling towers, since the dry tower is less sus-
ceptible to the problems of recirculation of discharged air (air and water vapor in
the case of the wet-type cooling towers), especially when the plant site is located
in a valley.  Fogging problems as a result of tower discharge are not encountered
with dry-type  cooling towers.

Effect of Elevation

        The  elevation of the plant site above sea level must  be taken into consider-
ation  in the design of dry-type cooling towers.  The same considerations which
affect dry tower design because of air temperature variations are also factors in the
tower design for different locations.

        Since a greater volume of air must  be moved through the tower at higher
elevations to achieve the same mass flow of air, provisions must be made to increase
the air flow for towers located at higher  elevations.  With natural-draft towers, this
is accomplished by increasing the height of the tower.  For mechanical-draft towers,
higher capacity fans must be  installed.

        According to studies made by  Dr. Heller, the capital cost of a natural-draft
tower is increased by approximately 4 to 4.5 percent for each 1,000-meter rise in
                                      94

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site elevation.  Figure 32 shows the relation between required height of a natural-
draft, dry-type cooling tower at various elevations and height of the tower at sea
level  to achieve equal heat rejection performance, as plotted from Dr.  Heller's
studies.
                                       95

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ir
UJ
                                NOTE • HEAT REJECTION IS
                                     CONSTANT
         -10
0       10      20      3O

   AIR  TEMPERATURE, *C
     FIGURE 32 — RELATION  OF NATURAL-DRAFT DRY-TYPE
      COOLING TOWER HEIGHT AT VARIOUS ELEVATIONS TO
                 HEIGHT AT SEA  LEVEL
                          96

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                                  SECTION IV

                         STRUCTURES AND MATERIALS


General

       Basically, there are two types of structures used for nature I-draft, dry-type
cooling towers:  reinforced concrete structures and  structural steel  framework
covered with thin siding material.

       The type of structure used for mechanical-draft,  dry-type  cooling  towers
consists of modular cooling cells of prefabricated components which are assembled
at the site.   The cells consist of heat exchanger coils, fans, motors and structural
steel supports.  Generally, the height of mechanical-draft towers is below 100feet,
and the supporting structures are relatively  light as compared to natural-draft
towers.

       The natural-draft tower consists of a shell of either cylindrical or hyper-
bolic shape,  having a height and diameter sized to the air-moving requirements of
the particular design.  Generally,  natural-draft, dry-type cooling towers require
less ground area than  mechanical draft, dry-type cooling towers of equivalent heat
rejection capacity.  (See Appendix E for cost data.)

Reinforced Concrete Structure,
Natural-Draft ToweT

       Since the cost of a reinforced-concrete,  natural-draft tower of hyperbolic
shape is generally less than the cost of an equivalent tower of cylindrical shape-
especially in the larger sizes above 400 mw—concrete natural-draft towers are
usually hyperbolic in shape.

       The hyperbolic concrete tower has a  relatively thin concrete shell of vary-
ing thickness which is greatest at the base.   The shell is terminated at the top of
the cooling coils and is supported from the ground by a cross-bracing structure
which  serves as the supporting columns and also provides the shell  opening for air
flow.  The shell must be stiffened at the top  and base with a ring beam to take the
concentration of stress at these points.  The columns are supported by a continuous
ring beam, and piling is provided under each column.

Structural Steel Natural-Draft Towers

       The structural steel tower would be of cylindrical shape using prefabricated
welded elements for the skeleton and covered with aluminum siding material.
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        The hyperbolic shape does not seem to be suitable for steel construction
 because of the difficulty in covering a hyperbolic shell with siding.  Also, the hy-
 perbolic shape is difficult  to analyze structurally by the membrane theory.

        The cylindrical tower would consist of prefabricated structural steel sections
 to form the main stack of the tower; either bolted or welded into place.  The stack
 must be stiffened at the top, bottom, and at intermediate points by trussed stiffen-
 ing rings to insure stability and to prevent the stack from becoming oval during wind
 loading.  The main columns should be supported  by reinforced concrete pads and
 piling, since piling Is required to counteract the upward  force caused by wind loads.

        The tower includes a delta roof structure to enclose the additional area re-
 quired for the base diameter of the cooling coil arrangement.

        The steel structure  can  be either galvanized or painted.  A cost comparison
 of painting versus galvanizing indicates that in the United States the galvanized
 structure would cost more initially,  but would be cheaper throughout the life of the
 structure, taking into account reduced maintenance and painting costs.

        The steel tower would be erected by  means of a crane which operates on
 rings inside the tower, a technique developed by Professor Heller's group. The
 rings can be left in place as  stiffeners.

 Design Loadings

        The design live loads for steel structures are controlled by wind load on the
 structure.  Seismic loads are not critical because of the relatively light dead load
 of the steel structure and aluminum siding .  The normal wind load is based upon a
 100-mph wind velocity at approximately 30 feet above ground level with variation
 of pressures according  to heights. This load  should be considered for all areas of
 the United States except for  locations subjected to hurricanes.  A wind velocity of
 120 mph should be considered for these areas.

        The design live loads for concrete structures are generally controlled by
wind loads on the structure, except in heavy earthquake areas (Zone 3) as defined
 by the Uniform Building Code.   The  wind loads are the same as defined for the
 steel structures.

        Hurricane  loads (120 mph wind) develop approximately the same maximum
 stress condition as for heavy earthquake loadings.

 Cost Comparison

        Total construction cost of cylindrical steel structures for unit sizes as used
 in this report will generally be  from  15 to 18 percent lower than costs for hyper-
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bolic concrete towers. The cost differential increases as the tower size increases.
Also, when tower structures, alone,  are  considered,  it would  be considerably
cheaper to build one large-capacity  tower than two smaller sizes for either steel or
concrete. The decisive factor in choosing the type of construction is solely the in-
vestment costs for a given locale which,  depending on labor rates and material
prices, may favor concrete or steel construction.

       Tower costs of either material type would be increased from 13 to 15 percent
in areas which are subject to hurricanes and heavy earthquake zones.

Corrosion of Coils and Fins

       In the  design and construction of dry-type cooling towers, particular atten-
tion must be given to the possibility of corrosion of the external surfaces of the fins
and  tubes as a result of atmospheric contaminants,  salt-laden fog, or catalytic
action between dissimilar metals.

       Before selection  of the  tube and fin  material is made,  a comprehensive
study and survey should be completed at the plant site under consideration in order
to obtain information as to the ability of various materials to withstand corrosion.

       The Marley Company of  Kansas City has recently conducted a series  of cor-
rosion and fouling tests to determine which  tube and fin  materials best withstood
exposure at a number of typical  power plant sites.  To provide accelerated testing,
the tube samples were not carrying heated fluid.  As a result, the corrosion rate
was greater than would be experienced during normal operation of a dry-type cool-
ing tower.

        Permission has been obtained  from  the Marley Company  to  include the
following summary of their test program results:

                                MARLEY  COMPANY

                                    "SUMMARY

                     CORROSION AND FOULING OF DRITOWER
                           HEAT EXCHANGER SURFACES

             "Corrosion  of fins  and tubes in the  dry cooling  tower at Rugeley
          Station of the Central Electricity Generating Board in England alerted
          us to the possibility of corrosion and fouling  of Dri tower heat exchanger
          surfaces in the USA.  To check this possibility, cooperative test pro-
          grams were established with American Electric Power and Jersey Central
           Power and Light.  The program with Jersey Central Power and Light en-
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 tailed only external corrosion tests  but  the  program with American
 Electric Power also included an internal corrosion test.

   "The external corrosion test consisted of specially constructed test
 units in which numerous combinations of alloys and coatings could be
 simultaneously exposed to a constantly  moving air stream.  The fins
 were made of aluminum alloys 1100,  7200,  and 3003, and the tubes
 were of aluminum alloys 3003, 6061, and 5052, and also of carbon
 steel, copper and admiralty .  Some of the samples were electrocoated
 with acrylic both  before and after finning.  Five external units were
 constructed and exposed at sites ranging from  sea coastal and  heavy
 industrial to clean rural midwestern.  At intervals of 8 and 16 months
 representative samples were removed, examined, cleaned, and re-
 examined and the results recorded .

   "Internal corrosion tests consisted of  sample tube strings of  copper,
 admiralty, and aluminum alloys 5052, 3003, 6061, and welded Alclad
 3004. Condensate at 140 F from a supercritical unit  was passed through
 the tube strings at 5 feet per second.  At  intervals, sections from each
 string were removed, cleaned, weighed, and  the corrosion rates calcu-
 lated.

   "Corrosion of uncoafed aluminum fins was severe in the sea  coastal
 and  heavy industrial exposures and moderate in the others.  There was
 no significant corrosion of external tube surfaces in any of the expo-
 sures.  Fouling was slight to moderate except  at the sea coastal,  heavy
 industrial and clean  rural exposure sites.  Vegetation was the  sole cause
 of fouling at the rural midwestern  site but corrosion products were an
 important cause of fouling at the other two.  The sections electrocoated
 after finning showed little corrosion and little fouling at the sea coastal
 and  heavy industrial sites but electrocoating had little effect  on fouling
 at the rural  site.

   "Internal  corrosion rates were significant  in all aluminum tubes carry-
 ing condensate.  Total corrosion was as high as 5 1/2 mills (.0055inches)
 for some alloys after 12 months exposure.  There was little corrosion of
 the copper and admiralty tubes in  the same period of exposure.

   "Significance

   "The external corrosion test units were unheated and corrosion effects
were greatly accelerated.  Therefore, the same effects would  not be
anticipated  in operating units.  However, some of the effects  could  be
 expected before startup or during periods of shutdown. The extent of
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          corrosion and fouling emphasized the importance of testing potential
          materials in the expected environment  before building the final operat-
          ing unit.

            "The corrosion of aluminum tubes in  the internal test was not antici-
          pated.  Further work  would be needed  before aluminum could be con-
          sidered satisfactory for service in the type condensate used in the test."

       Hudson Projects Corporation of Houston, Texas was asked to provide their
opinion of the corrosion problems which might be  expected with dry-type cooling
towers, based upon their extensive  experience in  the chemical, petroleum and
natural gas industries.

       Reproduced below  is the answer received  from Hudson.

                              HUDSON PRODUCTS

                                  "SERVICE LIFE
                           ALUMINUM FINNED TUBES

          "1 .   Process Industry Air Cooled Heat Exchanger
                Experience Record

                  "The air-cooled heat exchange industry has been in existence
                for nearly 40 years.  Industrial air coolers were first used in the
                gas pipeline industry about 35 years ago as shown below . With
                time,  its applications have  grown and continue to grow.

                                                   Approximate  Year
                          Application             First Placed in  Service

                      Gas Pipelines                        1935
                      Natural Gas Plants                   1940
                      Petroleum Refineries                  1945
                      Chemical Plants                      1950

                   "We have air coolers in service today that are over 25 years old.
                The aluminum  fins have some surface corrosion but  the air coolers
                continue to function and will for many  more years.  In  the last 20
                years, we, and our licensees, have manufactured over  600 million
                square feet of extended surface.  Of this amount, we estimate
                that less than  three per cent of  these bundles had to be replaced
                for reasons of  corrosion. The replacements were split about
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      equally between internal tube corrosion/erosion and external fin
      corrosion caused mainly by the presence of acid or halogen gases.

        "Our statistical records on air-side fin corrosion are very meager
      because this has never been a serious problem in the process indus-
      try.  Of those cases that we are aware, the problem occurred
      because of a known chemically corrosive gas atmosphere in the
      area of the air coolers.  The towers were comparatively small and
      installed close to the ground where the corrosive mist  was most
      highly concentrated.

        "We could provide you with specifications and a  list of our
      world-wide air cooler installations serving  the process industries
      but we question  its value.  It would be an impressive  list of  ' big-
      name1 companies using air cooled heat exchangers in  all types of
      services,  but we do not believe  it would answer your  needs.

"2.   Extended Surface Materials and  Corrosion
      Resistance Properties

        "In the first 10 to 15 years of  industrial air cooler manufacture,
      copper fins were commonly used as the extended cooling surface.
      Since then, aluminum has replaced copper  because of its large
      price advantage.  The aluminum specification generally used for
      this fin  stock application is shown below.

                                                   Plate or
      Type  Fins:                   Extruded       Tension Wound
      Aluminum Designation:      B-241-67         B-209-67
                               Alloy 6063-0    Alloy 1100-T-24

      Al -  %                   98.35-97.50    99.0 Minimum
      Si + Fe  -  %                 0.55-  0.95     1 .0 Maximum
      CU-%                       0.1          0.2
      Mn-%                       0.1          0.05
      Zh-%                        0.1          0.10
      Mg - %                    0.45-  0.09
      Cr-%                         0.10
      Ti - %                         0.10
      All Others-%                 0.15        0.15 (NMT - .05 ea)

        "One of the outstanding  features of aluminum is its  resistance to
      corrosion.  Aluminum has a great affinity for oxygen with which it
      combines almost instantaneously  to form a protective coating of
                         102

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     aluminum oxide.  The base aluminum is thus covered by a coating
     of aluminum oxide which prevents further oxidation and corrosion
     in the normal air atmosphere.

        "Reynolds Metal Company rates the relative cold-test corrosion
     resistance of aluminum (air-cooler fin-stock material) in various
     atmospheric surroundings as follows:

                                 B-241-67         B-209-67
     Aluminum Designation:     Alloy 6063-0    Alloy 1100-1-24

     Rural                      Very Good        Excellent
        (Inland areas away
        from smoke, fumes
        and industrial dust)

     Industrial                    Good           Very Good
        (Areas contaminated
        by smoke, chemical
        fumes and other in-
        dustrial dusts)

     Marine                      Good             Good
        (Areas ranging  up to
        one mile from the
        sea coast subject  to
        intermittent salt mists)

     Reference:  "Structural  Aluminum Design"; Pages 91 and 113.
                 Reynolds Metal Company - 1968

        "Cold corrosion tests by the aluminum manufacturers (Alcoa  and
     Reynolds) show Alloy 1100 being slightly better than Alloy 6063
     due  to purity .  Plate fin material for power plant service will be
     B-209-67/Alloy 1100-T-24.

"3.  Aluminum Fin Corrosion and Its Prevention

        "Aluminum fin corrosion occurs when operating with moisture in
     a corrosive atmosphere. It is further aggravated when this  mois-
     ture does not run off but lies on the horizontal fin surfaces allow-
     ing the corrosive liquid to work away and destroy the protective
     aluminum-oxide coating.
                          103

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  "There is lltHe that can be done to clear the air of corrosive
gases in some chemical plant installations.  However, something
can be done about the design and construction of the air cooler
and its operation to minimize corrosion.

  "Here are our recommendations:

  "A.   Install  induced-draft fans with fan-ring rain-gutters.
         Induced-draft fans installed on the top of the tube bundle
         protect the fin surface from direct rain exposure.  Fan-
         ring rain-gutters carry the centrifuged water away from
         the tube bundles.  A forced-draft fan installation, on
         the other hand, has the entire top face of the bundle ex-
         posed to rain water.

   "B.   Install  horizontal tube bundles with the extended fin-
         surface vertical, or near vertical.  Moisture (as a result
         of dew, mist, fog or rain) will collect and flow off the
         fin surface  rather than lie on it as would be the case with
         the fins positioned horizontally.

   "C.   Use variable-pitch, reversible-flow, fans for fluid tem-
         perature control.   Keep all the tubes in service and
         warm (10 F to 15 F above ambient air) at all times there-
         by preventing water condensation on the fin surface.
         Rather  than removing bundles from  service at low-loads
         and low ambient-air temperatures, the air flow can  be
         reduced with variable pitch fans and even reversed with
         reverse pitch.  Reverse pitch will provide co-current
         flow which ensures warm air at the inlet to  the fin tube
         bundles.

   "D.   Avoid natural-draft designs which have  uncontrolled air
         flow.  Depending upon tower design, ambient air tem-
         perature, wind velocity and fluid temperature, some fin
         surface temperatures can drop below the dew point tem-
         perature as a result of air channeling inside the tower
         structure.  This will cause condensation that could pro-
         mote fin corrosion.

  "We believe that the air-cooler corrosion experienced at
Rugeley Station  was  a combination of several factors:
                    104

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        "A.   The air-cooler sections were Installed vertically around
              the periphery of the tower with the plate fins in a hori-
              zontal position.  Moisture from the atmosphere (rain,
              dew,  fog,  etc.) and spray from the adjacent wet cooling
              towers which  lay on the fin surface provided the environ-
              ment  for the atmospheric corrosive pollutants.  An air-
              polluting,  ash-sintering,  plant, which fabricates building
              blocks from the plant-ash, is built adjacent to the
              Rugeley Station.

        "B.   The air-cooler plate fins are joined with compression
              collars. The  resulting tube-to-plate  fin joints are not
              water-tight and hence are subject to  corrosion.

        "C.   The natural-draft hyperbolic tower has no air-control
              means such as louvers.  Under certain operating condi-
              tions  some of  the tube temperatures could fall below the
              dew point  and cause condensation .

"4.   Protective Coatings

        "Protective coatings on fins such as epoxy,  phenolics, etc.,
      may have a place in the process industry but we firmly believe
      they  should not be used in power plant application.  We have
      used, on rare occasions, protective coatings on process-plant air-
      coolers installed in  known corrosive atmospheres.  But only the
      most  exceptional power plant locations would  ever be subject to
      such  surroundings.

        "An industrial-type power plant serving a chemical complex
      might but certainly  not a normal electric utility plant.

        "Our  principal objections to the  use of protective coatings on
      the extended surface are high cost and degradation of overall  heat
      transfer! All effective corrosion resistant coatings are dielectric
      in their  properties;  hence they are inferior heat-transfer materials.

        "Fin coating is a poor solution to the problem from an engineer-
      ing viewpoint  because it is preventing the effect  and not the cause.
      A lower-cost installation could be achieved by moving the pro-
      posed power plant site away from an existing chemical complex or
      a sea-coast area.  If the power plant stack gases  are suspect of
      being a  potential source of the problem,  then  stack orientation
      and stack height can be optimized.
                          105

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         "English Electric Company is contemplating the use of an epoxy
       resin at Rugeley Station but this is to solve a specific problem in
       a specific situation which they have to live with.  We all  have
       learned a great deal since the Rugeley tower was designed about
       ten years ago.

 "5.    Simulated Corrosion Tests
         "Simulated cold-tube corrosion tests provide a relative corro-
      sion resistance evaluation of alternate materials.  It has been our
      experience that  cold-tube corrosion tests are of no value in pre-
      dicting and evaluating fin service life. A cold metal will gener-
      ally condense moisture on its surface at least once in every 24-
      hour period during certain seasons whereas an operating air-cooler
      may not be exposed to such  a moist condition once in a year. It
      is  the moisture which is the  catalyst that is operating in conjunc-
      tion with the corrosive gases that breaks down the protective
      aluminum oxide  film.

"6.   Fin Surface Fouling

         "We have experienced fin-surface fouling from cottonwood and
      poplar lint  in a few specific installations.  When it occurs, it can
      be vacuum  cleaned from the fin surfaces.

         "If it is known in advance that the area foliage  in the vicinity
      of  the power plant does produce  this nuisance,  then screens of
      about 10 mesh size can be installed  (during the lint season) below
      the air cooler bundles.  The screens can be cleaned as required
      and removed during most of the year.

"7.   Power Plant Operation

        "Air-cooler fin corrosion in power plants should  be less than
      that experienced in process plants for the following reasons:

       "A.   With the exception of the boiler stack gases, there
              should be no corrosive gas producers in the power  plant
              area.  Process plants generally have many more poten-
              tial corrosive gas sources.

       "B.   Power plant stacks are many times higher  than the 40 to
              100 ft.  stack heights found in process plants.  Hence
              corrosive gases and  particulate matter are more effec-
              tively dispersed away from  the immediate  plant area.
                         106

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                  "C.   Large dry cooling towers serving power plants will have
                         air-coolers installed about 40 ft. (and higher) above
                         grade which is higher than that generally found on the
                         smaller process-plant units.  Heavier-than-air chemical
                         gases and pollutants will be of negligible concentration
                         at these heights.

                  "D.   Nuclear  power plants presently have a large exclusion
                         area and they produce no corrosive stack gases.  Their
                         atmospheric surroundings could be classified  as excellent.
                         In the future it is proposed to build nuclear plants close
                         to or in urban areas.  From a corrosive gas and particu-
                         late release standpoint, future urban areas could also be
                         classified as good or excellent as regards aluminum fin-
                         life expectancy."

Effect of Corrosion on  Performance  of Coils

       Corrosion of cooling coils and  fins as a result of atmospheric contamination,
salt spray, or other causes would result in loss of heat rejection performance of the
dry-type cooling tower if the corrosion were severe  enough to change the heat
transfer characteristics of the design.

       Probably the greatest loss of heat transfer would be suffered if corrosion
should destroy the bond between the fins and the tubes so that the path of heat con-
duction from the tube wall to the fins is broken.

       Loss of fin metal by corrosion would reduce  the area of heat transfer surface
in contact with the air.  The products of corrosion,  such as metal  oxides or sul-
phides, would have poorer heat conductivity than the pure metal and would impede
heat flow.  Surface corrosion may also affect the outside film factor and reduce the
heat transfer from the fluid  inside the tubes to the tube wall  because of the poorer
conductivity of the metal oxides inside the tube.

       Corrosion that  resulted in perforation of the tube walls would result in  leak-
age of water in the indirect system  and the admission of air in the direct system
since the  direct-type condensing coils would be at a pressure less than atmospheric.
In addition to possibly lowering the heat transfer capability,  the air in the coils
would result in poorer  turbine performance, since it would have  the effect of rais-
ing the back pressure against which the turbine operates.  This is caused by partial
pressure of the air in the steam-air mixture.
                                     107

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                                 SECTION V

                            AUXILIARY EQUIPMENT


Genera!

        The principal components of the auxiliary equipment associated with an in-
direct dry-type cooling tower are:

        1 .    Direct-contact condenser.

        2 .    Circulating water pumps.

        3.    Water turbines.

        4.    Air-removal equipment.

        The direct-type, air-cooled condensing system does not utilize circulating
water pumps or a direct-contact condenser since  the exhaust steam from the turbine
is conveyed to the cooling coils in the tower and is condensed directly by the air
flowing past the coils.

Condensers
        In the indirect-type system, condensation of exhaust steam from the turbine
is accomplished in the condenser by direct mixing of the circulating water and the
exhaust steam.  Several designs of direct-contact condensers have been developed
by European manufacturers, and at least one United States manufacturer is working
on a direct-contact condenser design.

       A well-designed condenser must condense the steam with a minimum of sub-
cooling of the condensate below the saturated temperature corresponding to the
turbine exhaust pressure, and must also deaerate the condensate and provide for
removal of the air and other noncondensable gases.  Adequate storage space for
the circulating water and condensed steam must be  provided in the condenser hot-
well .  In order to reduce circulating water pumping power requirements, the pres-
sure drop across the spray-water nozzles should be low—in the order of 1 .5 psi .

        Figure 33 shows a cross-sectional view of the  English Electric Company
condenser installed at Rugeley Station with  a 120-mw unit equipped with a Heller-
type dry tower (31) .
                                    108

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                                                                 20* dia bore atmospheric
                                                                 exhaust branch pipe
                                                                     12* bore air suction
                                                                        to *'r
                                                                     Main spray nozzles
                                                                  Auxiliary spray nozzles
                                           Exhaust
                                          chamber
Bled steam pipe to
No. 2 LP heater
12* dia bore air
suction pipe
to air ejectors
Baffle plate to prevent
impingement of sprays
on condenser shell
                               Normal water level
Expansion joint
                       52* dia bore pipe
                                        52* dia bore pipe
                                                                    Support springs
                                                              Condensate and cooling water
                                                              outlet to circulating water
                                                              extraction  pump
 FIGURE 33— CROSS SECTION OF DIRECT- CONTACT CONDENSER
                     USED AT  RU6ELEY  STATION  (31)
                                        109

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        The Rugeley condenser is a single-shell type, constructed of mild steel .
 Water boxes at each end supply circulating water to 24 spray-nozzle header pipes
 running across the shell .  The circulating water sprays are directed  towards the
 steam flow, and condensation is achieved as the steam flows downward past the
 nozzles.  The nozzles are constructed of stainless steel.

        A direct-contact condenser designed by M.A.N. is shown in cross section
 in Figure  34.

        The M.A.N. condenser is designed for minimum restriction  in the exhaust-
 steam-flow area in order to achieve a low steam-pressure drop from  the turbine
 exhaust flange to the condenser hotwell.  Circulating water from the cooling tower
 enters the condenser at 1 to an annular header and passes through several distribu-
 tion pipes, 2, to an inner header, 3.  The circulating water is sprayed out through
 a number of nozzle headers,  4,  impinging against baffles, 5, and drips downward
 over a series of plates, 6.  The exhaust steam flows downward through the con-
 denser to  the hotwell level and turns upward through the cascading  circulating
 water and condensed steam .  Air is drawn off across small surface condensers, 7,
 by water-powered air ejectors, 8.

 Air Removal Equipment

        For proper operation of a steam condenser, it is necessary to continuously
 vent off the noncondensable gases and air which are present in the condenser.
 Also, it is necessary to evacuate the air from the steam space of the condenser
 prior to putting the condenser into service. The foregoing operating requirements
 hold true for the direct-contact condenser used with dry-type cooling  towers as
 well as for conventional surface condensers.

        Multistage steam-jet ejectors and water-type ejectors have  both been used
 with dry tower installations.  Because the air removal capacity of a steam-jet
 ejector, measured in pounds of air per hour, is reduced when condenser pressure  is
 low, the water-type ejector which uses water as motive power rather than steam  is
 preferred by some European manufacturers of dry tower equipment for the reason
 that the air-removal capacity of the water-type jet,  measured in pounds of air per
 hour, remains fairly constant over a wide condenser pressure range.

        For evacuation of air during start-up, special  high-capacity jets which ex-
 haust directly to the atmosphere are used. The start-up jets are shut down when
sufficient air has been evacuated for the operating jets, which generally return th
steam or water to the cycle, to handle the task.
e
                                    110

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                                              4
                    OUTLET
FIGURE 34— M.A.N.  DIRECT-CONTACT CONDENSER (9)
                       111

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Pumps

        Since in the indirect-type dry cooling tower system the circulating water
from the tower and the exhaust steam from the turbine are mixed together,  the cir-
culating pumps  must be able to remove water that is at, or very close to, the boil-
ing point corresponding to the condenser pressure.  These pumps must have  the same
design features  as conventional condensate pumps; i .e., the suction passages must
be generously sized to achieve low velocities because of the high lift against which
the pump operates, and adequate provisions must be made to prevent the leakage of
air into the parts of the pump under suction pressure.

        Because of the large volume of circulating water that is mixed with the con-
densed steam, the circulating water pumps must handle  from approximately 40 to  70
times the amount of water that a  conventional condensate pump would handle for
the same size unit installed with  a surface condenser, depending,  of course, upon
the temperature rise of the circulating water through the condenser. As an example,
an 800-mw turbine-generator equipped with a conventional surface-type condenser
requires removal of approximately 7,400 gpm of condensate from the condenser hot-
well .  The same size unit with an indirect-type dry cooling tower and direct-contact
condenser and designed for a 30°F rise in circulating water temperature requires
that 300,000 gpm of water be removed from the hotwell.

       Although such large  pumps designed for removing water from a chamber
under high vacuum have not been constructed in the United States, the technology
and design experience is readily  available.  Circulating water pumps for large, in-
direct, dry-type systems will,  in effect, be conventional circulating water pumps,
either of split-case horizontal  configuration, or of the vertical type, modified for
operation with high suction lift.  The pump head would be approximately 80 to 100
feet.

       Since rhe direct-type, air-cooled condensing systems do not require circu-
lating water pumps, the condensate pumps used  with a direct system would be  sim-
ilar to the condensate pumps used with conventional surface condensers.

       In order to reduce the plant water make-up to a minimum, we would expect
that the circulating and condensate pumps used with dry-type towers would utilize
mechanical seals in place of shaft packing.

Recovery Turbines

       For large generating units equipped with dry-type cooling  systems,  it is
economical to recover the excess head imposed on the tower to maintain positive
pressure on  the cooling coils.  For this purpose, recovery turbines  would be in-
stalled in the circulating water piping between the tower and the direct-contact
                                    112

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condenser.  Approximately 20 to 40 percent of the totgl power required to pump
the circulating water through the system can be recovered.

       The  recovery turbines would be conventional,  low-head, hydraulic turbines,
probably of  the Francis type, and would be similar to the turbines of current design
and manufacture used with hydroelectric generators.

       For the 800-mw fossil-fueled unit considered in this study,  the total capa-
bility of  the recovery turbines would be approximately 3,000 horsepower,  operat-
ing at a head of approximately 30 to 40 feet.

       Power from the recovery turbines could be utilized by either of two methods:

       1 .    Direct connection to the shaft of the motor-driven circulating
             water pumps; or,

       2.    Connection  to a generator which would produce electrical
             power for certain auxiliaries.

Auxiliary Cooling

       The  coolers of the generators, turbines, and auxiliary equipment of a steam-
electric generating plant are generally designed for use with  cooling water having
q maximum incoming temperature of 95°F.  Since the temperature of the circulating
water from a dry-type cooling tower will exceed 95°F during much of the year,
depending upon  the ITD selected for the tower, it is necessary to provide auxiliary
cooling water from a source other than the main tower.  For an 800-mw unit,  the
auxiliary cooling is approximately 50 million Btu per hour—approximately 1 .3 per-
cent of the condenser heat rejection requirement.

       There are several methods by which cooling water of an appropriate tem-
perature  could be provided for cooling generators and auxiliaries.  It would be
necessary to make an economic evaluation of these different systems, described
below, for each particular plant before determining which would be the proper
selection.

       1 .    A  small wet-type tower could be used.  Because the  tempera-
             ture of cooling water circulated through a wet-type cooling
             tower approaches the wet-bulb  temperature of the ambient air
             and since the wet-bulb temperatures would not  exceed approx-
             imately 85  F at any location In the United States, it is possi-
             ble to obtain 9o F cooling water during hot weather.
                                    113

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       2.    An evaporative spray for cooling ambient air could be used
             in connection with a small mechanical-draft, dry-type
             tower sized to handle the auxiliary heat rejection load.
             The air to the cooling coils would be cooled by the evapor-
             ation of the spray water.  Since it is possible to cool  the
             air to within a  few degrees of the wet-bulb temperature  of
             the ambient air, an auxiliary cooling-water supply of 95°F
             could be obtained with an air cooling coil of low ITD design.

       3.    A third method of supplying 95°F cooling water for plant
             auxiliaries would be to use a small mechanical-draft, dry-
             type cooling tower which would operate on  the wetted-fin
             principle.  The surface of the cooling coils  would  be  wetted
             with water as the air passes over the coils.  With proper
             selection of the  ITD of the cooling coils, it would be possi-
             ble to obtain 95°F auxiliary cooling water.  Special provi-
             sions would have to be made to clean the coils of scale which
             might accumulate.

       4.    Mechanical refrigeration could be used to cool a portion of
             the main circulating water supply to 95°F for auxiliary  cool-
             ing purposes during periods when the main supply exceeds
             95°F.  Standard water-chilling equipment could be adopted
             for this use.

       5.    In 1958 at the World Power Conference, Professor  Heller
             presented a method of providing cooling water  for  auxiliaries
             during periods when the dry tower circulating water exceeded
             temperature limits suitable for plant auxiliary cooling by
             using steam-jet refrigeration  (34) . The use of steam-jet
             refrigeration has been well-established  in the refrigeration
             and air-conditioning industry and, although this method is
             not commonly employed, steam-jet refrigerating systems have
             been used for over 50 years—especially in certain  process
             industries.

       Figure 35 shows a schematic diagram of the cycle  presented by Professor
Heller as it would apply to a dry-type tower plant for providing  auxiliary cooling
water during periods of high ambient air temperature.  Condensate-purity circulat-
ing water is pumped through the auxiliary cooling system and sprayed into the
evaporator where a small portion of the water flashes  into steam  because of the low
absolute pressure maintained in the evaporator by the steam-jet ejector. Approxi-
mately 1 percent of the auxiliary cooling water is flashed to steam for every 10°F
                                    114

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                                        STEAM TURBINE
Ol
            EXTRACTED  STEAM
       DIRECT CONTACT
       CONDENSER
                            STEAM JET EJECTOR
CIRCULATING
WATER PUMP
             AUXILIARY
             COOLING LOAD
  DRY
COOLING
 TOWER
                                 AUXILIARY COOLING
                                 WATER PUMP
                  FIGURE 35—AUXILIARY COOLING BY STEAM-JET REFRIGERATION

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of temperature drop in the evaporator.  The pressure in the evaporator would be
held at 1 .66 inches Hg to maintain a temperature of 95°F.

        Motive steam  to operate the jet would be obtained from steam extracted
from the turbine.  The steam-jet ejector maintains the required low absolute pres-
sure in the evaporator and discharges the mixture of motive steam and flashed water
to the main condenser where it is condensed and pumped to the cooling tower along
with the main circulating water supply. The flashed auxiliary cooling water is re-
placed from the cooled circulating water supply.

        Since the system  is closed, no water is lost to the cycle with this type of
refrigeration system.
                                    116

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                                 SECTION VI
                       DRY-TYPE COOLING TOWER USE
                           WITH BINARY CYCLES
Genera]

       The use of two working fluids having different temperature-pressure charac-
teristics in a power plant cycle is identified by the term "binary cycle".   In  the
binary cycle, a fluid of relatively low vapor pressure is used in the higher temper-
ature (top) section of the cycle and a fluid of high vapor pressure is used  in the
lower temperature (bottom) section.  A number of binary cycles using various fluids,
usually with steam as the bottom  fluid, have been proposed and a small plant using
the mercury-steam binary  cycle was constructed in 1930 by the Hartford  Electric
Light Company.  Other fluids investigated as the top fluid  in  a binary  cycle are
diphenyl, diphenyloxide, aluminum bromide and zinc ammonium chloride.

       Slusarek (35)  has presented a study of the economic feasibility of a binary
cycle using steam as the top fluid and ammonia as the bottom fluid, which has par-
ticular appeal for use  with a dry-type cooling tower. Other studies of binary cycles
with dry-type cooling towers using commercial refrigerants as the low temperature
fluid, are currently underway by European manufacturers.

Description of Steam-Ammonia Binary Cycle

       Figure 36, from (35), shows the temperature-entropy (T-S) relationship and
the basic flow diagram of the steam-ammonia binary cycle. In the upper part of the
T-S  diagram, steam is the working medium and is expanded through the steam tur-
bine from temperature T-i to T2 ,  flowing into the steam-to-ammonia heat exchanger
where the steam  is condensed" as it  boils the ammonia, which is the  working fluid
in the lower part of the cycle.   The  temperature difference, At,  is necessary to
transfer heat from the  condensing steam to the boiling ammonia in the heat ex-
changer.  The ammonia vapor at temperature To expands through the ammonia tur-
bine to temperature T* where it is condensed and flows to the heatexchanger  for
recycling.

Conclusions

       There are a number of theoretical advantages in the use of the steam-
ammonia cycle.  The ammonia turbine  is much smaller than a low-pressure steam
turbine which would be required for a condensing steam cycle since the specific
volume of ammonia vapor  is lower than steam at corresponding temperature.   By
terminating the steam cycle at a pressure above atmospheric (34.8 psia in the study
                                    117

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                       ffftfSW    N
                       &»T4   N  /
                       iv^-ii!	T^x
                    ENTROPY

  BINARY  CYCLE TEMPERATURE-ENTROPY  DIAGRAM
BOILER


T2
TURBINE
T2

r
r1
X
H20
1
                             ,H._J
                                   r-
                                   IAMMONIA i  .
                                   n-URBINEl"""^
                                   r-*~ _   '
                     HEAT EXCHANGER
                                            DRY CCK)LING
                                            TOWER
                                            CONDENSER
FIGURE 36  — FLOW DIAGRAM OF BINARY  CYCLE
          WITH  DRY  COOLING TOWER
                     118

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by Slusarek), vacuum pressures lower than atmospheric are not used in the cycle,
because the saturated pressure of ammonia at the lowest probable ambient tempera-
ture is above atmospheric.  The  low specific volume of  the  ammonia  vapor  ex-
hausted  from the ammonia turbine would permit direct condensation in a dry-type
cooling  tower with smaller equipment.  Thus there is opportunity for reduced cost
of a smaller ammonia turbine plant to offset the added cost of a dry-type cooling
process.  Also, the temperature at which ammonia will freeze (— 103°F) is so low
that there is no danger of freezing in an air-cooled  condensing plant.

       The efficiency of a dry ammonia turbine stage is given as 85 percent (35) .
Mechanical losses, stage losses,  moisture losses and exit losses must also be sub-
tracted from the stage efficiency.  The resulting ammonia stage operating efficiency
is in the order of 74  percent for design conditions and reduces further for off-design
performance conditions.  A further reduction occurs in the temperature  gap in  the
isothermal  heat  exchanger where the condensing steam gives up heat to evaporate
ammonia, the vapor  of which is used to drive the ammonia turbine.  The extent of
this temperature gap determines this loss of  efficiency which amounts to approxi-
mately 1 percent per 7.7°F of temperature gap  (35) .  The temperature  gap can be
reduced  by a more expensive heat exchanger, but an optimum must be selected
which considers the  higher cost of heat exchanger versus higher fuel consumption.

        A total steam-ammonia binary cycle generating plant promises an over-all
plant efficiency of  approximately 42 percent with  further improvements to result
from feedwater regenerative heating which  was not included in the analysis  by
Slusarek (35) .  In addition to making available a somewhat higher plant efficiency
than is normally obtained with a  standard steam plant using once-through or
evaporative-type cooling, the binary cycle uses a dry cooling tower which allows
mine-mouth plant location in arid areas.
                                     119

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                                  SECTION VII
                     METEOROLOGICAL CONSIDERATIONS
Possible Effects of Dry-Type Cooling Towers
on Local Meteorological  Conditions

       Air temperature.  Air temperatures in the exit plume from a dry-type cool-
ing tower of the size studied in this report will be increased by  18 to 36°F.  The
resulting plume of warm air will tend to stay aloft and will  produce a local change
in the vertical temperature structure which will be favorable to the dispersal of air
pollutants.

       This local warming will be confined to an area in the immediate vicinity of
the cooling tower as there will be a ten-to-one dilution of the warmair very shortly
after exiting from  the cooling tower.  The speed  of  dilution will increase with in-
creasing wind speed.

       The  resulting short- and long-term effects of releasing large amounts of heat
into the atmosphere  is a subject which should be studied extensively in the  near
future since the increase  in  temperature will  be  the  major change in the  local
micro-meteorology caused by the dry-type cooling tower.

       Cloudiness.  Studies of  the meteorological effects  of wet-type  cooling
towers by Dr.  Eric Aynsley (36) have  shown that the initiation of cumulus clouds
is a rare occurrence, and on such occasions clouds triggered by towers only precede
natural  cloud  formations.  Cumulus cloud initiation  by  dry-type  cooling towers
would be even less likely because of the lack of water vapor.  However,  the possi-
bility of cumulus cloud formation cannot be completely ruled out.

       Aynsley also  found  that  under stable conditions and high humidities, wet
plumes will persist after leveling off and appear downwind as stratus cloud coverage
or merge and reinforce existing cloud coverage.  Conversely, the dry, warm plume
exiting from a dry-type cooling tower will tend to disperse rather than augment low
stratus clouds.

       The effects of dry-type cooling towers on local cloudiness then would be
limited to an extremely rare initiation of cumulus clouds and a slight decrease  in
the local stratus cloud coverage.

       Fog.  The exit plume from dry-type cooling towers will tend to disperse
local fog.  Appleman and Coons  (37)  found  that the  use of the heat and mixing
properties of jet engine exhaust was quite successful in evaporating fog from an
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aircraft runway.  Dry-type cooling towers would be even more effective in fog dis-
sipation since the heat discharge from the dry-type cooling tower will be greater
and jet engine exhaust contains a significant amount of water vapor.

        Precipitation.  It is doubtful that the precipitation pattern of the surround-
ing area will be altered by the operation of a dry-type  cooling tower.  There may
be a slight decrease in precipitation in the immediate vicinity of the cooling tower,
but it would probably not be  measurable.

        Air currents.  Fritchen et al (38)  found in their study of the meteorological
effects of a forest fire with a heat release of the order of magnitude of that of a
large dry-type cooling tower that the air currents in the immediate vicinity of the
fire were  significantly altered.

        In the case  of the dry-type cooling tower, a convergence zone would be
formed over the tower which  would redirect and alter the speeds of local winds.
This effect would vary with local wind conditions and be most pronounced with low
wind speeds.

        The strength of the updraft will also depend on  the local micro-meteorology
and  will be specifically related to stability, wind speed, and ambient air tempera-
ture.  Because of the turbulence encountered in strong  updrafts, the area should be
avoided by aircraft.

Pry-Type Cooling Towers and Air Pollution

        If there is an effluent discharge from the power plant associated with a dry-
type cooling tower, the  best place to vent the effluent would be in the updraft from
the cooling tower.   This would carry the pollutants up into zones of higher winds
where the particulates and gases would be greatly diluted and dispersed through a
larger volume of air.  This redistribution of the pollutant load would be beneficial
locally,  but will still add the same amount of contamination to the total pollution
problem.

        Under certain meteorological conditions, the updraft may break through an
inversion and disperse the pollutants above a layer they may have otherwise been
trapped beneath.  This also would be a beneficial local effect.

        How large an area surrounding the plant will be vented by the updraft de-
pends upon the  local micro-meteorological parameters  and the location  and emis-
sion factors associated with other sources.
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Comparative Effects of Various Cooling Methods

       Once-through cooling.  The most undesirable side effect of once-through
cooling in some installations is the heat addition to natural waters and possible
damage to the marine environment of the recipient body of water.  The extent of
such damage will depend upon the size and mixing properties of the lake,  river  or
ocean into which the warm water is being released.

       In addition to the possible environmental damage to the marine life, this
method of cooling also produces  large amounts of water vapor.  Depending on local
climatic conditions,  this can become a source of fog and mist downwind and cause
serious icing problems on adjacent towers and transmission lines.

       Cooling ponds.  Cooling ponds are another source of large amounts of water
vapor and will produce the same undesirable side effects  associated with water
vapor described for once-through cooling.

       Wet  (evaporative)  type cooling towers.  In addition  to the  problems asso-
ciated with water vapor production,  the wet-type cooling towers add to the air
pollution problem through  drift losses.  Waselkow (39)  also experienced maintrans-
mission line  flash-overs due to cooling tower drift losses.

       If production of SOo is associated with a wet-type cooling  tower,  the mix-
ing of the two effluents wilt cause a major pollution problem.  The  rateof  oxidation
of SOo to sulphuric acid is enhanced by increased relative humidity.  According to
Aynsley (36), the rate of oxidation increases rapidly when the relative humidity
reaches 80 percent.  Thus, a release of SO2 into the effluent from  a wet-type cool-
ing tower can produce deleterious results.

       Natural-draft versus mechanical-draft towers.  Pollution concentrations and
temperature  increases will be lower with natural-draft than with mechanical-draft
towers.  The reason for this is that the updraft from hyperbolic natural-draft towers
persists longer and go higher than plumes from mechanical-draft towers. Thus,
pollutants and temperature changes will be diluted in larger  volumes of air.

       Underlying causes  of the higher plume, according to Aynsley (36)  are the
following:

       1 .    The release area of natural-draft towers is higher than that
             of mechanical-draft towers.

       2.    The natural-draft tower is constructed in a  manner which
             complements the natural flow.
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       3.    In mechanical-draft towers, fans produce small scale tur-
             bulence which tends to break up the plume.

       The original speed of ventilation with natural-draft towers is not as great as
with mechanical-draft towers, but the speed Is more quickly dissipated from the
mechanical eddies formed by the  fans.

Conclusions

       Before the meteorological effects of a dry-type cooling tower can be pre-
dicted for any given set of conditions,  a thorough model  of the meteorological  im-
plications of such system must  be  developed through analysis of actual observations.
It is hoped that the first available tower will be  used for a complete pilot study.
Measurements of temperature (using both horizontal and vertical grids), wind speed
and direction, and humidity should be taken in and around dry-type cooling tower
sites before and  after plant start-up.  Efforts should also  be made to collect long-
term meteorological data from the area  to determine if any changes in weather pat-
terns can be  identified which are related to the operation of the dry-type cooling
tower.

       It is our general conclusion that the release of heat into the atmosphere from
a dry-type cooling tower will  be  much  less harmful to the environment than the
combined release of heat and water vapor associated with other cooling methods.
In addition,  harmful effluents  would be effectively dispersed by inclusion into  the
updraft from  a dry-type cooling tower,  whereas the combination of certain pollut-
ants with wet plumes would compound rather than  alleviate the pollution problem.

       While it is our opinion that a dry-type cooling tower will not produce a
measurable effect on a region's climatology, the worldwide buildup of thermal re-
leases and the resultant climatological effects remain important considerations.
Thus, on either a local or global  scale  the subject of the meteorological effects of
releasing  large amounts of heat into the atmosphere raises many unanswered ques-
tions and should be investigated extensively in the next  few decades.
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                                  SECTION VIII
                      DISCUSSION WITH MANUFACTURERS
 Introduction

        During the course of preparation of this report, the authors met with manu-
 facturers and engineers who are currently engaged in dry-type cooling tower work,
 either in production of components or in development engineering, and also with
 manufacturers who, although not  engaged in actual production of dry  tower compo-
 nents, are conducting  research and studies on the matter.

        In order to obtain first-hand information  as  to the basic principles, con-
 struction and operation costs,  operational  history and  current state of  the art of dry
 towers with steam-electric power plants, conferences were held with  various indi-
 viduals  and representatives of manufacturers as hereinafter described.

 Dr. Laszlo Heller and  Hoterv

        Dr. Laszlo Heller, of Budapest, Hungary, serves as the Head of the Depart-
 ment of Energetics of the  University of Budapest and as Technical Director of
 Hoterv—a 1,200-man engineering firm charged with the development  of the dry-type
 cooling tower and the design of industrial  plants in  Hungary.  Dr.  Heller presented
 the initial concept of the indirect dry-type cooling tower  system at the World Power
 Conference in 1956, and, subsequently, has been responsible for the  design either
 in toto  or as  a special consultant for  the Heller-type cooling towers designed to
 date. Dr. L.  Forgo, who serves as assistant to Dr.  Heller, developed the cooling
 coil used with the Heller system.  The marketing of the dry tower system components
 manufactured is under the direction of  the Hungarian firm Transelektro, which is
 also responsible for the manufacture and sale of all  electrical equipment.

        Hoterv  has designed  a series of cooling towers in sizes up to  900 mw and
 conducts computerized studies for manufacturers and utilities throughout the world.

        Dr. Heller furnished  basic performance data of nature I-draft, dry-type
 Heller towers to the authors.  These data were very helpful in the development of
 the computer program used in the  determination of the optimum ITD for various geo-
graphical locations in the United  States.

       An interesting development by Hoterv is a natural-draft dry tower with a
steel structural frame and  an aluminum skin.  The  tower is cylindrical in shape,
rather than hyperbolic, and is estimated to be somewhat less expensive to construct
than  the reinforced concrete hyperbolic towers.  A special erection technique has
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been developed  by Hoterv utilizing steel  rings inside  the  tower on which a con-
struction crane operates to hoist pieces of the structure.  The rings are raised by
hydraulic jacks and certain rings are left in place permanently for stiffening the
structure.  After erection of the tower has been completed, the crane and erection
ring are lowered by helicopter.

       During the author's trip to Hungary,  a  visit was made to  the factory where
the Forgo coils are manufactured.   The factory is  located  near  the Town  of
Jaszbereny and is a part of the Hungarian government1 s program to bring industry to
rural areas.  Besides the Forgo coils,  home refrigerators and truck radiators are pro-
duced at the factory.

       Aluminum tubes and aluminum strips for the fins are  received from another
factory and constitute the raw material for manufacturing the  Forgo  coil.   Much of
the labor of putting the coils together is done by hand,  in keeping with the program
of providing jobs for unskilled persons, and,  for that reason, automation is  less than
would normally be expected.

       The coil components—consisting of the  tubes; the fin sections, approximately
2 feet long,  which have been cut from rolls of  aluminum strip and punched  to re-
ceive the tubes and spacer rings; and aluminum spacer rings, which  fit  between the
tubes and fins—  are assembled by hand on a rack and pressed together by a  hydrau-
lic ram.   After the components are  pressed together, expanding mandrels are pulled
through the tubes to make a mechanical bond between the coils,  fins,  and collars.
The coils are dipped into an alkaline solution to form an oxide coating on  all sur-
faces for corrosion protection .  The water boxes of the coils are made of aluminum
and are of welded construction.

        Either two or three of the coil sections  are  joined together to make a
"column" and two columns are joined into a  "delta", fitted into a supporting steel
frame and tested hydrostatically for leaks.  The 3-coil delta,  approximately 45 feet
long, is  shipped as an integral unit and handled and erected at the plant site by
means of a specially designed carrier.

        Dr. Heller and  his associates  have developed many  techniques and devices
for control and operation  of dry-type towers as a result of over 30 years'experience,
and are the holders of over 20 patents applying to dry tower systems.

        Dr. Heller has also performed studies of locating a generating plant inside
the shell of a nature I-draft, dry-type cooling tower in order to take advantage of
the uplift from the tower discharge of warm air to disperse stack gases and to over-
come inversions, thereby  reducing air pollution.
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M.A.N. (Maschlnenfabrik Augsburg-Nurnberg)

       M.A.N. is one of the largest manufacturers of power, heavy mechanical
and transportation  equipment in West Germany with 7 plants and  approximately
50,000 employees. The M.A.N. factory at Nurnberg produces steam and gas tur-
bines, boilers, condensers, heat exchangers and associated steam plant equipment,
and has sold a 200-mw, non-reheat steam turbine and dry tower plant (1,139 x 10°
Btu/hr. exhaust heat) to ESCOM, a large electric utility in South Africa, for their
Grootvlei Power Station, which is scheduled for start-up in early 1971 .

       M.A.N. does not manufacture  the cooling  coils, but takes responsibility for
engineering and procurement of the complete system and the economic selection of
the turbine and cooling tower combination.  M.A.N. feels that the turbine and dry
cooling tower system should, at this stage of the development,  be considered as an
integral  unit  rather than selected  as two separate components.

       M.A.N. will offer either a direct- or indirect-type air cooling system, de-
pending upon the economics of the particular situation studied.

       The design  of the natural-draft  tower as delivered by M.A.N. to ESCOM
has the cooling coil tubes in a horizontal position  inside the tower and M.A.N.
indicates that they expect less wind influence than if the coils were in a vertical
position.  M.A.N. also states that locating the coils inside the tower shell  in-
creases the heat load capability of a tower since the inside area is proportional to
the square of the tower diameter, whereas the area available for a circumferential
heat exchanger coil installation is only directly proportional to the diameter of the
tower.

       According  to M.A.N., optimization of the dry tower system  usually dic-
tates a turbine back pressure above 3.5 inches  Hg. They,  therefore,  eliminate the
last row of blades of the standard  turbine design and place the shaft bearings out-
side the low-pressure casings.  The turbine  capability is maintained  over a wide
temperature range  by providing a  second admission  point after the initial stages of
the turbine and increased boiler capacity for use during periods of high back pres-
sure .

       M.A.N. advised that they are  prepared to offer steam turbines and dry-type
cooling  towers up to 1,000 mw in size.

GEA - Gesellschaft Fur Luftkondensation

       GEA  Airexchangers, Inc. of Bochum, West Germany produces finned air-
cooling coils for industry and power, and manufactures a direct, air-cooled, con-
densing system for  power stations.  GEA is also a licensee for construction of the
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Heller system and supplied the dry tower for the 150-mw unit at Ibbenburen,
Germany.  GEA has furnished a direct, air-cooled, condensing system for a 160-
mw generating unit  at Utrillas, Spain (Figure 37), and also furnished the direct,
air-cooled, condensing system for the 20.18-mw unit of Black Hills Power and Light
Company at Wyodak,  Wyoming.

       The following  estimating  figures for approximate cost of the dry-type cool-
ing tower systems in the United States were given to the author by Mr. Hans H.
Von Cleve, Chief Engineer of GEA, for the Heller system (indirect) using a natural-
draft concrete tower up to 450 feet high (over 450 feet, a steel natural-draft tower
would be used).   With a distance  of 300 feet between the tower and the turbine,
the cost is estimated to be:

             $520,000 x   "eqHoad,106Btu/hr.
                            (ITD, OF)1'25

       The above cost is for an erected system covering all condensing system com-
ponents from the turbine flange outward,  tower,  pumps, piping, foundations, etc.
According to Mr. Von Cleve, GEA is prepared to offer a cooling tower system  up
to 1,000 mw in size,  and has actually quoted 450- and 900-mw sizes to United
States utilities on the  above basis.

       The independent cost estimates made for the 800-mw, fossil-fueled plant
used in the computer program of this report corresponded closely with the foregoing
GEA cost estimating formula.

       Mr. Von Cleve1 s estimate for a direct, air-cooled, condensing system with
mechanical draft for sizes up to 200 or 300 mw is:

             *OIA  nnn   heat load, 10   Btu/hr.     D  .  c ..
             $210,000x	'-	  — Basic Estimate
                               ITD, °F

to which must be added:

             0.12 x the basic estimate for the steel structure
             0.08 x the basic estimate for the exhaust trunk
             0.12 x the basic estimate for erection

       The required land area for the direct system is:

             200 x  heqt load JO6 Btu/hr.      f).
                           ITD, °F
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  FIGURE 37—DIRECT CONDENSING SYSTEM
UTRILLAS POWER STATION, SPAIN (GEA PHOTO)

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       Installed fan power is:

             250 x  heat load,  IP6 Btu/hr.   kw
                          ITD, °F

       The power consumption of the fans is approximately 2 percent of the gross
output. Mr. Von Cleve estimates that the cost difference between a plant equipped
with a conventional tower and a  GEA direct, air-cooled, condensing  system is
approximately $5 to $6 per kw for sizes up to 300 mw.

       For sizes larger than  approximately 200 to 300 mw maximum, GEA would
offer the indirect system only.

English Electric Company

       English Electric Company,  now English  Electric - AEI Turbine Generators
Limited as a result of a recent merger,  has undertaken development  work on the
indirect dry tower system for approximately 10 years. English Electric has a license
from Transelektro for marketing  and constructing the  Heller tower  in  the  United
Kingdom and furnished the dry tower for the 120-mw unit at Rugeley  Station, which
went into service in 1961 . At that time, it was believed that most of the large
generating stations in England would be constructed at mine sites  and at inland
locations close to fuel  supplies, and it was anticipated that cooling water make-up
for evaporative towers would  soon be a problem.  However,  it now appears more
likely that future large generating plants in England will be nuclear or oil-fired and
located on the sea coast,  with the result that cooling water for once-through con-
densing systems will not be a problem and the  need for dry cooling towers in  England
may not  be as imminent as had once been thought.

       Conferences were  held with Mr. W. H. P.  Wolff,  Technical  Director of the
Willans Works at Rugby, now Director of British  Nuclear Design and  Construction
Ltd., and Messrs.  D. W.  Crane, P. J.  Christopher and J. L. Daltry, who have
been engaged in the dry cooling program.

        Studies made by English Electric indicate that capital costs are increased
approximately $12 to $14  per kw  for a dry tower plant, and that the average bus-
bar costs, taking into account fixed costs and fuel costs,  are increased approxi-
mately 6 percent as compared to an evaporative  tower system.  English Electric
considers that the components of a dry tower system cost about 1-1/2 times the cost
of an equivalent wet tower system.

        English Electric have concluded that a fully cost-optimized dry cooling
tower  scheme would  require a somewhat higher back pressure on the turbine  than
with a water-cooled condenser.   In the majority of cases that they  have studied,
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 the optimized back pressure  increases relative  to  wet tower cooling by an amount
 broadly in the region of 1 .0 inches Hg.   For such a case,  they would consider that
 modification of the last-stage blades of a turbine designed for other cooling systems
 would be practicable.

        In the case of Rugeley,  the  turbine  exhaust pressure for the dry  cooling
 tower was specified to be the same as for the other turbines in the station associated
 with  evaporative towers.  In this respect,  Rugeley is not a cost-optimized scheme
 and the question of turbine design for higher back pressure did  not arise.

 Brown Boverf Corporation

        A visit was made to the  Brown Boveri plant at Baden, Switzerland to discuss
 the operation of large turbine generators at back pressures higher than 3.5inches Hg.
 Brown Boveri is  one of the major manufacturers of power-generating equipment in
 the industry and is presently  constructing turbine-generator units up to 1,300 mw.
 Discussions were held with Mr.  W. Hossli,  Head of the Turbine Department Design
 & Calculation; Mr. H. Muhlhauser,  Head of the Turbine Performance Section; and
 other Brown Boveri engineers.

        Brown Boveri  is  interested in the use of dry cooling towers for generating
 plants and has participated in studies of dry towers for utilities.  Brown Boveri does
 not produce dry tower equipment, but has used Dr. Heller as a consultant.

        Brown Boveri  believes that if there is a demand for a  large number  of
 turbines to operate with  dry cooling  towers at high back pressures, a new design
 will be developed. They estimate that a turbine designed for 6 inches  Hg  back
 pressure would cost approximately 15 percent less than a turbine  designed for 2
 inches Hg back  pressure.  However,  a high-back-pressure turbine would probably
 not be designed  until  there were enough  units foreseen to absorb  the development
 costs. Until  that time, a standard modified design having a shorter  last-stage blade
 (eliminating the last stage) or reducing the number of low-pressure turbines could
 be used.

 United States Turbine Manufacturers

       Discussions and correspondence were held with the two  major manufacturers
of large turbine  generators in the United States—the General Electric Company and
the Westinghouse Electric Corporation—to obtain their respective opinions as to the
feasibility of operating large turbines at high back  pressures.  The results of these
discussions are covered in Section III  of this report.

       General  Electric believes that until  there is a sufficient demand for  a
specially designed high-back-pressure turbine, modifications would  be  made to
existing units for operation at back pressures up to 15 inches Hg.
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       Westinghouse advises that it is possible to design and manufacture a turbine
for high back pressure application.  Modifications of an existing or standard design
unit may or may not be economically practical, depending upon the particular unit
involved.

Hudson Products Corporation

       Hudson  Products Corporation of Houston, Texas has manufactured air-cooled
heat exchanger equipment for chemical and refinery industries for over 25 years.
Today, they and their foreign licensees are the world1 s largest manufacturers of in-
dustrial  air-cooled  heat exchangers. At one time, they also manufactured con-
ventional mechanical-draft, wet-type cooling towers for the  process and power
industries.  As  a  result of extensive research and  development in air-cooled heat
transfer surfaces,  hyperbolic tower shells and large fans designed  specifically for
power plant application, Hudson Products  Corporation  is now  offering a dry-type
cooling tower to the utility industry.  Hudson Engineering Corporation, a subsidiary
of J. Ray McDermott, as is also  Hudson  Products Corporation,  is prepared to offer a
dry  cooling tower system package completely engineered and installed.  The system
starts at the turbine exhaust flange and includes the direct-contact condenser, cir-
culating  water pumps, dry-type cooling tower, piping, valves  and  controls.

       Conferences were held with Mr. Ennis C. Smith, Vice President and General
Manager, and Mr. Michael W. Larinoff, Vice President, to obtain cost estimating
information and tower performance data for use in this report.  An excel lent summary
of Hudson's performance and economic studies  is contained in "Power  Plant Siting,
Performance and Economics with Dry Cooling Tower Systems" by Smith and Larinoff,
presented at the 1970 American Power Conference  (13) .

The Marley Company

       The Marley Company of Kansas City, Missouri is one of the largest manufac-
turers of evaporative-type cooling towers in the world, with operations  in many
foreign countries as well as in  the United States.

       Conferences were held with Marley  engineers, including  Mr.  Joe  Ben
Dickey/  Jr., Vice President of Engineering; Mr. J. O. Kadel, Vice President  of
Major Projects; Mr. Robert E. Gates, Senior Evaluations Engineer;  Mr. John A.
Nelson,  Senior Metallurgist;  Mr.  Edward  P. Hansen, Vice  President of Marfab
Radiator Division; and Mr. Joel Blake, Consulting Engineer to  DriTowerCommittee.
During these conferences, much information was obtained for use in this report.
The following excerpt,  taken from "Managing  Waste Heat with the Water Cooling
Tower",  by  Joe Ben Dickey, Jr. and Robert E. Cates of the Marley Company, sums
up Marley's work on the dry-type cooling tower:
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                                   "DRITOWERS*

            "For the past four years The Marley Company has contributed its Divi-
          sion Managers of Major Projects;  the L. T. Mart Company of  England;
          Marfab,  its Radiator Company; Engineering Division;  its Senior  Evalua-
          tions Engineer; its  Senior Metallurgist; and a distinguished alumnus of an
          experienced operating company as a  team  constituting  its  DriTower
          Committee.  The first year, five  exterior corrosion  induced draft coil
          sample demonstrations were installed at operating plants of the American
          Electric  Power and General Public Utilities  Systems.  The second year
          internal corrosion test tables were  installed with  all  manner of suitable
          tube alloys circulating actual plant deionized water.   In  the  second
          winter extensive outside  freeze tests  on  a  full  scale model were con-
          ducted at the Marley Research Laboratories.  During the third year the
          disappointing results of the first exterior corrosion studies resulted in  the
          commissioning of a second generation induced  draft wrapped tube study.
          Throughout this three year period  continuous re-evaluation of all Amer-
          ican and foreign wrapped  fins and core sections were conducted in the
          dry lab heat transfer wind tunnel in Kansas City.  Foul factors, tube
          spacing, and boundary-layer turbulence were analyzed.  While outside
          of the scope of this paper, the authors  may briefly comment that for the
          freezing  North American latitudes the  natural  draft hyperbolic Dry Tower
          of the style built abroad would have freezing problems, start-up and
          shut-down problems, and corrosion problems that would magnify both  the
          operating cost and technique beyond the conception of present market
          acceptability in America.  In  the southern climates of the United States,
          to permit the difference between dry bulb  and  steam operating  tempera-
          ture, the economic usage of a DriTower on large power plants would  re-
          quire turbines and open condensers not foreseeably available in this
          decade.  The vexing problem of managing the  large steam quantities  in
          the generating plant sizes planned in future years, caused this  Committee
          to abandon the possibility of direct steam condensers of any consequence
          in North American latitudes.  Induced  Draft DriTowers scaled-up in size
          from those designs commonly proven most operable and most economic in
          the hydrocarbon and petrochemical industries of America were  found  by
          the DriTower Committee to offer the best interim solution in  this country.
          Even then, the designer must be prepared for considerable study and
          attention to controls,  signal monitors, dampers, and dumping mechanisms
          which would result in equipment having much higher risk of problems
          during a rapid scram,  and much more manpower devoted to fine tuning
          than an American market will readily digest.   As a final project in cal-
          endar 1969, the Committee was a partner of a  prominent Eastern


* DriTower is Registered Trademark.
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       Architect-Engineer in the development of a full-fledged proforma plant
       design using both Natural Draft and Induced Draft DriTower on an 800
       megawatt machine.   The Committee would observe that these estimates
       are extremely complicated because of many parenthetical  matters that
       are completely foreign to a normal estimate.  Because of the vast ori-
       ginal work  required for a rigorous study,  simple rule of thumb estimates
       on Induced Draft units,  only, are recommended.   The committee's ob-
       servations,  together with information now available on pioneer installa-
       tions in the Eastern hemisphere teach us that the materials section, the
       maintenance labor,  and the extensive controls required by a Northern
       DriTower plant,  form a  very sober undertaking.   For very clean non-
       freezing sites, given very low cost fuel,  economic transmission,  and tax
       plus ecological incentive, DriTowers will deserve study as rotating and
       condensing hardware becomes suitable."

Ingersoll-Rand Company

       Ingersoll-Rand Company of Phillipsburg, New Jersey is making studies of
the cost of direct-contact condensers for use with dry-type cooling towers.  Cost
estimating data of  direct-contact condensers furnished by Ingersoll-Rand were used
in the optimization studies  in this report.

GKN Birwelco Limited

       GKN Birwelco Limited is a subsidiary of Guest,  Keen and Nettlefolds, a
large English-based international engineering group.  They are specialists in the
design and execution of substantial contracts involving heat transfer equipment
and were responsible for the complete process, mechanical and civil design, pur-
chasing, inspection and construction of the 200-megawatt dry cooling tower which
is now being commissioned for Escom, a large electricity utility in South Africa.
They offer complete construction of both direct and indirect condensing systems
using natural or mechanical draft systems.  GKN Birwelco, through  its New York
subsidiary, GKN International, Inc., offers these installations using complete
supply of materials and  services from United States manufacturers and is currently
performing studies using United States subcontractors for dry cooling towers  up to
1,000 megawatts in size.

        GKN  Birwelco  uses cooling sections in a horizontal position inside the base
of the natural draft tower shell rather than the vertical arrangement previously de-
scribed.  They have performed wind tunnel tests at the National  Physical Laboratory
which they state have shown enhanced performance of the horizontal sections during
windy conditions.
                                    133

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                                   SECTION IX
                          OPTIMIZATION PROGRAMS
 Introduction

        The state of the art of dry cooling for steam-electric generation is such that
 relatively little information is presently available to assist electric utilities ineval-
 uating the economics of dry cooling.  In contrast, the various methods of wet cool-
 ing have been in use by the electric utility industry long enough so that  the sizing,
 performance and cost of the economically optimum, or near optimum, cooling sys-
 tem can be  established with relative ease.

        A method of analysis and a computer program  were developed to  determine
 the cost and performance of a range of sizes of dry-type cooling systems  for specific
 sets of conditions and to select the economically optimum size for  those conditions.
 The measure of dry-type cooling system size used in the analysis is the initial  tem-
 perature difference (ITD) which is defined as the temperature difference  between
 the turbine  exhaust steam  and the  ambient air.   The concept of initial temperature
 difference is discussed in Section II of this report. The specific conditions which
 affect the selection of the  economically optimum dry  cooling system include such
 factors as the relationships of performance and capital cost to ITD, the fixed-charge
 rate,  fuel cost,  air temperature, the amount of generating capability lost at high
 ambient air temperature, and the cost  of replacing the lost capability. These  var-
 ious factors and  their effects on the optimization of the dry cooling system are dis-
 cussed in detail  later in this section.

        Two computer programs were developed to facilitate the analysis. The first
 program determines the optimum tower  size for a given ITD.  The second  chooses
 the optimum ITD and, consequently, tower size with consideration given to all costs
 of construction and operation .

        The  method of analysis and the computer programs yield information  as to
 the size, cost and performance of the economically optimum dry cooling  system for
 a specific set of conditions, but does not provide the information necessary to com-
 pare the relative economics of dry cooling versus other cooling methods.  Although
 it was not within the scope of this study to compare the economics  of dry cooling
with other methods, some preliminary economic comparisons of dry cooling  systems
 versus wet tower cooling systems were made in order to indicate the factors which
 must be considered in such  a  comparison and to establish the order  of magnitude of
 the cost differences which may  result if a dry-type cooling system  is utilized in lieu
of a conventional wet tower cooling system.  These preliminary economic compari-
sons are discussed in SectionXII of this report.
                                    134

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Method of Analysis and  Description
of Tower Optimization Program

        An  important preliminary step in determining the economically optimum ITD
for a dry-type cooling system is the selection of the design parameters of the cool-
ing tower to choose the  optimum design to use in the studies.  The basis  for the
method of optimizing  the tower design was described in Section III  — Performance.
The purpose of this optimization is to determine the most economically optimum
balance of tower and  equipment to achieve the  lowest construction capital cost.

        Since a given ITD can be  achieved by a number of combinations of air flow,
water flow, water temperature range and approach to the ambient air temperature,
it was  necessary to evaluate their interaction.  Five parameters are considered in
the tower optimization.  They are range, amount of heat rejected, quantity of water
flowing in  the coils, ambient air  temperature and tower elevation.  Output of the
program is  tower height, stack diameter at the top of the tower, diameter at its base
and its cost.

        For a given heat rejection at a given ITD,  the range and water flow are
varied to provide the  minimum cost of the tower.

        The cost estimate is divided into tower structure,  condenser, piping and
controls.  The cooling coils are included with the  tower cost.   A sample computer
printout is  shown in Table 7.

        The basic capital cost of the dry cooling systems which  were developed are
assumed to be average United States costs.  Applicable construction cost indexes
have been  analyzed and capital cost multipliers have been determined for each of
the 27 sites to approximately reflect changes in capital costs which may  be expected
from area to area.

        In addition, structural analyses of the natural-draft cooling tower indicate
that the capital  cost of  the natural-draft, dry-type cooling system should be in-
creased by about 2 percent to reflect the higher cost tower structure necessary in
areas subject to hurricanes.  This 2 percent adjustment is reflected in  the capital
cost multipliers applied to natural-draft systems for two of the  sites investigated.

        A procedure,  comparable to the one described above,  was  followed in the
optimization of mechanical-draft towers.

        The physical sizes of the  dry cooling towers corresponding to the ITD values
are shown  in Figure 38 for natural-draft towers  and Figure 39 for mechanical-draft
towers.
                                     135

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                             TABLE 7
   COMPUTER PRINTOUT - NATURAL-DRAFT COOLING TOWER SYSTEM-
                  SIZING AND COSTING PROGRAM
 DRY COOLING TOWER SIZING AND COST EVALUATION
 DESIGN PARAMETERS
          ITD =   60     RANGE -   30
          HEAT REJECTION  =   4.0E+09
          WATER FLOW PER HOUR =   2.2E-K)5
          AMBIENT AIR TEMP -   50    ELEVATION =
   3000
TOWER SIZING
          TOWER HEIGHT -  539.1
          UPPER DIAMETER =  346.8
          BOTTOM DIAMETER  =  450.6
          GALLONS PER MINUTE =  266549
COST EVALUATION
    TOWER STRUCTURE
          STACK COST             2120830
          SHED COST               493872
          COIL COST              4408000
          TOTAL STRUCTURE
    CONDENSER
          CONDENSER COST
    PIPING, VALVES, ETC.
          PIPE COST               1306835
          VALVE COST              833500
          PUMP COST              1200000
          FILLER PUMP COST           40000
          STORAGE TANK COST        28720
          TOTAL PIPING FACILITIES
    CONTROLS
          CONTROL COST
COMPLETE TOWER FACILITIES
          TOTAL TOWER COST
TOTAL TOWER COST AND CONTINGENCIES
 7022702
  832000
 3409054

  500000

11763756
14704696
                               136

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                                                     30  40   50   60   70  80
                                            INITIAL  TEMPERATURE  DIFFERENCE
                          FIGURE  38-COOLING TOWER DIMENSIONS AS A FUNCTION OF INITIAL TEMPERATURE
                          DIFFERENCE AND ELEVATION FOR NATURAL-DRAFT COOLING  TOWERS- STEEL AND
                                   ALUMINUM CONSTRUCTION-80OMW GENERATING CAPACITY

-------
   18
  16
  14
              	
                         COOLING TOWER
                         NUCLEAR-FUELED
                         GENERATING UNIT
               COOLING TOWER
               FOSSIL-FUELED
               GENERATING UNIT
            40        50        60        70
           INITIAL TEMPERATURE DIFFERENCE  (°F)
FIGURE 39—GROUND AREA REQUIREMENT AS A  FUNCTION
OF INITIAL TEMPERATURE DIFFERENCE FOR MECHANICAL—
     DRAFT,DRY COOLING TOWERS- 800 MW UNIT
                       138

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        Figure 40 shows the capital  cost, in dollars per kw,  of natural-draft and
mechanical-draft systems for both fossil- and nuclear-fueled plants. The costs are
shown for ITD values of 30° to 80 F and for ground elevations of 0, 3,000  and
6,000 feet.

Factors  Affecting the Economic Optimization
of Dry-Type Cooling Towers

        The economically optimum dry cooling system for a specific set of condi-
tions is  that which results in the lowest annual cost. The annual cost must  reflect
all  costs incurred on an annual  basis, such as operation, maintenance, total plant
fuel costs and the annual capital costs.

        The performance of dry  cooling systems has been discussed  in Section III  of
this report and the key factors affecting the economic optimization stem from those
performance characteristics and the capital cost of the cooling system.  These key
factors  are:

        1 .    The effect of increasing the  ITD and/or the ambient air
              temperature is to increase the temperature of the turbine
              exhaust steam and, therefore, the turbine back pressure.
              An increase in  turbine back  pressure results in poorer fuel
              economy and in loss of generating capability.

        2.    The physical size and,  therefore,  the capital  cost of the
              dry-type cooling system decreases with increasing ITD.

        The combination of the above factors indicates that a dry cooling system
could be:  1)  a low-lTD, high-capital-cost system with good fuel economy and
little or no loss of generating capability at high ambient air temperatures;  or, 2) a
high-lTD, low-capital-cost system  with poorer fuel economy and a significant loss
of generating capability at high ambient air temperatures;  or, 3) some intermediate-
size cooling system.  The economically optimum dry cooling system for a specific
location and specific set of conditions must reflect the effect of a number of varia-
bles. Those variables which affect the economic optimization are discussed below.

        Performance related to ITD.  The effect of increasing the  ITD is to increase
the exhaust steam temperature  for a given air temperature and,  therefore,  the tur-
bine back pressure. This results in poorer fuel economy and loss of generating cap-
ability.

        Capital cost of the dry cooling system.  The physical size and the  capital
cost of the dry cooling system decrease with increasing ITD.
                                     139

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  60
                         (o) NATURAL DRAFT
                             FOSSIL  FUEL
                                                                         (b) NATURAL  DRAFT  TOWER
                                                                             NUCLEAR  FUEL
                                                                     6000 FT.
                                                                     3000 FT.     .
                                                                     SEA LEVEfc—it
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<40
LEGEND:
I. STEEL CYLINDRICAL CONSTRUCTION
2. COSTS ARE FOR AVERAGE US CONDITIONS.
  COSTS IN AREAS SUBJECT TO HURRICANE WINDS
  WOULD BE APPROXIMATELY 2 % HIGHER
LEGEND:
I. STEEL CYLINDRICAL CONSTRUCTION
2. COSTS ARE FOR AVERAGE U.S CONDITIONS.
  COSTS IN AREAS SUBJECT TO HURRICANE WINDS
  WOULD BE APPROXIMATELY  2 % HIGHER
                          (C)
                   MECHANICAL  DRAFT TOWER
                   FOSSIL FUEL
      (d)  MECHANICAL  DRAFT TOWER
          NUCLEAR FUEL
  10
     30        40        50        60        70        80      30        40        50

                                       INITIAL TEMPERATURE  DIFFERENCE (°F)

                             FIGURE 4O-RELATIONSHIP OF DRY COOLING  SYSTEM CAPITAL COST
                                TO ITD AND ELEVATION-BOO MW  GENERATING PLANT
                                                  (1970  COST LEVEL)
                                                                                             60
                                                                                                        70
                                                                                                                  80

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       Elevation.  The effect of increasing ground-level elevation  is to increase
the capital cost of the dry cooling system since the reduced air density makes it
necessary to move  a greater volume of air past the cooling elements in order to
achieve the same mass flow rate of air.

       Fixed-charge  rate. The fixed-charge rate is a percentage rate applied to
the capital cost which reflects the following items as defined by  the Bureau of
Power of the Federal Power Commission  (40):

       1 .    Interest, or cost of money.

       2.    Depreciation, or amortization .

       3.    Interim  replacements.

       4.    Insurance,  or payments in lieu of insurance.

       5.    Taxes (federal, state and  local),  or payments in lieu
             of taxes.

       The effect on  the economic optimization of  an increase in the fixed-charge
rate is to give more weight to capital costs and less weight to annual operation,
maintenance and fuel costs.

       Ambient air temperatures. The  effect of higher ambient  air temperatures is
to increase the turbine back pressure resulting in  poorer fuel  economy and loss of
generating capability.  The full range of annual air temperatures at the site affect
the fuel  economy, but it is the extreme high temperature which  has the more signi-
ficant economic effect.  The extreme high temperature, in combination with the
cooling system ITD and the turbine characteristics,  sets the maximum loss of gener-
ating capability which would be experienced during the year.

        Fuel  costs. The  effect of increasing the unit cost of fuel is to  increase the
weight given to fuel economy and decrease the weight given to  capital cost con-
siderations.  Therefore,  increasing the  fuel cost would  tend to reduce the optimum
ITD, or, in other words, would tend toward a higher capital  cost cooling system.

        Turbine  performance.  The shape of the turbine performance curve of heat
rate versus back pressure is important in that it affects the relative  importance of
fuel economy and  loss of generating capability.  As shown on Figure 28 in Section
III of this report,  a conventional turbine modified to operate at  high back pressures
would have a relatively  low  heat rate at low back pressures, and would have a
poorer heat rate with resulting loss of both economy and generating capability  at
high back pressures.  On the other hand, the high-back-pressure turbine would
                                     141

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 have a poorer heat rate at low back pressures than the modified conventional tur-
 bine, but the loss of capability at high back pressures would not be as pronounced.

        Auxiliary power requirements.   The power requirements for pumps and fans
 decrease with increasing ITD.

        Replacement of capacity losses.  Since some loss of generating capability
 can  be expected at high ambient air temperatures with a dry cooling system,  the
 replacement of this capability is an important consideration in the economic anal-
 ysis. The relative importance of this capability loss may vary from area to area and
 would,  of course,  be of the greatest concern in an area where the  annual peak
 electrical demand occurs in the summer rather than in the winter.  In this instance,
 the capability lost would need to be replaced from some other source.  A utility
 having a winter peak system demand would not be as much affected  by the loss of
 generating capability on a hot summer  day, with respect to meeting its  own demands,
 but may still  be economically interested in the lost capacity since that utility may
 have the opportunity to sell surplus capacity to other interconnected systems.  Once
 it has been determined whether or not the replacement of the lost capacity is nec-
 essary,  then  the  cost of replacing that capacity  must be determined.  In the opti-
 mization, the economic impact of  the capacity loss increases as the cost of replac-
 ing that capacity increases.  Therefore, the significance of lost capacity is much
 greater if the  lost capacity is replaced at a capital  cost of $150 per kw than if it is
 replaced at a capital  cost of $100  per kw.

 Method of Analysis  and Description of
 the Economic Optimization Program

       The method  of analysis which was applied in the determination of the eco-
 nomically optimum dry cooling system for various conditions is based on an analysis
 of all costs which would be affected by the choice of size, or ITD, of the dry cool-
 ing system.  Therefore,  the costs  reflected  are the plant fuel  cost and all  costs
 related to the dry cooling system which is  defined  as those facilities from the tur-
 bine flange outward.  These  facilities would include the condenser, the cooling
 system piping, water storage facilities, pumps,  valves, controls,  recovery turbine
 if used, and  the  cooling tower with its heat exchanger equipment.  The  analysis
does not include consideration of the other generating plant costs since those costs
would not vary with the selection of the dry cooling system ITD.  Also included  in
 the analysis is the economic consideration of the generating capability lost at high
ambient air temperatures.

       For the purposes of this analysis,  fossil-  and  nuclear-fueled  generating
 plants of 800-mw size were assumed.   The  results  of the analysis, as evaluated  on
a cost-per-kw basis, should be generally applicable to generating plants in the size
range of 600 mw to  1,000 mw, or  perhaps  over a somewhat larger range of sizes.
                                    142

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It is recognized that it may not be practical to build an 800-mw unit at some of the
sites which were selected for analysis.  The sites included in the analyses were not
selected as being particularly likely sites for the construction of an 800-mw  unit,
but were selected to represent a variety of air temperature, elevation and fuel cost
conditions so that the effect of these factors on the economic optimization could be
analyzed.

        The method of analysis and the computer program were  developed to  handle
both nuclear- and fossil-fueled generating units and both natural-draft and mechan-
ical-draft dry cooling systems.  Although a dry-type cooling tower has not yet been
used with a nuclear plant, there is a great need for such combination due to the
relatively large amount of waste heat rejected by the turbine and  consequent heat
addition to natural bodies of water as compared to fossil-fueled plants.  However,
before a dry-type tower  can be built with a nuclear plant, important questions in-
volving shielding requirements,  necessitated as a  result of the direct mixing  of
turbine exhaust steam and circulating water, must be resolved by  the agencies hav-
ing jurisdiction over these matters.

        On the basis of information obtained as to the sizing and performance char-
acteristics of existing dry cooling systems, it was  determined that the analysis
should cover a range  of ITD values and that range was established as 30°F to 80°F.
The design value of ITD  for a dry  cooling system is related to a specific  value of
heat rejection, as discussed in Section II of this report.  For this analysis, the
design ITD is that which occurs at a nominal heat rejection of  4 x 10' Btu per hour
for a fossil-fueled plant, and at a nominal value of 6 x 10  Btu per hour for a
nuclear-fueled plant. As shown in Figure 24 of Section III of  this report, the heat
rejection capability of the cooling system varies with turbine back pressure.  The
fossil plant nominal heat rejection value of 4 x 10  Btu per hour and the nuclear
plant nominal heat value rejection of 6 x 10  Btu per hour both occur at a turbine
back pressure of approximately 8  inches Hg.  The following tabulation shows the
rates of heat rejection requirements for the fossil and nuclear plants  for several
specific back pressures.   The  heat rejection values shown in the table are based on
full throttle flow  performance and indicate the reduced generating capability at
elevated back pressures.
                                     143

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                                    TABLE 8

                        Heat Rejection Versus Back Pressure
                          for an 800-Mw Generating Unit
                          (Full Throttle Flow Performance)
    Turbine                    Fossil                          Nuclear
 BackPressure      Output       Heat Rejection     Output      Heat Rejection
   (in.  Hg)           (mw)        (1 09 Btu/hr.)       (mw)         (TO9
      2.0            809             3.80           814             5.68
      4.0            796             3.85           790             5.76
      6.0            771             3.93           751              5.89
      8.0            750             4.01            718             6.01
     10.0            728             4.08           692             6.09
     12.0            709             4.14           672             6.16
     14.0            692             4.20           655             6.22


        The performance of the cooling system and the turbine have been discussed
 previously in Section III of this report.  Figure 24 of Section III illustrates the in-
 terrelationship of the tower and turbine performance curves.

        As a result of preliminary analysis, a standard design assumption which re-
 sults in a dry-type cooling system very close in cost to trie cooling system that
 would  be selected by much more detailed analysis was established.  The more de-
 tailed  analysis would consist of an evaluation  for each set of conditions of the
 economic effect of varying range, approach, airflow and water flow. The approach
 of this study has been to establish the over-all economics of dry cooling systems for
 a large number of combinations of conditions.  This has  been accomplished.   Once
 a specific site has been selected for a detailed analysis, it would, of course, be
 necessary to thoroughly investigate the effect  of these other variables  in order to
 refine  the cooling system design.

        The economic optimization program consists of an analysis of the annual
 costs which are affected by the size, or ITD, of  the dry cooling system  for each
 1°F differential of ITD between ITD values of  30°F and  80°F.  The costs evaluated
are the annual  capital cost of the dry cooling  system, calculated as the capital cost
 times the fixed-charge rate; the annual operation and maintenance cost of the dry
cooling system; the annual fuel cost of the 800-mw unit; the annual cost of power
and energy required by the cooling system pumps and fans; and the annual cost of
replacing the capacity and energy lost due to  high-back-pressure operation.  These
annual costs are summed and the minimum value of that sum within the  range of
 ITD values analyzed  defines the economically  optimum size of dry cooling system
                                    144

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Each of the analyses made reflects certain specific assumptions as to site elevation,
annual air temperatures, season of system peak demand, type  of fuel (fossil  or
nuclear), fuel cost,  turbine characteristics, method of operation of the generating
plant,  type of cooling tower (natural draft or mechanical draft),  cooling system
capital costs,  fixed-charge rate, cost of auxiliary power and energy,  and cost of
replacing lost capacity.

        For the purposes of the economic optimization analyses which are sum-
marized  in Section X of this report,  27 sites within the United States were chosen
to represent a range  of air temperature conditions, ground-level elevations and fuel
costs.  The  annual  air temperature data were obtained from the U.S. Weather
Bureau Bulletin 82, "Climatography of the United States" (41), which summarizes
the frequency of occurrence of air temperatures.  The ground-level elevations used
in the analyses were the weather station elevations rounded to  the nearest TOO feet
above sea level.

       Analyses were made for each of these 27 sites for both fossil- and nuclear-
fueled generating units,  natural-draft and mechanical-draft dry cooling systems,
with fixed-charge rates ranging from 8 percent to 18 percent and a range of fuel
costs.

        It is believed that the range of fixed-charge rates of 8  percent to 18 percent
represents the range of values which would be applicable to electric utilities in the
United States for new construction.

        The fuel costs selected are  generally representative of existing fuel  costs.
In some cases, the highest value  of fuel  cost investigated may be somewhat higher
than current fuel  costs, but no attempt has been made to predict future fuel prices,
just as no attempt has been made to predict the future capital  cost of the  dry-type
cooling system. In general, all costs used in the analyses are current (1970) prices.

        In all  cases, an allowance for the operation  and maintenance cost of  the
dry-type cooling system was estimated at 1 percent of the capital cost of the  dry-
type cooling system.

        For  this analysis, it was assumed that the generating plant would  operate
7,500 hours per year, 50 percent of that time at  full  throttle and 50 percent of that
time at 75 percent load, which is equivalent to 600 mw.  The annual generation
required by the system from this 800-mw unit under the assumptions stated would,
therefore, be 5,250,000 mwh.  In  the analyses,  the total energy production of the
800-mw unit was computed, reflecting  both capacity gains and losses. The energy
gains and losses were then computed and considered in the economic  analysis.  A
credit for the energy gains, reflecting the fuel cost of generating that energy, was
calculated  and the cost of replacing  the energy  losses was  also calculated,  as
                                     145

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described later.  The net result is that the analyses reflect the cooling system costs
and  the costs of supplemental peaking generation necessary to produce a combined
output of 800 mw at the high air temperatures, as defined later, and a combined
energy output of 5,250,000 mwh per year.

       As discussed previously, the economic impact of the loss of generating capa-
bility at high ambient air temperatures is more significant if the period of maximum
demand occurs  in the summer rather than in the winter.  For the basic analyses,  a
summer peak was  assumed for all sites except the one located in Alaska.  Some sup-
plemental analyses were made to indicate the effect of assuming a winter peak.   It
is recognized that not all  of the sites investigated lie in summer peak areas,  but it
was  assumed that  utilities constructing plants in these areas would have the oppor-
tunity to market excess generating capability to other utilities in the summer and,
therefore, would  have an economic interest in the loss of generating capability.

       For the  purposes of these analyses, the loss of generating  capability was
evaluated at the ambient air temperature for the  site  which is equalled or exceeded,
on the average, only 10 hours each year.   It would not be  reasonable to evaluate
the lost capacity  for the extreme maximum temperature. The peak electrical system
load may not occur at the coincidental time of maximum hourly temperature.  In
fact, the use of the temperature equalled or exceeded only 10 hours each year may
be somewhat severe and it may be reasonable to evaluate that loss of capability at
a  lower air temperature.  A possible alternative to the  10-hour temperature would
be that temperature which is equalled or exceeded only 1 percent of the time dur-
ing the 4-month period June through September.  That  temperature duration would
be 1 percent of 2,928 hours, or about 29 hours.  This temperature duration is com-
monly used in the analysis of wet-bulb temperatures for wet cooling tower design.

       It was assumed that the loss of capability which was experienced at the air
temperature which is equalled or exceeded only 10 hours per year would be re-
placed by generation from another source.  The basic analyses are based on the
assumption  that the source of the replacement capacity  would be peaking units
having a capital cost of $100 per kw.  The loss of energy generated due to high-
back-pressure conditions would also be replaced by these peaking units.  The  cost
of energy from these peaking units reflects a heat rate of 15,000 Btu per kwh and,
in most cases, a fuel cost of $0.40 per million Btu.  In  some cases, it was assumed
that  natural gas would be available to operate the peaking and, therefore, a some-
what lower fuel cost was assumed.

       The cost of the auxiliary power and energy required for the cooling system
pumps and fans was calculated assuming  that incremental steam plant capacity
could be provided for a cost of $150 per kw for fossil-fueled units and $225 per kw
for nuclear-fueled units,  and that the energy cost for the auxiliaries would be the
average fuel cost  in mills per kwh of the 800-mw plant  plus an allowance for oper-
ation and maintenance costs of 0.1 mills per kwh.
                                    146

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       Perhaps the method of analysis can best be described by summarizing the
results of one of the analyses as presented on the computer printout.  Table 9 shows
the computer printout for a natural-draft dry cooling system associated with a fossil-
fueled plant at Burlington, Vermont for a plant fuel cost of 25$ per million Btu, a
peaking fuel cost of 40$ per million Btu and an annual fixed-charge rage of 15 per-
cent .

       Referring to Table 9, the first column shows the initial temperature differ-
ence (ITD).  The second column shows the gross energy generation of the 800-mw
unit reflecting both the capacity gains at back pressures less than 3.5 inches Hgand
capacity losses at back pressures above 3.5 inches Hg.  The third column shows the
amount of the excess energy due to operation at back pressures under 3.5 inches Hg.
The fourth column shows the energy associated with capacity losses at back pres-
sures above 3.5 inches Hg.  The column headed "Auxiliary Energy" shows the
annual energy requirement of the cooling system pumps.  The column headed "Loss
of Capacity" shows the capacity lost at the air temperature which is equalled or
exceeded only  10 hours each year.  The column headed "Maximum Auxiliary Power"
shows the maximum capacity required for the cooling system pumps.  For the
mechanical-draft analyses,  the auxiliary power and  energy requirements would, of
course,  also reflect the cooling system fan requirements.

       The next seven columns of Table 9 show annual  cooling system costs. The
column "Annual Capital and O&M Cost of Dry Cooling System"  is the capital cost
of the dry cooling system multiplied by the fixed-charge rate plus the 1  percent
allowance for operation and maintenance cost.  In this case, the column is com-
puted at the capital cost times 16 percent.  The next column shows the total annual
fuel cost of the 800-mw unit and reflects the gross energy generation as shown in
column two.  The column headed "Credit for Excess  Energy" is a fuel-cost credit
related to the excess energy amounts shown in the third column.  This credit re-
flects the fuel cost of  energy generated by the 800-mw  unit.  The column headed
"Capacity Penalty Cost" reflects the costs of replacing both the  capacity and
energy losses due  to operation and  at back pressures  above 3.5 inches Hg.  The
column headed "Auxiliary Cost" reflects the cost of  providing the power and energy
necessary to supply the cooling system pumps.

       The total annual cost in dollars  is the sum  of the preceding five columns
and the total annual cost in mills per kwh is that sum divided by 5,250,000 mwh.
The optimum  ITD is that which produces the lowest total annual  cost and in this
case is 57°F, which results in a total annual cost of  $15,203,529, equivalent to
2.8959 mills per kwh.
                                    147

-------
00
               TABLE 9—COMPUTER PRINT-OUT, ECONOMIC OPTIMIZATION.800 MW,FOSSIL-FUELED
                    GENERATING UNIT, NATURAL-DRAFT TOWER, BURLINGTON .VERMONT
                CAPITAL COST FACTORS!      PLANT - 15 0/0
                PLANT FUEL COST - 35 cENTS/io««6 BTU
                PEAKING CAPITAL COST - 100 S/KM
 PEAKING CAPACITY - 15 0/0  AUXILIARIES - 15 0/0
PEAKING FUEL COST - 40 CENTS/10»«6 BTU
 AUXILIARY CAPITAL COST - 150 S/KW
INIT.
TEMP.
OIFF.
(OEG )
< F )
30
31
32
33
34
35
36
37
38
39
40
41
42
43
44
45
46
47
46
49
50
51
52
53
54
55
56
OPT 1 Ml
57
58
59
60
61
6?
63
64
65
66
67
68
69
70
71
11
73
74
75
76
77
78
79
DO
EXCESS
GBOSS ENERGY
ENERGY DUE TO CAPACITY MAXIMUM
BOO MH EXTRA PENALTY AUXILIARY LOSS OF AUXILIARY
UNIT CAPACITY cuirorw ruroftv rADiriTv DAyro
IMHH) (HUH)
5285897
5285716
S28SS11
5285289
5285049
5284789
5284495
5284161
5283799
5283413
5282994
5282538
5282023
5281468
5280876
5280236
5279550
5278787
5277968
5277095
5276159
S275160
5274069
5272901
5271663
5270351
5268947
JM:
5267444
5265842
5264153
5262376
5260488
5258491
5256376
5254158
525 IB 39
5249400
5246839
5244144
5241331
5238413
5235365
5232185
5228856
5225404
5221834
5218126
5214269
5210265
5296124
5201851
35901
35723
35521
35303
35067
34818
34548
34239
33903
33544
33166
32808
32398
31953
31476
30959
30536
30049
29521
28955
28341
27863
27354
26800
26204
25563
25043

24548
24002
23413
22784
22239
21772
21251
20687
20085
19526
19088
18592
18052
17475
16898
16478
15992
15463
14895
14288
13862
13374
12841
12268
(MWH)
3
7
11
15
19
29
53
78
104
131
171
271
375
485
600
724
986
1262
1553
1860
2182
2703
3286
3899
4541
5213
6096

7104
8160
9260
10408
11751
13281
14875
16529
18246
20125
22249
24447
26721
29063
31533
34292
37136
40059
43061
46162
49S92
53109
56718
60417
(MHH) (KM)
9337S
90624
87966
85401
82929
80550
78264
76071
73971
71964
700SO
68229
66501
64866
63324
61875
60519
59256
58086
57009
56025
54878
53778
52734
51717
50756
49842

48974
48153
47378
466SO
4S861
45099
44364
43656
429 7S
42321
41694
41094
40521
39975
39402
38847
38309
37788
37284
36798
36129
35877
35442
35025
249
1077
1942
2846
3791
4819
5944
7144
8426
9791
1124Z
12r«6
14429
16160
17939
19797
21737
23690
25669
27694
29717
J1711
33691
35673
37664
39648
41724

43S78
46089
48134
50623
52939
55279
57635
59995
62387
64850
67357
69907
72509
75182
77916
80676
83528
86455
89420
92425
95445
98306
101180
104J33
(KM)
12450
12083
11729
11387
11057
10740
10435
10143
9863
9595
9340
9097
8867
8649
8443
8250
8069
7901
7745
7601
7470
7317
7170
7030
6896
6767
6646

6530
6420
6317
6220
6115
6013
5915
5821
5730
5643
5559
5479
5403
5330
5254
5180
5108
5038
4971
4906
4844
4784
4726
4670
ANNUAL
CAPITAL
AND 0»M
COST
OF DRY
COOLING
SYSTEM
(»l
5004988
4825202
4651822
4484847
4324277
4170112
4022353
3880999
3746051
3617507
3495369
3379636
3270309
3167387
3070870
2980758
2897052
2819750
2748855
2684364
2626279
2560865
2499359
2441762
2388073
2338292
2292419

2250455
2212399
2178251
2148011
2109274
2072190
2036759
2002981
1970856
1940383
1911563
1884396
1858882
1835020
1806150
1777721
1749735
1722189
1695086
1668424
1642204
1616425
1591088
1566191
ANNUAL
FUEL
COST
OF
800 MK
UN I T
(»)
11969738
11969991
11970270
11970573
11970904
11971255
11971636
11972053
11972508
11973001
11973532
11974096
11974704
11975365
11976081
11976852
11977681
11978568
11979502
11980516
11981603
11982775
11984035
11985380
11986787
11988291
11989895

11991618
11993459
11995419
11997481
11999626
12001916
12004350
12006932
12009669
12012557
12015513
12018648
12021940
12025418
12029066
12032928
12036879
12041004
12045298
12049788
12054518
12059507
12064623
12069865
CREDIT
FOR
EXCESS
ENERGY
($>
80808
80411
79961
79472
78944
78385
77781
77089
76335
75529
74680
73878
72959
71958
70886
69724
68773
67681
66493
65219
63837
62762
61620
60372
59031
57587
56417

55303
54076
52751
S1333
50106
49056
47885
46614
45258
43997
43015
41899
40685
39385
38084
37140
36049
34859
33580
12212
11254
10158
28960
27669
CAPACITY
PENALTY AUXILIARY
COST fner
(Sl
4055
17490
31525
46193
6JS27
78244
96610
116199
137124
159400
183149
208757
236000
26470!
294213
325053
358052
391350
425158
459800
494510
529939
565508
601297
637400
673574
712506

753445
795601
838571
882539
928120
975204
1022939
1071092
1120144
1171322
1224678
1279177
1334972
1392323
1451437
1512705
1575968
1640924
1706971
1774260
1843763
1911211
1479421
2049456
t>ua I
(S)
500907
486156
471919
458171
444914
432169
419915
408175
396925
386165
375918
366162
3S69JO
348168
339907
332160
324903
318161
311910
306150
300904
294779
288900
283290
277927
272787
267939

263315
258938
254831
250970
246786
242736
238841
235102
231496
228046
224728
221567
218562
215690
212688
209775
206948
204209
201580
199039
196609
194267
192013
189846
TOTAL ANNUAL
COST OF COOLING
SYSTEM AND TOTAL
PLANT FUEL
(S) (MILL/KNHI
17398879
17218428
17045574
16880312
16722677
16573395
16432735
16300338
16176272
16060545
15953289
15854774
15764975
15683663
15610184
15545098
15488916
15440149
15398932
15365612
15339459
15305596
15276182
15251357
15231156
15215357
15206342

15203529
15206321
lb?14321
15227669
15233700
15242990
15255004
15269493
15286906
15308311
15333467
15361889
15393670
15429066
15461258
15495989
15533461
15573467
15615356
15659299
1570S841
15751252
15798185
15847691
3.3141
3.2797
3.2468
3.2153
3.1853
3.1568
3.1300
3.1048
3.0812
3.0592
3.0387
3.0200
3.0029
2.9874
2.9734
2.9610
2.9503
2.9410
2.9131
2.9268
2.9218
2.9154
2.9097
2.9050
2.9012
2.898?
2.8964

2.8959
2.8964
2.89HO
2.9005
2.9017
2.9034
2.9057
2.9085
2.9118
2.9159
2.9207
2.9261
2.9321
2.9389
2.9450
2.9516
2.9588
2.9664
2.9744
3. 9827
2. 9916
3.0002
3.0092
3.0186

-------
                                  SECTION X
                 RESULTS OF THE ECONOMIC OPTIMIZATION
        Economic optimization  analyses of dry-type cooling systems for electric-
generating plants were made for 27 selected sites in the United States,  including
one site each in the states of Hawaii and Alaska.  The  sites were selected  to repre-
sent a range of annual  air temperatures, ground-level elevation, and fuel  cost, all
of which have some effect on the  economic optimization.

        Four basic sets  of economic optimization analyses were performed for each
of the 27 sites.  These  basic analyses were for the following conditions:

        1 .    Fossil-fueled generating plant,  natural-draft tower.

        2.    Fossil-fueled generating plant,  mechanical-draft tower.

        3.    Nuclear-fueled generating plant, natural-draft tower.

        4.    Nuclear-fueled generating plant, mechanical-draft tower.

        Fifteen analyses were made for each site for each of the 4 basic conditions
summarized above.  These 15 analyses reflect the combination of 5 fixed-charge
rates and 3 fuel cost assumptions.  Therefore,  a total of 4  times 15, or  60  analyses
were made for each site reflecting the basic assumptions.  In addition,  as  discussed
in Section XI,  some supplemental analyses were made to illustrate the effect of
varying certain parameters over a wider range.

        The basic assumptions used in the economic optimization analyses have been
previously discussed in Section  IX of this report.

        Table 10 shows the 27 sites which were analyzed; summarizes the site air
temperature conditions and ground elevations; and summarizes  the assumptions made
as to fuel cost and capital cost  multipliers.

        As discussed in Section  IX of this report, the computer printouts show,  for
each analysis reflecting a specific set of assumptions, the  total annual  cost for those
cost items which are affected by the selection of the dry cooling tower size, orlTD.
As expected, the shape of the curve of annual cost versus  ITD  is affected by the
assumption as to the season of peak electrical demand.  Figure 41 shows typical
curves of annual cost versus ITD for a summer peaking assumption and for a winter
peaking assumption. Under the winter peaking assumption, the annual cost con-
tinually decreases with increasing  ITD to the 80°F ITD limit established, reflecting
                                    149

-------
                                                                                 TABLE 10



                                                                       Economic Optimization Analysis

                                                             Summary of Sites,  Site Data and Study Assumptions (1)
Oi
o

Site
No.
1
2
3
4
5
6
7
8
9
10
11
12
13
14
15
16
17
18
19
20
21
22
23
24
25
26
27


Area and
Pacific:


Mountain:





West North Central:


West South Central:


East North Central:



East South Central:
New England:
Mid-Atlantic:
South-Atlantic:


Hawaii:
Alaska:


City
Seattle, Wash.
San Francisco, Calif.
Los Angeles, Calif.
Great Falls, Mont.
Boise, Ida.
Casper, Wyo.
Reno, Nev.
Denver, Colo.
Phoenix, Ariz.
Bismarck, N. Dak.
Minneapolis, Minn.
Omaha, Neb.
Little Rock, Ark.
Midland, Tex.
New Orleans, La.
Green Bay, Wis.
Grand Rapids, Mich.
Detroit, Mich.
Chicago, III .
Nashville, Tenn.
Burlington, Vt.
Philadelphia, Penna .
Charleston, W. Va.
Atlanta, Ga.
Miami, Fla.
Honolulu, Hawaii
Anchorage, Alas.
Ground-Level
Elevation (2)
(ft.)
400
0
100
3,700
4,000
5,300
4,400
5,300
1,100
1,600
800
1,000
300
2,900
0
700
700
600
600
600
300
0
1,000
1,000
0
0
100
Ambient Air Temp. (°F)
Annual
Median
50
56
62
46
50
45
48
51
72
43
47
53
65
65
71
44
49
51
50
64
46
56
57
64
77
76
38

lOhrs. (3)
91
89
93
94
101
96
101
97
114
100
97
103
105
105
97
92
96
97
96
103
92
99
96
100
97
92
77
Capital
Cost
Multiplier (4)
1.0
1.05
1.0
1.0
1.0
0.95
1.05
0.95
1.0
1.0
1.0
1 .0
0.90
0.95
0.95/0.97
1.0
1.0
1 .05
1.05
0.90
0.95
1.0
1.0
0.95
1.0/1.02
1.10
1.50
Fuel Cost Range (C/10 Btu)
800-Mw Unit
Fossi 1
25-40
25-40
25-40
15-30
20-35
10-25
25-40
20-35
20-35
12-25
25-40
25-40
25-40
20-35
20-35
25-40
25-40
25-40
25-40
18-30
25-40
25-40
15-30
25-40
25-40
30-45
30-45
Nuclear
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
11-20
Peaking
40
40
40
40
40
15-30
40
40
40
40
40
40
25-40
20-35
20-35
40
40
40
40
40
40
40
40
40
40
40-45
30-45
              Note:  See footnotes on following page.

-------
                            TABLE 10 FOOTNOTES


(1)    General assumptions:

      a.    Unit size:  800 mw .

      b.    The following basic conditions were studied for each site:

                 Fossil-fueled unit, natural-draft tower
                 Fossil-fueled unit, mechanical-draft tower
                 Nuclear-fueled unit, natural-draft tower
                 Nuclear-fueled unit, mechanical-draft tower

      c.    Fifteen analyses were made at each  site  for  each of the 4 basic
           conditions. The 15 analyses reflect the combination of 3 fuel cost
           assumptions and 5 assumptions as to fixed-charge rates.

      d.    The fixed-charge rates assumed were 8%,  10%,  12%,  15%, and 18%.

      e.    A summer  peak was assumed for all  sites other than  No.  27,
           Anchorage, Alaska.

      f.    The capital cost of peaking capacity necessary to replace capacity
           lost at high back pressures  was assumed to be  $1 00/kw.

      g.    The incremental capital cost of the generating capacity necessary for
           cooling  system pumps and  fans was assumed to be $150/kw for fossil-
           fueled  units and $225/kw for nuclear-fueled units.   The auxiliary
           energy cost was assumed equal to the  fuel  cost of energy from the 800-
           mw unit.

(2)    Weather station  elevation rounded to the nearest 100 feet above sea  level.

(3)    The air temperature equalled or exceeded 10 hours per year.

(4)    Reflects approximate construction cost differences and, for two sites, the
      additional cost of natural-draft towers in areas subject to  hurricane  winds.
      The lower multipliers shown for New  Orleans and Miami are applicable to
      mechanical-draft cooling systems and the higher multipliers are applicable
      to  natural-draft cooling systems.

-------
3
2
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v>
o
o
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)TE. DOES NOT INCLUDE
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COOLING SYSTEM FROM ^
EXHAUST FLANGE OUT
b. OPERATION AND MAINTEN
OF PLANT OTHER THAN
COOLING SYSTEM AND TC
-4-.- ,.- 	 -




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         30        40       50        60        70
            INITIAL  TEMPERATURE DIFFERENCE - °F
80
      LEGEND :  I. FUEL COST USED - 40 {  / I06  BTU
               2. FIXED CHARGE RATE - 15%
               3. CAPITAL COST MULTIPLIER
                SUMMER CURVE-1.0
                WINTER CURVE - 1.5

  FIGURE 41-TYPICAL CURVES OF TOTAL ANNUAL  COST (COOLING
  SYSTEM, PEAKING CAPACITY LOSS PENALTY AND TOTAL PLANT
  FUEL)  VARIATION  WITH ITD  FOR SUMMER AND  WINTER  PEAK-
  ING ASSUMPTIONS
                           151

-------
the capital cost versus ITD relationship.  If this limit  had not been arbitrarily
established, the curve would eventually turn upward, indicating an optimized  ITD
selection.  In the case of the summer peaking assumption, the  annual cost declines
with increasing ITD, up to a certain point, after which the economic effect of the
assumptions as to the replacement of lost generating capability causes the curve to
turn upward.

        For those sites where  a summer peak was assumed, it was found that the
bottom of the optimization curve of annual  cost versus  ITD was fairly flat and  that
a range of ITD values could be defined for which the total annual cost of the cool-
ing system was very close to the total annual cost at the optimum point.   For the
purposes of these analyses, the range of ITD values which are close to the optimum
value has been defined as those values for which the total annual cost is within
0.01 mills per kwh of the cost at the optimum point.

        The results of the economic  optimization analyses reflecting the basic as-
sumptions summarized in Table 10 are presented in the figures and tables described
below.

        The economically optimum values of ITD are summarized on Figures  42
through 45.  These figures show, on a map of the United States, the optimum ITD
values found  for the 15 combinations of fuel cost and fixed-charge  rates which were
investigated for each of the 27 sites.  Figure 42 shows this information for the com-
bination of a fossil-fueled generating unit and natural-draft tower.  The other  3
basic sets of analyses—fossil-fueled  unit, mechanical-draft tower; nuclear-fueled
unit, natural-draft tower; and nuclear-fueled unit, mechanical-draft tower—are
shown on Figures 43, 44 and 45, respectively.

        Referring to Figure 42,  it is noted that the range of economically optimum
ITD values found for Chicago was 55°-57°F. The range of ITD values which were
near the optimum (within 0.01 mills per kwh) was found to be 51°-63°F.  In con-
trast, the range of economically optimum values for Miami was found to be  48°-53°F
and the range of ITD values near the optimum was found to be 44°-56°F.

        For Anchorage, Alaska where a winter peak was assumed,  it was found that
the total annual cost of the dry cooling system was lowest at the largest value of
ITD investigated, 80°F, and,therefore, the dry cooling system  was  not optimized for
the Anchorage site.

        The results which are presented in this section will be discussed in detail  in
Section XI.

        The generating capacity losses which would be  experienced at the ambient
air temperature equalled or exceeded 10 hours per year for the ITD values summa-
                                    152

-------
                                            OPTIMUM  ITD-FOSSIL FUEL-NATURAL-DRAFT
LEGEND:
I  BASED ON THE SITE DATA AND STUDY ASSUMPTIONS SUMMARIZED IN TABLE 9 v
  WITH SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT I> lOCVKW ASSU
2 THE  UPPER FIGURES INDICATE THE RANGE OF ECONOMICALLY OPTIMUM ITD
  VALUES FOUND FOR THE 15 COMBINATIONS OF FUEL COSTS AND Fl
  RATES WHICH WERE ANALYZED FOR EACH SITE
» TUC  i ni»/FR FIGURES IN PARENTHESES INDICATE THE RANGE OF ITD VALUES
  FOR  ™THEUTOETSA"^ANNUAL COST OF PLANT OPERATiON IS WITHIN O.OI MILLS/KWH

(i) NOTom^zED7 THE" LOWEST "COST' WAS FOUND AT THE HIGHEST ITD VALUE INVESTIGATED, eo •
                                            FIGURE 42 —ECONOMICALLY OPTIMUM VALUES OF  INITIAL
                                               TEMPERATURE DIFFERENCE W — FO SSIL-FU ELED
                                                GENERATING UNIT-NATURAL-DRA FT TOWER

-------
                                            OPTIMUM  ITD-FOSSIL FUEL-MECHANICAL-DRAFT
  BASED ON THE SITE DATA AND STUDY ASSUMPTIONS SUMMARIZED IN TABLE 9
  WITH SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT « 100/KW ASSIJ
2. THE UPPER FIGURES INDICATE THE RANGE OF ECONOMICALLY OPTIMUM ITD
  VALUES FOUND FOR THE 15 COMBINATIONS OF FUEL COSTS AND FIXED CHARGE
  RATES WHICH WERE ANALYZED FOR EACH SITE
3. THE LOWER FIGURES,IN PARENTHESES, INDICATE THE RANGE OF ITD VALUES
  FOR WHICH THE TOTAL ANNUAL COST OF PLANT  OPERATION IS WITHIN 0 Ol MILLS/KWH
  OF THE COST AT THE OPTIMUM POINT
(I) NOT OPTIMIZED. THE  LOWEST COST WAS FOUND AT THE HIGHEST  ITD VALUE INVESTIGATED, 80 °F
                                           FIGURE  43— ECONOMICALLY OPTIMUM VALUES OF INITIAL
                                              TEMPERATURE DIFFERENCE (eF) — FOSSIL-FUELED
                                               GENERATING UNIT- MECHANICAL-DRAFT TOWER

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                                            OPTIMUM ITD-NUCLEAR FUEL-NATURAL-DRAFT
LEGEND:
I. BASED ON THE SITE DATA AND STUDY ASSUMPTIONS SUMMARIZED IN TABLE
  WITH SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT 8 IOO/KW ASSU
2. THE UPPER FIGURES INDICATE THE RANGE OF ECONOMICALLY OPTIMUM ITD
  VALUES FOUND FOR THE 15 COMBINATIONS OF FUEL COSTS AND FIXED CHARGE
  RATES WHICH WERE ANALYZED  FOR EACH SITE
3. THE LOWER FIGURES, IN PARENTHESES,  INDICATE THE RANGE OF ITD VALUES
  FOR WHICH THE TOTAL ANNUAL COST OF PLANT OPERATION IS WITHIN 0.01
  OF THE  COST AT THE OPT I MUM POINT
(I) NOT OPTIMIZED  THE LOWEST COST  WAS  FOUND AT THE HIGHEST ITD  VALUE  INVESTIGATED, 80 °F

                                   FIGURE 44— ECONOMICALLY OPTIMUM VALUES OF INITIAL
                                     TEMPERATURE DIFFERENCE (°F>— NUCLEAR-FUELED
                                        GENERATING UNIT— NATURAL-DRAFT  TOWER

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                                                                OPTIMUM ITD-NUCLEAR  FUEL-MECHANICAL-DRAFT
Oi
o-
                    LEGEND:
                    I.  BASED ON THE SITE DATA AND STUDY ASSUMPTIONS SUMMARIZED IN TABLE 91
                      WITH SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT 9 IOO/KW ASS
                    2. THE UPPER FIGURES INDICATE THE RANGE OF ECONOMICALLY OPTIMUM ITD
                      VALUES FOUND FOR THE 15 COMBINATIONS OF FUEL COSTS AND FIXED CHARGE
                      RATES WHICH WERE ANALYZED FOR EACH SITE
                    3. THE LOWER FIGURES, IN PARENTHESES, INDICATE THE RANGE OF ITD VALUES
                      FOR WHICH THE TOTAL ANNUAL COST OF PLANT OPERATION IS WITHIN 0.01 MILLS/KWH
                      OF THE COST AT THE OPTIMUM POINT
                    (I) NOT OPTIMIZED. THE LOWEST COST WAS FOUND AT THE HIGHEST ITD VALUE INVESTIGATED ,80 °F
                                                               FIGURE 45—ECONOMICALLY OPTIMUM VALUES OF INITIAL
                                                                  TEMPERATURE DIFFERENCE (°F)—NUCLEAR-FUELED
                                                                   GENERATING UNIT—MECHANICAL-DRAFT TOWER

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rized in Figures 42 through 45 are also summarized on United States base maps, and
these figures are designated as Figures 46 through 49.

       To illustrate, Figure 46 shows that for Chicago the loss of capacity for the
range of economically optimum ITD values when using a natural-draft cooling tower
with a fossil-fueled generating unit would be on the  order of 5.0 to 8.5 percent of
the rated generating capacity.

        Figures 50 through 53 summarize the capital cost of the dry cooling system
for the range of the economically optimum  ITD values. Again to illustrate,  Figure
50 shows this range  of capital cost, when using a natural-draft cooling tower with a
fossil-fueled generating unit, for Chicago to be  $17.6 to $22.2  per kw.

        Figures 54 through 57 indicate the sum of the total capital cost of the dry
cooling system and the capital cost of the  required peaking capacity.  The peaking
capacity cost is applied to the total capacity of  the 800,000-kw plant in order to
determine the  penalty per kw.  The capital cost  of the peaking capacity has been
evaluated at $100 per kw of peaking capacity required.  Figure  54  therefore shows
that the combined cost of the dry cooling system and peaking capacity is $26.1  to
$27.3 per kw for  Chicago when using a natural-draft cooling tower with a fossil-
fueled generating unit.

        Much of the information shown on  the United States  base maps has also been
summarized  in Tables 11  through 22, with  all dollar values per kw rounded to the
nearest whole  dollar. The information has been  tabulated by fixed-charge rate in
order to illustrate the effect of the fixed-charge rate on the economic optimization.

        The  economically optimum values of ITD are shown in Tables 11 through 14.
For each site, 3 fossil-fuel costs and 3 nuclear-fuel costs were assumed. In  many
cases, for a given fixed-charge rate, the fuel  cost variation did not have sufficient
effect on the economic optimization to change the optimum  ITD  by a  full degree F.
In some cases, however, the  fuel cost did  affect the optimum and this is indicated
in the tables.  For example,  as shown in Table 11,  the value of  the economically
optimum ITD at Seattle for a  10 percent fixed-charge rate varied from 57° to 58°F
for the range of fuel costs analyzed when using a natural-draft cooling tower with
a fossil-fueled generating unit.

        The  capital  cost of the dry cooling system is  tabulated in Tables 15 through
18 for the range of  optimum ITD values.

        Tables 19 through 22  show the combined capital cost of the  dry cooling sys-
tem and the required peaking capacity.
                                     157

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Oi
00
                                                         % CAPACITY  LOSS-FOSSIL FUEL- NATURAL-DRAFT
                  LEGEND:

                  I.  THE LOSSES SHOWN ARE THOSE WHICH WOULD OCCUR AT THE AMBIENT
                     AIR TEMPERATURE EQUALLED OR EXCEEDED 10 HOURS PER YEAR

                  2.  SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT 8 100/KW ASSUMED

                  '*  °TDT»HE L°WEST C°ST *AS FOUND AT THE HIGHEST ITD VALUE
                                                    FIGURE 46- GENERATING CAPACITY LOSSES AS PERCENT OF RATED LOAD
                                                    FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD  SHOWN ON

                                                    FIGURE 42-FOSSIL-FUELED GENERATING UNIT-NATURAL-DRAFT TOWER

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                                      %  CAPACITY LOSS- FOSSIL FUEL-  MECHANICAL-DRAFT
LEGEND:
I   THE LOSSES SHOWN ARE THOSE WHICH WOULD OCCUR AT THE AMBIENT
   AIR TEMPERATURE EQUALLED OR EXCEEDED 10 HOURS PER YEAR
2  SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT « IOO/ KW ASSUMED
(I)  NOT OPTIMIZED. THE LOWEST COST WAS FOUND AT THE HIGHEST  D VALUE
   INVESTIGATED, 80 *f
                             FIGURE 47 —GENERATING CAPACITY  LOSSES AS PERCENT OF RATED LOAD FOR THE
                                 RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 43
                                    FOSSIL-FUELED GENERATING UNIT— MECHANICAL-DRAFT TOWER

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                                     %  CAPACITY  LOSS-NUCLEAR FUEL- NATURAL-DRAFT
LEGEND:
I.  THE LOSSES SHOWN ARE THOSE WHICH WOULD OCCUR AT THE AMRIFNT
  AIR TEMPERATURE EQUALLED OR EXCEEDED 10 HOURS PEF1 YEAR
2. SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT 8 100/KW ASSUMED
'             ™ L°WEST C°ST **S F°UND *T ™E HIGHEST ^D VALUE
                             FIGURE 48-GENERATING CAPACITY LOSSES AS PERCENT OF RATED LOAD FOR THE
                                ^fJSSS^S^SS^ VALUESOFI^ SHOWN ONDF"GURE 44
                                    NUCLEAR-FUELED  GENERATING UNIT- NATURAL-DRAFT TOWER

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                                      % CAPACITY LOSS-NUCLEAR  FUEL-MECHANICAL-DRAFT
LEGEND:
I.  THE LOSSES SHOWN ARE THOSE WHICH WOULD OCCUR AT THE AMBIENT
   AIR TEMPERATURE EQUALLED OR EXCEEDED 10 HOURS PER YEAR
2.  SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT 8 IQO/KW ASSUMED
(I)  NOT OPTIMIZED.  THE LOWEST COST WAS FOUND AT THE  HIGHEST ITD VALUE
   INVESTIGATED, 80°F
                           FIGURE 49 —GENERATING CAPACITY LOSSES AS PERCENT OF RATED LOAD FOR THE
                               RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 45
                                 NUCLEAR-FUELED GENERATING  UNIT—MECHANICAL-DRAFT TOWER

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                                           CAPITAL COST- FOSSIL FUEL- NATURAL-DRAFT
LEGEND:

I.  INCLUDES CAPITAL COSTS  OF THE CONDENSER; COOLING
   SYSTEM PIPING, PUMPS, VALVES AND CONTROLS, AND
   THE COOLING TOWER
2.  SUMMER PEAKSIEXCEPT ANCHORAGE) AND PEAKING CAPACITY
   AT 8 IOO/KW ASSUMED
(I)  NOT OPTIMIZED. THE LOWEST COST WAS FOUND AT THE
   HIGHEST ITD VALUE INVESTIGATED, 80 • F
                                    FIGURE 50—CAPITAL COST OF THE DRY COOLI N6 SYSTEM («/KW)
                                FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 42
                                       FOSSIL-FUELED GENERATING UNIT— NATURAL-DRAFT TOWER

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                                           CAPITAL COST-  FOSSIL  FUEL- MECHANICAL-DRAFT
LEGEND:
I.  INCLUDES  CAPITAL COSTS OF THE CONDENSER; COOLING
   SYSTEM PIPING, PUMPS .VALVES AND CONTROLS; AND
   THE COOLING TOWER
2.  SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY
   AT * IOO/KW ASSUMED
(I)  NOT OPTIMIZED.  THE  LOWEST COST  WAS FOUND AT THE
   HIGHEST ITD VALUE INVESTIGATED, 80° F
                                     FIGURE 51 —CAPITAL COST OF THE DRY COOLING SYSTEM (8/KW)
                                FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 43
                                      FOSSIL-FUELED GENERATING UNIT— MECHANICAL- DRAFT TOWER

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                                          CAPITAL  COST-NUCLEAR  FUEL-NATURAL-DRAFT
LEGEND:
I.  INCLUDES CAPITAL COSTS OF THE CONDENSER; COOLING
   SYSTEM PIPING, PUMPS, VALVES AND CONTROLS; AND
   THE COOLING TOWER
Z.  SLIMMER PEAKSIEXCEPT ANCHORAGE) AND PEAKING CAPACITY
   AT » 100/KW ASSUMED
(I)  NOT OPTIMIZED. THE LOWEST COST  WAS FOUND AT THE
   HIGHEST ITD VALUE INVESTIGATED, 80' F
                                  FIGURE  52 —CAPITAL COST OFTHE DRY COOLING SYSTEM(&/KW)
                             FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 44
                                   NUCLEAR-FUELED GENERATING UNIT— NATURAL-DRAFT TOWER

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                                          CAPITAL COST-NUCLEAR  FUEL-MECHANICAL-DRAFT
                                                                                             _ — —.-»	r
                                                                                       _-r~. NASHVILLE  -J     C»BOLI»»
LEGEND:
I. INCLUDES CAPITAL COSTS OF THE CONDENSER; COOLING
  SYSTEM PIPING, PUMPS, VALVES AND CONTROLS; AND
  THE COOLING TOWER
2. SUMMER PEAKSfEXCEPT ANCHORAGE) AND PEAKING CAPACITY
  AT 9 100/KW ASSUMED
(I) NOT OPTIMIZED. THE LOWEST COST  WAS FOUND AT THE
  HIGHEST ITD VALUE INVESTIGATED, 80° F
                                  FIGURE  53—CAPITAL COST OF THE DRY COOLING SYSTEM (8/KW)
                             FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 45
                                  NUCLEAR-FUELED GENERATING UNIT- MECHANICAL-DRAFT TOWER

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                                  COOLING AND PEAKING CAPITAL COST-FOSSIL FUEL-NATURAL-DRAFT
LEGEND:

I.  SUMMER PEAKSIEXCEPT ANCHOR AGE) AND PEAKING CAPACITY
   AT 8|00/KW ASSUMED
(I)  COOLING SYSTEM COST ONLY SINCE NO LOSS OF CAPACITY
   OCCURS AT THE TIME OF SYSTEM PEAK
                              FIGURE 54—CAPITAL COST OF THE DRY COOLING SYSTEM(8/KW)
                                  PLUS  CAPITAL COST OF PEAKING CAPACITY  (B/KW)
                         FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 42
                                FOSSIL-FUELED GENERATING UNIT— NATURAL-DRAFT TOWER

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                                 DOLING AND PEAKING CAPITAL COST-FOSSIL  FUEL-MECHANICAL-DRAFT
 LEGEND:
 I.  SUMMER PEAKSIEXCEPT ANCHORAGE) AND PEAKING CAPACITY
   AT tlOO/KW ASSUMED
(I)  COOLING SYSTEM COST ONLY SINCE NO LOSS OF CAPACITY
   OCCURS AT THE TIME OF SYSTEM PEAK
                               FIGURE  55— CAPITAL COST OF THE DRY COOLING SYSTEM (tt/KW)
                                   PLUS CAPITAL COST OF PEAKING CAPACITY  (I/KW)
                           FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 43
                                 FOSSIL-FUELED GENERATING UNIT—MECHANICAL-DRAFT TOWER

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                                COOLING AND PEAKING CAPITAL COST-NUCLEAR FUEL-NATURAL-DRAFT
LEGEND:
I.  SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY
   AT BlOO/KW ASSUMED
(I)  COOLING SYSTEM COST ONLY SINCE NO LOSS OF CAPACITY
   OCCURS AT THE TIME OF SYSTEM PEAK
                                FIGURE 56— CAPITAL COST OF THE DRY COOLING SYSTEM18/KW)
                                    PLUS CAPITAL COST OF PEAKING CAPACITY («/KW)
                            FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 44
                                 NUCLEAR-FUELED GENERATING UNIT—NATURAL-DRAFT TOWER

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                                        AND PEAKING CAPITAL COST-NUCLEAR FUEL-MECHANICAL-DRAFT
 LEGEND:
 I.  SUMMER PEAKSIEXCEPT ANCHORAGE) AND PEAKING CAPACITY
   AT llOO/KW ASSUMED
(I)  COOLING SYSTEM COST ONLY SINCE NO LOSS OF CAPACITY
   OCCURS AT THE TIME OF SYSTEM PEAK
                              FIGURE 57 —CAPITAL COST OF THE DRY COOLING SYSTEM (J/KW)
                                  PLUS CAPITAL COST OF PEAKING CAPACITY (8/KW)
                         FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 45
                              NUCLEAR-FUELED  GENERATING UNIT—MECHANICAL-DRAFT TOWER

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                                   TABLE 11

        Economically Optimum Values of Initial Temperature Difference (°F)
               Fossil-Fueled Generating Unit, Natural-Draft Tower


                              	Initial Temperature Difference (°F)	
          Fixed-Charge Rate:     8%      10%      12%      15%      i~8%"

 PLANT SITE

     Seattle,  Wash.              57      57-58      58        58       58
     San Francisco, Calif.         57      57-58     57-58      58       58
     Los Angeles,  Calif.          55      55-56      56        56      56-57
     Great Falls, Mont.         57-58      58       58        59      59-61
     Boise, Ida.                  56      56-57     56-57      57       58
     Casper, Wyo.             58-59     58-62     58-62    59-62     59-62
     Reno, Nev.                 57       57       58        58       58
     Denver,  Colo.               56       57       57        58       58
     Phoenix, Ariz.             47-48      48       49      51-52     52-53
     Bismarck, N. Dak.           56       56       57        57       57
     Minneapolis,  Minn.          55       56       56        57       57
     Omaha,  Neb.               54       55       55      55-56      56
     Little Rock, Ark.           49-52     51-53     51-53    52-54     53-54
     Midland, Tex.             52-54     53-55     54-55    54-56     55-56
     New Orleans, La.          51-53     52-54     53-55    53-55     54-55
     Green Bay, Wis.             57       57       57      57-58      58
     Grand Rapids, Mich.       55-56      56       56        57       57
     Detroit, Mich.             55-56      56       56        57       57
     Chicago, III.               55-56      56       56        57       57
     Nashville,  Tenn.             52      52-53      53        54      54-55
     Burlington, Vt.              56       56      57-58      57       57
     Philadelphia,  Penna.         54      54-55     55-56    55-56      56
     Charleston, W. Va.          55       55      55-56      56      56-57
    Atlanta,  Ga.                 52       53       54      54-55      55
    Miami, Fla.               48-49     49-51      51        52       53
     Honolulu, Hawaii             49      51-52     53-54      53      53-54
    Anchorage, Alas.  (1)         80       80       80        80       80


     Note:  Based  upon the site data and study assumptions summarized in Table 10
           with summer  peaks  (except Anchorage), peaking capacity at $100Aw
           and pump replacement capacity at $150/kw.


(1)   Not optimized — winter peak assumed.
                                   170

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                                  TABLE 12

       Economically Optimum Values of Initial Temperature Difference (°F)
             Fossil-Fueled Generating Unit, Mechanical-Draft Tower
          Fixed-Charge Rate:
                                     Initial Temperature Difference (°F)
PLANT SITE
    Seattle, Wash.
    San Francisco, Calif.
    Los Angeles, Calif.
    Great Falls, Mont.
    Boise, Ida.
    Casper, Wyo.
    Reno, Nev.
    Denver, Colo.
    Phoenix, Ariz.
    Bismarck,  N. Dak.
    Minneapolis, Minn.
    Omaha, Neb.
    Little Rock, Ark.
    Midland, Tex.
    New Orleans, La.
    Green Bay, Wis.
    Grand Rapids, Mich.
    Detroit, Mich.
    Chicago, III.
    Nashville, Tenn .
    Burlington, Vt.
    Philadelphia, Penna.
    Charleston, W. Va.
    Atlanta, Ga.
    Miami, Fla.
    Honolulu,  Hawaii
    Anchorage, Alas.  (1)
8%
62-63
62
59
60-61
58
60-62
59
58
47
59
59
55
51-53
52-55
52-55
60
59
59
59
51-52
60
56
56
53
50
51
79-80
10%
63
63
60
61
59
60-62
60
59
49
60
60
56
52-54
53-56
53-56
61
60
60
60
53
60
57-58
58
55
51
52-53
80
12%
64
63
60
62
60
61-63
60
60
50
60
60
57
53-55
54-57
54-58
61-62
60
60
60
54
61
59
59
55-56
51-52
53-54
80
15%
64
64
60
63
60
62-63
61
60
51
60
60-61
58
54-56
55-58
56-59
62
60-61
61
61
55
61
60
60
56-57
53
55
80
18%
64-65
64-65
61
63
60
62-63
61
61
52
61
61
59
55-57
61-63
57-60
63
61
61
61
56
62
60
60
57-58
54
56
80
     Note:  Based upon the site data and study assumptions summarized in Table 10
            with summer peaks  (except Anchorage),  peaking capacity at $100/kw;
            fan and pump replacement capacity at $150/kw-
 (1)  Not optimized — winter peak assumed.
                                    171

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                                   TABLE 13

        Economically Optimum Values of Initial Temperature Difference (°F)
              Nuclear-Fueled Generating Unit, Natural-Draft Tower
                                      Initial Temperature Difference (°F)
                                 8%
10%
12%
15%
18%
          Fixed-Charge Rate:

PUNT SITE

    Seattle, Wash.
    San Francisco, Calif.
    Los Angeles, Calif.
    Great Falls, Mont.
    Boise,  Ida.
    Casper, Wyo.
    Reno, Nev.
    Denver, Colo.
    Phoenix, Ariz.
    Bismarck,  N. Dak.
    Minneapolis, Minn.
    Omaha, Neb.
    Little Rock, Ark.
    Midland, Tex.
    New  Orleans, La.
    Green Bay, Wis.
    Grand Rapids, Mich.
    Detroit, Mich.
    Chicago, III .
    Nashville, Tenn.
    Burlington, Vt.
    Philadelphia, Penna.
    Charleston, W. Va.
    Atlanta, Ga.
    Miami, Fla.
    Honolulu,  Hawaii
    Anchorage, Alas.  (1)
    Note:  Based upon the site data and study assumptions summarized in Table 10
           with summer peaks (except Anchorage), peaking capacity at $1 00/kw
           and pump replacement capacity at $225/kw.


(1)  Not optimized — winter peak assumed.
59-65
58-59
56
63-65
58-62
65-68
63-64
57-59
52-53
61-62
57-58
56-57
53-56
55-58
52-55
58
57-58
57-59
57-58
55
58
56
55-56
54-55
49
50-51
80
65
59-63
57
65
64
66-69
64-65
61-62
54-55
63-64
61
57-58
55-56
56-58
53-56
58-59
59-61
61
59-61
55-56
58
56
56-57
55
51
51-52
80
65
64-65
57-58
66
64-65
66-70
65-66
63
55
64-65
62
58
55-57
57-64
54-56
65
61-62
62
61-62
56
58
57-58
58
56
52-53
52-53
80
65-66
65
58
66
65
68-70
67-68
65-66
56
65
63-65
63-64
56-58
58-65
55-61
65
63-65
63-65
65
56-57
65
61-62
61
56-57
53
53-54
80
66
65
58-59
67
67
69-70
68-69
66
57-58
65-67
65-66
64-65
57-58
58-66
56-62
66
65-66
66
66
57
65
62-63
62-63
57-58
54
54-55
80
                                    172

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                                  TABLE 14

       Economically Optimum Values of Initial Temperature  Difference (°F)
           Nuclear-Fueled Generating Unit, Mechanical-Draft Tower
          Fixed-Charge Rate:
                                     Initial Temperature Difference (°F)
PLANT SITE
    Seattle, Wash.
    San Francisco, Calif.
    Los Angeles, Calif.
    Great Falls, Mont.
    Boise, Ida.
    Casper, Wyo.
    Reno, Nev.
    Denver, Colo.
    Phoenix, Ariz.
    Bismarck, N. Dak.
    Minneapolis, Minn.
    Omaha, Neb.
    Little Rock, Ark.
    Midland,  Tex.
    New Orleans, La.
    Green Bay, Wis.
    Grand Rapids, Mich.
    Detroit, Mich.
    Chicago,  III.
    Nashville, Tenn.
    Burlington, Vt.
    Philadelphia, Penna.
    Charleston, W.  Va.
    Atlanta, Ga.
    Miami, Fla.
    Honolulu, Hawaii
    Anchorage, Alas. (1)
8%
65
64
58-59
65
61
64-69
63
60-61
51
62
61
58
51-57
56-60
52-55
61
60-61
61
60
55
59
60
56-59
55
48
50
80
10%
66
65
60
65-66
64
66-69
64
61
54
64
61
61
56-69
57-63
53-60
65
61
61-62
61
56
61
61
60
56
50
51
80
12%
66
65
60
66
64
66-69
65-66
62
55
64-65
63
62
56-60
59-64
54-61
65
63
63
65
57
65
61
61
57-60
51
51-52
80
15%
67
66
61
67
65
68-70
68
65
58-60
67
66
64
58-64
60-65
60-62
66
66
66
66
59-60
65
63
62
60-61
53
53
80
18%
67
66
61-62
69
67
69-70
69
66
60-61
68
66
65
60-64
63-66
60-63
66
66
66-67
66-67
61
66
64
65
61
53
54-56
80
     Note:  Based upon the site data and study assumptions summarized in Table 10
            with summer peaks (except Anchorage), peaking capacity at $1 OOAW;
            fan and pump replacement capacity at $225/kw.

 (1)  Not optimized — winter peak assumed.
                                     173

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                                  TABLE 15

                 Capital Cost of the Dry Cooling System ($/Kw)
                     for the Economically Optimum Values
               of Initial Temperature Difference Shown in Table 11
              Fossil-Fueled Generating Unit, Natural-Draft Tower
                                                 $/K
w
          Fixed-Charge Rate:     8%      10%      12%      15%      18%

PLANT SITE

    Seattle, Wash.              19      18-19      18       18        18
    San Francisco, Calif.        19       19        19       19        19
    Los Angeles, Calif.          19       19        19       19      18-19
    Great Falls, Mont.         19-20     19        19       19      18-19
    Boise, Ida.                 20       20        20       20        19
    Casper,  Wyo.               19      18-19    18-19     18-19    17-19
    Reno, Nev.                 21       21        21        21       21
    Denver, Colo.              20       19        19       19        19
    Phoenix, Ariz.              23       23        22       21      20-21
    Bismarck,  N. Dak.          19       19        19       19        19
    Minneapolis, Minn.          19       19        19       19        19
    Omaha, Neb.               20       19        19       19        19
    Little Rock, Ark.           19-20    18-19    18-19      18        18
    Midland, Tex.             20-21     19-20    19-20     19-20      19
    New Orleans, La.          19-20    19-20      19       19        19
    Green Bay, Wis.             19       19        19      18-19      18
    Grand Rapids, Mich.         19       19        19       19        19
    Detroit,  Mich.              20       20        20       20       20
    Chicago, III.                20       20        20       20       20
    Nashville, Tenn.            19      18-19      18       18      17-18
    Burlington, Vt.              18       18        18       18       18
    Philadelphia,  Penna.         20      19-20      19       19       19
    Charleston, W. Va.          19       19        19       19       19
    Atlanta, Ga.                20       19        19      18-19     18
    Miami, Fla.               22-23    21-22      21        21       20
    Honolulu,  Hawaii            24       23        23       22       22
    Anchorage, Alas.  (1)        19       19        19       19       19

    Note:  Based upon the site data and study assumptions summarized in Table 10
          with summer peaks  (except Anchorage), peaking capacity at $100/kw
          and pump replacement capacity at $150/kw.

(1)  Not optimized — winter peak assumed.
                                  174

-------
                                 TABLE 16

                Capital Cost of the Dry Cooling System ($/Kw)
                    for the Economically Optimum Values
              of Initial Temperature Difference Shown in Table 12
            Fossil-Fueled Generating Unit, Mechanical-Draft Tower

                             	$/Kw	
         Fixed-Charge Rate:     8%      10%     12%     15%""     18%

PLANT SITE

    Seattle, Wash.               16       16       15        15       15
    San Francisco, Calif.         17       16       16        16       16
    Los Angeles,  Calif.          17       17       17        17       16
    Great Falls,  Mont.           17       17       16        16       16
    Boise, Ida.                  18       17       17        17       17
    Casper, Wyo.               16       16       16      15-16     15-16
    Reno, Nev.                 18       18       18        18       18
    Denver, Colo.               17       17       16        16       16
    Phoenix, Ariz.               22       21       21         20       20
    Bismarck, N. Dak.           17       17       17        17       16
    Minneapolis, Minn.          17       17       17        17       16
    Omaha, Neb.               19       18       18        17       17
    Little Rock, Ark.           17-18     17-18      17      16-17     16-17
    Midland, Tex.              18-19     18-19      18      17-18     17-18
    New Orleans, La.          18-19      18      17-18    16-18     16-17
    Green Bay, Wis.             17       16       16        16       16
    Grand Rapids, Mich.         17       17       17      16-17      16
    Detroit, Mich.               18       18       18        17       17
    Chicago, III.               18       18       18        17       17
    Nashville, Tenn.            18       17       17        17       16
    Burlington, Vt.              16       16       15        15       15
    Philadelphia, Penna.         18      17-18      17        17       17
    Charleston, W. Va.          18       17       17        17       17
    Atlanta, Ga.               18       18      17-18      17       17
    Miami, Fla.                 21        20       20        19       19
    Honolulu,  Hawaii            22      21-22      21        20       20
    Anchorage, Alas.  (1)         18       18       18        18       18

    Note:  Based upon the site data and study assumptions summarized in Table 10
           with  summer peaks (except Anchorage), peaking capacity at $100/kw;
           fan and pump replacement capacity at $150/kw.

(1)  Not optimized — winter peak assumed.
                                  175

-------
                                  TABLE 17

                 Capital Cost of the Dry Cooling System ($/Kw)
                     for the Economically Optimum Values
               of Initial Temperature Difference Shown in Table 13
             Nuclear-Fueled Generating Unit, Natural-Draft Tower
                                                  $/K
w
          Fixed-Charge Rate:     8%      10%      12%      15%     T8%"

PLANT SITE

    Seattle, Wash.             24-25     24       24       24        24
    San Francisco, Calif.        29      27-28      26       26        26
    Los Angeles, Calif.          29       28       28       28      27-28
    Great Falls, Mont.         25-26     25       25       25        24
    Boise,  Ida.                27-28     26      25-26    22-24    22-23
    Casper, Wyo.              23-26    23-24     23-24    22-24    22-23
    Reno, Nev.                 28      27-28     26-27     26        25
    Denver, Colo.              27      26-27      26      24-25      24
    Phoenix, Ariz.              32      30-31      30       29      28-29
    Bismarck,  N. Dak.         26-27    25-26      25       25      24-25
    Minneapolis, Minn.         28-29     27       26       25        24
    Omaha, Neb.               29      28-29      27       25        25
    Little Rock, Ark.           26-27    26-27     26-27    25-26    25-26
    Midland, Tex.             28-29    28-29     24-28    24-28    23-26
    New Orleans, La.          29-31     28-30     28-29    26-29    25-28
    Green Bay, Wis.            28       28       24       24        24
    Grand Rapids, Mich.        28-29    27-28     26-27    24-25      24
    Detroit, Mich.             29-30     28       27      26-27      25
    Chicago, III.              29-30     28       27       26      25-26
    Nashville, Tenn.            27      26-27      26       26      24-26
    Burlington, Vt.              27       27      23-27     23        23
    Philadelphia, Penna.         29       29      26-28     26        26
    Charleston, W. Va.         29-30     29      28-29     27      25-26
    Atlanta, Ga.               29       29       28       27        25
    Miami, Fla.                 35       33      32-33     32      29-30
    Honolulu,  Hawaii           36-37    35-36     34-35    33-34      33
    Anchorage, Alas.  (1)       30-31      30       30       30        30

    Note:  Based upon the site data and study assumptions summarized in Table 10
           with summer peaks  (except Anchorage), peaking capacity at $100/kw
           and pump replacement capacity at $225/kw.

(1)  Not optimized — winter peak assumed.
                                  176

-------
                                 TABLE 18

                Capital Cost of the Dry Cooling System ($/Kw)
                    for the Economically Optimum Values
              of Initial Temperature Difference Shown in Table 14
           Nuclear-Fueled Generating Unit, Mechanical-Draft Tower
                                                 $/K
.w
          Fixed-Charge Rate:     8%      10%     12%      15%      18%

PLANT SITE

    Seattle, Wash.              22        22        22      21-22      21
    San Francisco, Calif.        24        23        23        23       23
    Los Angeles, Calif.         25-26      25        25        24       24
    Great Falls, Mont.          23      22-23      22        22      21-22
    Boise, Ida.                 23        23        23        23       21
    Casper, Wyo.              20-22    20-22     20-22    20-21      20
    Reno, Nev.                25        25        24        23       22
    Denver, Colo.              24        24        23        22       22
    Phoenix, Ariz.              30        28        27      25-26    24-25
    Bismarck,  N. Dak.          24        23       22-23      22       21
    Minneapolis, Minn.         24        24        23        22       22
    Omaha, Neb.              26        24        24        23       22
    Little Rock, Ark.           24-27    23-24     22-24    20-23    20-22
    Midland, Tex.             24-26    22-25     22-24    22-24    21-22
    New Orleans, La.          26-27    23-27     23-26    22-23    22-23
    Green Bay, Wis.            24        22        22        22       22
    Grand Rapids, Mich.       24-25      24       23        22       22
    Detroit, Mich.              25        25       24        23       23
    Chicago, III.               26        25      23-24      23       23
    Nashville, Tenn.            25        24       24      22-23     22
    Burlington, Vt.             24        23       21        21       21
    Philadelphia, Penna.        25        24       24        23       23
    Charleston, W. Va.        25-27      25       24        24       22
    Atlanta, Ga.               26       26      23-25     23-24     23
    Miami, Fla.                31        30       30        28       28
    Honolulu,  Hawaii           33       32       32        31       29-31
    Anchorage, Alas.  (1)        27       27       27        27        27

    Note:  Based upon the site data and study assumptions summarized in Table 10
            with summer peaks (except Anchorage), peaking capacity at $100/kw;
            fan and pump replacement capacity at  $225/kw.

 (1) Not optimized — winter peak assumed.
                                    177

-------
                                TABLE 19

              Capital Cost of the Dry Cooling System ($/Kw) Plus
 Capital Cost of Peaking Capacity ($/Kw) for the Economically Optimum Values
             of Initial Temperature Difference Shown in Table 11
             Fossil-Fueled Generating Unit, Natural-Draft Tower
                                                $/Kw
         Fixed-Charge Rate:     "S%~"      10%

PLANT SITE

    Seattle, Wash.               24         24      24         24        24
    San Francisco, Calif.         24         24      24         24        24
    Los Angeles, Calif.          25       24-25     24         24        24
    Great Falls, Mont.           26         26      26         26      25-26
    Boise,  Ida.                  28       27-28     28         28        28
    Casper, Wyo.                26         26      26         26        26
    Reno, Nev.                 29         29      29         29        29
    Denver, Colo.               27         26      26         26        26
    Phoenix, Ariz.               32         32      32       31-32       31
    Bismarck,  N. Dak.           27         27      27         27        27
    Minneapolis, Minn.          26         26      26         26        26
    Omaha, Neb.                28         28      28         28        28
    Little Rock, Ark.           26-27        26      26         26        26
    Midland, Tex.               28         28      28         28        28
    New Orleans, La.          25-26        25      25         25        25
    Green Bay, Wis.             24         24      24         24        24
    Grand Rapids, Mich.       25-26        25      25         25        25
    Detroit, Mich.               27         27      27         26        26
    Chicago, III.                26         26      26         26        26
    Nashville,  Tenn.             26         26      26         26      25-26
    Burlington, Vt.               23         23      23         23        23
    Philadelphia, Penna.         26         26      26         26        26
    Charleston, W.  Va.          26         26    25-26        25        25
    Atlanta, Ga.                26         26      26         26        26
    Miami, Fla.                 27         27    26-27        26        26
    Honolulu,  Hawaii            28         27      27         26        26
    Anchorage, Alas. (1)         19         19      19         19        19

    Note:  Based upon the site data and study assumptions summarized in Table 10
           with summer peaks (except Anchorage), peaking capacity at $100/kw
           and pump replacement capacity at $150/kw.

(1)  Not optimized— winter peak assumed.
                                    178

-------
                               TABLE 20

             Capital Cost of the Dry Cooling System ''S/Kw) Plus
 Capital Cost of Peaking Capacity f$/Kw) for the Economically Optimum Values
              of Initial  Temperature Difference Shown in Table 12
            Fossil-Fueled Generating Unit, Mechanical-Draft Tower
                                                 $/K
w
         Fixed-C^rge Rate:     8%        10%     12%       15%      18%

PLANT SITE

    Seattle, Wash.              23         23      23         23        23
    San Francisco, Calif.        23         23      23         23        23
    Los Angeles,  Calif.         23         23      23         23        23
    Great Falls, Mont.          24         24      24         24        24
    Boise, Ida.                 26         26      26         26        26
    Casper,  Wyo.              24         24      24         24        24
    Reno, Nev.                27         27      27         27        27
    Denver, Colo.              25         24      24         24        24
    Phoenix, Ariz.              31         31      31         31        31
    Bismarck,  N. Dak,          26         26      26         26        26
    Minneapolis,  Minn.         25         25      25        24-25      24
    Omaha, Neb.              27         27      27         27        27
    Little Rock, Ark.            26        25-26   25-26       25        25
    Midland, Tex.              27         27      27         27        27
    New Orleans, La.          24         24      24          24        24
    Green Bay, Wis.            23         23      23          23        23
    Grand Rapids, Mich.        24         24      24          24        24
    Detroit, Mich.              25         25       25          25        25
    Chicago,  III.              25        25       25          25        25
    Nashville, Tenn.            25        25       25          25       25
    Burlington, Vt.             22        22       22          22       22
    Philadelphia, Penna.        25        25       25         25       25
    Charleston, W. Va.         25        25       25         25       25
    Atlanta, Go.              25        25       25         25       25
    Miami, Fla.                25        25       25         25       25
     Honolulu,  Hawaii          26        26     25-26        25       25
    Anchorage, Alas. (1)         18         18       18         18        18

     Note: Based upon the site data and study assumptions summarized in Table 10
           with summer peaks (except Anchorage), peaking capacity at $100/kw
           fan and pump replacement capacity at $150/kw.

 (1) Not optimized — winter peak assumed.
                                     179

-------
                                TABLE 21

              Capital Cost of the Dry Cooling System ($/Kw) Plus
  Capital Cost of Peaking Capacity ($/Kw) for the Economically Optimum Values
              of Initial Temperature Difference Shown in Table 13
             Nuclear-Fueled Generating Unit, Natural-Draft Tower

                              	        $/Kw	
          Fixed-Charge Rate-   ~8%"     T5%"    ~J2%      T5%"     Tg%

PLANT SITE

    Seattle, Wash.              36         36      36         36        36
    San Francisco, Calif.       37-38       37      37         37        37
    Los Angeles, Calif.         38         38      38         38        38
    Great Falls, Mont.        38-39       38      38         38        38
    Boise,  Ida.                41-42       41      41          41        41
    Casper, Wyo.               38         38      38         38        38
    Reno, Nev.                42-43       42      42         42      41-42
    Denver, Colo.              40        39-40    39         39        39
    Phoenix, ArizV             47-48      46-47    46         46      45-46
    Bismarck,  N. Dak.          40         40      40         40      39-40
    Minneapolis, Minn.        39-40       39      39       38-39      38
    Omaha, Neb.               42         42      42         41        41
    Little Rock, Ark.           40-41        40      40       39-40      39
    Midland, Tex.             42-43       42     41-42      40-42    40-42
    New Orleans, La.          39-40      39-40    39       38-39    38-39
    Green Bay, Wis.            37         37      36         36        36
    Grand  Rapids, Mich.        39         39      39         38        38
    Detroit, Mich.              41         40      40       39-40      39
    Chicago, III.              40-41        40      49         39        39
    Nashville, Tenn.            39         39      39         39        39
    Burlington, Vt.              36         36      36         35        35
    Philadelphia, Penna.        40         40      40         39        39
    Charleston, W.  Va.         40         40      39         39      38-39
    Atlanta, Ga.               40         40      40         40        39
    Miami, Fla.                43         42      41          41        41
    Honolulu, Hawaii            43        42-43   41-42        41        41
    Anchorage, Alas.  (1)       30-31        30      30         30        30

    Note   Based upon the site data and study assumptions summarized in Table 10
           with summer peaks (except Anchorage), peaking capacity at  $100/kw
           and pump replacement capacity at $225/kw.

(1)  Not optimized — winter peak assumed.
                                    180

-------
                                TABLE 22

              Capital Cost of the Dry Cooling System ($/Kw) Plus
  Capital Cost of Peaking Capacity ($/Kw) for the Economically Optimum Values
              of Initial Temperature Difference Shown in Table 14
           Nuclear-Fueled Generating Unit, Mechanical-Draft Tower
                                                 $/K
w
         Fixed-Charge Rate:     8%        10%     12%      T5%^8%

PLANT SITE

    Seattle, Wash.              34         34       34         34       34
    San Francisco, Calif.        35         34       34         34       34
    Los Angeles, Calif.          35         35       35         35       35
    Great Falls, Mont.          36         36       36         36       35
    Boise, Ida.                 39         38       38         38       38
    Casper,  Wyo.               36         36       36         36       36
    Reno, Nev.                40         39       39         39       39
    Denver, Colo.              36         36       36         36       36
    Phoenix, Ariz.              45         44       44         43       43
    Bismarck, N. Dak.          38         38       38         37       37
    Minneapolis, Minn.         37         37       36         36       36
    Omaha,  Neb.              39         39       39         39       39
    Little Rock, Ark.           37-38      37-38     37         37       37
    Midland, Tex.              39         39     38-39      38-39     38-39
    New Orleans, La.          36        35-36   35-36        35       35
    Green Bay, Wis.            35         34       34         34       34
    Grand Rapids, Mich.        36         36       36         36       36
    Detroit,  Mich.             38         38       38         37       37
    Chicago, III.               38         37       37         37       37
    Nashville, Tenn.            37        37       37       36-37     36
    Burlington, Vt.             34        33       33         33       33
    Philadelphia, Penna.        37        37       37         37       37
    Charleston, W. Va.         37         37       36         36        36
    Atlanta, Ga.               38         37      37         37        37
    Miami, Fla.                39         38       38         38        38
    Honolulu,  Hawaii           39         39       39         38        38
    Anchorage, Alas. (1)        27         27      27         27        27

     Note: Based upon the site data and study assumptions summarized in Table 10
           with summer peaks (except Anchorage), peaking capacity at $1 00/kw;
           fan and pump replacement capacity at $225/kw.

 (1) Not optimized — winter peak assumed.
                                      181

-------
        Tables 23 through 26 show, for optimized installations, the annual costs of
the cooling system (including the condensers and all other equipment associated
with the cooling system), the capacity necessary to replace loss of turbine capacity
at high ambient temperatures, the cooling system auxiliary capacity requirements,
and the total plant fuel cost. The above  annual costs are presented in mills per kw
for the 3 ranges of fuel cost used for each location and  for the 5 fixed-charge rates
considered for the optimization program.   The annual plant costs for other parts of
the generating  plant, except for the total  plant fuel cost which was included above,
were not incorporated in the figures listed on Tables 23 through 26. The optimiza-
tion program evaluated only  those parameters affected  by the dry-type cooling
system.  It is possible that the cost of the  turbine will be affected to some degree
by the varying  conditions studied, but it is assumed that the net result will be no
increase in cost from  a present-day standard  design.  Some foreign firms claim a
reduction in turbine cost due to  design for operation at  high back pressures.

        Tables 27 through 30 show the auxiliary capacity requirements,  in mw, for
optimized dry-type  cooling system installations for the 3  ranges of fuel costs used
and for the 5 fixed-charge rates used in the computer analysis program.
                                    182

-------
                                                                                             TABLE 23
                                                           Optimized Total Annual Costs (in Mills per Kwh) Influenced by the Cooling System
                                                                               800-Mw, Fossil-Fueled Generating Unit
                                                                            Nature I-Draft,  Dry-Type Cooling Tower System
8
        Fixed-Charge Rate:

PLANT SITE

    Seattle, Wash.
    San Francisco, Calif.
    Los Angeles, Calif.
    Great  Falls,  Mont.
    Boise,  Ida.
    Casper, Wyo.
    Reno,  Nev.
    Denver, Colo.
    Phoenix, Ariz.
    Bismarck,  N. Dak.
    Minneapolis, Minn.
    Omaha, Neb.
    Little Rock, Ark.
    Midland, Tex.
    New Orleans, La.
    Green  Bay, Wis.
    Grand  Rapids, Mich.
    Detroit, Mich.
    Chicago, III.
    Nashville,  Tenn.
    Burlington, Vt.
    Philadelphia, Penna.
    Charleston, W. Va.
    Atlanta, Ga.
    Miami, Fla.
    Honolulu,  Hawaii
    Anchorage, Alas.
                                                      Low Fuel Cost Range
8%
2.64
2.65
2.66
1.76
2.26
1 .30
2.73
2.23
2.35
1.50
2.68
2.72
2.70
2.27
2.23
2.65
2.67
2.69
2.69
2.05
2.64
2.69
1.76
2.70
2.73
3.19
3.04
10%
2.71
2.73
2.74
1 .84
2.35
1 .38
2.82
2.32
2.46
1.58
2.76
2.80
2.79
2.36
2.31
2.73
2.76
2.78
2.77
2.13
2.71
2.78
1.85
2.78
2.82
3.28
3.11
12%
2.79
2.80
2.82
1 .92
2.44
1.46
2.91
2.40
2.56
1.67
2.84
2.89
2.87
2.45
2.39
2.80
2.84
2.86
2.86
2.22
2.78
2.86
1.93
2.87
2.90
3.36
3.17
15%
2.90
2.92
2.94
2.04
2.57
1 .59
3.05
2.53
2.71
1.80
2.96
3.03
2.99
2.58
2.51
2.92
2.96
2.99
2.98
2.34
2.90
2.98
2.05
2.99
3.03
3.49
3.26
18%
3.02
3.03
3.05
2.16
2.70
1.71
3.19
2.65
2.86
1.92
3.09
3.16
3.12
2.71
2.63
3.04
3.08
3.11
3.11
2.46
3.00
3.11
2.17
3.12
3.15
3.61
3.35
                                                                                     Medium Fuel Cost Range
8%
3.56
3.57
3.58
2.67
3.18
2.22
3.65
3.15
3.28
2.23
3.60
3.64
3.63
3.20
3.16
3.57
3.59
3.61
3.61
2.70
3.55
3.61
2.69
3.62
3.66
4.12
3.97
10%
3.63
3.64
3.66
2.76
3.27
2.30
3.74
3.24
3.38
2.32
3.68
3.73
3.71
3.29
3.24
3.65
3.67
3.70
3.69
2.78
3.63
3.70
2.77
3.71
3.74
4.20
4.03
12%
3.71
3.72
3.74
2.84
3.35
2.39
3.83
3.32
3.48
2.40
3.76
3.81
3.80
3.38
3.32
3.72
3.76
3.78
3.78
2.86
3.70
3.78
2.85
3.79
3.83
4.29
4.09
15%
3.82
3.84
3.86
2.96
3.49
2.51
3.97
3.45
3.64
2.53
3.88
3.95
3.92
3.51
3.44
3.84
3.88
3.91
3.90
2.98
3.81
3.90
2.97
3.91
3.96
4.42
4.18
18%
3.93
3.95
3.97
3.08
3.62
2.64
4.11
3.57
3.79
2.66
4.00
4.06
4.05
3.64
3.56
3.95
4.00
4.03
4.03
3.11
3.92
4.03
3.09
4.04
4.08
4.54
4.27
High Fuel Cost Range
8%
4.01
4.03
4.04
3.13
3.64
2.68
4.11
3.61
3.74
2.69
4.06
4.10
4.09
3.66
3.62
4.03
4.05
4.07
4.07
3.16
4.01
4.07
3.15
4.08
4.12
4.58
4.43
10%
4.09
4.10
4.12
3.21
3.73
2.76
4.20
3.70
3.84
2.78
4.14
4.19
4.18
3.75
3.71
4.10
4.13
4.15
4.15
3.24
4.09
4.16
3.23
4.17
4.21
4.67
4.49
12%
4.17
4.18
4.20
3.30
3.81
2.85
4.29
3.78
3.95
2.86
4.22
4.27
4.26
3.84
3.79
4.18
4.22
4.24
4.24
3.32
4.16
4.24
3.31
4.25
4.29
4.76
4.55
15%
4.28
4.30
4.32
3.42
3.95
2.97
4.43
3.91
4.10
2.99
4.34
4.41
4.39
3.98
3.91
4.30
4.34
4.37
4.36
3.44
4.27
4.36
3.43
4.37
4.42
4.88
4.65
18%
4.39
4.41
4.43
3.54
4.08
3.10
4.57
4.03
4.25
3.12
4.46
4.54
4.51
4.11
4.03
4.41
4.46
4.49
4.49
3.57
4.38
4.49
3.55
4.50
4.55
5.01
4.74
             (1)  The costs shown in this table reflect the study assumptions as summarized in Table 10.

             (2)  The costs influenced by the cooling system are:  a) the annual capital and operating cost of the cooling system (from the turbine flange outward); b) the annual
                 cost of auxiliary power and energy required for the cooling system; c) the annual cost of replacing capacity and energy lost at high turbine back pressures; and
                 d) the total annual plant fuel cost.

             (3)  The costs shown in this table do not include the variable and fixed costs, except fuel, related to the basic generating plant (boiler, turbine-generator, auxil-
                 iary equipment associated with the boiler and turbine-generator,  step-up  transformer,  swirchgear equipment,  and associated structures and foundations).

-------
                                                                                TABLE 24
                                              Optimized Total Annual Costs (in Mills per Kwh) Influenced by the Cooling System
                                                                  800-Mw, Fossil-Fueled Generating Unit
                                                             Mechanical-Draft, Dry-Type Cooling Tower System
        Fixed-Charge Rate:

PLANT SITE

     Seattle,  Wash.
     San Francisco, Calif.
     Los Angeles, Calif.
     Great Falls, Mont.
     Boise, Ida.
     Casper,  Wyo.
     Reno, Nev.
     Denver,  Colo.
     Phoenix, Ariz.
     Bismarck, N. Dak.
     Minneapolis, Minn .
     Omaha, Neb.
     Little Rock, Ark.
     Midland, Tex.
     New Orleans, La.
     Green Bay,  Wis.
     Grand Rapids, Mich.
     Detroit, Mich.
     Chicago, III.
     Nashville,  Tenn.
     Burlington,  Vt.
     Philadelphia, Penna.
     Charleston, W. Va.
     Atlanta, Ga.
     Miami,  Fla.
     Honolulu, Hawaii
     Anchorage, Alas.
                                        Low Fuel Cost Range
Medium Fuel Cost Range
8%
2.68
2.69
2.71
1 .78
2.29
1.31
2.76
2.26
2.40
1.53
2.72
2.76
2.76
2.31
2.27
2.69
2.72
2.74
2.73
2.10
2.68
2.74
1 .80
2.76
2.78
3.26
3.07
10%
2.76
2.77
2.80
1.86
2.38
1.39
2.85
2.35
2.51
1.61
2.81
2.86
2.85
2.40
2.36
2.77
2.80
2.82
2.82
2.18
2.76
2.83
1.89
2.84
2.87
3.35
3.13
12%
2.84
2.85
2.88
1 .94
2.47
1.48
2.95
2.43
2.61
1.70
2.89
2.95
2.93
2.49
2.44
2.85
2.89
2.91
2.91
2.27
2.83
2.92
1.97
2.93
2.96
3.43
3.20
15%
2.95
2.97
3.00
2.07
2.60
1.60
3.09
2.56
2.77
1.83
3.02
3.09
3.07
2.63
2.56
2.97
3.01
3.04
3.03
2.40
2.95
3.05
2.10
3.06
3.09
3.57
3.29
18%
3.07
3.09
3.12
2.19
2.74
1.72
3.22
2.68
2.93
1.96
3.14
3.22
3.20
2.77
2.68
3.09
3.14
3.17
3.16
2.53
3.06
3.18
2.22
3.19
3.22
3.70
3.38
8%
3.61
3.63
3.65
2.71
3.22
2.24
3.69
3.19
3.34
2.27
3.65
3.70
3.70
3.25
2.22
3.62
3.65
3.67
3.66
2.75
3.61
3.68
2.73
3.69
3.73
4.20
4.00
10%
3.69
3.70
3.73
2.79
3.31
2.33
3.79
3.28
3.45
2.36
3.74
3.79
3.79
3.35
3.30
3.70
3.73
3.76
3.75
2.84
3.69
3.76
2.82
3.78
3.82
4.29
4.06
12%
3.77
3.78
3.81
2.87
3.40
2.41
3.88
3.36
3.56
2.44
3.82
3.88
3.88
3.44
3.39
3.78
3.82
3.84
3.84
2.92
3.76
3.85
2.90
3.87
3.90
4.38
4.13
15%
3.88
3.90
3.93
3.00
3.54
2.53
4.02
3.49
3.72
2.58
3.95
4.02
4.01
3.58
3.51
3.90
3.94
3.97
3.97
3.05
3.88
3.98
3.03
4.00
4.04
4.51
4.22
18%
4.00
4.02
4.05
3.12
3.67
2.66
4.15
3.61
3.87
2.71
4.07
4.15
4.14
3.72
3.64
4.02
4.07
4.10
4.09
3.18
3.99
4.11
3.16
4.13
4.17
4.64
4.31
                                                                                                                                       High Fuel Cost Range
8%
4.08
4.09
4.12
3.17
3.69
2.71
4.16
3.66
3.82
2.73
4.12
4.16
4.17
3.73
3.69
4.08
4.12
4.13
4.13
3.22
4.07
4.14
3.20
4.16
4.20
4.67
4.47
10%
4.16
4.17
4.20
3.25
3.78
2.79
4.25
3.74
3.92
2.82
4.20
4.26
4.26
3.82
3.78
4.16
4.20
4.22
4.22
3.31
4.15
4.23
3.29
4.25
4.29
4.76
4.53
12%
4.23
4.25
4.28
3.34
3.87
2.88
4.34
3.83
4.03
2.91
4.28
4.35
4.35
3.91
3.86
4.24
4.28
4.31
4.30
3.39
4.23
4.32
3.37
4.33
4.38
4.85
4.59
15%
4.35
4.37
4.40
3.46
4.00
3.00
4.48
3.95
4.19
3.04
4.41
4.48
4.48
4.05
3.99
4.36
4.41
4.44
4.43
3.52
4.34
4.45
3.50
4.46
4.51
4.99
4.69
18%
4.47
4.49
4.52
3.58
4.14
3.13
4.62
4.08
4.34
3.17
4.54
4.62
4.62
4.19
4.11
4.48
4.53
4.57
4.56
3.65
4.46
4.58
3.62
4.59
4.64
5.12
4.78
(1)  The costs shown in this table reflect the study assumptions as summarized in Table 10.

(2)  The costs influenced by the cooling system are: a) the annual  capital and operating cost of the cooling system (from the turbine flange outward); b) the annual
     cost of auxiliary power and energy required  for the cooling system; c) the annual cost of replacing capacity and energy lost at high turbine back pressures; and
     d) the total annual plant fuel cost.

(3)   The costs shown in this fable do not include  the variable and fixed costs, except fuel, related to the basic generating plant (boiler, turbine-generator, auxil-
    iary equipment associated  with the boiler and turbine-generator,  step-up transformer, switchgear equipment, and associated structures and foundations).

-------
CO
        Fixed-Charge Rate:

PLANT SITE

    Seattle, Wash.
    San Francisco, Calif.
    Los Angeles,  Calif.
    Great Falls, Mont.
    Boise, Ida.
    Casper,  Wyo.
    Reno, Nev.
    Denver, Colo.
    Phoenix, Ariz.
    Bismarck,  N. Dak.
    Minneapolis,  Minn.
    Omaha,  Neb.
    Little Rock, Ark.
    Midland, Tex.
    New Orleans,  La.
    Green Bay, Wis.
    Grand Rapids, Mich.
    Detroit, Mich.
    Chicago, III.
    Nashville,  Tenn.
    Burlington,  Vt.
    Philadelphia, Penna.
    Charleston, W. Va.
    Atlanta,  Ga.
    Miami, Fla.
    Honolulu, Hawaii
   Anchorage, Alas.
                                                                                               TABLE 25

                                                             Optimized Total Annual Costs (in Mills per Kwh) Influenced by the Cooling System
                                                                               800-Mw, Nuclear-Fueled Generating Unit
                                                                             Natural-Draft, Dry-Type Cooling Tower System
                                                       Low Fuel  Cost Range
8%
.72
.73
.75
.77
.83
.75
.84
1 .80
.97
1 .80
1.78
1.83
1.80
1.84
1.79
1 .74
1.78
.80
.80
.80
.72
.80
.79
.81
.86
.86
.62
10%
1.83
1.85
1 .88
1.89
1.96
1.88
1 .98
1 .92
2.12
1 .93
1.91
1.97
1 .93
1.97
1.92
1 .86
1.90
1 .93
1 .93
1.93
1.84
1.93
1.92
1 .94
2.00
2.00
1.71
12%
1.95
1 .97
2.00
2.01
2.09
2.00
2.11
2.05
2.27
2.05
2.03
2.10
2.06
2.10
2.05
1.98
2.03
2.06
2.06
2.06
1.95
2.06
2.05
2.07
2.14
2.13
1.81
15%
2.13
2.15
2.19
2.20
2.29
2.18
2.31
2.24
2.50
2.25
2.22
2.30
2.25
2.30
2.23
2.16
2.22
2.26
2.25
2.25
2.13
2.26
2.24
2.27
2.34
2.33
1.95
18%
2.30
2.33
2.37
2.38
2.49
2.36
2.51
2.43
2.72
2.44
2.41
2.50
2.45
2.50
2.42
2.33
2.40
2.45
2.44
2.44
2.30
2.45
2.43
2.46
2.54
2.53
2.10
                                                                                     Medium Fuel  Cost Range
8%
2.14
2.16
2.18
2.19
2.25
2.19
2.27
2.22
2.41
2.22
2.21
2.26
2.25
2.28
2.24
2.17
2.20
2.23
2.22
2.23
2.15
2.23
2.22
2.24
2.29
2.29
2.05
10%
2.26
2.28
2.30
2.32
2.39
2.32
2.41
2.35
2.56
2.35
2.33
2.40
2.38
2.42
2.37
2.29
2.33
2.36
2.36
2.36
2.26
2.36
2.35
2.37
2.43
2.43
2.14
12%
2.38
2.40
2.43
2.44
2.52
2.44
2.54
2.48
2.71
2.48
2.46
2.53
2.51
2.56
2.50
2.41
2.46
2.49
2.49
2.49
2.38
2.49
2.48
2.50
2.57
2.56
2.24
15%
2.55
2.58
2.61
2.62
2.72
2.62
2.74
2.67
2.93
2.67
2.65
2.73
2.70
2.76
2.69
2.58
2.64
2.68
2.68
2.68
2.56
2.68
2.67
2.70
2.77
2.77
2.38
18%
2.73
2.76
2.80
2.81
2.91
2.81
2.94
2.86
3.15
2.87
2.83
2.93
2.89
2.96
2.88
2.76
2.83
2.87
2.87
2.87
2.73
2.88
2.86
2.89
2.97
2.97
2.53
High Fuel Cost Range
8%
2.68
2.69
2.71
2.73
2.79
2.73
2.80
2.76
2.95
2.76
2.74
2.80
2.79
2.83
2.78
2.70
2.74
2.76
2.76
2.77
2.68
2.76
2.76
2.78
2.83
2.83
2.58
10%
2.79
2.81
2.84
2.85
2.92
2.86
2.94
2.89
3.10
2.89
2.87
2.93
2 92
2.97
2.91
2.82
2.86
2.89
2.89
2.90
2.80
2.90
2.88
2.91
2.97
2.97
2.68
12%
2.91
2.93
2.96
2.97
3.05
2.98
3.08
3.01
3.25
3.02
2.99
3.07
3.05
3.10
3.04
2.94
2.99
3.02
3.02
3.03
2.91
3.03
3.01
3.04
3.11
3.11
2.78
15%
3.09
3.11
3.15
3.16
3.25
3.17
3.28
3.20
3.47
3.21
3.15
3.27
3.25
3.31
3.23
3.12
3.18
3.22
3.21
3.22
3.09
3.22
3.20
3.23
3.31
3.31
2.92
18%
3.26
2.29
3.33
3.34
3.45
3.35
3.48
3.39
3.70
3.40
3.37
3.47
3.44
3.51
3.42
3.30
3.36
3.41
3.40
3.41
3.26
3.41
3.39
3.43
3.51
3.52
3.07
             (1)  The costs shown in this table reflect the study assumptions as summarized in Table 10.


             (2)  coTt oTaUmaTyToier and er^r'"9^uTdT If? *' T^0' T^"' ""' °Per°tin9 "*' °f the C°°lin9 T"*"1 (ft°m the turbine flan9e °utw^d); b) the annual


             (3)  The costs shown in this table do not include the variable and fixed costs, except fuel, related to the basic generating plant (boiler, turbine-generator  auxil-
                 ,ary equ.pment assorted w,th the boiler and turbine-generator,  step-up transformer,  switchgear  equipment, and associated structures and  foundation! .

-------
                                                                                              TABLE 26
                                                            Optimized Total Annual Costs (in Mills per Kwh) Influenced by the Cooling System
                                                                              800-Mw, Nuclear-Fueled Generating Unit
                                                                          Mechanical-Draft, Dry-Type Cooling Tower System
00
        Fixed-Charge Rate:

PLANT SITE

    Seattle, Wash.
    San Francisco, Calif.
    Los Angeles, Calif.
    Great  Falls, Mont.
    Boise,  Ida.
    Casper, Wyo.
    Reno,  Nev.
    Denver, Colo.
    Phoenix, Ariz.
    Bismarck, N.  Dak.
    Minneapolis, Minn.
    Omaha, Neb.
    Little Rock, Ark.
    Midland, Tex.
    New Orleans, La.
    Green  Bay,  Wis.
    Grand  Rapids, Mich.
    Detroit, Mich.
    Chicago, III.
    Nashville,  Tenn.
    Burlington, Vt.
    Philadelphia, Penna.
    Charleston,  W. Va.
    Atlanta, Ga.
    Miami, Fla.
    Honolulu, Hawaii
    Anchorage, Alas.
                                                      Low Fuel Cost Range
8%
.77
.78
.81
.81
.87
.79
.88
1.83
2.03
.84
1.83
1 .88
1.86
1.89
1 .85
1.79
.82
1.85
.84
.86
.77
.86
.85
.87
.91
.91
.63
10%
1.89
1.90
1.94
1.93
2.00
1.92
2.02
1.96
2.19
1.97
1.96
2.02
2.00
2.03
1.98
1.91
1.96
1.98
1 .98
2.00
1 .89
1.99
1.98
2.01
2.05
2.05
1.73
12%
2.01
2.03
2.07
2.06
2.14
2.04
2.16
2.10
2.35
2.11
2.09
2.17
2.14
2.16
2.11
2.04
2.09
2.12
2.11
2.14
2.02
2.13
2.11
2.15
2.19
2.20
1.82
15%
2.20
2.22
2.26
2.26
2.35
2.23
2.37
2.29
2.58
2.31
2.29
2.37
2.34
2.37
2.30
2.22
2.28
2.32
2.31
2.34
2.20
2.33
2.31
2.35
2.41
2.41
1.97
18%
2.38
2.40
2.46
2.45
2.55
2.42
2.57
2.49
2.81
2.51
2.48
2.58
2.54
2.57
2.49
2.41
2.48
2.52
2.51
2.54
2.38
2.53
2.51
2.55
2.62
2.62
2.11
                                                                                      Medium Fuel Cost Range
8%
2.20
2.22
2.25
2.24
2.30
2.24
2.31
2.27
2.47
2.27
2.26
2.32
2.32
2.34
2.30
2.22
2.26
2.28
2.28
2.30
2.20
2.29
2.28
2.31
2.35
2.35
2.06
10%
2.32
2.34
2.38
2.37
2.44
2.36
2.45
2.40
2.63
2.41
2.39
2.47
2.46
2.49
2.44
2.35
2.39
2.42
2.42
2.44
2.33
2.43
2.42
2.45
2.50
2.50
2.16
12%
2.45
2.47
2.51
2.50
2.58
2.49
2.59
2.53
2.79
2.54
2.53
2.60
2.60
2.63
2.57
2.47
2.52
2.56
2.55
2.57
2.45
2.56
2.55
2.59
2.64
2.64
2.26
15%
2.63
2.65
2.70
2.69
2.78
2.68
2.80
2.73
3.02
2.74
2.72
2.81
2.80
2.84
2.77
2.66
2.72
2.76
2.75
2.78
2.63
2.77
2.75
2.79
2.85
2.85
2.40
18%
2.81
2.84
2.89
2.88
2.99
2.87
3.01
2.92
3.25
2.94
2.92
3.02
3.00
3.05
2.96
2.84
2.91
2.96
2.95
2.98
2.81
2.97
2.94
2.99
3.06
3.07
2.55
                                                                                                                                                    High Fuel Cost Range
8%
2.74
2.76
2.79
2.78
2.85
2.79
2.86
2.81
3.03
2.82
2.80
2.87
2.88
2.90
2.86
2.76
2.80
2.83
2.82
2.85
2.75
2.84
2.83
2.86
2.91
2.91
2.61
10%
2.87
2.89
2.92
2.91
2.99
2.91
3.00
2.94
3.19
2.95
2.94
3.01
3.01
3.05
2.99
2.89
2.94
2.96
2.96
2.99
2.87
2.98
2.96
3.00
3.05
3.06
2.71
12%
2.99
3.01
3.05
3.04
3.12
3.04
3.14
3.08
3.35
3.09
3.07
3.15
3.15
3.19
3.13
3.01
3.07
3.10
3.09
3.12
2.99
3.11
3.10
3.13
3.19
3.20
2.80
15%
3.17
3.20
3.25
3.23
3.33
3.23
3.35
3 27
3.58
3.29
3.27
3.36
3.36
3.40
3.33
3.20
3.26
3.30
3.29
3.33
3.17
3.31
3.29
3.34
3.40
3.42
2.95
18%
3.36
3.38
3.44
3.42
3.53
3.43
3.55
3.47
3.81
3.49
3 46
3.56
3.56
3.61
3.53
3.39
3.46
3.50
3.49
3.52
3.36
3.51
3.49
3.54
3.61
3.63
3.09
             (1)  The costs shown in this table reflect the study assumptions as summarized in Table 10.

             (2)  The costs influenced by the cooling system are: a) the annual capital and  operating cost of the cooling system (from the turbine flange outward); b) the annual
                  cost of auxiliary power and energy required  for the cooling system; c) the annual cost of replacing capacity and energy lost at high turbine back pressures; and
                  d) the total annual plant fuel cost.

             (3)  The costs shown in this table do not include  the variable and fixed costs, except fuel, related to the basic generating plant (boiler, turbine-generator, auxil-
                 iary equipment associated with the boiler and turbine-generator, step-up  transformer,  switchgear equipment, and associated  structures and foundations).

-------
                                                                                             TABLE 27
                                                                     Auxiliary Capacity Required (in Mw) for Cooling System Pumps
                                                                   at the Optimum ITD for an 800-Mw, Fossil-Fueled Generating Unit
                                                                            Natural-Draft, Dry-Type Cooling Tower System
00
                                                      Low Fuel Cost Range
                                                                                     Medium Fuel Cost Range
                      Fixed-Charge Rate:
8%
6.5
6.5
6.8
6.4
6.6
6.3
6.5
6.6
7.7
6.6
6.8
6.9
7.2
6.9
7.0
6.5
6.6
6.6
6.6
7.2
6.6
6.9
6.8
7.2
7.6
7.6
4.7
10%
6.4
6.4
6.6
6.4
6.5
6.0
6.5
6.5
7.7
6.6
6.6
6.8
7.0
6.8
6.9
6.5
6.6
6.6
6.6
7.0
6.6
6.8
6.8
7.0
7.3
7.2
4.7
12%
6.4
6.4
6.6
6.4
6.5
6.0
6.5
6.5
7.6
6.5
6.6
6.8
7.0
6.8
6.8
6.5
6.6
6.6
6.6
7.0
6.5
6.8
6.6
6.9
7.3
7.2
4.7
15%
6.4
6.4
6.6
6.3
6.5
6.0
6.4
6.4
7.2
6.5
6.5
6.6
6.9
6.6
6.8
6.4
6.5
6.5
6.5
6.9
6.5
6.6
6.6
6.8
7.2
7.0
4.7
18%
6.4
6.4
6.5
6.1
6.4
5.9
6.4
6.4
7.0
6.5
6.5
6.6
6.9
6.6
6.8
6.4
6.5
6.5
6.5
6.8
6.5
6.6
6.5
6.8
7.0
6.9
4.7
PLANT SITE

    Seattle,  Wash.
    San Francisco, Calif.
    Los Angeles, Calif.
    Great Falls, Mont.
    Boise, Ida.
    Casper, Wyo.
    Reno, Nev.
    Denver,  Colo.
    Phoenix, Ariz.
    Bismarck, N. Dak.
    Minneapolis, Minn.
    Omaha,  Neb.
    Little Rock, Ark.
    Midland, Tex.
    New Orleans, La.
    Green Bay, Wis.
    Grand Rapids, Mich.
    Detroit,  Mich.
    Chicago, III.
    Nashville, Tenn.
    Burlington, Vt.
    Philadelphia, Penna.
    Charleston, W.  Va.
    Atlanta, Ga.
    Miami,  Fla.
    Honolulu,  Hawaii
    Anchorage, Alas.


(1)  The tabulated values reflect the maximum net power demands required by the pumps to overcome head loss and to circulate
    sufficient water through  the cooling system to remove the heat rejected by the turbine for the range and other operating
    conditions established at  the optimum ITD for the fuel  costs and fixed-charge rates used at each location.

(2)  Recovery water turbines were assumed to be directly connected  to  each pump-motor shaft to  recover excess pressure head
    and to control  the water pressure in the condenser spray nozzles.
8%
6.5
6.5
6.8
6.4
6.6
6.4
6.5
6.6
7.7
6.6
6.8
6.9
7.3
7.0
7.2
6.5
6.8
6.8
6.8
7.2
6.6
6.9
6.8
7.2
7.7
7.6
4.7
10%
6.4
6.5
6.8
6.4
6.5
6.4
6.5
6.5
7.7
6.6
6.6
6.8
7.2
6.9
7.0
6.5
6.6
6.6
6.6
7.0
6.6
6.8
6.8
7.0
7.6
7.3
4.7
12%
6.4
6.4
6.6
6.4
6.5
6.3
6.5
6.5
7.6
6.5
6.6
6.8
7.2
6.9
7.0
6.5
6.6
6.6
6.6
7.0
6.5
6.8
6.6
6.9
7.3
7.2
4.7
15%
6.4
6.4
6.6
6.3
6.5
6.3
6.4
6.4
7.2
6.5
6.5
6.8
7.0
6.8
6.9
6.4
6.5
6.5
6.5
6.9
6.5
6.8
6.6
6.9
7.2
7.0
4.7
18%
6.4
6.4
6.6
6.3
6.4
6.0
6.4
6.4
7.2
6.5
6.5
6.6
6.9
6.8
6.8
6.4
6.5
6.5
6.5
6.9
6.5
6.6
6.6
6.8
7.0
6.9
4.7
                                                                                                                                                   High Fuel Cost Range
8%
6.5
6.5
6.8
6.5
6.6
6.4
6.5
6.6
7.9
6.6
6.8
6.9
7.6
7.2
7.3
6.5
6.8
6.8
6.8
7.2
6.6
6.9
6.8
7.2
7.7
7.6
4.7
10%
6.5
6.5
6.8
6.4
6.6
6.4
6.5
6.5
7.7
6.6
6.6
6.8
7.3
7.0
7.2
6.5
6.6
6.6
6.6
7.2
6.6
6.9
6.8
7.0
7.6
7.3
4.7
12%
6.4
6.5
6.6
6.4
6.5
6.4
6.4
6.5
7.6
6.5
6.6
6.8
7.2
6.9
7.0
6.5
6.6
6.6
6.6
7.0
6.5
6.8
6.8
6.9
7.3
7.2
4.7
15%
6.4
6.4
6.6
6.3
6.5
6.3
6.4
6.4
7.3
6.5
6.5
6.8
7.0
6.9
7.0
6.5
6.5
6.5
6.5
6.9
6.5
6.8
6.6
6.9
7.2
7.0
4.7
18%
6.4
6.4
6.6
6.3
6.4
6.3
6.4
6.4
7.2
6.5
6.5
6.6
7.0
6.8
6.9
6.4
6.5
6.5
6.5
6.9
6.5
6.6
6.6
6.8
7.0
7.0
4.7

-------
                                                                                             TABLE 28
                                                                Auxiliary Capacity Required (in Mw) for Cooling System Pumps and Fans
                                                                  at the Optimum  ITD for an 800-Mw,  Fossil-Fueled Generating  Unit
                                                                         Mechanical-Draft, Dry-Type Cooling Tower System
00
00
        Fixed-Charge Rate:

PLANT SITE

    Seattle, Wash.
    San Francisco, Calif.
    Los Angeles,  Calif.
    Great Falls, Mont.
    Boise,  Ida.
    Casper, Wyo.
    Reno, Nev.
    Denver, Colo.
    Phoenix, Ariz.
    Bismarck, N.  Dak.
    Minneapolis, Minn.
    Omaha, Neb.
    Little Rock, Ark.
    Midland, Tex.
    New Orleans, La.
    Green Bay, Wis.
    Grand Rapids, Mich.
    Detroit, Mich.
    Chicago, III .
    Nashville,  Tenn.
    Burlington, Vt.
    Philadelphia, Penna .
    Charleston, W.  Va,
    Atlanta, Ga.
    Miami, Fla.
    Honolulu, Hawaii
    Anchorage, Alas.
                                                     Low Fuel Cost Range
8%
16.3
16.3
17.2
16.6
17.6
16.3
17.2
17.6
22.2
17.2
17.2
18.7
19.5
18.7
18.7
16.9
17.2
17.2
17.2
20.4
16.9
18.3
18.3
19.5
20.8
20.4
12.0
10%
15.9
15.9
16.9
16.6
17.2
16.3
16.9
17.2
21.3
16.9
16.9
18.3
19.1
18.3
18.3
16.6
16.9
16.9
16.9
19.5
16.9
18.0
17.6
18.7
20.4
19.9
12.0
12%
15.6
15.9
16.9
16.3
16.9
15.9
16.9
16.9
20.8
16.9
16.9
18.0
18.7
18.0
17.6
16.6
16.9
16.9
16.9
19.1
16.6
17.2
17.2
18.7
20.4
19.1
12.0
15%
15.6
15.6
16.9
15.9
16.9
15.9
16.6
16.9
20.4
16.9
16.6
17.6
18.3
17.6
17.2
16.3
16.6
16.6
16.6
18.7
16.6
16.9
16.9
18.3
19.5
18.7
12.0
18%
15.4
15.4
16.6
15.9
16.9
15.9
16.6
16.6
19.9
16.6
16.6
17.2
18.0
17.6
16.9
15.9
16.6
16.6
16.6
18.3
16.3
16.9
16.9
18.0
19.1
18.3
12.0
                                                                                     Medium Fuel Cost Range
8%
15.9
16.3
17.2
16.6
17.6
16.9
17.2
17.6
22.2
17.2
17.2
18.7
20.4
19.5
19.5
16.9
17.2
17.2
17.2
20.4
16.9
18.3
18.3
19.5
20.8
20.4
12.0
10%
15.9
15.9
16.9
16.6
17.2
16.6
16.9
17.2
21.3
16.9
16.9
18.3
19.9
19.1
19.1
16.6
16.9
16.9
16.9
19.5
16.9
18.0
17.6
18.7
20.4
19.5
12.0
12%
15.6
15.9
16.9
16.3
16.9
16.6
16.9
16.9
20.8
16.9
16.9
18.0
19.1
18.7
18.7
16.6
16.9
16.9
16.9
19.1
16.6
17.2
17.2
18.3
19.9
19.1
12.0
15%
15.6
15.6
16.9
15.9
16.9
16.3
16.6
16.9
20.4
16.9
16.9
17.6
18.7
18.4
18.3
16.3
16.9
16.6
16.6
18.7
16.6
16.9
16.9
18.0
19.5
18.7
12.0
18%
15.4
15.6
16.6
15.9
16.9
15.9
16.6
16.6
19.9
16.6
16.6
17.2
18.7
18.0
17.6
15.9
16.6
16.6
16.6
18.3
16.3
16.9
16.9
18.0
19.1
18.3
12.0
High Fuel Cost Range
8%
15.9
16.3
17.2
16.9
17.6
16.9
17.2
17.6
22.2
17.2
17.2
18.7
20.4
19.9
19.9
16.9
17.2
17.2
17.2
19.9
16.9
18.3
18.3
19.5
20.8
20.4
12.0
10%
15.9
15.9
16.9
16.6
17.2
16.9
16.9
17.2
21.3
16.9
16.9
18.3
19.9
19.5
19.5
16.6
16.9
16.9
16.9
19.5
16.9
17.6
17,6
18.7
20.4
19.9
12.0
12%
15.6
15.9
16.9
16.3
16.9
16.6
16.9
16.9
20.8
16.9
16.9
18.0
19.5
19.1
19.1
16.3
16.9
16.9
16.9
19.1
16.6
17.2
17.2
18.3
19.9
19.5
12.0
15%
15.6
15.6
16.9
15.9
16.9
16.3
16.6
16.9
20.4
16.9
16.9
18.0
19.1
18.7
18.3
16.3
16.9
16.6
16.6
18.7
16.6
16.9
16.9
18.0
19.5
18.7
12.0
18%
15.6
15.6
16.6
15.9
16.9
16.3
16.6
16.6
19.9
16.6
16.6
17.2
18.7
18.3
18.0
15.9
16.6
16.6
16.6
18.3
16.3
16.9
16.9
17.6
19.1
18.3
12.0
             (1)  The  tabulated values reflect the maximum net power demands required by the fans to pass sufficient air through the heat
                  exchangers and by the pumps to overcome head loss and to circulate sufficient water in the cooling system to remove the
                  heat rejected by the turbine for the range and other operating conditions established at the optimum  ITD for the fuel cost
                 and fixed-charge rates used at each location .

             (2)  Recovery water turbines were assumed to be  directly connected to each pump-motor shaft to recover excess pressure head
                 and to control the water pressure in the  condenser spray nozzles.

-------
                                                                               TABLE 29

                                                      Auxiliary Capacity Required (in Mw) for Cooling System Pumps
                                                   at the Optimum ITD for an 800-Mw,  Nuclear-Fueled Generating Unit
                                                          Natural-Draft, Dry-Type Cooling Tower System
        Fixed-Charge Rate:

 PLANT SITE

     Seattle, Wash.
     San Francisco, Calif.
     Los Angeles, Calif.
     Great Falls, Mont.
     Boise, Ida.
     Casper,  Wyo.
     Reno, Nev.
     Denver, Colo.
     Phoenix, Ariz.
     Bismarck, N. Dak.
     Minneapolis, Minn.
     Omaha, Neb.
     Little Rock, Ark.
     Midland, Tex.
     New  Orleans, La.
     Green Bay,  Wis.
     Grand Rapids, Mich.
     Detroit, Mich.
     Chicago, III.
     Nashville, Tenn.
     Burlington,  Vt.
     Philadelphia, Penna.
     Charleston,  W.  Va.
     Atlanta, Ga.
     Miami, Fla.
     Honolulu, Hawaii
     Anchorage,  Alas.
                                        Low Fuel Cost Range
                                                                                     Medium Fuel Cost Range
High Fuel Cost Range
8%
8.6
9.5
10.0
8.6
9.0
8.2
8.7
9.5
10.5
9.0
9.5
9.8
10.0
9.6
10.2
9.6
9.6
9.5
9.6
10.2
9.6
10.0
10.0
10.2
11.4
11.0
7.0
10%
8.6
8.9
9.8
8.6
8.7
8.1
8.6
9.0
10.2
8.7
9.2
9.6
10.0
9.6
10.0
9.5
9.2
9.2
9.2
10.0
9.6
9.8
9.8
10.2
11 .0
10.8
7.0
12%
8.6
8.6
9.6
8.5
8.6
8.0
8.5
8.9
10.2
8.6
9.0
9.6
9.8
8.7
10.0
8.6
9.0
9.0
9.0
10.0
9.6
9.6
9.6
10.0
10.5
10.5
7.0
15%
8.5
8.6
9.6
8.5
8.6
8.0
8.2
8.5
10.0
8.6
8.6
8.7
9.6
8.6
9.2
8.6
8.6
8.6
8.6
9.8
8.6
9.0
9.2
9.8
10.5
10.3
7.0
18%
8.5
8.6
9.5
8.3
8.4
8.0
8.1
8.5
9.6
8.3
8.5
8.6
9.6
8.5
9.0
8.5
8.5
8.5
8.5
9.8
8.6
8.9
8.9
9.6
10.3
10.2
7.0
8%
8.6
9.5
10.0
8.9
9.2
8.5
8.9
9.6
10.5
9.2
9.6
9.8
10.2
10.0
10.5
9.6
9.6
9.6
9.6
10.2
9.6
10.0
10.0
10.2
11.4
11.0
7.0
10%
8.6
9.0
9.8
8.6
8.7
8.5
8.6
9.0
10.3
8.8
9.2
9.8
10.2
9.8
10.3
9.6
9.2
9.2
9.5
10.2
9.6
10.0
9.8
10.2
11.0
10.8
7.0
12%
8.6
8.7
9.6
8.5
8.6
8.3
8.5
8.9
10.2
8.7
9.0
9.6
10.0
9.6
10.3
8.6
9.0
9.0
9.0
10.0
9.6
9.8
9.6
10.0
10.8
10.5
7.0
15%
8.5
8.6
9.6
8.5
8.6
8.1
8.2
8.6
10.0
8.6
8.9
8.7
9.8
9.6
10.0
8.6
8.6
8.9
8.6
9.8
8.6
9.2
9.2
9.8
10.5
10.3
7.0
18%
8.5
8.6
9.5
8.5
8.3
8.1
8.1
8.5
9.8
8.6
8.5
8.7
9.8
8.7
10.0
8.5
8.5
8.5
8.5
9.8
8.6
8.9
9.0
9.6
10.3
10.2
7.0
8%
9.5
9.6
10.0
8.9
9.6
8.6
8.9
9.8
10.8
9.2
9.8
10.0
10.5
10.2
10.8
9.6
9.8
9.8
9.8
10.2
9.6
10.0
10.2
10.3
11 .4
11 .2
7.0
10%
8.6
9.5
9.8
8.6
8.7
8.5
8.7
9.2
10.3
8.9
9.2
9.8
10.2
10.0
10.5
9.6
9.5
9.2
9.5
10.2
9.6
10.0
10.0
10.2
11.0
11.0
7.0
12%
8.6
8.7
9.8
8.5
8.7
8.5
8.6
8.9
10.2
8.7
9.0
9.6
10.2
9.8
10.3
8.6
9.2
9.0
9.2
100
9.6
9.8
9.6
10.0
10.8
10.8
7.0
15%
8.6
8.6
9.6
8.5
8.6
8.2
8.3
8.6
10.0
8.6
8.9
8.9
10.0
9.6
10.2
8.6
8.9
8.9
8.6
10.0
8.6
9.2
9.2
10.0
10.5
10.5
7.0
18%
8.5
8.6
9.6
8.3
8.3
8.1
8.2
8.5
9.8
8.6
8.6
8.7
9.8
9.6
10.0
8.5
8.6
8.5
8.5
9.8
8.6
9.0
9.0
9.8
10.3
10.3
7.0
(1)  The tabulated values reflect the maximum net power demands required by the pumps to overcome head loss and to circulate
    sufficient water through the cooling system to remove the heat rejected by the turbine for the range and other operating
    conditions established at the optimum ITD for the fuel costs and fixed-charge rates used at each location.
(2)  Recovery water turbines were assumed to be directly connected to each pump-motor shaft to recover excess pressure
    and to control the water pressure in the condenser spray nozzles.
                                                                                                                head

-------
                                                                                TABLE 30
                                                   Auxiliary Capacity Required (in Mw) for Cooling System Pumps and Fans
                                                   at the Optimum ITD for an 800-Mw, Nuclear-Fueled Generating Unit
                                                            Mechanical-Draft, Dry-Type Cooling  Tower System
         Fixed-Charge Rate:

 PLANT SITE

      Seattle, Wash.
      San Francisco, Calif.
      Los Angeles, Calif.
      Great  Falls, Mont.
      Boise,  Ida.
      Casper, Wyo.
      Reno,  Nev.
      Denver, Colo.
      Phoenix, Ariz.
      Bismarck,  N. Dak.
      Minneapolis, Minn.
     Omaha, Neb.
      Little Rock, Ark.
     Midland, Tex.
      New Orleans, La.
      Green  Bay, Wis.
      Grand  Rapids, Mich.
      Detroit, Mich.
     Chicago, III.
      Nashville,  Tenn.
     Burlington, Vt.
     Philadelphia, Penna.
     Charleston, W. Va.
     Atlanta, Go.
     Miami, Fla.
     Honolulu,  Hawaii
     Anchorage, Alas.
                                         Low Fuel  Cost Range
8%
23.5
23.9
26.8
23.5
25.3
21 .9
24.4
25.3
30.9
24.8
25.3
26.8
27.4
25.8
28.5
25.3
25.3
25.3
25.8
28.5
26.3
25.8
27.9
28.5
33.0
31.6
18.3
10%
23.1
23.5
25.8
23.1
23.9
21.9
23.9
25.3
29.1
23.9
25.3
25.3
26.3
24.4
25.8
23.5
25.3
24.8
25.3
27.9
25.3
25.3
25.8
27.9
31 .6
30.9
18.3
12%
23.1
23.5
25.8
23.1
23.9
21.9
23.1
24.8
28.5
23.5
24.4
24.8
25.8
23.9
25.3
23.5
24.4
24.4
23.5
27.4
23.5
25.3
25.3
25.8
30.9
30.3
18.3
15%
22.7
23.1
25.3
22.7
23.5
21.6
22.3
23.5
25.8
22.7
23.1
23.9
23.9
23.5
24.8
23.1
23.1
23.1
23.1
25.8
23.5
24.4
24.8
25.3
29.7
29.7
18.3
18%
22.7
23.1
25.3
21.9
22.7
21.6
21.9
23.1
25.3
22.3
23.1
23.5
23.9
23.1
24.4
23.1
23.1
22.7
22.7
25.3
23.1
23.9
23.5
25.3
29.7
27.9
18.3
Medium Fuel Cost Range
8%
23.5
23.9
26.3
23,5
25.3
23.5
24.4
25.8
30.9
24.8
25.3
26.9
28.5
27.9
29.7
25.3
25.8
25.3
25.8
28.5
26.3
25.8
27.9
28.5
33.0
31 .6
18.3
10%
23.1
23.5
25.8
23.5
23.9
23.1
23.9
25.3
29.1
23.9
25.3
25.3
27.9
26.8
29.1
23.5
25.3
24.8
25.3
27.9
25.3
25.3
25.8
27.9
31.6
30.9
18.3
12%
23.1
23.5
25.8
23.1
23.9
22.3
23.5
24.8
28.5
23.5
24.4
24.8
26.8
25.8
28.5
23.5
24.4
24.4
23.5
27.4
23.5
25.3
25.3
26.3
30.9
30.3
18.3
15%
22.7
23.1
25.3
22.7
23.5
21.9
22.3
23.5
26.3
22.7
23.1
23.9
26.3
24.4
25.8
23.1
23.1
23.1
23.1
25.8
23.5
24.4
24.8
25.3
29.7
29.7
18.3
18%
22.7
23.1
24.8
21.9
22.7
21 .9
21.9
23.1
25.3
22.3
23.1
23.5
25.3
23.9
25.3
23.1
23.1
23.1
23.1
25.3
23.1
23.9
23.5
25.3
29.7
27.9
18.3
                                                                                                                                      High Fuel Cost Range
8%
23.5
23.9
26.3
23.5
25.3
23.9
24.4
25.8
30.9
24 8
25.3
26.8
30.9
27.9
30.3
25.3
25.8
25.3
25.8
28.5
26.3
25.8
26.3
28.5
33.0
31 .6
18.3
10%
23.1
23.5
25.8
23.5
23.9
23.1
23.9
25.3
29.1
23.9
25.3
25.3
27.9
27.4
29.7
23.5
25.3
25.3
25.3
27.9
25.3
25.3
25.8
27.9
31.6
30.9
18.3
12%
23.1
23.5
25.8
23.1
23.9
23.1
23.5
24.8
28.5
23.9
24.4
24.8
27.9
26.3
29.1
23.5
24.4
24.4
23.5
27.4
23.5
25.3
25.3
27.4
30.9
30.9
18.3
15%
22.7
23.1
25.3
22.7
23.5
22.3
22.3
23.5
26.8
22.7
23.1
23.9
26.8
25.8
25.8
23.1
23.1
23.1
23.1
26.3
23.5
24.4
24.8
25.8
29.7
29.7
18.3
18%
22.7
23.1
24.8
21.9
22.7
21.9
21.9
23.1
25.8
22.3
23.1
23.5
25.8
24.4
25.8
23.1
23.1
23.1
23.1
25.3
23.1
23.9
23.5
25.3
29.7
29.1
18.3
(1)  The tabulated values reflect the maximum net power demands required by the fans to pass sufficient air through the heat
     exchangers and by the pumps to overcome head loss and to circulate sufficient water in the cooling system to remove the
     heat rejected by the turbine for the range and other operating conditions established at the optimum ITD for the fuel cost
     and fixed-charge rates used at each location.

(2)  Recovery water turbines were assumed to be directly connected to each pump-motor shaft to recover excess pressure head
    and to controf the water pressure in the condenser spray nozzles.

-------
                                 SECTION XI

                           DISCUSSION OF RESULTS


General

       The results of the economic optimization of dry-type cooling systems  have
been presented in Section X of this  report, and those results reflect the effects of
the basic study assumptions and  the method of analysis.

       As shown on Figures 42 and 43, the economically optimum ITD values,  in °F,
for fossil-fueled  generating units were found to range from the mid 50's to the low
60' s in the cooler portions of the United States.  In the warmer areas of the country,
these optimum ITD values were found  to lie in  the mid 40's to mid 50's.   For
nuclear-fueled plants,  Figures 44 and 45 show the optimum ITD values to range from
the high 50" s to  mid 60' s in the cooler areas and from the high 40' s to the mid 60" s
in the warmer areas.

       The optimum ITD values for fossil-fueled plants shown on Figures 42 and 43
may be compared to the design  ITD values which are summarized in  Table 31 for
one United States generating plant and four  European plants which are now in oper-
ation .

       The economic optimization analyses  indicate that in the United States the
optimum result would be  obtained by sizing the dry cooling system so that some loss
of capacity would be experienced in hot weather.   The analyses  indicate that it
would be more economical to replace  the lost capacity  from  other  generating
sources than to increase the size of the cooling system to reduce the capacity losses.
As shown on Figures 46 and 47, the capacity losses at the optimum ITD for fossil-
fueled units were generally on the order of 5 to 10 percent of rated  capacity and
the maximum value  found for the sites studied was 12.6 percent.  The capacity
losses at the optimum ITD were  found to be somewhat higher for nuclear-fueled units
ranging up near  15  percent in many cases.   The maximum capacity loss at an opti-
mum ITD was found  to  be 19.9 percent for the sites studied.

       The capital  costs of the dry cooling system at optimum  ITD as summarized on
Figures 50 and 51 for  fossil-fueled  units were found to range  from  slightly  below
$14 per kw to about $25 per kw for mechanical-draft systems and were found to be
slightly higher for natural-draft systems, $15 to $27 per kw.  The capital cost of
dry cooling systems at  optimum ITD for nuclear-fueled units is summarized  on
Figures 52 and 53.  These nuclear plant values are about 50 percent higher per kw
than the figures  for the fossil-fueled plants reflecting the greater heat rejection of
the nuclear units.  The dry cooling system costs include all costs of the  generating
                                     191

-------
                                               TABLE 31

                          Initial Temperature Differences of Dry Cooling Systems
                                      Existing Installations Visited
        Name
 Rugeley
     (England)

 Ibbenburen
     (Germany)


 Volkswagen
     (Germany)

 Gyongybs
     (Hungary)

 Neil Simpson
     (Wyoming, U.S.)
                       Power Plant
                              Dry-Cooled Generating Units
                                                 Cooling System
No.
 2
 2
Capacity per Unit
              120
                  mw
              150 mw
               50 mw
    100 mw
    200 mw

     20 mw
                                                                           Type
Natural draft
Indirect cooling system

Natural draft
Indirect cooling system


Mechanical draft
Direct condensing system

Natural draft
Indirect cooling system

Mechanical draft
Direct condensing system
                                                                 Design ITD
                                                                   35(1)
                                                                   50
                                                          51
46
47

55
(1)  Not optimized.

-------
plant from the turbine flange outward, and therefore include condenser costs, the
costs of cooling system pumps, piping and  valves, and cooling tower costs.

       Although the capital cost of the mechanical-draft cooling systems was found
to be slightly lower than the capital  cost of the natural-draft cooling systems,  the
economic analyses indicated that the annual cost of the natural-draft systems would
be slightly  lower than the annual cost of the mechanical-draft systems due to the
power and energy requirements of the mechanical-draft fans.  This cost difference
in favor of the natural-draft systems was very small,  less than 0.1 mills per kwh.
A detailed  study for a particular site and a particular set of conditions would be
required in order to select the type of dry cooling system to be used at that site. In
some cases, particularly if there is a shortage of capital, a mechanical-draft system
may be selected over a  natural-draft system on the basis of the capital cost differ-
ence .

        The combined capital cost of the dry cooling system and  the required peak-
ing capacity, both  at optimum ITD, as shown on Figures 54 and 55 for fossil-fueled
units is generally in the range of $22 to $28 per kw. The  corresponding values for
nuclear-fueled units are shown on Figures 56 and 57.

        The effects  of the various parameters on  the economic optimization analyses
have been  studied and are discussed  below.

Effect of Fixed-Charge Rate

        The effect of increasing  the fixed-charge rate is to increase the value of
the  economically optimum ITD and,  therefore,  to reduce the cooling system invest-
ment.   This effect is clearly shown in Tables  11, 12, 13 and  14.  In these tables,
the  optimum ITD at a fixed-charge rate of 18 percent is a few degrees higher than
the  optimum ITD at a fixed-charge rate of 8 percent.

Effect of Fuel Cost

        The effect of increasing the plant fuel cost is to reduce the optimum ITD.
This reduction in the optimum ITD increases the cooling system investment and  im-
proves the plant efficiency.  The range of fuel costs investigated in our analyses
had only a minor effect on the optimum ITD.  In many cases,  as shown in Tables 11
through 14, the optimum ITD was not sufficiently affected by the fuel cost to change
the  ITD by a full degree F.   In other cases, the fuel cost difference caused the op-
timum ITD value to vary over a  range of 1°F to  4°F for a given fixed-charge rate.

        In  order to  test the effect of varying the fuel cost over a somewhat wider
range, some supplemental runs were made for the Chicago site assuming fossil-fuel
costs ranging from  10
-------
                                    TABLE 32

                       Effect of Fuel Cost on Optimum ITD
               (Chicago,  fossil-fueled plant,  15% fixed-charge rate)

              Fuel Cost         	Optimum ITD (°F)
             (
-------
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PHOENIX
92






JD: 1. FIXED CHARGE RATE = 15%
2. FUEL COST=250/IO« BTU
3 ACTUAL SITE ELEVATION
4. CAPITAL COST MULTIPLIER
APPLICABLE TO SITE
  80      85      9O     95      IOO      IO5     110      115
AIR TEMPERATURE EQUALLED OR EXCEEDED TEN HOURS PER YEAR (°F)

FIGURE 58-RELATIONSHIP OF ECONOMICALLY OPTIMUM INITIAL
 TEMPERATURE DIFFERENCE TO AMBIENT AIR TEMPERATURES
 FOR THE SITES STUDIED	NATURAL-DRAFT DRY COOLING
 SYSTEM FOR A FOSSIL-FUELED 800 MW  GENERATING UNIT
                        195

-------
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OPTIMUM ITD VALUES
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- • ' ,





Z. FUEL COST= 25 0/10 6 BTU
3. ACTUAL SITE ELEVATION
4. CAPITAL COST MULTIPLIERS
APPLICABLE TO SITE

   80      85      90     95      100      105     110      115
 AIR TEMPERATURE EQUALLED OR EXCEEDED TEN HOURS PER YEAR (»F)
 FIGURE 59-RELATIONSHIP OF ECONOMICALLY OPTIMUM INITIAL
  TEMPERATURE  DIFFERENCE TO AMBIENT AIR TEMPERATURES
FOR THE SITES STUDIED	MECHANICAL-DRAFT DRY COOLING
 SYSTEM FORA FOSSIL-FUELED 800 MW GENERATING UNIT
                         196

-------
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66


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65
65
OREEN BAY

OPTIMUM ITD VALUES
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•56


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9
61


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63 ••SB
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PHOENIX
56


























LEGEND: 1 FIXED CHARGE RATE= 15%
2. FUEL COST: 1507 10 • BTU
3. ACTUAL SITE ELEVATION




4. CAPITAL COST MULTIPLIER
APPLICABLE TO SITE





  60      85      90     95     100      105      110      115
AIR TEMPERATURE EQUALLED OR EXCEEDED TEN HOURS PER YEAR (°F)

FIGURE 60-RELATIONSHIP OF ECONOMICALLY OPTIMUM INITIAL
 TEMPERATURE DIFFERENCE TO AMBIENT AIR TEMPERATURES
 FOR THE SITES STUDIED — NATURAL-DRAFT DRY COOLING
SYSTEM FOR A NUCLEAR-FUELED 800 MW GENERATING  UNIT
                        197

-------
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OPTIMUM ITD VALUES
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PHOENIX
59

























40 \. FIXED CHARGE RATE= 15%
Z. FUEL COST: |5£/IO« BTU
3. ACTUAL SITE ELEVATION
* CAPITAL COST MULTIPLIER
APPLICABLE TO SITE [
  60      85      90     95     100      105      110      |)5
AIR TEMPERATURE EQUALLED OR EXCEEDED TEN HOURS PER YEAR (°F)

FIGURE 61-RELATIONSHIP OF ECONOMICALLY OPTIMUM INITIAL
 TEMPERATURE  DIFFERENCE TO AMBIENT AIR TEMPERATURES
FOR THE SITES STUDIED—MECHANICAL-DRAFT DRY COOLING
SYSTEM FOR A NUCLEAR-FUELED 800 MW GENERATING UNIT
                       198

-------
  80
  75
  70
£  65
UJ
(T
CC.
UJ
a.
2
ui
t-
jr
UI
  60
  50
   45
   40
               ECONOMICALLY OPTIMUM  ITD  VALUES

                             92

                       ±.
                      93
                          96
                            V
                              • 56
                                  96
                                       • 99
                                     ..96
                                    • 96
                                    \"
                                     X56
                                                        91
                                                        52
                       LEGEND: i. FIXED CHARGE RATE = is %
                               2. FUEL COST: 250/IO* BTU
                               3. ELEVATION = 0
                               4. CAPITAL COST MULTIPLIERS* 1.0
                                           -t-
   35
    80      85       90      95      IOO      105      110      115
  AIR TEMPERATURE EQUALLED OR EXCEEDED TEN HOURS PER YEAR (°F)


  FIGURE 62-RELATIONSHIP OF ECONOMICALLY OPTIMUM INITIAL

   TEMPERATURE DIFFERENCE TO AMBIENT AIR TEMPERATURES

  AT SEA-LEVEL ELEVATION—NATURAL-DRAFT DRY COOLING
   SYSTEM FOR A FOSSIL-FUELED 800 MW  GENERATING UNIT
                            199

-------
 80
             ECONOMICALLY OPTIMUM  ITD VALUES
                     LEGEND:  i FIXED CHARGE RATE = is %
                             2. FUEL COST= 251/10 « BTU
                              ELEVATION^ 0
                             4. CAPITAL COST MULTI PLIERS = 1.0
  80      85      90      95      100      105     110      115
AIR TEMPERATURE EQUALLED OR EXCEEDED TEN HOURS PER YEAR («F)

FIGURE 63-RELATIONSHIP OF ECONOMICALLY OPTIMUM INITIAL
 TEMPERATURE  DIFFERENCE TO AMBIENT AIR TEMPERATURES
AT SEA-LEVEL ELEVATION— MECHANICAL- DRAFT DRY COOLING
SYSTEM FOR A FOSSIL-FUELED 800 MW GENERATING UNIT
                         200

-------
                 I       I        I       I
             ECONOMICALLY OPTIMUM 1TD VALUES
                          53
                   53     ~"
                    LEGEND: I  FIXED CHARGE RATE = 15%
                            2. FUELCOST= I50/I06 BTU
                            3. ELEVATION = 0
                            4. CAPITAL COST MULTIPLIERS* 1.0
  80       85      90      95      100     105      110      115
AIR TEMPERATURE EQUALLED OR EXCEEDED TEN HOURS PER YEAR (°F)

FIGURE 64-RELATIONSHIP OF ECONOMICALLY OPTIMUM INITIAL
 TEMPERATURE DIFFERENCE TO AMBIENT AIR TEMPERATURES
AT SEA-LEVEL ELEVATION— NATURAL- DRAFT DRY COOLING
SYSTEM FOR A NUCLEAR-FUELED 800 MW GENERATING UNIT
                         201

-------
   80
   75
   7O
   65
o

LU
<

LU
Q.
2
LU
 CC 55
   60
 LU
   50
   45
   40
                ECONOMICALLY OPTIMUM ITD VALUES
                      54
                      •
                         I 60
                     62*
                     62»
                       62-64
                               • 5«
                              • 60
                                60
                                     60
                                             55-57
                                                       51 '
                       LEGEND: i. FIXED CHARGE RATE-i5%
                               2. FUELCOST= I5^/IO« BTU
                               3. ELEVATION = 0
                               4. CAPITAL COST MULTIPLIERS* 1.0
   35-
    80      85      90      95       100     105      110      115
  AIR  TEMPERATURE EQUALLED OR EXCEEDED TEN HOURS PER YEAR(°F)

  FIGURE 65-RELATIONSHIP OF ECONOMICALLY OPTIMUM INITIAL
   TEMPERATURE  DIFFERENCE TO AMBIENT AIR TEMPERATURES
AT SEA-LEVEL ELEVATION—MECHANICAL-DRAFT DRY COOLING
  SYSTEM FOR A  NUCLEAR-FUELED 800 MW GENERATING UNIT
                           202

-------
replaced at all sites other than the Anchorage, Alaska site and that the cost of that
replacement capacity is $100 per kw.  The effect of varying the cost assumption is
indicated by the results of supplemental analyses made for the Chicago site. Assum-
ing a base condition for a fossil-fueled plant using a fuel cost of 35$ per million Btu
and a fixed-charge rate of 15 percent, the effect of three different peaking capacity
costs—$75 per kw, $100 per kw and  $150 per kw—was studied. The results of these
analyses are indicated  in Table 33.

                                    TABLE 33

                 Effect of  Peaking Capacity Cost on Optimum ITD
                          (fossil-fueled plant, Chicago,
             fuel cost  - 35$ per million  Btu, fixed-charge rate - 15%)

               Peaking
            Capacity Cost        	Optimum ITD (°F)	
               ($/kw)           Natural Draft      Mechanical Draft

                 75                   58                  64
                100                   57                  61
                150                   54                  55

        Any method of  reducing the capital cost of  replacing the lost capacity such
as the use of dual inlet turbines, or the bypassing of feedwater heaters would tend
to increase the optimum ITD and,  therefore, reduce the cooling system investment.
It is recognized that it may be possible to replace the lost capacity for considerably
less than the lowest cost shown in Table  33, $75 per kw.

        The  effect of varying the assumption as to winter or summer peak was also
investigated.  If a winter peak is  assumed for Chicago, and therefore the lost capa-
city is not replaced, the optimum ITD for the fossil-fueled plant with a natural-
draft tower is at some point above 80°F as compared to the 55°-5/  F range, and the
optimum for a fossil-fueled plant with mechanical-draft tower is 78°F as compared
to a 59°-61  F range.   On the other hand, if a summer peak is assumed for
Anchorage, Alaska, the optimum  ITD fora fossil-fueled plant with natural-draft
tower becomes 68°F as compared to a range of 80°F and above, and the optimum
point for the mechanical-draft tower  is 71 F as compared to a range of 79°-80°F.
                                    203

-------
                                  SECTION XII

               ECONOMIC COMPARISON OF THE DRY-TYPE AND
                  THE EVAPORATIVE-TYPE COOLING SYSTEMS
        In this report information has been presented as to the theory of dry-type
cooling as it would apply to steam-electric generating plants; the operating results
have been summarized for several existing dry cooling tower installations; the com-
ments of equipment manufacturers have been summarized; and economic analyses
have been made to indicate the economic factors which should be considered i-n the
determination  of the type of cooling system to be used for future generating plants.

        To illustrate,  the capital cost, based  on  1970  prices, of a dry-type,
mechanical-draft cooling system for use with  a fossil-fueled generating plant plus
the capital cost of the required supplemental  peaking  capacity would be  approxi-
mately  $25 per kw for the Chicago area, as indicated on Figure 55.  In contrast,
the cost for a wet-type, mechanical-draft cooling  system may be on the order of
$11 per kw and there  would be no requirement for additional peaking capacity to
provide comparable station output capacity.  This  would indicate a capital cost
penalty for the dry tower installation of $14 per  kw which is equivalent to approxi-
mately 0.34 mills per kwh based on a 15 percent fixed-charge  rate, a 1 percent
operating and  maintenance charge,  and the  energy generation assumed for these
analyses. The analyses of the dry cooling systems indicate  that the higher operat-
ing back pressures  of such systems coupled with the requirement for some energy
generation  by  peaking units during periods of hot weather would  result in an in-
crease in annual fuel  cost of approximately 1  .8 percent as compared to a wet cool-
ing tower installation. Assuming a fuel cost of 35$ per million Btu and a plant heat
rate of 9,000 Btu per  kwh, the 1 .8 percent difference is equivalent to about 0.06
mills per kwh.  In  addition, the fan power requirements of the  dry cooling tower in-
stallation would be somewhat greater than for a wet tower and  this may result in an
additional penalty of  about 0.08 mills per kwh for  a total cost difference of about
0.48 mills per kwh.  Table 34 lists a comparison  of bus-bar costs of a dry-type ver-
sus evaporative-type cooling system.  This cost difference, if not offset by some of
the factors discussed below, is equivalent to about 7 to 10 percent of the cost of
oower and energy at the generating station bus bar.  This is approximately 2  or 5
aercent of the cost of power and energy at retail, reflecting all costs of generation,
transmission and distribution.  This indicates that,  even without the benefit of po-
tential cost savings discussed below, the impact  of dry cooling on retail electric
bills would be small.  A 2 to 5 percent increase in  a $20 monthly electric bill is
equivalent to only 40£ to $1 .00 and an  increase  of even this magnitude would not
occur unless all generating plants of a given utility were cooled  by a dry-type
cooling system.
                                      204

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                                    TABLE 34

                          Comparison of Bus-Bar Costs
              Dry-Type and Evaporative-Type Cooling Tower Systems
            Mechanical-Draft, 800-Mw  Fossil-Fueled Generating Unit
                                 Chicago Area

                           	Mills per Kwh	
                            Dry-Type       Evaporative Type
                             System         _ System            Difference

Plant Fuel Cost  .........     3.210               3.153               .057

Cooling System Auxiliary
Power Costs ............     0.140               0.062               .078

Cost of Capacity
Replacement ............     0.193               0.000               .193

Cooling System Capital,
Operation and Mainten-
ance  Costs  .............     0.418               0.268               .150
                                                         Net Diff:   ~478~

(1)  The costs  shown in this  table reflect the study assumptions as summarized in
    Table 10.

(2)  The annual average turbine heat rate with a dry-type tower is estimated to be
    approximately 1 .8% higher than with an evaporative-type tower due to higher
    average back pressure operation, 9,170 Btu/kwh compared to 9,010 Btu/kwh.
    The above plant fuel  costs reflect  this difference.

(3)  The mechanical-draft, dry-type cooling system capital cost used to develop
    annual  costs in  this  table is $17.15/kw and the evaporative cooling system
    capital  cost — $1 1 .
(4)  The  cost figures in this  table are based upon a 15% fixed-charge rate,
    10°  Btu steam turbine fuel cost and 40$/10° Btu gas turbine fuel cost.

(5)  Weather data used is listed in Table A- VI of Appendix B.

(6)  The  evaporative-type system analyzed is based upon a  range of 24 F, an
    approach of 18^F and a terminal temperature difference of 6°F.

(7)  The  dry-type system initial temperature difference is 61°F.
                                   205

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        As shown above, the cost penalty of the dry-type cooling tower system is a
small percentage of the total cost of power and energy delivered to the customer,
even if no offsetting cost savings are assumed.  It should be pointed out, however,
that, since the dry cooling system requires very little water, an electric utility
would have  much more flexibility in locating its generating plant and would have
the opportunity for certain cost savings.   For example, a saving in fuel cost of
approximately 5<£ per million Btu would entirely offset the cost difference of 0.48
mills per kwh computed above.

        The  greater flexibility which is afforded by  the dry cooling system may make
possible transmission savings which would offset a portion or all of the  cost differ-
ence between the wet and dry systems, or may permit an additional generating unit
to be built at an existing station even though there  is not adequate water for an
additional wet cooling tower.  This would  permit the utility to realize the econo-
mies of an additional unit at an existing  facility.

        The  most obvious cost saving is that related  to cooling water.   If it is as-
sumed that cooling water costs $100 per acre-foot (about 31 cj: per thousand gallons),
water cost savings, alone,  for the dry tower installation would approximate 0.2
mills per kwh.

        Based on the above, when all factors are considered, it appears that in many
cases the  dry cooling system would be economically competitive with wet cooling
tower systems, and in some cases  the dry system may have a decided economic ad-
vantage.  The economic  advantage would be most pronounced in those cases where
the use of a  dry-type cooling tower would allow the utilization of  low-cost fuel in
a water-short area.

        In some cases, the relative economics of dry versus wet cooling will be
overshadowed by  pollution control considerations, and in these instances the dry
cooling tower would, of  course, have an advantage.  The closed cycle of the  dry
cooling system means that thermal pollution of lakes, rivers, streams and the ocean
from power plant waste heat would not occur. In addition, since there is no evap-
oration  of water from the dry cooling system to increase  the concentration of solids
in the cooling water, there is no need for blowdown and,  therefore,  no danger of
discharge  of pollutants to the waterways .
                                    206

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                                SECTION XIII
                                 REFERENCES
1 .    "National  Power Survey", A Report by the Federal Power Commission, 1964,
      U.S. Government Printing Office.

2.    "Considerations Affecting Steam Power Plant Site  Selection", A Report spon-
      sored by the Energy  Policy Staff Office of Science and Technology in coop-
      eration with Atomic  Energy Commission, Department of Health,  Education
      and Welfare, Department of the Interior;  Federal  Power Commission,  Rural
      Electrification  Administration,  Tennessee Valley Authority, December,  1968.

3.    "Industrial Waste Guide  on  Thermal  Pollution", U. S. Department  of  the
      Interior, Federal Water Pollution Control Administration, Northwest Region
      Pacific Northwest Water Laboratory, Corvallis, Oregon,  September,  1968.

4.    Olds, F. C. "Thermal Effects, A Report on Utility Action", Power Engineer-
      ing, April, 1970.

5.    Hauser, L. G. and Oleson, K. A. "Comparison of Evaporative Losses in
      Various Condenser Cooling Water Systems",  American Power Conference,
      1970.

6.    Parker,  Frank L. and Krenkel, Peter A.  "Thermal Pollution: Status of the
      Art", Vanderbilt University, prepared for the Federal Water Pollution Con-
      trol Administration,  December, 1969.

7.    "The Conservation Foundation  Letter 3-70", March, 1970.

8.    "Report of the  Committee on Water Quality Criteria", Federal Water Pollu-
      tion Control Administration, 1968.

9.    Heeren, Hermann and Holly,  Ludwig . "Air Cooling for Condensation and
      Exhaust Heat Rejection in Large Generating Stations", American Power
      Conference, 1970.

10.  Heller, Prof.  Dr. Sc. Techn.  L. and Forgo,  Ing.  L., Budapest.
      "Betriebserfahrungen mit einer Kraftwerks-Kondensationsanlage mlt
      luftgekuhltem  Kuhlwasserkrelslauf und die Moglichkeiten der
      Weiterentwicklung", World Power Conference, Vienna,  1956.
                                      207

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11 .    "Why CPI is Warming to Air Coolers", Chemical Week, July 5, 1969.

12.    Private Correspondence with E. C. Smith and W. F. Berg, Hudson Products
      Corporation,  Houston, Texas.

13.    Smith, Ennis C. and Larinoff,  Michael W.  "Power Plant Siting Performance
      and Economics with Dry Cooling Tower  Systems",  American Power Confer-
      ence, 1970.

14.    Mathews, Ralph T.  "Some Air Cooling Considerations", American  Power,
      Conference, 1970.

15.    "Dry Cooling Tower Condensing  Plant",  English Electric Company,  Publica-
      tion ST/120.

16.    Gardner, K.  A.   "Efficiency of  Extended Surface", Transactions ASME
      Volume 67.

17.    Kays, W. L.  and  London, A.  L., Stanford University; "Compact Heat
      Exchangers", published by National Press.

18.    "ASHRAE Guide and Data Book, Fundamentals and  Equipment", published by
      American Society of Heating,  Refrigerating and Air Conditioning Engineers.

19.    Cheshire, L.  J. and Daltry, J. L.  "A Closed Circuit Cooling System for
      Steam Generating Plant", The South African Engineer,  February, 1960.

20.    Private Correspondence with R. E. Cates,  Senior Evaluations Engineer, The
      Marley Company, Kansas City, Missouri.

21 .    Bowman, R. A., Mueller, A.  C. and Nagle, W. M.  "Mean Temperature
      Difference in Design",  Transactions ASME, May, 1940.

22.    McAdams, W. H.  "Heat Transmission", 1942, published by McGraw-Hill.

23.    Dukler, A. E. CEP Symp. Series 56 (30) 1-10 (1960).

24.    Kirkbride. Transactions of AICHE 30, 170-186 (1933-1934).

25.    Akers, W. W., Deans, H. A. and Grosser, O.K.  CEP Symp.  Series 55
      (29) 171-176(1959).

26.    Bartlett, R. L.  "Steam Turbine Performance and Economics", 1958, pub-
      lished by McGraw-Hill.
                                  208

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27.   Babcock and Wilcox Steam Book.

28.   Schroder, Karl.  "Grosse Dampfkraftwerke,  Plannung,  Ausfuhrung und Bau",
      Drifter Band, Tell B., 1968, published by Springer-Verlag.

29.   "Surface Water Temperature and  Salinity — Atlantic Coast — North and
      South America", C and GS Publication 31-1, Second Edition, 1965, U.S.
      Department of Commerce.

30.   Heller,  Prof. Dr. Sc. Techn. L.   "Series Connection of Jet Condensers on
      the Cooling Water Side", World Power Conference,  1968.

31.   Christopher, P. J.  "The Dry Cooling Tower System at the  Rugeley  Power
      Station of the Central Electricity Generating Board", English Electric Journal,
      February, 1965.

32.   Christopher,  P. J. and Forster,  V. T. "Rugeley Dry Cooling Tower System",
      The Institution of Mechanical Engineers—Steam Plant Group, October, 1969.

33.   Goecke, Direktor Dipl.-lng. Ernest; Gerz, Dipl.-lng.  Hans-Bernd; Schwarze,
      Dipl.-lng.  Win fried; and Scherf,  Dipl.-lng.  Ottokar.   "Die  Kondensation-
      sanlage des 150-Mw-Blocks im  Kraftwerk  Ibbenburen  der  Preussag AG",
      V.I .K. — Berichte — Nr. 176, May,  1969, published by Vereinigung Indus-
      trie! le Kraftwirtschaft (V.I .K.) 43 Essen, Richard-Wagner-Strassee 41 .

34.   Heller, Prof. Dr. Sc . Techn. L.  "The Possibilities Offered by Artificial
      Cooling for Increasing the Capacity of Electric Generators", World Power
      Conference, 1958,  Montreal, Canada.

35.   Slusarek, Z. M. "The  Economic  Feasibility of the Steam-Ammonia Power
      Cycle", Franklin Institute Research Laboratories,  Philadelphia,  Pa., pre-
      pared for the Office of  Coal Research, Department of  the Interior.

36.   Aynsley, E. "Cooling Tower Effects: Studies Abound", Electric World,
      May  11, 1970.

 37.  Appelman,  H.  S. and Coons, F.  G. "The Use of Jet  Aircraft Engines to
       Dissipate Warm  Fog", Journal of  Applied Meteorology, June, 1970.

 38.   Fritschen, L.,  Bovee, H., Buettner, K. and others.  "Slash Fire Atmos-
       pheric  Pollution",  USDA Forest Service Research  Paper PNW-97, 1970.
                                     209

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39.   Waselkow, C.  "Design and Operation of Coofing Towers",  Federal Water
      Pollution Control Administration and Vanderbilt University sponsored sym-
      posium on thermal pollution,  1968.

40.   "Hydroelectric Power Evaluation", Federal Power Commission, 1968.

41 .   "Climatography of the United States No. 82 —Decennial Census of United
      States Climate  —  Summary of Hourly Observations", U.S. Department of
      Commerce, Weather Bureau,  1962-1963.
                                  210

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                            APPENDIX FOREWORD
       This appendix contains background  information and data which were used in
the analyses made for this study.

       Appendix A  describes visits to  existing steam-electric generating plants
which  are  equipped with dry-type cooling towers to obtain first-hand information
relative to the operation, maintenance and construction costs of these plants.  Five
stations equipped with dry-type cooling towers were visited—the Rugeley Station in
England; the Ibbenburen and Volkswagen plants in Germany; the Gybngyos Station
in Hungary, and the Neil Simpson Station in Wyoming. At the time of these visits,
these stations had the largest electric utility operating units utilizing dry-type cool-
ing towers.  The operating experience gained from these existing cool ing systems can
be of value to those contemplating future installations.

       Appendix B summarizes the ambient air  temperature data  utilized in the eco-
nomic  optimization analyses  for  the 27 United States sites considered, and refers to
the source of this data.  Temperatures were analyzed from  -40°F to+119°F in 5°
increments to determine their effect upon turbine back pressure and plant operating
efficiency.

        Important considerations to determine the economic choice of cooling system
were summarized in Appendix C,  "General Specifications  for Dry-Type  Cooling
System Applications", as a guide  for future installations.

       Appendix D covers testing aspects for completed dry-type tower installations.

        Procedures for developing cooling system costs used  in  the  report are out-
lined in Appendix E.
                                     211

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                             APPENDIX FIGURES

                                                                        Page

Al     Rugeley Generating Station, Site Layout  	     217

A2     Rugeley Power Station  	     218

A3     Diagrammatic Arrangement of Water Circuit	     219

A4     Design Performance Charts for Rugeley Natural-Draft
       Cooling Tower 	     221

A5     Inside View,  Natural-Draft, Dry-Type Cooling Tower —
       Rugeley Station (English Electric Photo) 	     224

A6     Sector Valves, Valve House Located Inside Natural-Draft,
       Dry-Type Cooling Tower- Rugeley  Station (English Elec-
       tric Photo)  	     226

A7     Wind  Effect Upon Performance  	     230

A8     Air Mass Velocity Variation Through the Coolers With a
       20-mph Wind	     230

A9     Plan View of  Preussag Power Station, Ibbenburen  	     233

A10   Cooling Water Circuit Diagram— 150-Mw, Ibbenburen
       Generating Station 	     235

Al 1    Direct-Contact Condenser	     237

A12   Ibbenburen Plant— Natural-Draft Dry-Type Cooling Tower
       Turbine Back  Pressure Variation With Ambient Air Tem-
       perature 	     238

A13   Hyperbolic Concrete Dry-Type Cooling Tower Installation at
       Ibbenburen— 150-Mw Generating Plant 	     240

A14   Operational Control Instrument	     241

A15   Wind  Effect Upon Natural-Draft Tower Performance	     244

A16   Dry-Type, Natural-Draft Cooling Tower:  Ibbenburen Plant
       Performance Test Results	     246
                                      212

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                             APPENDIX FIGURES

                                                                       Page

A17    Volkswagen Plant With Direct-Type Air-Cooled Condenser
       Units on Plant Roof (V-W Photo)	     250

A18    Exhaust Steam and Condensate Plant of Air-Cooled Condensing
       System — Volkswagen Plant 	     252

A19    Calculated Operating Characteristics (Predicted Performance)
       for the Direct Air-Cooled Condensing  System — Block "C" of
       Volkswagen Plant (from GEA)	     254

A20    Gyongyos Power Station  — Two Reinforced Concrete Dry-Type
       Cooling Towers for 100-Mw Generating Units	      266

A21    Reinforced Concrete Tower for First of Two 200-Mw Generat-
       ing Units in the Gyongyos  Power Station  	      267

A22    Water Circuit for Heller Dry Tower, Gyongyos Station  	      268

A23    3,000-Kw Pilot Plant Direct-Type, Air-Cooled Condenser
       Installation — Neil Simpson Plant,  Wyodak, Wyoming  	      275

A24    Side View of A-Frame Direct-Type Air-Cooled Condensing Unit —
       20-Mw Generating Unit, Neil  Simpson Plant,  Wyodak, Wyoming      275

A25    Side Walls Erected Around Direct-Type Air-Cooled Condensing
       Unit — 20-Mw Generating Unit, Neil Simpson Plant,  Wyodak,
       Wyoming  	     277

A26   Steam Headers and Hail  Screens — Direct-Type, Air-Cooled
       Condensing Unit — 20-Mw Generating Unit, Neil  Simpson Plant,
       Wyodak, Wyoming;	     277

A27    Fan Arrangement for Direct-Type Air-Cooled Condensing System —
       20-Mw Generating Unit,  Neil Simpson Plant, Wyodak, Wyoming      279

A28   Outline of Natural-Draft Tower (for a Dry-Type Cooling System for
        Use With an 800-Mw Fossil-Fueled Generating Plant at 6,000 Feet
        Elevation) Using Steel and Aluminum  Construction  	     318
                                      213

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                              APPENDIX TABLES

                                                                        Page

A-l      Operating Data — Rugeley Power Station,  120-Mw Turbine
         Generator - Unit No. 3 With Natural-Draft Dry Cooling
         Tower 	      227

A-l I      Operating Data — Preussag-Kraftwerk, 150-Mw Turbine
         Generator — Ibbenbiiren, Natural-Draft Dry Cooling
         Tower 	      247

A-lll     Operating Data — Power Station  "Wolfsburg" of the
         Volkswagenwerk AG., 49-Mw Automatic-Extraction
         Turbine-Generator and Air-Cooled Condenser	       255

A-IV     Operating Data — Neil Simpson Station, 20-Mw Turbine-
         Generator With Mechanical-Draff, Direct Air-Cooled
         Condensing System  	      281

A-V     Economic Optimization Analysis, Site  Summary	      284

A-VI     Annual Distribution of Air Temperatures  	      285
                                     214

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                                 SECTION XIV

                                 APPENDICES


                                  Appendix A

                  Field Trips to Dry Cooling Tower Installations


                              RUGELEY STATION


Introduction

       On December 15, 1969  John P. Rossie, accompanied by Mr. D.  W. Crane,
Project Engineer for English Electric Company, visited Rugeley Station.  They were
escorted  by Mr. Platt, Assistant Superintendent of the station.  At the time of the
visit, Unit No. 3, which is equipped with the dry-type cooling tower, was in opera-
tion and carrying approximately 80 mw load.  The air temperature was approxi-
mately 40°F and vacuum was approximately 1 .5inches Hg on Unit No.  3.  Although
Rugeley Station was operated  as a base-load plant for the first few years after com-
pletion, it is now operated on a load-factor basis since larger, more efficient units
operate base loaded. Unit No. 3 is called  upon to operate at 120 mw during system
peaks.

Description of Station

       Rugeley Station of the Central Electricity Generating Board is located in the
West Midlands Division adjacent to the Town of  Rugeley, England.  The original
station, now designated as Rugeley Station "A", has a total generating capability
of 600 mw, comprising five 120-mw units.   All  units are designed for an over-all
thermal efficiency of 34.2 percent (9,980  Btu per kwh) and have throttle steam
conditions of 1,500 psi, 1,000°F/1,000°F.

       The station is at the site of the Lea  Hall Colliery, which has been in opera-
tion for 600 years.   Construction  of Rugeley  Station was started in July, 1955 and
was completed in December,  1962.  Since  the completion  of Rugeley  Station "A",
a new station—Rugeley "B", with two 500-mw units—has been constructed at  the
same site,  but is not physically  connected  to the  original plant.

        With the exception of Unit No. 3 of Station "A", all the turbine-generators
are equipped with surface condensers and evaporative-type cooling towers with re-
inforced, concrete,  natural-draft towers of hyperbolic form.  Make-up water for
                                      215

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 the  evaporative-type cooling towers is  pumped  from the River Trent, which flows
 past the site.

        The 120-mw Unit No. 3,  which was commissioned in  December, 1961  and
 has been in operation since,  is equipped  with a dry-type cooling tower of the Heller
 system design utilizing a concrete, hyperbolic, natural-draft tower.  At the time of
 its construction, Rugeley Unit No. 3 was the largest Heller-type dry tower to be
 built.   Up to that time, the largest such unit was a small pilot plant in Hungary.

        Figure Al  (3A)  illustrates the site layout of the station  and  Figure A2 is an
 aerial  view of the plant.  Note the difference in physical size  of the  dry tower on
 the right as compared to the size of conventional  cooling towers serving units of the
 same mw rating.  Approximately three times more air is moved through the dry tower
 than through each of the evaporative towers. The four evaporative-type cooling
 towers are each 350 feet high with base diameter of 216 feet.   The dry-type cooling
 tower, which at the time of its construction was the largest concrete tower shell in
 the world, is 350 feet high and 325 feet in diameter at the  base.

 Water  Circuit
        Figure A3 shows a diagrammatic arrangement of the circuit in which  62,000
gpm of condensate quality water circulates (2A).  There are four sectors in the cool-
ing coils of the tower, each of which can be independently drained and filled with
the other sectors in operation.

       Two half-capacity circulating water pumps are provided to pump the  water
to the  tower, and also to handle the  small percentage of the flow which goes to the
boiler  feedwater system,  amounting to approximately 3 percent of the total flow.
The circulating water is conveyed to the tower through 60-inch-diameter pipes and
is directed to each of the four equal  cooling coil quadrants through specially de-
signed  sector valves.

       From the sector valves, the water passes through the 48-foot-high columns
of coolers.  The Forgo coil used  in the  Heller system has a depth of six  rows of
tubes; water flow is upward in the inner three rows from the bottom of the column to
the top, and the flow direction is reversed in the top water box downward  through
the outer three rows of tubes.  Since the cooling air is flowing horizontally across
the vertical tubes and comes into contact first with the lower temperature water,
the system is designated as cross-counterflow.

       After leaving the tower,  the  cooled water again passes through  the sector
valves  and then through the two  half-capacity  recovery turbines which  are con-
nected to  the same shaft as the circulating water pumps and motors. The purpose of
the recovery turbines is to furnish a portion of the work necessary to drive  the main
                                     216

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to
         FIGURE Al
                                                         RUGELEY GENERATING STATION
                                                                   SITE LAYOUT
(3A)

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FIGURE A2—RUGELEY  POWER STATION

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                     STEAM TURBINE (I2O MW)
SPRAY VALVES
WATER
TURBINE-*.^
                                            NATURAL DRAUGHT1
                                            COOLING TOWER
                            AUXILIARIES
         -^CIRCULATING WATER
           EXTRACTION PUMPS
    QUADRANTS
      SECTOR
      VALVES
    •TO BOILER WATER
    EXTRACTION PUMPS
TRANSFER
VALVE-^
                        EMERGENCY DRAIN VALVE
                                   BYPASS
                                   VALVE
                                             COOLING WATER STORAGE TANK
          FIGURE  A3—DIAGRAMMATIC  ARRANGEMENT
                      OF WATER  CIRCUIT  (2A)
                                219

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circulating water pumps.  An excess pressure of a few pounds per square inch is
maintained at the top of the 45-foot cooling coil columns in order to have positive
water pressure on the coils to prevent ingress of air in case of leaks. The recovery
turbines utilize the major part of the excess pressure head to help drive the circu-
lating water pumps.

        From  the recovery turbines, the water flows to the direct-contact condenser,
through the spray nozzles where it condenses the exhaust steam from the turbine, and
then is recycled through the cooling circuit.

        In order to provide for quick drainage of the cooling circuit, a necessity in
case operation of the unit is curtailed during freezing weather, an underground stor-
age tank is located inside the base of the cooling tower.  Two transfer pumps are
used for transferring  condensate and filling the coil sectors.

Design Parameters

       Apparently,  the dry-type cooling tower was constructed at Rugeley for the
purpose of obtaining experience with cooling towers which do not require a large
amount of make-up water in anticipation of a shortage of water for  power plant use
in England.   The desire to be able to construct power generating stations near a
source of fuel or near a load center without being dependent upon an adequate sup-
ply of make-up water for a wet-type cooling tower was also a factor in the decision
to obtain operating experience with a dry tower  in England.

       The turbine back pressure design at Rugeley No.  3 is 1 .3  inches Hg with
52°F ambient air temperature, which is the same as the design of  the other four 120-
mw units equipped with evaporative-type cooling towers. The design initial  tem-
perature difference between the saturated steam  temperature and the ambient air
temperature is therefore 35°F, since the saturated steam temperature corresponding
to an absolute pressure of 1 .3 inches Hg is 87°F.  The tower design heat rejection
load is 587 million Btu per hour.  Apparently the back pressure was not optimized
but was  selected so that generating plant equipment similar to the conventional 120-
mw units could be utilized with the dry tower.

       Figure A4 shows the design performance curves of the dry  tower for one, two,
three and  four quadrant operation.

Capital  Costs of the  Dry Tower

       No figures are available as to the construction costs of the dry tower  system
at Rugeley Station.  However, representatives of the English Electric Company (the
contractors for the equipment) advise that,  in general, the components of a dry-type
tower are  from one and one-half to two times the cost of the components of an
evaporatiye-type tower.
                                      220

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  0  10  20 30  4O SO  60 70  SO »O
  AMBIENT AIR TEMPERATURE (°F)
   ONE QUADRANT OPERATION
0  10  20  30 40  50 80  70 80  90
 AMBIENT AIR TEMPERATURE (°F)
 TWO QUADRANT OPERATION
120
100
                             ISO
   AMBIENT AIR TEMPERATURE (°F)
  THREE QUADRANT OPERATION
   10 20  30 40 90 60 70 80 90
 AMBIENT  AIR TEMPERATURE (°F)

 FOUR QUADRANT OPERATION
NOTE:
SOLID LINES REFER TO OPERATION WITH TWO PUMPS  AND
DASHED  LINES WITH  ONE PUMP.
Tw = 45° F  REFERS TO THE MINIMUM  AVERAGE COOLER
WATER OUTLET TEMPERATURE PERMITTED TO SAFEGUARD
THE  COOLERS FROM FREEZING.
    FIGURE A4 —DESIGN  PERFORMANCE CHARTS
  FOR RUGELEY NATURAL-DRAFT COOLING TOWER (2A)
                        221

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Manpower Requirements of the Tower

       There are no special manpower requirements associated with operation of the
dry tower at Rugeley because of the  central control system which is installed with
the tower. With the exception of the steam-jet air ejectors, all starting and con-
trol functions are carried out from the control  room of the station. The time neces-
sary to bring the dry tower system to operating vacuum level is approximately 10
minutes, which  is the same time required for the units equipped with surface con-
densers .

       There are no problems associated with  establishing an air flow through the
tower.  Water is first circulated through the windward cooling section, or through
two opposite sections if there is no wind.  After the tower is in service, no operat-
ing functions are required unless the low-temperature alarm on the circulating water
sounds,  at which time the operator initiates the coil drainage sequence to take one
or more cooling  coil sections out of service.   When conditions permit, the tower
sections are returned to service from the control room.   The process of removing sec-
tors from service, restoring them to service and starting  up and shutting down the
dry tower system is accomplished by  the automatic sequential control  system which
is initiated by control-button operation.

Winter Operation

       Except for some minor instances of coils freezing as a result of automatic
vent valves not operating,  and  faulty low-temperature alarms, operation of  the dry
tower during freezing weather has been satisfactory.  Despite the freezing problems
during the first winter' s operation, Unit No.  3 was able to generate up to 137 mw
in very severe weather at a time when  the other four units with evaporative-type
cooling towers were having operational difficulties because of tower icing.

       Although the climate at Rugeley is  not as severe as in continental  Europe,
temperatures below freezing are experienced regularly in winter.  The lowest re-
corded temperature at Rugeley is 9°F.

       Operation is controlled to keep the condensate  temperature leaving  the
tower above 45°F as a precaution against freezing. Temperature control isobtained
by taking cooling coil sectors out of service and draining the condensate in  the idle
section into the  storage  tank.  Limited temperature control can also be achieved by
taking one of the circulating pumps out of service.  The operation of the drainage
system to  take cooling coil sectors out  of service  is initiated manually, and  auto-
matic sequential operation  of valves and pumps follows.

       The Rugeley tower  is not equipped with louvers  to control the  flow of air
across the coils.
                                      222

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       Test's have been made which indicate that the coils can be safely filled and
drained during freezing weather if the operation is accomplished within 2 minutes.
The most critical  operation is filling the coolers after a previous  drainage when some
ice has already formed in the tubes.  Tests showed that water of 90°F inlet tempera-
ture would prevent freezing during filling with  temperatures  as low as  —30°F and
water of 100°F/ down to  -36°F. The procedure for placing the  dry tower coolers
into service during freezing weather is to bypass circulating  water around the coils
by means of the bypass valve  until the water is  heated up to  80°  to 90°F before
entering  the cooling coils.

Description of System Components

       Cooling coils.  There are 648 cooling coils of the  Forgo  design in the tower.
Each coil (element) is 16 feet high by 8 feet wide and 6 inches thick, containing 6
rows of 40 tubes in square pitch.  The coils are arranged in columns, three elements
high, with two columns joined together to form a "delta" .

       The cooling coils are  constructed of aluminum which is 99.5 percent pure.
The tubes, plate-type fins, and water boxes are all of aluminum  construction. The
total frontal area  of the coils is 80,000 square feet.

       Tower shell. The shell  is hyperbolic in shape and  is constructed of rein-
forced concrete of 5 inches minimum thickness, becoming  thicker where a reinforced
concrete ring beam takes the  thrust.  The tower is supported on reinforced concrete
legs 55 feet high, which  provide an opening for the air to pass through the coils  and
upward through the shell.  Figure A5 shows a view inside the tower and shows the
supporting structure and ring  beam at the base of the tower.   Tower dimensions are:
height -  356 feet; base diameter - 325 feet; throat diameter - 205 feet.

       Condenser.  The condenser (Figure 33) is a direct-contact, spray-type
designed by the English Electric Company and has a single steel  shell mounted
directly below the turbine receiving the exhaust steam from  the 3-flow, low-pres-
sure turbine cylinders. Cooling water is supplied through  water  boxes at each end
of the  condenser and is sprayed into the shell, mixing directly with the turbine ex-
haust steam.  There are 24  spray pipes, fed alternately from opposite ends of the
condenser through the water boxes. The spray nozzles are divided into four groups
and each of the groups is fitted with a spray control valve of the butterfly type.
The spray control  valves are automatically controlled to close in case the circulat-
ing water pumps fail and  there is a danger of flooding the  condenser.  The valve
closure is automatically controlled to prevent water hammer damage to the piping
system and coil sections. The original condenser design has been reworked to im-
prove the performance because of subcooling of the condensate .  When the unit was
first placed into service, subcooling of the condensate as  much as 15°F was  exper-
ienced, which was found to be the result of air leakage and difficulty in removing
                                      223

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FIGURE A5-INSIDE VIEW, NATURAL-DRAFT, DRY-TYPE COOLING TOWER —RUGELEY STATION
                         (ENGLISH ELECTRIC PHOTO)

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air from the condenser.  By redesigning the air-collection system and rearranging
some of the spray nozzles, the difficulties were eliminated and the subcooling was
reduced to less than l°F.

        Sector valves.  An interesting feature of the  Rugeley Station is the use of
sector valves to control  the flow of circulating water.  These valves are located in
the pump house inside the tower and are of the multi-port rotary  plug design. The
sector valves are  used for isolating individual cooling sectors of the tower, for fill-
ing and draining sectors and for normal operation.

        The sector valves were designed especially for the Rugeley dry  tower. The
other stations using  Heller-type towers which were visited did not use  this type of
valve, but relied upon individual valves to perform the various functions.

        Figure A6 shows a view of the sector valves in the pump house.

Auxiliary Power Requirements

        The  total  auxiliary power requirements for Unit  No. 3 are 8.9mw at full
load, or approximately  7.3 percent.  The power-using auxiliaries associated  with
the operation of the dry tower are the two half-capacity circulating  water pumps,
the power use of which  is compensated for in part by the energy  regained by the
water-recovery turbines.  The main circulating pumps are each 1 ,104 kw and the
recovery turbines, 324 horsepower (242  kw). The net pumping requirement at full
load with both pump and recovery turbines in operation  is approximately 1,723  kw,
or 1 .4 percent of output.  The  recovery turbines recover approximately 22  percent
of the  pumping power.

        Cooling water for the  generators,  oil coolers, bearing service and  other
auxiliary cooling requirements  is furnished by a small auxiliary or dry-type tower
which  uses mechanical draft for moving the air across the coils.

Turbine Cycle  Performance

        The turbine cycle design heat rate  for Unit No. 3 at Rugeley is the same as
for the other four units which are equipped with evaporative towers (3A).  The tur-
bine cycle efficiency of all units is 41 .3 percent, equivalent to 8,264 Btu per kwh
with 1 .3 inches Hg back pressure.  Because  of  the  higher  back pressure  actually
experienced with Unit No. 3,  station records made  available by the station  super-
intendent indicate the turbine  cycle heat rate is slightly higher than design.

        However, as reported in (1A), the performance at design point (vacuum 28.7
inches  Hg at 120 mw  and air temperature 52°F) has been met, and performance
throughout  the operating range  closely follows that predicted.  Table A-l  shows
                                       225

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K3
NJ
O>
        FIGURE A6—SECTOR  VALVES, VALVE HOUSE LOCATED INSIDE NATURAL-DRAFT,
         DRY-TYPE COOLING TOWER	RUGELEY STATION ( ENGLISH  ELECTRIC  PHOTO)

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TABLE A-l


Operating Data — Rugeley
120-Mw Turbine-Generator
Power Station
- Unit No. 3


with Natural-Draft Dry Cooling Tower
Ambient Air
Temperature
(°F)
43.3
51.4
41 -°
10
VI
36.0
29.8
22.5
26.2
32.6
38.0
27.9
Wind Velocity
(mph)
12 N.E.
18 N.W.
18 N.
4S.W.
8S.E.
8 N.
25 N.E./N.W.
10E.
20 E.
10W./N.E.
Condenser
Loading
(Ibs. of steam/hr.)
700,000
700,000
700,000
700,000
700,000
700,000
700,000
700,000
700,000
700,000
Back
Pressure
(in. Hg)
1.85
2.62
1 .71
1.52
1 .55
1.31
1.37
1 .53
1.69
1.50
Auxi 1 iary Power for
Cooling System
(mw)
1.7
1.7
1.7
1.7
1.7
1.7
1.7
1.7
1.7
1.7
Net
Output
(mw)
111 .1
111 .1
111.1
111 .1
111 .1
111 .1
111.1
111.1
111 .1
111.1

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 operating results as logged by station operating personnel with 10 operating  condi-
 tions selected at random from a period of time from  November, 1966 through
 February,  1969.

 Corrosion Problems

        The Rugeley Station is located at the site of a coal mine and is also adjacent
 to an ash-sintering plant which makes building blocks from the plant ash.  Also, the
 dry-type cooling tower is in close proximity to the four evaporative-type towers and
 subject to any drift of spray from these towers.

        Within a short time after start-up, serious corrosion was  found to have
 started in the crevices between the cooling coil fins and the spacer collars of the
 cooling sections, and also in the tube walls beneath the spacer collars.  The Forgo
 coil is constructed by placing an aluminum collar over each aluminum tube,
 followed by a section of the aluminum plate fin.  The collars and  fins are stacked
 alternately on the tubes until each  coil is completed,  at which time the fins and
 collars are tightly pushed together by a hydraulic press; then an expanding mandrel
 is drawn through the tubes, resulting in a tight mechanical bond between tubes,
 collars and fins.

        Apparently,  the combination of moisture in the air and pollutants, especially
 chlorides, was able to find its way  into the tiny crevices despite the tight mechani-
 cal joint between the fins and collars, setting up corrosion cells.

        Based upon experience gained  in Hungary with the Heller system, no corro-
 sion was expected at the Rugeley plant.  However, the corrosion advanced to the
 point where tube walls were perforated.  Also, the products of corrosion which were
 deposited in the fins caused damage to the  fin surfaces.  Damage was more severe
 on coil sectors which were out of service.

        A program of research was undertaken  by the Central Electricity Generating
 Board and the English Electric Company and a protective coating of epoxy resin was
 selected as the best method of corrosion prevention.  At the present time, a large
 number of the cooling coils at the Rugeley  Station have been treated with the epoxy
 coating.

 Effect of Wind on Tower Performance

        Since winds have an adverse effect upon the performance of a natural-draft
cooling tower (both the evaporative type and the dry type),  tests were made  at
Rugeley to measure the  effect of the wind.
                                      228

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       With wind velocity up to 10 mph, no effects were noticed, but loss of
vacuum was observed at wind speeds above  10 mph, with a vacuum deterioration of
approximately 0.36 inches Hg for a wind speed of 30 to 35 mph.  The average
annual wind speed at Rugeley is 13 mph; thus, the effect of wind on tower perform-
ance is not a significant factor.

       The air flow around the cylindrical  tower causes a reduction in air flow
through the coolers.  Figure A7 shows the effect  of wind on the tower performance
as observed during the  tests.

       Other tests were conducted to determine  the variation of air mass velocity
through the coolers around the tower.  Figure A8 shows  the results of tests made at
full load and wind of 20 mph.  The terms "upstream" and "downstream" refer to the
position of the coolers  in the deltas,  or V-shaped sections of coolers.  The survey
revealed that  the loss of vacuum was mainly a result of the blanketing of the down-
stream deltas of those coolers having tangential wind components, which more than
offset the increased air flow through  the upstream coolers (1A).

       Other effects of weather which have been observed are:  fog  improves per-
formance; rain reduces performance slightly; and intermittent sun produces a flicker
on the vacuum gauge.

Water-Side Chemistry

       The high purity aluminum tubes  used in the cooling coils require close  con-
trol of the pH of the circulating water.  In order to prevent corrosion  of the  alumi-
num water-side surfaces and to keep  the aluminum  from  going into solution in the
water and ultimately depositing in the turbine blades, a lower pH is carried  in the
tower circuit than in the boiler feedwater circuit of Unit No. 3.

       In order to protect the steel surfaces of the circulating water system piping
from the  low pH of the water, the inner surfaces were coated with plastic.

       Both the aluminum and iron content of the circulating water  has  remained
satisfactory; aluminum  0.01 to 0.02  ppm and soluble iron 0.02  ppm.   The pH of the
boiler feedwater is controlled by  morphaline.  The dissolved oxygen content of the
water in  the tower circuit is  0.1  to  0.3 ppm.   Oxygen content  of the boiler feed-
water after the deaerating feedwater heater is reported to be about 0.02 ppm, which
is considered satisfactory.   Neither  the boiler  nor the turbine have  experienced
deposi ts.

Maintenance

       Other than the repairs and cooling coil coating with epoxy which was nec-
essitated  by  the external  corrosion, there are  no extraordinary maintenance
                                      229

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           0.4
        111
        ac
        o
                            30   40
              AVERAGE WIND SPEED-MPH


FIGURE A7—WIND EFFECT UPON PERFORMANCE (IA)
           0 SO* 60» 90* ItO* *tf Wf tW 140* tTO»300«330« 3*0°


               POSITION  AROUND TOWER
     FIGURE A8—AIR  MASS VELOCITY  VARIATION

  THROUGH THE COOLERS WITH A 20 M.RH. WIND(IA)
                       230

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problems associated with the dry tower.  The station superintendent at Rugeley,
Mr. J .  E. Farrington,  reported that dry tower maintenance has been relatively low,
disregarding the corrosion of the cooling elements,  and is mainly associated with
venting valves and quadrant valves.

        It is reported that no cleaning of the coils has been necessary to remove dirt
and soot. A thin deposit forms on the exterior surfaces of the coils and fins and
reaches equilibrium with minor influence on performance.

Conclusion

        Although exterior corrosion has been a major problem, the Rugeley dry tower
is considered a success in that much useful  information was gained towards advanc-
ing the art of dry tower design, construction and operation.
                                       231

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                              IBBENBUREN PLANT
Introduction

       On Wednesday, December 3,  1969, John P. Rossie, accompanied by
Mr. Hans-Bernd Gerz of the Development Division of the Firm of GEA (Gesellschaft
fur Luftkondensation) of Bochum, West Germany, visited the Ibbenburen power plant
of the Preussag AG in Ibbenburen,  West Germany. The Ibbenburen plant is equipped
with a Heller dry-type cooling tower and  the equipment for the plant was supplied
by GEA.   During  the visit to the plant, the party  met with Mr. Ottokar  Scherf,
Superintendent of the plant, and was escorted around the plant by his assistant,
Mr. Hoffmann. At the time of the visit, the unit  equipped with the dry-type  cool-
ing tower was operating at design load of 150 mw. The turbine back pressure  was
1 .75 inches Hg and the ambient air temperature was 39°F.

Description of Plant

       The Ibbenburen plant is located in  the Town of Ibbenburen in the Ruhr Valley,
the area  where much of the heavy industry of West Germany is concentrated.  The
plant is located at the site of an underground coal  mine in an area where  mining  has
been undertaken for over 500 years.  Preussag, the corporation which owns and
operates  the plant, is engaged primarily in coal mining operations.

       The coal mined at the Ibbenburen plant is  anthracite, hard coal and forge
coal; over 2 million tons of coal  per year are mined. Much of the coal is for  home
fuel, but the fines and smaller granulated coal are sold to  industrial plants and
power plants and shipped via railroad  transportation. However, there is a certain
amount of the coal which is  high in ash and moisture and is not considered suitable
for sale.  In order  to utilize the low-grade coal at the site, a power plant was con-
structed by Preussag.  This plant went into operation in 1954.  The original  plant,
of 100-mw capability, is served by two natural-draft, evaporative-type cooling
towers of concrete construction and hyperbolic shape. The original plant consists
of four boilers, three of which have a capacity of 275,000 pounds of steam per hour
and one boiler with a capacity of 400,000 pounds of steam per hour,  with 1,100psi
pressure and 968 F serving two 21-mw automatic extraction turbine-generators and
one 50-mw regenerative cycle condensing  turbine-generator. Figure A9 illustrates
the Ibbenburen plant layout.

       The electrical output of the plant supplies the energy  requirement of the
mining operation; however,  the greater part  of the production is sold via the
German electrical  grid to various utilities.
                                     232

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         JMffiWMJ
        a.  MACHINE 'HOUSE (150 MW)
        b.  BOILER HOUSE ( 573, 200 #/ H R.)
        c.  ELECTRO FILTER
        d.  CONVEYOR BRIDGE
        e.  COOLING TOWER
        A.  NEW UNIT ( 150 MW)
        B.  OLD POWER STATION (APPROX.  100 MW)
FIGURE A9—PLAN  VIEW OF PREUSSAG POWER STATION
                     IBBENBlJREN  (5A)
                           233

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        As coal production increased and more high-ash coal  became available,
 Preussag decided to construct an addition to the generating plant to utilize the addi-
 tional low-grade coal.   In 1967,  a  150-mw unit equipped with a Heller dry-type
 cooling tower was placed into service.  Two 600,000-pound per  hour capacity boilers
 with steam conditions of 2,700 psi, 977°F/977°F  were installed to serve a  150-mw
 reheat turbine-generator.  The tower is reinforced concrete, natural-draft  type  of
 hyperbolic shape.

        The lack of suitable water for cooling tower make-up at a sufficiently low
 price was the reason for selecting a dry-type tower for the 150-mw unit at Ibbenburen.
 The  existing water plant has a maximum daily capacity of approximately 4 million
 gallons per day.   The wet-type cooling tower of  the original 100-mw plant,  alone,
 uses 2 million gallons per day of this supply; thus, it would have been necessary  to
 construct additional water-treating facilities and  develop a new water supply if an
 evaporative cooling tower were chosen, since the wet tower would require an addi-
 tional water make-up of approximately 3 million gallons per day.

        Engineering studies were made by Preussag to compare the economics of a
 generating unit equipped with a dry-type cooling tower, which would require no
 make-up water, and a generating unit with a conventional  evaporative tower. Con-
 sideration was given to the difference in initial construction cost, the cost of water,
 differences in operating efficiency because of higher back pressure with the dry
 tower, and other pertinent factors.  From these studies, Preussag concluded that the
 price of make-up water would have to be lower than the range of 27$ to 33$ per
 thousand gallons in order for the total annual operating costs of a wet tower to equal
 that of a dry  tower.  This compared with the actual price of water of 47$ per thou-
sand gallons from Preussag1 s water-treatment plant, and 65$ per  thousand gallons
 from an outside supply of water which would have had to be developed for a new
evaporative tower.

       In performing the studies,  Preussag investigated the Heller (indirect) system,
the direct, air-cooled, condensing system and a combination of various cooling
systems before selecting  the Heller system.

 Water Circuit
       Figure A10 shows the diagram of the water circuit through which approxi-
mately 66,000 gpm of condensate quality water is pumped.  The cooling tower is
divided into four sections, each of which can be independently drained and filled

       The water is circulated through the tower by means of two half-capacity
motor-driven circulating water pumps which are also equipped with Francis-type
recovery  turbines.
                                     234

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            Sinuntlfcihiiur BS
              260m1
r '-i
c
H8K
) (
HSH

1

HgH
)




VA2 • •
VAI il
t ;' ' ';l
j ^ 	 | t
Umwilt-
•Mttgat SucJ
•V
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va
•«
1-

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1
VAS
ll
\
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I
4
FIGURE AIO—COOLING WATER CIRCUIT DIAGRAM —150 MW
          IBBENBUREN GENERATING STATION (5A)
                           235

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        The Heller coils are of the same general construction and arrangement as at
the Rugeley plant, with 48-foot-high columns of coils joined into deltas and operat-
ing in cross-counterflow pattern.  However, sector valves are not used to control
water flow.  Instead, motor-operated butterfly valves direct the water flow for reg-
ular operation, drainage and filling.

        Two underground water-storage tanks are provided inside  the tower base.
One is sized to hold the water in one sector and the other can hold the water from
the entire system.  Two filling pumps  are provided with the storage  reservoirs.

        The direct-contact condenser  is of somewhat different design than the
Rugeley condenser.  Rather than the circulating water being sprayed from the indi-
vidual water spray pipes supplied from water boxes,  the circulating water flows
through four large water chambers in the  condenser and 2,600 water spray valves
are connected directly to the distribution chambers.  Subcooling  is  reported to be
less than 1°C.  The water-pressure drop in the condenser nozzles  is  approximately
8 feet.  Figure Al 1 shows a cross section of the direct-contact condenser.

Design Parameters

        The turbine back pressure design of the Ibbenburen plant is  1 .22  inches Hg
with 34.7°F ambient air temperature and a tower heat rejection of  645 million Btu
per hour.   The design initial temperature difference, which  is the most important
criterion affecting tower performance and initial cost, is 50.5°F.  This compares to
the Rugeley tower design of 1 .3 inches Hg with 52°F ambient air and an  initial tem-
perature difference of 35*^F, which resulted in a greater tower surface in proportion
to the heat rejection load for Rugeley.

        The Ibbenburen tower has 498 cooling elements as compared to 648 for
Rugeley, which reflects the higher back-pressure design resulting from the optimiza-
tion  studies made by Preussag before specifying the design parameters of  the dry-type
cooling tower.

        Figure A12 shows the operating characteristics of the Ibbenburen  dry tower
at various loads and ambient air temperatures.  This curve was replotted  from its
original version (5A) to English  units.

        It is interesting to note that the cooling tower coils were  formed  into 83
delta sections at the factory in Jaszbereny, Hungary and transported by rail to
Ibbenburen without damage.  A  repetition of the pressure test at the site  showed no
leaks.

        The Ibbenburen tower is  equipped with horizontal air-control  shutters. Some
90 percent  of the shutters  are operated by electric motors controlled from the central
                                       236

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I. STEAM  INLET
2. COOLING WATER INLET
3. SPRAY JETS
4. AIR EXTRACTION
5 COOLING WATER OUTLET
6.  FEED WATER OUTLET
7.  FINAL COOLER
8.  DEAERATOR
9.  INLET FOR DEAERATING STEAM
10.  COMPENSATE INLET TO DEAERATOR
  FIGURE  All— DIRECT  CONTACT  CONDENSER  (5A)
                             237

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ro
                                             50      60      70
                                         AMBIENT AIR TEMPERATURE  (°F)
                FIGURE  A12 — IBBENBUREN PLANT—NATURAL-DRAFT, DRY-TYPE COOLING TOWER

                   TURBINE  BACK  PRESSURE  VARIATION WITH  AMBIENT AIR TEMPERATURE

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control room by pushbuttons and the remainder are controlled manually.  Figure A13
shows the Ibbenburen tower with the shutters on the outside of the cooling coils.

Capital Costs

       Mr. Scherf reported that the capital  cost difference between the dry tower
installation and a comparable wet tower installation was estimated to be approxi-
mately $1,500,000, noting that almost $400,000 in water costs are saved annually
by the dry tower.  The total cost of the plant, including the dry tower, was
$22,800,000, making the cost of the dry tower plant approximately 7 percent higher
than a plant with a conventional tower.

Manpower Requirements of the Tower

        During the planning of the  150-mw unit, special attention was given to lim-
iting the number of operating personnel required for the expanded plant.  A criterion
was adopted that no more operators should be utilized with the new addition than
were required for a conventional wet tower installation.  To accomplish this purpose,
the components of the dry tower system were equipped for  centralized control  and
automated to a great extent.  All selections of pumps, draining and filling of coil
sectors, shutter operation, and valve operation are performed from the central con-
trol room.

        The 150-mw unit utilizes 8 men per shift as follows:

             1    Shift Foreman
             1    Turbine and Tower Control Operator
             2    Boiler Control Operators
             1    Roving operator who watches machinery,
                    including the tower and condenser
             1    Turbine-driven Boiler  Feed Pump Operator
             2    Boiler Auxiliary Operators and Ash Transporters

        Only one visit per day is made to the tower by the operator.

        Figure A14 shows the type  of instrument which is used by the tower operator
to determine when to shut down or restore a  circulating pump to service and when
to operate louvers.

        By observing in which zone of the indicator the mw pointer and the ambient
air temperature pointer cross, the operator is alerted as to when to operate louvers
or circulating pumps to keep  condensate  temperature up to a safe level.
                                      239

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FIGURE AI3— HYPERBOLIC CONCRETE DRY-TYPE COOLING
     TOWER INSTALLATION AT IBBENBUREN- 150  MW
                GENERATING  PLANT
                   (GEA  PHOTO)
                        240

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+2°-v
                            120
           + 10
-10
                                           MW
                                             50
    (I)  BOTH PUMPS RUNNING, LOUVERS OPEN
    (2). ONE PUMP RUNNING, LOUVERS OPEN
    (3). ONE PUMP RUNNING, LOUVERS CLOSED
    (4). BOTH PUMPS RUNNING, LOUVERS CLOSED
   FIGURE AI4-OPERATIONAL CONTROL I NSTRUMENK5A)
                       241

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 Winter Operation

        The Ibbenburen tower differs from the Rugeley installation  in that remotely
 operated louvers are installed to control  the flow of air through the coils.

        The 150-mw unit operates on the system load factor with output varying from
 30 mw to 150 mw.  Generally, the unit is held at rated load all day and drops off in
 load during the  night and on weekends.

        The means available to control the operation of the tower in winter to pre-
 vent freezing are:

        1 .    Louver operation.

        2 .    Taking one circulating pump out of service.

        3 .    Draining of cooling sectors when  the condensate
             temperature falls below a level where there is
             danger of freezing.

        Drainage of sectors is initiated by a pushbutton with automatic sequential
 operation after the initial signal.  It was reported that during the coldest weather
 and at loads as low as 30 mw, it has not been necessary to drain a sector to prevent
 freezing of the coils during operation. Apparently, at loads below 30 mw, it would
 be necessary to drain sectors during freezing weather.  The  lowest temperature re-
 corded at Ibbenburen is -4  F.  The 50-year average temperature is 48 F and there
 is an average of 30 hours per year with air temperatures above 77°F.

        During the initial winter operation, coils were damaged by freezing when an
 automatic vent valve failed  to open to allow the column to drain to the storage tank.
 The trouble was corrected by replacing the vent valves by a new design insulating
 and protecting them  from freezing with electric  heating cable.  The cooling coil
 column was replaced and no further trouble was  encountered with freezing.

       In placing the tower into service during  freezing weather,  the condensate
 from the condenser is recirculated, bypassing the cooling towers until the water  is
 heated to a temperature high enough to safely fill the coils.  The operation of fill-
 ing the sectors is done automatically by the control system with sequential interlocks.

       Draining of the tower is accomplished in 22 seconds and filling in 5 minutes.

Auxiliary Power Requirements

       The total auxiliary power requirements of the 150-mw Ibbenburen generating
 unit are approximately 8 percent of generator output.
                                      242

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       The auxiliary power-using equipment associated with the dry-type tower are
the two circulating pumps, which are equipped with water-recovery turbines to re-
cover the excess head imposed upon the cooling coils and to maintain approximately
3 psi pressure at the top of the coils. The reason for maintaining the excess pressure
is to have a positive pressure on all  parts of the large area of cooling coils so that,
in case of coil leaks,  air will not be drawn into the system.

       The  total pumping power required by  the circulating pumps is  I7640kw,
which  is reduced by 550 kw recovered in the water turbines, for a net pumping power
requirement of 1,090 kw, or 0.72 percent of the plant output.  The water turbines
recover 33 percent of the pumping power.

Turbine Cycle Performance

       Since the steam conditions and heater cycle of the old and new sections of
the Ibbenburen  plant are different, it is not possible to make a  direct comparison be-
tween  the units served by a wet tower and  those served by a dry tower.

       The design heat rate  for the new unit is 8,400 Btu per kwh (2,100 K calories
per kwh)  at 1 .2  inches Hg as compared to 11,500 Btu per kwh (2,900 K calories per
kwh) for the older low-pressure units without reheat.

Corrosion Problems

       Contrary to the experience at Rugeley  Station, no external  corrosion  prob-
lems with the cooling coils have been experienced at Ibbenburen.

       As at Rugeley, the Ibbenburen dry  tower is  located next to a wet tower and,
presumably, also subject to drift of water spray from that source,  although no special
coating on the fins or tubes was applied at Ibbenburen.

Effect  of Wind on Performance

       The same adverse effect of the wind noted at  the Rugeley Station was ob-
served at Ibbenburen, with the exception that  tests made at Ibbenburen indicate
that wind speeds as low as 2.24 mph (1 meter per second) influenced tower perform-
ance,  whereas the Rugeley tower was not influenced  up  to 10 mph.

       Tests made at Ibbenburen to verify  tower performance show that for 6.7 mph
(3 meters per second) wind velocity, the cooling effect is reduced by 2.7°F(1 .5°C)
and for 9 mph (4 meters per second) the cooling effect is reduced by 5.5°F (3°C).

       Figure A15 shows a curve of the deviation of cold water temperature  from
that obtained under ideal conditions (optimum  performance of the cooling tower at
                                      243

-------
to
        UJ UJ
        Q >
        Si
        H CO
          Z
        tr o

        pt

        *§
          o
        -I O
        o cr
        o u.
        o
6



5
            _ 1
                                    5                    10

                                WIND VELOCITY-FEET PER  SECOND


                                FIGURE  AI5—WIND EFFECT  UPON

                            NATURAL-DRAFT  TOWER PERFORMANCE (4A)
                                                                 15

-------
zero wind velocity).  Figure A16 shows test data taken at Ibbenburen for 28 hours
continuously.  The bottom two curves show the marked effect of wind speed on the
tower performance.

       Since it rained occasionally during the testing periods, the observers were
able to actually  measure the effect of rain on the tower performance.  Rain was
found to worsen the cooling effect of the tower.  This was attributed to the fact that
rain cools the air inside the tower, reducing the thermal lift which results in a lower
air flow across the coils.. Table A-ll shows operating results as logged by station
operating personnel  with five operating conditions selected at random.

Water-Side Chemistry

       Because of lack of experience with aluminum in the condensate circuit and
the effects of high purity condensate on the  life  of aluminum, Preussag conducted a
series of  laboratory  tests before placing the dry-type cooling tower  into operation.
During the laboratory tests, with pH held from 7.0 to 9.0 it was observed that the
aluminum content in the water became very  high, ranging up to 4 mg per liter.

       It was determined that the high solubility of  the aluminum was a result of a
brass pump in the test installation and the presence of copper ions caused the alumi-
num to dissolve.  For that reason, the use of copper  and copper alloy products was
avoided in the thermal cycle and the cooling tower cycle of the 150-mw unit.

       When the unit was first placed into service,  the condensate pH was held be-
tween 8.5 and 8.7, but experience showed  that when 8.5  pH was exceeded the solu-
bility of aluminum became too  high.  Based upon that experience,  condensate is
controlled to a pH value of 7.8 to 8.0 by the addition of hydrazine.  Aluminum
content is held to 0.002 mg per liter with the pH at  7.8 to 8.0.

       The  expected  life of the aluminum tubes is estimated to be  from 20 to 30
years based  upon the solubility now experienced.

        In order to prevent aluminum in the  condensate from reaching the boiler,
the portion of the condensate which is returned  to the thermal cycle is passed through
special screening filters coated with asbestos, reducing aluminum content of the
feedwater to 0.01 milligramsper liter.  After the screen,  the condensate to the
boiler passes through a cation and anion demineralizer polisher where the aluminum
content is further reduced to 0.002 milligrams per liter.

        The oxygen content of  the circulating water is from 0.1  to  0.3 milligrams
per liter; the oxygen content in the  thermal condensate is  0.01 to 0.02 milligrams
per liter.
                                       245

-------
  50
  30
                WARM WATER TEMPERATURE
                COLD WATER TEMPERATURE
             (DEAL VALUE OF THE COLD
             WATER TEMPERATURE
             AMBIENT AIR TEMPERATURE
 + 4
                DEVIATION OF MEASURED
                COLD WATER TEMPERATURE
                FROM IDEAL VALUE  . |  ,
(m/$)


'




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J


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WIND VELOCITY











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10 12 14 16 18 2O 22 24 2 4 6







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^MH


8 10 12 14 HOUR
    FIGURE AI6— DRY-TYPE, NATURAL-DRAFT
COOLING TOWER: IBBENBUREN PLANT PERFORMANCE
              TEST RESULTS (4A)
                    246

-------
                                                  TABLE A-ll

                                      Operating Data — Preussag-Kraftwerk
                                    150-Mw Turbine-Generator — Ibbenburen
                                      with Natural-Draft Dry Cooling Tower


KJ
VJ

Ambient Air
Temperature
60.3
60.4
63.3
63.7
65.1
Wind Velocity
(mph)
2.2
9.0
2.5
2.9
5.2
Back Pressure
(in. Hg)
3.1
3.5
2.9
3.0
2.2
Auxiliary Power
for Pumps
(mw)
1.0
1.0
1.0
1.0
1.0N
Net Output
(mw)
152.5
147.2
135.2
126.3
71.4
The above information supplied by the Stein Kohlenberg-Werke Ibbenburen
with back pressures calculated from the saturation water temperature.

-------
       After 1 year of operation, the unit was taken out of service and inspection
of the circulating water system and the condenser did not reveal any pitting or cor-
rosion .

Maintenance

       After 2-1/2 years of operation, there is a slight coating of coal dust and soot
on tne external surfaces of the fins and tubes which is not considered to be of signif-
icance as far as  adversely affecting performance.  It is planned to remove the dirt
coating with detergent and water on the next scheduled shutdown.

       There have been no extraordinary maintenance problems associated with the
dry tower.

Conclusion

       Based upon  the  continuous operation of the  150-mw unit at Ibbenburen, it
can be concluded  that the operation of the Heller-type tower at Ibbenburen  has
been successful.
                                     248

-------
                            VOLKSWAGEN PLANT


Introduction

       On December 4, 1969 John P. Rossie, accompanied by Mr. Hans-Bernd
Gerz of GEA, visited the steam-electric generating plant of the Volkswagen manu-
facturing plant in Wolfsburg.  During the visit, they interviewed Mr.  F.  Wehrberger,
Plant Engineer for Utilities, and his assistant, Mr. Erich Kirchhubel.

       At the time of the visit, all three 50-mw units equipped with dry  towers were
carrying rated load.  The turbine back pressure was 1 .2 inches Hg  (0.04  atmos-
pheres) with ambient air temperature approximately 36°F.

Description of Station

       The power station Wolfsburg of the Volkswagenwerk AG plant in  Wolfsburg,
West Germany supplies all electrical power and steam for the automobile manufac-
turing processes and also to the Town of Wolfsburg, which is heated by the steam
extracted from the plant turbines.  Approximately 24,000 people are employed at
the plant, and the Town has a population of approximately 85,000.  Many of the
plant workers live in neighboring towns.

       The power plant is divided into two sections; the old section is equipped with
evaporative-type cooling towers to cool  condenser circulating water and the new
section is equipped with direct condensing dry towers.  The old section has five
automatic-extraction type turbines, each of 8.5-mw capability for a total of 42.5
mw. The new section has three 48-mw units.

       The reason for constructing the dry-type towers with the three 48-mw units
rather than to continue the use of wet towers was the shortage of water at the plant.
There was not enough water available for evaporative towers without bringing  it  in
from a long distance at a high price.

       Contrary to  the dry tower installations at Rugeley and Ibbenbiiren,  the dry
towers at the Volkswagen generating plant are the direct condensing type in which
exhaust steam is conveyed from the turbines through large-diameter pipes and is con-
densed in the cooling coils. The V-W dry towers are of the mechanical-draft  type,
manufactured by GEA of Bochum, Germany.

       Figure A17 shows the mounting of the direct, air-cooled condenser on  the
roof of the turbine house.
                                      249

-------

  FIGURE AI7 —VOLKSWAGEN  PLANT WITH  DIRECT-TYPE,
AIR-COOLED CONDENSER UNITS ON PLANT ROOF(V-W PHOTO)

-------
       All the turbine-generators in the plant are of the automatic-extraction type
which draw off steam from the turbines for processing and heating at a constant pres-
sure over the varying turbine-load range.  Since the demand for extracted steam is
highest during the winter months because of the steam-heating load,  the heat rejec-
tion duty of the cooling towers is lowest during the cold-weather months and highest
during the warm months. The reason for this is that the amount of steam passing
through  the turbine to the condenser, for any given electrical  load, varies with the
amount of steam extracted for process.  Consequently, the operating characteristics
of the condensing system of an automatic-extraction type turbine-generator plantare
quite different from the typical utility steam-electric generating plant where only
enough steam  for feedwater heating is extracted from the turbine, and the heat re-
jection load from the turbine exhaust steam is  almost directly proportional to the
electrical load on the unit.   Other  than the foregoing, the problems associated with
the two types  of plants are the same, and the experience of the V-W plant can be
utilized by prospective  purchasers of dry tower equipment.

       Throttle steam conditions of the new section of the plant are  1,600 psi and
977°F.  Three fuels—coal, natural gas, and residual oil—are burned in the plant.

       The  first 48-mw unit went into service in 1961, and the last unit in 1966.

Condensation  Circuit

       The  exhaust steam from each of the 48-mw turbines is conveyed through
pipes 10 feet in diameter to the air-cooled condenser units located above the  tur-
bine room.  The air-cooling  coils are arranged in the form of inverted V-shaped
sections, similar to the deltas of the Heller system, except for the tube orientation
and the position of the deltas or V-shaped tube bundles.  Each of the three 50-mw
generating units has its individual block of condensers,  independent from the other
two.  Each 48-mw unit has 12 individual coils,  and each coil is equipped with a
2-speed fan located beneath the coil, for a total of 12 fans per turbine  unit.

        Figure A18 shows a diagrammatic view of the piping from the turbine to the
condenser.  Note the motor-operated valves which can be operated  remotely to
take individual sections of cooling  coils out of service in each block, while the  re-
mainder of the condensing coils remains in service.  This feature is necessary for
cold-weather operation, as described  later.

        Exhaust steam from the turbine condenses directly in the cooling coils and
drains to a receiver where it is pumped back to  the boiler.  In contrast to the in-
direct system, no cooling water is used.

        There are two different types of cooling coils in each  block of condensers.
One  type is designated as the standard air-cooled condenser in which the steam
                                       251

-------
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'"1 , .//PUMPS
1 f f CONDENSATE /-< /
1 1 KH»C,I VbH j V V~1
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TO BOILER
INSTALLATION "A" FEED SYSTEM
(INSTALLATION "B" r-l — .
] ^ A ISTHE SAME AS "A" ) I 	 J TURB'NE
— I
_ j
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VcONDENSATE RETURN
_ ALINES

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VI i i J TURBINE
•"1 T V ' 	
^1 "K" DESIGNATES " KONDENSATORLUFTER " OR STANDARD COILS
— 1
_l_j "D" DESIGNATES " DEPHLEGMATORLUFTER" OR COUNTER- FLOW COILS
_ _J
                                   FIGURE AI8 —EXHAUST STEAM  AND CONDENSATE  PLANT
                                      OF AIR CONDENSING SYSTEM-VOLKSWAGEN PLANT

-------
enters the coil from the top of the coil and is condensed as it travels downward along
the sloping side of the V-section.  In the standard  condenser section, the flow of
steam and  of the water that forms during condensation is in  the same direction; that
is, from the top of the coil to the bottom.  The other type of coil is designated as the
counterflow condenser in which the steam  distribution trunk  is located along the
bottom of the cooling-coil section so that the flow of  steam into  the  coil  is upward
from the bottom to the top, while the flow of the condensate is downward.

        The main purpose of the counterflow  coil is to  prevent  the freezing of con-
densate during cold-weather  operation,  and also to  prevent subcooling which
results in a thermal  loss to the turbine cycle. Although the counterflow coils have
a lower heat transfer coefficient than the standard coil,  it is necessary that a
certain number of these be used during cold-weather operation.

        In German,  the  standard coils are  called  "Kondensatorlufter" and  the
counterflow coils, "Dephlegmatorlufter".  This  designation is important in under-
standing the  use  of the  operating diagram  (Figure A19), which  is explained on
page 259.

        The first two 48-mw units designated as  "A" and "B", as can be seen from
Figure A18, Groups A and  B,  each have  three standard coil  groups and one counter-
flow coil group, and each of the four groups can be shut off independently.  Group
C has a different  arrangement in which there are four  condenser groups, but each
group consists of two standard coils and one  counterflow coil.

        The coils  were designed and constructed by  GEA with elliptical-shaped tubes
and plate-type fins. The tubes and fins are  made of steel, and protection against
external corrosion and binding of the fin  to the  tube is accomplished by hot-dipped
galvanizing.

Design  Parameters

        The turbine design back pressure at the  V-W plant is 2.7 inches Hg at59°F
ambient temperature when condensing 242,000  pounds per hour of steam. This cor-
responds to an initial temperature difference (ITD) between saturated steam tempera-
ture in  the condenser and the  ambient air of 51 °F, which is  practically the same as
the 50.5 F ITD at Ibbenburen, and is representative of European design practice.

        The average ambient  temperature at  Wolfsburg is 47°F; highest temperature,
91  F; and lowest  temperature, —4 F.

        The greatest condensing load is in the warm weather when there is less de-
mand on the automatic-extraction type turbines.  This trend can be seen from
Table A-lll,  which was obtained from actual operating records of the  V-W plant.
                                      253

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                     Betriebsdiagramm der GEA-Luftkondensationsantage KW-Nord
                            Block C VW-Wolfsbum A-Nr 221/2B28
  FIGURE AI9 —CALCULATED OPERATING CHARACTERISTICS (PREDICTED PERFORMANCE) FOR
THE DIRECT AIR-COOLED CONDENSING SYSTEM-BLOCK 'C' OF VOLKSWAGEN PLANT (FROM GEA)

-------
                                                          TABLE A-lII
Ol
Oi
Operating Data — Power Station "Wolfsburg"
of the Vblkswagenwerk AG.
49-Mw Automatic-Extraction Turbine-Generator
and Air-Cooled Condenser
Ambient
Air Temp.
(°F)
82
73
70
66
60
25
19
14
3
Wind Vel .
(mph)
7.82
7.61
7.61
8.50
0.93
9.40
6.93
7.82
6.70
Condenser
Loading
(Ibs. of steam/hr.
198,414
101,191
162,920
198,414
251,276
48,501
48,501
33,069
55,115
Back
Pressure
) (in. Hg)
4.20
2.31
2.31
2.60
2.31
2.02
2.31
2.60
2.60
Auxiliary Power Net
for Fans Output
(kw) (kw)
850
120
830
860
490
18
27
20
16
30,000
13,600
25,000
30,000
20,000
15,000
40,000
35,000
20,000
Number of Fans in
Operation and Speed
12 (all
6
12
12
9
1
2
2
1
1 fans) Full Speed
Half Speed
Full Speed
Full Speed
3 - Half Speed
6 - Full Speed
Half Speed
Half Speed
Half Speed
Half Speed

-------
 Table A-lll also  illustrates  the  great amount of operating flexibility available with
 the 12 fans and 2-speed motors.

        An economic comparison of a wet tower and a dry-type cooling system was
 made before the final determination to construct the initial dry tower units (6A).
 Because of lack of space, a  wet-type cooling tower would have had to be located
 approximately 3,000 feet from the power station in order to achieve a distance far
 enough from the coal storage pile to avoid problems of coal dust blowing into the
 cooling water.

        The following assumptions were made in the analysis of the two towers:

                                               Seasons
Operating hours

Condensing  load
(metric tons/hr.)

Average air tem-
perature,  C

Average relative
humidity,  %
Summer
Day
1,785
110
15.8
77
Night
735
73.3
12.2
84
Intermediate
Day
1,428
70
7.1
82
Night
588
46.7
4.9
87
Winter
Day
1,071
20
1.2
85
Night
441
20
0.1
87
Summer:
Intermediate Seasons:
Winter:
Capital Cost Basis:
Cost of Power:
Cost of Water:
May to September
March, April, October,  November
December to February
12%
0.04 Deutschmarks per kwh (1 .0£per kwh)
0.06 DM per ton (5.7$ per 1,000 gallons)
       A comparison of costs of two types of wet towers with direct, air-cooled
condensing is as follows:
                                      256

-------
                                        Natural-Draft       Mechanical-Draft
                                         Wet Tower             Wet Tower

Increased auxiliary power, DM/year         +61,600               +52,000

Cost of make-up water                      -30,000               -24,000

Difference in maintenance costs             - 8,000               -  9,000

Difference in capital costs                  -33,000               -57,000

Total cost savings, dry  tower over
wet tower,  DM/year:                      -10,000               -38,600


               Note:  + sign  indicates penalty to dry tower;
                      - sign  indicates credit to dry tower.

       Based upon the  analysis made by the Volkswagen AG engineers betore the
selection of the initial  dry tower unit,  an estimated annual saving of 10,000 DM
($2,500) favored the dry tower over a natural-draft wet tower, and 38,600 DM
($9,650) over a mechanical-draft wet tower.

       Apparently, one of the significant economic factors in the selection of the
direct condensing system at the Volkswagen plant was the great distance that a wet-
type cooling tower would have had to be located from the plant because of space
limitations.

Manpower Requirements of the Tower

       There are no additional operators required to handle the air-cooled con-
densers.  Instruments are installed in the central  control room which enable the
turbine control  operator to oversee the tower operating conditions.  All tower oper-
ating functions such as opening or closing valves to take cooling-coil sections out
of service and fan speed changes  are done from the control room.  All  operations
are manually initiated  from the control room and no automatic functions are pro-
vided for the air-cooled condenser.

       Once each shift, an auxiliary operator checks the air-cooled condensers
and the fan  motors and gear box  lubrication.
                                      257

-------
Freezing Problems

        The mefhod of preventing freezing of the coils during winter operation is to
provide close control of the fan speed, number of fans and number of cooling-coil
sections in operation  for varying steam loads to the condenser,  as hereinafter ex-
plained .

        Freeze damage was experienced in A and B units after two winters of opera-
tion, when condensate pipes froze at the air aftercooler.  This freezing was at
probes which had been installed to determine air leaks.  The probes were removed
from A and B condensers and probes were not installed on  the C condenser.

        Condenser coils of installation A were frozen and  damaged after three win-
ters of operation at a time when air temperature was 10 F.  Several tubes were
frozen and four tubes were split, requiring replacement by welding in new sections.
Freezing also occurred during the next winter  in both A and B installations when the
air temperature was 12°F.  This time, a great  many tubes  were  frozen, but only  one
tube was split, which required repair.

        As a result of these freeze-ups, an extensive analysis of turbine and coil
performance was made and certain conclusions were reached as to  the reasons for
the freezing.  Improved operating methods were put into effect with  the  result that
no further freezing damage was experienced.

       The investigation disclosed that freezing of the cooling coils occurred under
either of the following two conditions:

        1 .    At a time when the extraction steam requirements were heavy
             and there was  low steam flow to the condenser, a change  in
             the electrical  load or steam demand caused  the turbine back
             pressure to rise because of greatly increased flow to the con-
             densers .

             The rise in back pressure often was particularly sharp because
             of the fact that under the above  conditions the counterflow
             section of either A or B installations was the only condensing
             surface in operation during light loads in cold weather. Al-
             though the counterflow sections have  better  characteristics
             with respect to prevention of freezing, the heat transfer is
             not as efficient as the standard coil sections because of the
             counterflow of steam and condensate.

             When the operators observed the rise in back pressure,  they
             placed additional condensers into operation, with the effect
                                      258

-------
              that too much cooling was achieved and certain coil sections
              froze.  The  freezing problem was compounded by the fact
              that when freezing started, the turbine exhaust pressure  in-
              creased because of the  loss of condensing surface, and the
             operators often reacted by starting up additional fan capacity
             since  the  condensate temperature did not immediately indi-
             cate that freezing had  occurred.

       2.    When  minimum flow of cooling steam was flowing to the con-
             denser, a condition common during cold weather and heavy
             extraction steam requirements, the exhaust steam was in a
             superheated state because it had been passed through the
             lower turbine blades for the purpose of cooling the blades
             and the expansion path of the steam, as  traced on a  Mollier
             Chart, was quite inefficient as compared to the condition
             where the exhaust steam has accomplished work in the tur-
             bine and had  approximately 10 percent moisture when it
             reached the condenser. Since superheated steam has lower
             heat transfer characteristics than saturated steam, or steam
             with moisture content,  any increase in the amount or tem-
             perature of cooling steam as a result of load changes, caused
             the turbine back pressure to rise and the operators to react by
             cutting in additional condensing surface or fan capacity,
             which often caused freezing.

       To overcome the freezing problems, a number  of air  temperature probes
were installed after the cooling coils and the operators were instructed to maintain
41  F when the air temperature reached 32°F.  In order to accomplish this, it was
necessary to keep the condenser loaded to approximately 90,000 to 100,000 pounds
per hour  of steam, and also to switch the fans on and off more frequently than had
been done before.  Subsequent tests have shown that the condenser steam loads can
be reduced to one-half of the above  figures without freezing.

       Figure A19, which is the predicted performance of Block C (the latest in-
stalled 48-mw unit)  was prepared by  GEA; this figure  shows the effect of air tem-
perature, condenser steam load and fan operation on the turbine back pressure.  The
table in the  upper right corner of Rgure A19 shows recommended fan operation of
the coil sections, which are shown in their arrangement by the designations:

                    K - Kondensatorlufter, standard  coil,  and

                    D - Dephlegmatorlufter, counterflow coil .
                                      259

-------
        The Roman numerals I,  II, III, and IV refer to the four zones shown on the
curve in different shades of cross-hatching.

              I    -   All 12 fans at full speed

              II   -   All 12 fans at half speed

              III   -   Four fans on parallel flow coils off;
                      2 fans on counterflow units at half speed

              IV  -   All fans off

        The guidelines illustrate two operating conditions:

        1 .     Design point - 110,000 metric tons of steam per hour with
              15°C ambient air and all fans in operation;  resulting back
              pressure - 0.09 atmospheres (2.69 inches Hg).

        2.     50,000 metric tons  per hour steam to condenser, with —7°C
              ambient air and only the two counterflow fans in operation;
              resulting back pressure — 0.08 atmospheres (2.39 inches  Hg).

        The area in the lower left portion of the curve is to be avoided to prevent
freezing.

        The same performance curve applies for the  condenser and turbine unit at
part load with cooling sections out of service.  With four of the  sections, the steam
loads as indicated would apply.  With three sections in service and one out of ser-
vice, the curve would apply when steam loads are 75 percent of those shown in the
curve;  that is,  the 110 tons per hour would be equivalent to 82.5 tons per hour.
With two sections in service, 50 percent of indicated steam load is used and with
only one section in service, 25 percent of indicated steam load is used to read the
curve.   The foregoing explanation further illustrates the flexibility which the opera-
tors have in preventing freezing even during extremely  cold weather and light con-
densing loads.

        Start-up of condensing units during cold weather has  not been a problem.
The condenser is put into service with limited  cooling coils operating and with fans
off, and cooling coils and fans are brought into service  as the condensing steam load
is increased by following the operating performance curve.

        The V-W plant has a peculiar winter operating  problem in freeze prevention
because the increase in extraction steam flow  which occurs in the winter results in
low turbine exhaust steam flow to the condenser.
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Auxiliary Power Requirements

       Each 48-mw turbine-generating unit has 12 two-speed motor-driven fans in-
stalled with the air-cooled condenser.  The 12 fans for each 48-mw unit have a total
full-speed power requirement of 860 to 890 kw.  The half-speed fan requirements are
1 07 kw.

       The fan power requirements  are affected by the air temperature, since at
colder temperatures the air density increases and more fan power is necessary.

       At full speed with all fans in operation, the auxiliary power requirements are
approximately 1  .85 percent of output.

       There are no circulating water pumping requirements associated with the
direct condensing system.

Turbine Cycle Performance

       Since the 48-mw turbines are automatic-extraction type, there is no gener-
ally accepted method of  comparing  the turbine cycle performance with a typical
regenerative turbine cycle.  However, the back pressure with 77°F air and full con-
densing load would be approximately  3.5 inches Hg,  which is somewhat higher than
the design of a typical wet tower installation.  The higher back pressure results in a
higher heat rate and it can reasonably be concluded that the turbine cycle heat rate
for the V-W units equipped with air-cooled condensers is slightly higher than if they
had  been equipped with  conventional wet towers.

       One characteristic of the air-cooled condenser operation at the V-W plant
which must be taken into account by the operators is the effect of the air in subcool-
ing the condensate  below the saturated temperature corresponding to the turbine back
pressure.  Because of the pressure drop in the exhaust steam trunk from the  turbine to
the air-cooled condenser,  there is a steam-pressure drop which accounts for approx-
imately 1  to 2  C (1 .8 to 3.6°F) subcooling; this results in a slight thermal  loss to
the cycle.  However, if close attention is not paid to operation of the fans when
taking cooling-coil sections out of  service with varying air temperatures and varying
steam condenser loads, the  subcooling effect can amount to from 10 to 12°C (18 to
22°F), which would have a marked effect on turbine efficiency .

 Corrosion Problems

        There have been no corrosion problems during the 8 years of operation of the
 air-cooled condensers, either on the exterior  surfaces exposed to the atmosphere or
 to the internal surfaces which are in contact with steam  and condensate.
                                       261

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        Although the exhaust steam from the turbines contains II percent moisture,
 no erosion of tubes or piping has been reported.

 Effect of Wind on Performance

        Because the air-cooled condensers are equipped with mechanical\y driven
 fans, the effect of wind on the cooling tower performance is not a significant factor
 as  it is in natural-draft towers.  However,  it is reported in (6A) that weather condi-
 tions have an influence on the condensing performance.  Sunshine, cloudiness, rain
 and vapors from the wet cooling towers of the plant are reported to have a consider-
 able influence on the cooling capacity of the direct condensers.

 Water-Side Chemistry

        The oxygen level in the  condensate from the air-cooled condensers ranges
 from 0.005 to 0.007 mg per liter.  The highest  recorded oxygen has been 0.010 mg
 per liter.

        During start-up, the oxygen content of  the condensate is 0.06 to 0.067 mg
 per liter, but is quickly reduced by the air ejection equipment.

        Hydrazine is used to control oxygen.  Since the condenser tubes are steel
 the V-W plant  does not have any special problems in controlling a pH that is satis-
 factory for both aluminum and steel in contact with the condensate.

        Although the entire cooling coil is under vacuum,  no air leakage has been
 experienced, except for a leak during initial start-up, which was found weeks later
 to be in a condensate drain line  under a pipe clamp and, consequently, very diffi-
 cult to locate.

 Maintenance

        There have been no special maintenance problems associated with the air-
cooled condensers.  Normal maintenance is given to the fan gears and motors.

        There have been no problems of dirt fouling the cooling-coil surfaces.  The
coils are cleaned once each year with water, under 200 psi pressure, and the time
required to clean the condenser for one 48-mw unit is  3 to 4 hours.  Before using
high-pressure water for cleaning, a method of cleaning with compressed air was
tried, but was abandoned for the water method.  Since placing the condensing sys-
tem into service, the lubrication system has been changed  from the original design
(6A).  When first operated, summer- and winter-weight lubrications were used dur-
ing the different seasons, but was changed to an all-weather weight when trouble
in overloading the motor-driven  lubricating oil  pump was encountered during the
                                      262

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intermediate seasons. The lubricating oil  pumps are operated continuously since the
idle fans rotate even when the motor drives are shut off, because of the wind effect
through the coils, and thus require continuous lubrication.  On the  latest installation
("C") as a result of experience gained with A and B installations, the oil pump is
driven directly by the fan gears so that lubrication is available whenever the fan
blades are turning.

        It was reported that maintenance requirements and expenses of the dry towers
have been less than that for the wet towers installed in the old section of the power
plant.

Conclusion

        Although  the extraction steam  operation of the Volkswagen plant at
Wblfsburg causes operational  problems with respect to freezing during cold weather,
a method of operation has been evolved which has resulted in successful winter
operation.

        Acceptance tests made indicated that the guaranteed vacuum was met with a
1°C margin of safety.

        Keeping the coil surfaces free of dirt and preventing air leaks in the  coils,
which would impair operating efficiency,  has not been a maintenance problem.
                                       263

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                            GYONGYOS STATION
 Introduction

        On March 24, 1970 John P. Rossie, accompanied by Dr. L. Forgo of Hoterv
 and Mr. Isrvan Lindner of Transelekrro, visited the Gyongyos steam-electric  power
 station  located at Gyongyos, Hungary, approximately 60  miles east  of Budapest.
 The Gyongyos Station will have 600 mw of capacity served by dry towers when com-
 pleted in 1972, and will be  the largest station with dry towers to date.   However,
 there is also under construction in Razdan, Soviet Armenia, a new station with three
 200-mw units with dry-type  towers, scheduled for completion in 1972.

        At the time of the visit, one of the 100-mw units was  in operation carrying
 approximately 40 mw (being limited because of ash conveyor operation) and the other
 1 00-mw unit was  out of service because of a feedwater heater leak, which was re-
 paired quickly and the unit returned to service that day.

 Description of Station

        The initial units of the Gyongyos Station went into operation  in 1969, and
 additional generating units are currently under construction.   Units 1  and 2, which
 were completed in 1969, are both 100 mw in capacity and are equipped with
 natural-draft, Heller-type dry towers.  There are two 200-mw generating units under
 construction, one of which is equipped with a conventional wet tower utilizing
 mechanical draft  and the other with a natural-draft Heller tower.  A  third 200-mw
 unit using a dry tower is planned to complete the station.

        The Gyongyos plant is a mine-mouth plant, located at the site of a newly
 opened  strip-mining operation.  Originally,  all of the units, totalling 800-mw of
 generating capacity, were planned to be equipped with dry towers, but subsequent
 studies indicated  that there was sufficient make-up water from the mining operation
 for a wet tower for one 200-mw unit.  Plans were then changed and a conventional
wet-type cooling  tower and surface condensers were installed  with one 200-mw unit.

        There are  approximately 150 million metric tons of lignite at the site, with
 heating value between  1,300 and 1,450 K calories per kg  (2,340 to 2,620 Btu  per
pound); a moisture content of 33 to 34 percent; and an ash content of 22 to 30 per-
cent.

        Because of the extremely low quality of the coal, much difficulty has been
encountered with  the start-up of the plant, mainly in connection with the boilers
and ash-handling  systems. In view of the low-grade coal at the station, which is
of much lower heating value than  that which has been utilized in power generation
                                      264

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heretofore anywhere in the world to our knowledge,  such  problems  are  to  be
expected.

       At the time of the visit, many of the problems inherent to starting up a new
plant had been solved and it appeared that the shakedown trouble would be over-
come within a short time.

       Figure A20 shows a  photograph of the Gyongyos Station with a view of the
two 100-mw size reinforced-concrete towers.  The station is  of outdoor construction.
Note that the concrete natural-draft towers serving the two 100-mw units are  cylin-
drical  in shape rather than  hyperbolic.

       Figure A21 shows the construction of the first 200-mw Heller tower, which
was partially constructed at the time the photograph was taken. Note the  slip-form
construction of the concrete hyperbolic tower.

       The two 100-mw steam turbines were manufactured in Hungary by Lang
Engineering Works.  The first 200-mw turbine will be a type  L.M.2 manufactured
in the USSR, and the second 200-mw turbine will  be constructed in Hungary by
Lang under a Brown Boveri  Corporation license.

       The generators are of Hungarian manufacture by the Ganz Electrical Works
with water-cooled srator windings.  The boilers are of outdoor design and were
manufactured by the  Hungarian Shipyards and Crane Factory.

       The power plant is of modern design with centralized controls.  Digital data
logging is installed for continuous supervision of operation.  The mining operations
are quite extensive and lignite is delivered directly to the boiler bunkers by a con-
veyor system. The power output of the station is delivered to the Hungarian electri-
cal grid over 120-kv and 220-kv  transmission lines.

       The turbines are covered by a thin-shelled reinforced-concrete building of
half elliptical shape with telescoping sections which are equipped with wheels
which run on separate parallel tracks so that the various sections of the turbine-
generator can be uncovered for crane handling or dismantling.

Water Circuit

        Figure A22 shows a diagrammatic arrangement of the circulating water cir-
cuit of the Heller system at the Gyongyos Station. Approximately 42,000 gpm are
circulated through the 100-mw unit towers and 93,000 gpm through the 200-mw
unit towers.
                                      265

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     FIGURE A20—GYONGYOS  POWER STATION
  TWO REINFORCED CONCRETE  DRY-TYPE COOLING
TOWERS FOR 100 MW GENERATING UNITS (HOTERV PHOTO)

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   FIGURE A2I—REINFORCED CONCRETE TOWER FOR FIRST OF TWO
200 MW GENERATING UNITS IN THE GYONGYOS POWER STATION(HOTERV PHOTO)

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                                           NATURAL DRAFT
                                           TOWER
         STEAM
         TURBINE
                                                                              COIL
                                    WATER RECOVERY
                                    TURBINE
                                                          I'll  I::
                                                   m**   **t**m
     
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       Two half-capacity circulating water pumps are provided for circulating
water to the tower and separate condensate pumps return the required amount of
condensate to the boiler feedwater circuit.

       The circulating water is carried to the dry towers through underground steel
pipes and is directed into four cooling-coil sectors which can be individually
drained or filled. The valves are  of the butterfly type and are automatically con-
trolled by the sequential  tower control system.

       An  underwater circulating water storage tank and filling pumps are provided
with each tower.

       A head-recovery  turbine with adjustable  vanes is provided to convert the
excess pressure head maintained in the cooling coils to electrical energy.

       The auxiliary equipment of the units with dry towers is cooled by circulating
water taken from the evaporative-type cooling tower serving the 200-mw  unit.

Design Parameters

       The design temperature difference between  ambient air temperature and
steam-condensing temperature of the dry-type cooling system at the Gyongyos Sta-
tion is 25.4°C,  or 45.7°F for the two 100-mw units and 26°C/ or 46.8°F for the
200-mw units.  These design temperature differences compare with 35°F for the
Rugeley tower and 50.5°F for the Ibbenburen tower.  The air temperature range
throughout the year is from approximately —10 F to 90  F.

       The design heat rejection  loads to the towers are 425 million and 900million
Btu per hour, respectively, for the 100-mw and 200-mw sizes.

       The towers at Gyongyos are all equipped with  adjustable air louvers which
are remotely controlled by the operators.  The operating control mechanism of the
louvers has a spring-loaded actuator between the driving motor and each individual
louver so that the binding of one individual louver will not hinder  the movement of
the operating rod in its control of the remaining  louvers served by that control motor.
Dr.  Heller advised  that he considers louvers desirable  with dry tower installations at
any location where  below-freezing temperatures are encountered.

       The height of the 100-mw natural-draft towers is 367 feet with a base dia-
meter of 176 feet.  The height of the 200-mw natural-draft towers  will  be 380 feet
with a base diameter of 357 feet.

       The 100-mw dry towers each have 59 cooling deltas that are  15 meters  in
height, and the  200-mw  towers have 119 deltas each.  Since each delta consists of
6 individual Forgo  coils, the 100-mw towers each have 354 cooling coils and the
200-mw towers have 714 coils.
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        The cooling coils are of aluminum and are of the same design and construc-
 tion as the coils of the  Ibbenburen and Rugeley power stations.

 Capital Costs of the Dry Tower

        No figures are available as to the construction costs of the natural-draft, dry
 tower system at the Gybngybs Station.

 Manpower Requirements of the Tower

        Since the tower filling and draining is completely automated, there are no
 additional manpower  requirements for the dry-type cooling  tower.  Centralized con-
 trol  is  provided with  sufficient operating information available for the control
 operator to make all operating changes necessary to the tower from the control room.

 Winter Operation

        The start-up of  the Gybngyos Station afforded an excellent opportunity  to
 observe the performance of a Heller-type tower under extremely adverse conditions
 caused by the starting and stopping of the generating unit frequently during freezing
 weather.

        During the visit, it was reported that the two 100-mw units had been started
 up and taken out of service a total of approximately 80 times, and many of the starts
 and  stops  were during freezing weather.

        The station  director stated that none of the outages were caused by the dry
 towers and that no  trouble was experienced in either draining or filling the towers
 during freezing weather. Automatic control  provides rapid  draining to the storage
 tank upon turbine shutdown in freezing weather and also provides automatic bypass-
 ing of  the cooling sections during start-up  in  order to heat the entire charge  of
 circulating water to a sufficiently high temperature to safely fill the coils.

        Dr. Forgo explained the procedure for filling the cooling coils of the Heller
 system to  insure that all'air is removed from the empty coils during the  filling process.
 Although  each cooling-coil column is  equipped with an automatic vent valve,
 trouble had been  experienced in the past with air being trapped in certain coil sec-
 tions and  preventing water circulation,  which lead to danger of freezing during
 cold weather.

        In order to prevent freezing during the filling process, it is necessary that
 the coils be filled rapidly and that all air be vented from the coils.  When the coils
 are filled  from the inlet side of the tubes in the usual direction of water flow  (up-
ward in the inner three  rows of tubes to the top water box where the flow direction
                                       270

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is reversed downward through the outer three rows of tubes), the air is pushed out of
the tubes ahead of the water into the top water box where most of the air is vented
out through the automatic vent valve located on the top water box.  However, some
of the air may be carried over into the downflow tubes instead of being discharged
through the vent, and may be trapped in the coil  to prevent water circulating
through individual tubes.

        To prevent the air from  being carried  over into the downflow tubes, the
following filling method has been developed and is used on the Gyongyos towers:

             The delta sections  are filled so that water enters the bottom of
        both  the inlet and outlet headers; that is, water flows upward in  the
        three tubes that normally carry water downward, and also upward in
        the three tubes which carry water upward during operation.  During
        the filling process, the water in the three downflow tubes is main-
        tained several  inches higher  than  in  the  three upflow tubes so that
        the two flow sections of  the coil are filled simultaneously with the
        downflow section leading in level.  The water level difference  is
        automatically controlled by the recovery turbine vanes.  The pur-
        pose of this filling procedure is to insure that all air from the down-
        flow side is either vented directly from the top water box, or into
        the few inches of air space ahead of the rising water  in the upflow
        tubes.  The rising water  level in the upflow tubes pushes the air into
        the water box and out the vent.  The air cannot re-enter the down-
        flow  tubes since they are filled with water and are sealed off to
        entry of air.

        Dr.  Forgo also explained that,  although drainage of the coils must  be
accomplished rapidly during cold weather, the rate of drainage must be controlled,
rather than to permit the 45-foot-high columns to drain as  rapidly as free flow with
the top vent  and drain valve open would permit.   It has been found by experience
that too rapid drainage of the columns during freezing weather causes the water
column in the tubes to break, slowing down the water flow and permitting the water
in the coil to freeze into ice crystals inside the tube.  This freezing, in  itself, was
not harmful to the tubes,  and, since no damage was evident, the freezing was  un-
detected until the tower was filled and placed into operation..  The ice crystals from
the draining  operations caused restriction of water flow, and serious freeze-up  oc-
curred  as soon as the tower was  placed into operation.

        Experiments were made to determine an acceptable draining velocity and
the rate of drainage is now automatically controlled.  Drainage of cooling coils at
the Gyongyos Station is accomplished in 1 to 1-1/2 minutes during cold  weather.
                                       271

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Turbine Cycle Performance

       Turbine cycle heat rate is reported to be 2,030 K calories per kwh for the
100-mw unit and 1,900 K calories per kwh for the 200-mw unit equipped with dry
towers. This is equivalent to 8,050 Btu per kwh and 7,520 Btu per kwh,  respec-
tively.  The turbine cycle heat rate of the 200-mw unit served by the evaporative
tower and surface condenser is reported to be 1,960 K calories per kwh, or 7,750
Btu per kwh.

       All turbine units are reheat type and will have six feedwater heaters in the
cycle with final feedwater temperature of 446 F for the 100-mw units, 46/ F for the
200-mw units served  by dry towers, and 446°F for the 200-mw unit on the conven-
tional tower.

       It  is noted  that the throttle steam conditions of the 200-mw units with the dry
towers are different than  the steam conditions of the 200-mw units with conventional
tower and surface condenser.  Listed below are the turbine cycle conditions:

                       Gyongyos Station Design Conditions

                                Dry Tower Unit         Conventional Tower Unit
                             100 mw       200 mw               200 mw
Number of units                2            2                     1

Throttle steam pres-
sure, psi                     1,850        2,350                 1,850

Steam temperature, ° F      995/995     1004/1004            1058/1058

Final feedwater tem-
perature, °F                  446°          468°                 446°

       Since the plant has not been completed nor have the 100-mw units  been
placed into full operation, no comparison can be made as to operating  results.

Corrosion Problems

       The location of the Gyongyos Station is on the Hungarian plains where the
weather is generally dry.  No corrosion problems are expected to be encountered in
the operation.  The initial units have been in operation less than 1 year, so no def-
inite conclusion can be drawn; but,  to date, the  dirt and coal dust have not been a
problem nor is there indication that they will be.
                                      272

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Conclusion

       The use of the  Heller-type dry tower with the Gyongyos Station  made  it
possible for the Hungarian Civil  Construction Enterprise to make use of large lignite
deposits for electrical power generation which otherwise could not have been used
because of lack of cooling water for evaporative-type towers.  The low  calorific
value of the lignite at Gyongyos made it infeasible to transport the coal  to a plant
site where cooling tower make-up water was available. Thus, the coal resources
would have remained unavailable without the use of the dry towers.
                                       273

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                           NEIL SIMPSON STATION


Introduction

       On May 21,  1970 Edward A. Cecil and Clarence  J.  Steiert visited the
Neil Simpson power plant of the Black Hills Power and Light Company at Wyodak,
Wyoming,  located about 6 miles  east of Gillette,  Wyoming,  at an elevation of
4,600 feet above sea  level.  They were  escorted on a tour of the plant  by
Mr. William Craig, Assistant Plant Superintendent.   Because of the  low fuel cost
associated with the Company's coal mine, all four machines  at this generating plant
are operated at base  load.

Description of Station

       The plant is located at a coal  mine owned by the Company and operated by a
subsidiary, the Wyodak Resources Development Corporation.   The plant consists  of
four small generating  units. The first  two units, rated 1,500 and 2,000 kw, respec-
tively, are small conventional machines with standard condensers cooled by evapor-
ative-type cooling  towers. Unit No. 3 is  an old  3,000-kw,  450-psig,  750°F
condensing turbine-generator set which was moved from a retired plant to Wyodak
for the purpose of experimenting with  an air-cooled  condensing system at the
Wyodak coal mine site.  Figure A23 shows the air-cooled condenser with louvers
open.  No additional water is available at this plant site,  but there is an abundance
of low-cost coal in a  seam  70 ,to 90 feet thick with an overburden ranging generally
in depth  from 5 to 20  feet.

       After a number of years of successful operation with air-cooled condensation
by the experimental  unit, the Company installed Unit No. 4, a 20-mw turbine-
generator set, utilizing a direct, air-cooled steam condenser supplied by  GEA of
Germany.  Figure A24 presents a view along the side of  one of the A-frame con-
denser units.  The turbine is rated at 20,180 kw nominal with 850 psi, 900°F steam.
A 72-inch pipe conducts the exhaust steam from the  turbine flange to  two A-frame
air-cooled condensing units mounted  adjacent to the turbine  room.   The  exhaust
steam enters the condensing.sections from the top and passes once through to the
bottom of the heat exchanger where the condensate  is collected  in a  header and
flows by gravity to a collecting tank.  The air passes over the finned tubes by cross-
flow.   The tubes, because of rugged design, are protected only by coarse hail
screens.  There are six fans, each of which  is driven by two motors through  aspecial
gear reducer.  The  large motor is rated 150 horsepower  constant speed for high-
speed operation and the small motor is a two-winding unit  requiring approximately
45 horsepower for half-speed fan operation (full-speed motor operation) and 10
horsepower for quarter-speed fan operation (half-speed motor operation).  The fans
are 20.8 feet in diameter and have six blades each.   The fan speed is varied to
provide the required turbine back pressure.
                                     274

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FIGURE A23—3,000 KW PILOT  PLANT DIRECT-TYPE, AIR
COOLED CONDENSER INSTALLATION-NEIL SIMPSON PLANT,
               WYODAK. WYOMING
FIGURE A24—SIDE VIEW OF A-FRAME, DIRECT- TYPE AIR-
 COOLED CONDENSING UNIT-20 MW  GENERATING UNIT
      NEIL SIMPSON PLANT.  WYODAK, WYOMING

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        The plant requires 5.7 gpm of make-up water which includes that used for
boiler blowdown and soot blowing, and is supplied from a pump rated at 10 gpm.

        The 20-mw unit was first started up on September 11, 1969, and went into
commercial operation on  December 15, 1969.

Design Parameters

        The air-cooled condenser has a design exhaust steam rate of 167,301 pounds
per hour with steam enthalpy of 1,027 Btu per pound to provide a turbine back pres-
sure of 4.5 inches Hg with an ambient temperature of 75°F and a heat rejection rate
of 155 x 106 Btu per hour. The design ITD is, therefore, 54.8°F.

Capital Cost

        The 20-mw plant  addition with an air-cooled condensing system costs ap-
proximately $315 per kw, with the air-cooled condensing system accounting for
approximately 11  percent of the total.  A standard cooling system with an evapora-
tive-type  tower would have cost approximately 3 percent less than the  one that was
used.

Manpower Requirements

        The only control elements associated with the air-cooled condensing system
as provided by GEA for the 20-mw generator unit are fan-motor control switches
located at the main turbine-generator control  board.  There are no operating or
maintenance requirements for which special manpower is needed. The  Company's
experience with the experimental 3,000-kw air-cooled unit indicates that dust
should be  blown from the  coolers approximately once per year.  Other  than dust
blowing, no regular maintenance was required.

Winter Operation

       Although the original 3,000-kw air-cooled condensing unit at Wyodak was
equipped with louvers for cold-weather operation, experience indicated that they
were unnecessary. Louvers were therefore omitted from the  20-mw installation.
Sidewalls  shown in Figure A25 were provided around the periphery of the heat ex-
changer units to provide some measure of protection from the emission of the low
stacks of other small units. To date,  no serious problems due to freezing have oc-
curred, although temperatures down to —33°F have been experienced.  During
severe cold-weather start-up, the operators alternately operate fans from the two
condenser sections to maintain proper vacuum and  to prevent coil freeze-up.
                                     276

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FIGURE A25—SIDE WALLS  ERECTED AROUND DIRECT-TYPE,
  AIR-COOLED CONDENSING  UNIT-20 MW  GENERATING UNIT,
        NEIL SIMPSON PLANT, WYODAK , WYOM ING
FIGURE A26-STEAM HEADERS AND HAIL SCREENS-DIRECT-
TYPE, AIR-COOLED CONDENSING UNIT-20 MW GENERATING UNIT,
       NEIL SIMPSON PLANT, WYODAK, WYOMING

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       The turbine, although chiefly a standard unit, is capable of quick start-up
and quick shutdown, a characteristic which is important during severely coldweather.

       The 2-motor-operated fans are operated at one-quarter, one-half, full speed
and some fans are taken out of operation, depending upon the  temperature and tur-
bine-generator loading requirements.  The manufacturer, GEA of Germany,  has
furnished an operating diagram  (Figure 20 in Section III) which outlines the required
speed of each fan for any combination of ambient temperatures and steam loading.

Description of System Components

       Air-cooled condensation system.  The  GEA air-cooled system conducts the
exhaust steam from the turbine through a 72-inch duct which branches under the air-
cooled units to deliver steam to the top headers of the four condenser sections con-
sisting of inclined heat exchanger finned-tube sections arranged in an inverted V-
form similar in shape to the A-frame type of cottage construction.  This arrangement
is shown in Figure A26.  Air is supplied to the center of this inverted V-arrangement
by large fans, shown in  Figure A27, similar to those used in conventional evapora-
tive-type cooling towers. The steam enters the top  headers of the cooling sections,
passes down through the finned tubes and into  the lower headers.  From the lower
headers, where the  mixture now consists of steam and liquid condensate,  the re-
maining steam and noncondensables pass upward  through additional aftercooling sec-
tions which are provided with separate fans.  The condensables are drawn off from
the connecting header between the condensing-and aftercooling sections. The non-
condensables are drawn off from the top of the aftercooling sections.

       Four condensing and two aftercooling  sections,  each with its own fan, are
provided for the 20-mw unit. There are two rows of inverted V-type cooling sec-
tions,  each consisting of two condensing sections with an aftercooling section in the
middle. The  fan speeds are  controlled individually  to minimize power consumption
and prevent freezing problems during cold-weather or light-load operation.  The
condensate flows by gravity  to two condensate pumps which return the condensate  to
the boiler feedwater cycle.

       Cooling coils. The GEA heat exchanger sections are fabricated from ellip-
tical -shajped^arEonTteeI tubes arranged in staggered rows for  best air flow charac-
teristics.  Steel fins are galvanized to the tubes  for  better heat transfer and corro-
sion protection.  Spacers at  the four corners of the fins assist the steel fin collars,
which also  act as  spacers at  the tube itself, to provide construction rigidity.

       Auxiliary  power requirements.  The maximum fan  power requirement of the
air-cooled  condensing system at Wyodak is 816 horsepower.  This condition occurs
at full-load operation with high ambient temperatures.  The fan power requirements
reduce in stages down to a minimum with reduced load and/or  cold-weather opera-
                                      278

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 FIGURE A27—FAN ARRANGEMENT FOR DIRECT-TYPE,
AIR-COOLED CONDENSING SYSTEM-20 MW GENERATING UNIT,
      NEIL SIMPSON  PLANT, WYODAK, WYOMING
                        279

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 tion which could mean that all fans would be off.  However, the operators  as  yet
 have not operated with all fans off at Wyodak.  Figure 20 in Section III" is a graph of
 calculated  operating characteristics for  the  Neil  Simpson  air-cooled condensing
 plant.

 Turbine  Cycle Performance

        The design conditions for the 20-mw unit at Wyodak call for a  turbine back
 pressure of 4.5 inches Hg with an exhaust steam flow of 167,301 pounds per hour at
 an ambient temperature of 75°F.   The turbine back pressure can be operated down to
 2 .0 inches Hg back pressure and up to as  high as 7.0 inches Hg.  The turbine trip is
 set at a  back  pressure of 7.5 inches Hg.  To date, the  generating unit  has operated
 at temperatures up to 90°F, and has exceeded design specifications. Table A-IV
 provides operating data recently tabulated by station operating personnel.

 Corrosion Problems

        The Neil Simpson plant is located at the site of a coal strip mine. Coal  is
 processed at the plant site for shipment to other power  stations of the Black Hills
 Power and Light Company.  However, no corrosion problems have been experienced
 to date, and none are anticipated.

 Effect of Wind on Cooling Tower  Performance

        Wind  has no apparent effect on the performance of this mechanical-draft,
 air-cooled condensing system except that  due  to low-stack emission from the smaller
 plant units when  the wind is in a  southwesterly direction.  This effect is of a very
 minor nature and will not be present in any future plant construction.

 Maintenance

       Maintenance requirements for the air-cooled condensing system is practically
 nil  and consists of normal fan-motor maintenance and air cleaning of the finned coils
 annually.  No additional plant labor is required for this small  maintenance.   The
 air-cooled units are preferred  by plant personnel over evaporative-type systems be-
 cause of their  small maintenance requirements.  Plant water treatment costs are also
 very low for the air-cooled units.

 Conclusion

       The air-cooled condensing systems have been successfully utilized by  the
 Black Hills Power and Light Company in Wyoming where a shortage of coo I ing water
 exists at  the coal mine site.   Both  the original 3,000-kw experimental  unit, which
went into operation in the early sixties, and the 20-mw unit, started up in 1969,
                                      280

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                                            TABLE A-IV
Ambient Air
Temperature       Gross Output
                     (kw)
    16             20,687
    30             20,479
    44             19,604
    52             20,400
     4             20,600
    43             17,500
   - 8             19,667
   -22             21,560
   -18             21,267
    46             20,600
    26             22,600
    74             22,000
    80             21,600
)perating Data -
- Neil Simpson Station
Turbine-Generator with Mechanical-Draft,
ect Air-Cooled

Steam Flow
(Ibs.Ar.)
203,760
204,000
190,000
200,000
206,000
172,000
186,200
204,200
203,333
205,000
211,000
205,000
203,000
Condensing System
Condenser
Loading
(Ibs.Ar.)
149,920
149,400
144,000
147,000
152,000
130,000
146,160
150,000
150,667
145,000
156,000
153,000
153,000
(6A)
Turbine
Back Pressure
(in. Hg)
4.33
4.38
4.22
4.26
4.71
4.14
5.56
3.68
4.76
2.70
3.41
3.99
4.68

Design
Back Pressure
(in. Hg)
6.2
3.7
4.8
4.5
5.2
4.1
5.3
5.8
6.4
3.6
3.5
3.99
4.45

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have been operating at base load satisfactorily.  The 3,000-kw unit enabled the
operators to become acquainted with handling air-cooled systems and made the in-
stallation and start-up of the larger unit a routine matter.  The Company has plans
for a larger unit of approximately 150 mw when their system requirements or arrange-
ments with neighboring systems will make a plant of this size feasible.  The proposed
larger station will be a few hundred yards from the present plant site so as to elimi-
nate any adverse effects from the low-stack arrangements of the  present  plant site.
                                     282

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                                  Appendix B

                           Engineering Weather Data
       The air temperature data utilized in the economic optimization analyses were
developed from information contained in the series of U. S. Weather Bureau publica-
tions entitled "Climatography of the United States No.  82,  Decennial  Census of
United States Climate, Summary of Hourly Observations".

       Economic optimization analyses were made for the 27 sites shown in Table
A-V.  The annual distribution of air temperatures for each of the 27 sites is summa-
rized in Table A-VI.
                                      283

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                                 TABLE A-V
      Area
Pacific:
Mountain:
West North Central:
West South Central:
East North Central:
East South Central:

New England:   '

Mid-Atlantic:

South-Atlantic:



Hawaii:

Alaska:
Economic Optimization Analysis
        Site Summary

            City                     Site Number

  Seattle,  Washington                    1
  San Francisco,  California               2
  Los Angeles, California                 3

  Great Falls,  Montana                   4
  Boise, Idaho                           5
  Casper, Wyoming                       6
  Reno, Nevada                          7
  Denver,  Colorado                      8
  Phoenix, Arizona                       9

  Bismarck, North Dakota                10
  Minneapolis, Minnesota                11
  Omaha,  Nebraska                     12

  Little Rock, Arkansas                  13
  Midland, Texas                       14
  New  Orleans,  Louisiana               15

  Green Bay, Wisconsin                  16
  Grand Rapids,  Michigan               17
  Detroit,  Michigan                     18
  Chicago, Illinois                      19

  Nashville, Tennessee                  20

  Burlington,  Vermont                   21

  Philadelphia, Pennsylvania             22

  Charleston, West Virginia              23
  Atlanta, Georgia                      24
  Miami, Florida                        25

  Honolulu,  Hawaii                     26

  Anchorage, Alaska                    27
                                     284

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00
Ol
                                               TABLE A-VI

                                    Annual Distribution of Air Temperatures


Site:                         No. 1, Seattle, Washington

Weather Station Location:      Seattle-Tacoma Airport

Period:                       1951-60


Air Temperature Range (°F)         Percent of Time         Air Temperature Range (°F)         Percent of Time

        119/115                            0                       39/35                       10.43
        114/110                            0                       34/30                        4.87
        109/105                            0                       29/25                        1.19
        104/100                            0                       24/ 20                        0.44
         99/ 95                         0.02                       19/ 15                        0.23
         94/ 90                         0.07                       14/ 10                        0.03
         89/ 85                         0.27                        9/5                           0
         84/ 80                         0.71                        4/0                           0
         79/ 75                         1.40                     - I/- 5                           0
         74/ 70                         2.94                     - 6/-10                           0
         69/65                         5.11                     -11/-15                           0
         64/60                         8.56                     -16/-20                           0
         59/ 55                       14.51                     -21/-25                           0
         54/50                       16.68                     -26/-30                           0
         49/45                       16.48                     -31/-35                           0
         44/40                       16.06                     -36/-40                           0

Total Percentage   =  100.

-------
                                                   TABLE A-VI  (continued)


          Site:                         No. 2, San Francisco, California

          Weather Station Location:      International Airport

          Period:                       1951-60


          Air Temperature Range (°F)'        Percent of Time         Air Temperature Range C*F)         Percent of Time

                   119/115                           0                      39/35                        1.13
                   114/110                           0                      34/30                        0.11
                   109/105                           0                      29/ 25                           0
M                  104/100                           0                      24/ 20                           0
£                   99/ 95                        0.01                      19/ 15                           0
                    94/ 90                        0.06                      14/10                           0
                    89/85                        0.17                       9/5                           0
                    84/ 80                        0.46                       4/0                           0
                    79/ 75                        1.13                     - I/- 5                           0
                    74/ 70                        3.25                     - 6/-10                           0
                    69/65                        7.59                     -11/-15                           0
                    64/60                       14.42                     -16/-20                           0
                    59/ 55                       26.70                     -21/-25                           0
                    54/50                       26.70                     -26/-30                           0
                    49/45                       13.15                     -31/-35                           0
                    44/40                        5.12                     -36/-40                           0


          Total  Percentage  =  100.

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00
•V4
                                          TABLE A-VI  (continued)


 Site:                         No. 3, Los Angeles, California

 Weather Station Location:      International Airport

 Period:                       1951-60


 Air Temperature Range fr)          Percent of Time          Air Temperature Range (r)         Percent of Time

         119/115                           0                      39/35                      0.11
         114/110                           0                      34/30                          0
         109/105                           0                      29/25                          0
         104/100                        0.01                       24/20                          0
          99/95                        0.05                      19/  15                          0
          94/ 90                        0.08                      14/  10                          0
          89/ 85                        0.32                       9/5                          0
          84/ 80                        1 .34                       4/0                          0
          79/ 75                        4.33                     - I/- 5                          0
          74/ 70                       10.05                     - 6/-10                          0
          69/65                       18.86                     -11/-15                          0
          64/ 60                       25.01                      -16/-20                          0
          59/ 55                       21.72                      -21/-25                          0
          54/50                       12.02                      -26/-30                          0
          49/45                        4.88                     -31/-35                          0
          44/ 40                        1.22                      -36/-40                          0


Total Percentage  =  100.

-------
CO
00
                                         TABLE A-VI (continued)


Site:                         No. 4, Great Falls, Montana

Weather Station Location:      International Airport

Period:                       1951-60


Air Temperature Range (°F)         Percent of Time         Air Temperature Range (°F)         Percent of Time

         119/115                          0                       39/35                       9.27
         114/110                          0                       34/30                       7.96
         109/105                          0                       29/25                       6.08
         104/100                       0.01                       24/20                       4.04
         99/95                       0.06                       19/15                       2.49
         94/ 90                       0.52                       14/  10                       1.90
         89/ 85                       1 .29                        9/   5                       1.55
         84/80                       2.13                        4/   0                       1.35
         79/ 75                       3.38                      - I/-  5                       1.15
         74/ 70                       4.64                      - 6/-10                       0.78
         69/65                       5.93                      -11/-15                       0.58
         64/60                       7.25                      -16/-20                       0.49
         59/55                       8.60                      -21/-25                       0.17
         54/ 50                       9.37                      -26/-30                       0.05
         49/45                       9.47                      -31/-35                          0
         44/40                       9.49                      -36/-40                          0


Total Percentage  =  100.

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00
•o
                                          TABLE A-VI  (continued)


 Site:                        No. 5,  Boise, Idaho

 Weather Station Location:      Boise Air Terminal

 Period:                      1951-60


 Air Temperature Range (°F)          Percent of Time         Air Temperature Range (°F)         Percent of Time

         119/115                           0                       39/35                      10.01
         114/110                           0                       34/30                       9.46
         109/105                       0.01                        29/25                       5.95
         104/100                       0.09                       24/20                       3.50
          99/95                       0.50                       19/15                       1.69
          94/90                       1.55                       14/10                       0.61
          89/ 85                       2.62                        9/5                       0.30
          84/80                       3.60                       4/ 0                       0.16
          79/ 75                       4.28                      - I/- 5                       0.07
          74/ 70                       5.61                      - 6/-10                       0.02
          69/65                       6.56                      -11/-15                          0
          64/ 60                       7.33                      -16/-20                          0
          59/ 55                       8.01                      -21/-25                          0
          54/ 50                       8.96                      -26/-30                          0
          49/45                       9.10                      -31/-35                          0
          44/40                      10.01                      -36/-40                          0


Total Percentage   =  100.

-------
                                         TABLE A-VI  (continued)


Site:                         No. 6, Casper, Wyoming

Weather Station Location:      Casper Air Terminal

Period:                       1956-60


Air Temperature Range (°FJ         Percent of Time          Air Temperature Range (°F)         Percent of Time

        119/115                          0                       39/35                      9.48
        114/110                          0                       34/30                      9.19
        109/105                          0                       29/25                      7.79
        104/100                          0                       24/20                      5.64
          99/ 95                        0.04                       19/ 15                      3.69
          94/90                        0.75                       14/10                      2.28
          89/ 85                        2.29                        9/  5                      1.32
          84/ 80                        3.23                        4/  0                      0.83
          79/ 75                        3.96                      - I/- 5                      0.51
          74/ 70                        4.82                      - 6/-10                      0.34
          69/65                        6.07                      -11/-15                      0.17
          64/ 60                        6.75                      -16/-20                      0.04
          59/ 55                        7.32                      -21/-25                      0.02
          54/ 50                        6.91                      -26/-30                          0
          49/ 45                        7.64                      -31/-35                          0
          44/ 40                        8.92                      -36/-40                          0


Total Percentage  =  100.

-------
N>
O
                                          TABLE A-VI (continued)


 Site:                         No. 7, Reno, Nevada

 Weather Station Location:      Municipal Airport

 Period:                       1956-60


 Air Temperature Range (°F)         Percent of Time          Air Temperature Range (°F)         Percent of Time

        119/115                           0                       39/35                       9.45
        114/110                           0                       34/30                       8.36
        109/105                           0                       29/25                       6.04
        104/100                       0.02                       24/20                       4.41
          99/ 95                       0.40                       19/ 15                       2.59
          94/ 90                       1.37                       14/ 10                       1.15
          89/ 85                       2.77                        9/  5                       0.42
          84/ 80                       3.80                        4/  0                       0.17
          79/ 75                       4.23                      - I/- 5                       0.05
          74/ 70                       4.77                      - 6/-10                       0.01
          69/ 65                       5.44                      -11/-15                          0
          64/ 60                       6.52                      -16/-20                          0
          59/ 55                       7.87                      -21/-25                          0
          54/ 50                       9.64                      -26/-30                          0
          49/ 45                      10.37                      -31 /-35                          0
          44/40                      10.15                      -36/-40                          0


Total Percentage   =  100.

-------
                                         TABLE A-VI  (continued)


Site:                         No. 8, Denver, Colorado

Weather Station Location:      Stapleton Airfield

Period:                       1951-60


Air Temperature Range (°F)         Percent of Time          Air Temperature Range (°F)         Percent of Time

        119/115                          0                       39/35                      8.18
        114/110                          0                       34/30                      8.22
        109/105                          0                       29/ 25                      6.31
        104/100                       0.01                       24/20                      4.09
         99/ 95                       0.11                       19/ 15                      2.46
         94/90                       1.18                       14/10                      1.36
         89/85                       2.69                        9/  5                      0.89
         84/ 80                       3.79                        4/  0                      0.41
         79/ 75                       4.98                      - I/- 5                      0.25
         74/ 70                       6.26                      - 6/-10                      0.07
         69/65                       7.80                      -11/-15                      0.01
         64/60                       8.93                      -16/-20                          0
         59/55                       8.34                      -21/-25                      0.01
         54/ 50                       7.73                      -26/-30                          0
         49/ 45                       8.03                      -31/-35                          0
         44/40                       7.89                      -36/-40                          0


Total Percentage  =  100.

-------
                                                    TABLE A-VI  (continued)


           Site:                         No.  9, Phoenix, Arizona

           Weather Station Location:      Sky Harbor Airport

           Period:                       1951-60


           Air Temperature Range (°F)         Percent of Time          Air Temperature Range (°F)         Percent of Time

                  119/115                       0.01                       39/35                       2.08
                  114/110                       0.17                       34/30                       0.65
                  109/105                       1.47                       29/25                       0.09
^                 104/100                       3.69                       24/20                          0
2                   99/95                       5.78                       19/ 15                          0
                    94/ 90                       6.97                       14/10                          0
                    89/ 85                       7.96                        9/5                          0
                    84/80                       9.10                        4/0                          0
                    79/ 75                       8.83                      - I/- 5                          0
                    74/ 70                       8.69                      - 6/-10                          0
                    69/65                       8.85                      -11/-15                          0
                    64/60                       8.75                      -16/-20                          0
                    59/55                       8.77                      -21/-25                          0
                    54/ 50                       7.52                      -26/-30                          0
                    49/45                       6.16                      -31/-35                          0
                    44/ 40                       4.46                      -36/-40                          0


          Total Percentage   =  100.

-------
                                         TABLE A- VI  (continued)
Site:

Weather Station Location:

Period:


Air Temperature Range (°F)         Percent of Time          Air Temperature Range (°F)         Percent of Time

         119/115                          0                       39/35                       6.89
         114/110                          0                       34/30                       7.45
         109/105                       0.01                       29/25                       6.27
         104/100                       0.07                       24/20                       5.40
          99/95                       0.32                       19/15                       4.23
          94/90                       0.89                       14/10                       3.85
          89/85                       1.72                        9/  5                       3.33
          84/80                       2.87                        4/  0                       3.17
          79/ 75                       4.07                      - I/- 5                       2.37
          74/70                       5.18                      - 6/-10                       1.49
          69/65                       6.46                      -11/-15                       0.88
          64/60                       7.00                      -16/-20                       0.54
          59/55                       6.91                      -21/-25                       0.23
          54/50                       6.42                      -26/-30                       0.10
          49/45                       5.93                      -31 /-35                       0.03
          44/ 40                       *  01                      -36/-40                       0.01


Total Percentage  =  100.
No. 10, Bismarck, North
Municipal Airport
1951-60
Percent of Time
0
0
0.01
0.07
0.32
0.89
1.72
2.87
4.07
5.18
6.46
7.00
6.91
6.42
5.93
5.91
Dakota


Air Temperature Range (°F)
39/ 35
34/ 30
29/ 25
24/ 20
19/ 15
14/ 10
9/ 5
4/ 0
- I/- 5
- 6/-10
-11/-15
-16/-20
-21/-25
-26/-30
-31 /-35
-36/-40

-------
                                          TABLE A-VI  (continued)


 Site:                         No. 11, Minneapolis, Minnesota

 Weather Station Location:       Minneapolis-St.  Paul International Airport

 Period:                       1951-60


 Air Temperature Range (°F)          Percent of Time          Air Temperature Range (°F)         Percent of Time

         119/115                          0                        39/35                       6.39
         114/110                          0                        34/30                       7.21
         109/105                          0                        29/25                       6.94
         104/100                          0                        24/20                       5.86
          99/ 95                       0.09                        19/ 15                       4.37
          94/ 90                       0.61                        14/ 10                       3.55
          89/85                       1.68                         9/  5                       2.81
          84/80                       3.36                         4/0                       2.12
          79/ 75                       5.34                       - I/- 5                       1 .36
          74/ 70                       7.08                       - 6/-10                       0.71
          69/65                       7.87                       -11/-15                       0.35
          64/60                       7.93                       -16/-20                       0.11
          59/ 55                       6.86                       -21/-25                       0.05
          54/50                       6.13                       -26/-30                       0.02
          49/45                       5.50                       -31 /-35                          0
          44/ 40                       5.70                       -36/-40                          0


Total Percentage  =  100.

-------
>0
o»
                                          TABLE A-VI (continued)


Site:                         No. 12, Omaha, Nebraska

Weather Station Location:      Eppley Airfield

Period:                       1951-60


Air Temperature Range (°F)         Percent of Time          Air Temperature Range (°F)         Percent of Time

        119/115                          0                       39/35                       7.47
        114/110                          0                       34/30                       7.56
        109/105                       0.01                       29/25                       5.83
        104/100                       0.15                       24/20                       4.45
          99/ 95                       0.50                       19/ 15                       3.27
          94/90                       1.70                       14/10                       2.16
          89/ 85                       3.29                        9/  5                       1.54
          84/80                       5.08                        4/  0                       1.06
          79/ 75                       6.96                      - I/- 5                       0.46
          74/ 70                       8.28                      - 6/-10                       0.17
          69/65                       8.22                      -11/-15                       0.03
          64/ 60                       6.91                      -16/-20                          0
          59/ 55                       6.37                      -21/-25                          0
          54/50                       6.15                      -26/-30                          0
          49/45                       6.19                      -31/-35                          0
          44/40                       6.19                      -36/-40                          0


Total Percentage   =  100.

-------
                                          TABLE A-VI  (continued)


 Site:                        No. 13, Little Rock, Arkansas

 Weather Station Location:      Adams Field

 Period:                      1951-60


 Air Temperature Range (°F)         Percent of Time          Air Temperature Range (°F)         Percent of Time

         119/115                           0                      39/35                       5.80
         114/110                           0                      34/30                       4.14
         109/105                       0.01                       29/25                       1.96
         104/100                       0.27                      24/20                       0.57
          99/95                       1.22                      19/15                       0.26
          94/ 90                       3.57                      14/ 10                       0.06
          89/ 85                       5.76                       9/  5                       0.01
          84/ 80                       7.93                       4/0                          0
          79/ 75                      10.80                     - I/- 5                          0
          74/ 70                      10.72                      - 6/-10                          0
          69/65                       9.16                     -11/-15                          0
          64/ 60                       8.30                     -16/-20                          0
          59/55                       7.66                     -21/-25                          0
          54/ 50                       7.27                      -26/-30                          0
          49/ 45                       7.63                      -31/-35                          0
          44/ 40                       6.90                     -36/-40                          0


Total Percentage   =  100.

-------
                                         TABLE A-VI  (continued)


Site:                         No. 14, Midland, Texas

Weather Station Location:      Midland Air Terminal

Period:                       1956-60


Air Temperature Range (°F)         Percent of Time          Air Temperature Range (°F)         Percent of Time

         119/115                          0                       39/35                       5.14
         114/110                          0                       34/30                       3.84
         109/105                       0.01                       29/25                       1.86
         104/100                       0.32                       24/20                       0.91
          99/95                       2.02                       19/15                       0.26
          94/90                       4.86                       14/10                       0.06
          89/85                       6.07                        9/  5                       0.01
          84/80                       7.68                        4/0                          0
          79/ 75                       9.86                      - I/- 5                          0
          74/ 70                       10.42                      - 6/-10                          0
          69/65                       9.04                      -11/-15                          0
          64/60                       8.21                      -16/-20                          0
          59/55                       7.73                      -21/-25                          0
          54/50                       7.20                      -26/-30                          0
          49/45                       7.63                      -31/-35                          0
          44/40                       6.87                      -36/-40                          0


Total Percentage  =   100.

-------
                                                    TABLE A-VI  (continued)


           Site:                         No. 15,  New Orleans, Louisiana

           Weather Station Location:       Moisant International Airport

           Period:                       1951-60


           Air Temperature Range (°F)          Percent of Time         Air Temperature Range (°F)         Percent of Time

                   119/115                          0                       39/35                       1.46
                   114/110                          0                       34/30                       0.54
                   109/105                          0                       29/25                       0.10
K>                 104/100                          0                       24/20                       0.02
3                  99/ 95                       0.14                       19/ 15                          0
                    94/ 90                       2.61                        14/ 10                          0
                    89/ 85                       7.07                        9/5                          0
                    84/80                      11.17                        4/0                          0
                    79/ 75                      19.06                      - I/- 5                          0
                    74/ 70                      13.56                      - 6/-10                          0
                    69/65                      11.26                      -11/-15                          0
                    64/ 60                       9.70                      -16/-20                          0
                    59/ 55                       7.89                      -21/-25                          0
                    54/ 50                       7.08                      -26/-30                          0
                    49/45                       5.12                      -31/-35                          0
                    44/ 40                       3.22                      -36/-40                          0


          Total Percentage  =  100.

-------
                                                   TABLE A-VI  (continued)


          Site:                         No. 16, Green Bay,  Wisconsin

          Weather Station Location:      Austin Straubel  Airport

          Period:                       1956-60


          Air Temperature Range (°F)         Percent of Time          Air Temperature Range (°F)         Percent of Time

                   119/115                          0                       39/35                       7.40
                   114/110                          0                       34/30                       9.35
                   109/105                          0                       29/25                       7.85
                   104/100                          0                       24/20                       5.87
§                  99/ 95                          0                       19/  15                       4.26
                    94/ 90                       0.10                       14/  10                       3.66
                    89/'85                       0.75                        9/5                       2.64
                    84/80                       2.01                        V  0                       1.83
                    79/75                       3.78                      - I/- 5                       1.08
                    74/70                       5.40                      - 6/-10                       0.48
                    69/65                       7.50                      -H/-15                       0.22
                    64/60                       8.64                      -16/-20                       0.03
                    59/55                       8.21                      -21/-25                          0
                    54/50                       6.82                      -26/-30                          0
                    49/55                       5.95                     -31/-35                          0
                    44/40                       6.17                     -36/-40                          0


           Total Percentage  =   100.

-------
                                                    TABLE A-VI (continued)


           Site:                         No. 17, Grand Rapids, Michigan

           Weather Station Location:      Kent County Airport

           Period:                       1951-60


           Air Temperature Range (°F)         Percent of Time          Air Temperature Range (°F)         Percent of Time

                   119/115                          0                       39/35                       8.46
                   114/110                          0                       34/30                      10.70
                   109/105                          0                       29/25                       7.87
                   104/100                          0                       24/20                       5.35
2                   99/95                       0.06                       19/15                       3.34
                    94/ 90                       0.47                       14/10                       1 .96
                    89/ 85                       1.56                        9/  5                       0.89
                    84/ 80                       3.27                        4/  0                       0.35
                    79/75                       5.15                      - I/- 5                       0.11
                    74/ 70                       7.23                      - 6/-10                       0.01
                    69/65                       8.42                      -11/-15                       0.01
                    64/60                       8.12                      -16/-20                          0
                    59/55                       7.37                      -21/-25                          0
                    54/ 50                       6.52                      -26/-30                          0
                    49/ 45                       6.45                      -31/-35                          0
                    44/ 40                       6.33                      -36/-40                          0


           Total Percentage  =  100.

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                                                    TABLE A-VI  (continued)


          Site:                         No. 18,  Detroit, Michigan

          Weather Station Location:      City Airport

          Period:                      1951-60


          Air Temperature Range (°F)          Percent of Time         Air Temperature Range (°F)         Percent of Time

                   119/115                          0                       39/35                      9.22
                   114/110                          0                       34/30                     10.08
                   109/105                          0                       29/25                      7.05
                   104/100                          0                       24/20                      4.30
o                   99/95                       0.01                       19/15                      2.83
                    94/90                       0.54                       14/10                      1.49
                    89/ 85                       1 .69                        9/  5                      0.70
                    84/80                       3.58                        4/0                      0.19
                    79/ 75                       5.88                      - I/- 5                      0.05
                    74/70                       8.22                      - 6/-10                      0.01
                    69/65                       8.93                      -11/-15                          0
                    64/ 60                       7.93                      -16/-20                          0
                    59/55                       7.22                      -21/-25                          0
                    54/ 50                       6.75                      -26/-30                          0
                    49/ 45                       6.45                      -31/-35                          0
                    44/ 40                       6.79                      -36/-40                          0


           Total Percentage   =   100.

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CO
                                          TABLE A-VI (continued)


 Site:                         No. 19, Chicago, Illinois

 Weather Station Location:      O1 Hare International Airport

 Period:                       1956-60


 Air Temperature Range (°F)          Percent of Time          Air Temperature Range (°F)         Percent of Time

         119/115                          0                       39/35                       8.41
         114/HO                          0                       34/30                       9.80
         109/105                          0                       29/25                       7.29
         104/100                          0                       24/20                       4.25
          99/95                       0.02                       19/15                       2.53
          94/ 90                       0.57                       14/ 10                       1.96
          89/ 85                       1 .85                        9/ 5                       1 .30
          84/ 80                       3.79                        4/0                       0 91
          79/ 75                       5.61                       - I/- 5                       0.52
          74/ 70                       8.28                      - 6/-10                       0.23
          69/ 65                       8.85                       -11/-15                       0.09
          64/ 60                       7.90                       -16/-20                         0
          59/ 55                       6.99                       -21/-25                          0
          54/ 50                       6.27                       -26/-30                         0
          49/45                       6.13                       -31/-35                          0
          44/40                       6.45                       -36/-40                          0


Total Percentage  =  100.

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                                          TABLE A- VI  (continued)


Site:                         No. 20, Nashville, Tennessee

Weather Station Location:      Berry Reid

Period:                       1951-60


Air Temperature Range (°F)          Percent of Time          Air Temperature Range (°F)         Percent of Time
               5                         0                       39/35                       645
         114/110                         0                       34/30                       5.29
         109/105                      0.01                       29/25                       3.00
         104/100                      0.13                       24/ 20                       1  51
          99/95                      0.75                       19/15                       0.77
          94/ 90                      2.59                       U/ 10                       0.32
          89/ 85                      5.06                        9/  5                       0.10
          84/ 80                      6.64                        4/  0                       0.03
          79/ 75                      9.28                      - I/- 5                       0.01
          74/ 70                      10.64                      - 6/-10                       0 01
          69/65                      9.55                      -11/-15                          0
          64/60                      8.41                      -16/-20                          0
          59/ 55                      7.97                      -21/-25                          0
          54/50                      7.26                      -26/-30                          0
          49/45                      7.06                      -31/-35                          0
          44/40                      7.16                      -36/-40                          0
Total Percentage   =  100.

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8
                                         TABLE A-VI  (continued)


Site:                         No. 21, Burlington, Vermont

Weather Station Location:      Municipal Airport

Period:                       1956-60


Air Temperature Range (°F)         Percent of Time          Air Temperature Range ( F)         Percent of Time

        119/115                          0                       39/35                       8.16
        114/110                          0                       34/30                       8.58
        109/105                          0                       29/25                       6.40
        104/100                          0                       24/ 20                       5.60
         99/ 95                       0.01                        19/ 15                       3.79
         94/90                       0.10                       14/10                       3.10
         89/ 85                       0.60                        9/  5                       2.46
         84/80                       2.16                        4/0                       1.54
         79/ 75                       4.13                      - I/- 5                       0.92
         74/ 70                       6.53                      - 6/-10                       0.44
         69/65                       7.64                      -11/-15                       0.19
         64/ 60                       8.02                      -16/-20                       0.06
         59/55                       7.91                       -21/-25                       0.02
         54/50                       7.47                      -26/-30                       0.01
         49/ 45                       6.88                      -31/-35                          0
         44/40                       7.28                      -36/-40                          0


Total Percentage   =  100.

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                                          TABLE A-VI (continued)


Site:                         No. 22, Philadelphia, Pennsylvania

Weather Station Location:      International Airport

Period:                       1951-60


Air Temperature Range (°F)         Percent of Time         Air Temperature Range (°F)         Percent of Time

         119/115                           0                       39/35                      9.33
         114/110                           0                       34/30                      7.46
         109/105                           0                       29/25                      3.82
         104/100                        0.01                        24/20                      2.16
          99/ 95                        0.19                       19/ 15                      1 .14
          94/ 90                        0.84                       14/ 10                      0.36
          89/ 85                        2.57                        9/  5                      0.10
          84/ 80                        4.79                        4/0                          0
          79/ 75                        7.47                      - I/- 5                          0
          74/ 70                        9.85                      - 6/-10                          0
          69/65                        9.22                      -11/-15                          0
          64/60                        8.38                      -16/-20                          0
          59/55                        8.10                      -21/-25                          0
          54/ 50                        7.56                      -26/-30                          0
          49/ 45                        8.00                      -31/-35                          0
          44/ 40                        8.65                      -36/-40                          0


Total Percentage  =  100.

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                                          TABLE A-VI  (continued)


 Site:                        No. 23, Charleston, West Virginia

 Weather Station Location:      Kanawha Airport

 Period:                      1956-60


 Air Temperature Range (°F)         Percent of Time         Air Temperature Range (°F)         Percent of Time

         119/115                           0                      39/35                       7.22
         114/110                           0                      34/30                       7.18
         109/105                           0                      29/25                       4.06
         104/100                           0                      24/20                       2.88
          99/ 95                       0.03                      19/  15                       1.54
          94/ 90                       0.65                      14/  10                       0.83
          89/ 85                       3.08                       9/  5                       0.25
          84/ 80                       5.38                       4/  0                       0.08
          79/ 75                       6.92                     - I/- 5                       0.01
          74/ 70                      10.40                     - 6/-10                          0
          69/65                      10.82                     -11/-15                          0
          64/60                       8.75                     -16/-20                          0
          59/ 55                       7.85                     -21/-25                          0
          54/ 50                       7.54                     -26/-30                          0
          49/ 45                       7.60                     -31/-35                          0
          44/ 40                       6.93                      -36/-40                          0


Total Percentage   =  100.

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                                         TABLE A-V1 (continued)


Site:                         No. 24, Atlanta, Georgia

Weather Station Location:      Municipal Airport

Period:                       1951-60


Air Temperature Range (°F)         Percent of Time         Air Temperature Range (°F)         Percent of Time

         119/115                          0                       39/35                       5.34
         114/110                          0                       34/30                       3.09
         109/105                          0                       29/25                       1.28
         104/100                       0.02                       24/20                       0.50
          99/ 95                       0.33                       19/  15                       0.22
          94/ 90                       2.02                       14/  10                       0.09
          89/ 85                       4.57                        9/   5                       0.02
          84/ 80                       7.13                        4/0                          0
          79/ 75                      10.07                      - I/- 5                          0
          74/ 70                      13.52                      - 6/-10                          0
          69/65                      10.56                      -11/-15                          0
          64/ 60                       9.39                      -16/-20                          0
          59/ 55                       8.94                      -21/-25                          0
          54/ 50                       8.38                      -26/-30                          0
          49/ 45                       7.71                      -31/-35                          0
          44/ 40                       6.82                      -36/-40                          0


Total  Percentage  =  100.

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                                          TABLE A-VI  (continued)
 Site:                         No. 25, Miami,  Florida

 Weather Station Location:      International Airport

 Period:                      1951-60
Air Temperature Range f*F)         Percent of Time         Air Temperature Range (°F)         Percent of Time

         119/115                          0                      39/35                       0.05
         114/110                          0                      34/30                          0
         109/105                          0                      29/25                          0
         104/100                          0                      24/ 20                          0
         99/ 95                      0.02                      19/ 15                          0
         94/90                      1.42                      14/10                          0
         89/85                     10.13                       9/5                          0
         84/ 80                     20 .47                       4/0                          0
         79/ 75                     28.09                     - I/- 5                          0
         74/ 70                     19.48                     - 6/-10                          0
         69/65                      9.24                     -11/-15                          0
         64/60                      5.15                     -16/-20                          0
         59/55                      3.16                     -21/-25                          0
         54/50                      1.68                     -26/-30                          0
         49/  45                      0.81                      -31 /-35                          0
         44/40                      0.30                     -36/-40                          0


Total Percentage   =  100.

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                                          TABLE A-VI  (continued)


Site:                         No. 26, Honolulu, Hawaii

Weather Station Location:      International Airport

Period:                       1951-60


Air Temperature Range (°F)         Percent of Time         Air Temperature Range (°F)         Percent of Time

        119/115                           0                      39/35                       0
        114/110                           0                      34/30                       0
        109/105                           0                      29/ 25                       0
        104/100                           0                      24/20                       0
         99/  95                           0                      19/ 15                       0
         94/  90                        0.01                       14/ 10                       0
         89/  85                        1.34                        9/5                       0
         84/  80                      17.70                        4/0                       0
         79/  75                      39.44                     -  I/-  5                       0
         74/  70                      31.81                      -  6/-10                       0
         69/65                        8.24                     -11/-15                       0
         64/60                        1.39                     -16/-20                       0
         59/  55                        0.07                     -21/-25                       0
         54/  50                           0                     -26/-30                       0
         49/  45                           0                     -31/-35                       0
         44/  40                           0                     -36/-40                       0


Total Percentage  =  100.

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                                          TABLE A-VI  (continued)


 Site:                         No. 27, Anchorage, Alaska

 Weather Station Location:      International Airport

 Period:                       1956-60


 Air Temperature Ronge (°F)          Percent of Time         Air Temperature Range (°F)         Percent of Time

        119/115                           0                      39/35                       8.13
        114/110                           0                      34/30                       9.48
        109/105                           0                      29/25                       9.06
        104/100                           0                      24/20                       7.55
         99/95                           0                      19/  15                       6.39
         94/ 90                           0                      14/  10                       4.28
         89/ 85                           0                       9/5                       3.25
         84/ 80                           0                       4/0                       2.41
         79/ 75                       0.09                     - I/- 5                       1 .42
         74/ 70                       0.57                     - 6/-10                       1.07
         69/65                       2.05                     -11/-15                       0.52
         64/60                       5.68                      -16/-20                       0.32
         59/ 55                      10.61                      -21/-25                       0.06
         54/50                      11.35                      -26/-30                          0
         49/ 45                       8.45                      -31/-35                          0
         44/ 40                       7.26                      -36/-40                          0


Total Percentage   =  100.

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                                  Appendix C

                           General  Specifications for
                      Dry-Type Cooling System Applications
       It is believed that dry-type cooling towers should be considered as part of the
complete cooling and condensing system including pumps,  fans (if mechanical draft),
and condenser, in order that  an economic evaluation may be made of the possible
tower selection.  Since the turbine-generator  performance  is most important in
selecting the optimum dry-type tower for a particular plant,  turbine-generator char-
acteristics over the back pressure range expected for the  various  tower selections
must also be considered.

       The following factors should be a part of the economic analysis made to de-
termine the size and  type of the dry-type cooling systems.

       1 .    Cooling tower capital cost versus  ITD for the design heat
             rejection.

       2.    Fixed-charge rate. The components of fixed-charge  rate
             are: interest or cost of money,  depreciation, interim re-
             placements, insurance and taxes.

       3,    Cost of fuel.

       4.    Operation and maintenance  costs.

       5.    Differences in turbine-generator heat rates and capital
             costs for the various back pressure designs available.

       6.    Auxiliary  power  requirements of pumps and fans including
             head-recovery turbines.

       7.    Loss of turbine-generator capability during elevated am-
             bient air  temperatures.

       8.    Cost of replacing  lost  capability  and energy.  This cost
             would consider the capital  investment, heat rate, fuel
             cost and operation and maintenance costs of the replace-
             ment capacity.

       9.    Air temperatures at the plant site.
                                      312

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       10.   Elevation of the site above sea level .

       11 .   Time of system peak electrical load demand (winter
             or summer).

       12.   Annual operation pattern of the plant.

       13.   Comparison of mechanical-draft and natural-draft towers.

       14.   Range of cooling water temperatures.

       15.   Quantity of cooling air.

       16.   Method of evaluating ground-area requirements for towers.

       After the heat load for the optimum ITD and the type of tower (mechanical-
draft or nature I-draft) have been determined, specifications should be drawn up to
cover requirements.  Included would be the following:

       1.    Wind loading and seismic design for the site.

       2.    Maximum noise level  (for mechanical-draft towers).

       3.    Corrosion protection for cooling coils and fins,  if required.

       4.    Material specifications for pumps, condenser, fan blades,
             and other components of the system .

       5.    Type of tower structure (concrete or structural steel), if
             natural draft.

       6.    Motor specifications (enclosures,  voltages, type of
             insulation).

       7.    Means  of modulation of air flow  (louvers, fan speed changes,
             fan pitch variation or other available methods).

       8.    Extent  of automation of operation and freeze protection
             desired.

       9.    Wind conditions at site (velocity and direction versus
             average hours per year).

       10.   Extent  of shop assembly of components desired.
                                      313

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1 1 .   Capacity of cooling water storage and drain and refill provi-
      sions.

12.   Hail protection required.

13.   Means of removing air and noncondensable gases from
      system.

14.   Allowable tower pumping head.

15.   Provisions for handling coils and equipment at site during
      erection and maintenance periods.

1 6.   Hydrostatic shop tests of cooling coils required.

17.   Performance guarantees and tests to be made for accept-
      ance of equipment after the plant is in operation.
                             314

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                                  Appendix D

                      Testing Upon Completion of Project


       Although there is no accepted code as yet developed for testing the perform-
ance of a dry-type cooling tower installed with a steam-electric generating plant,
the ASME Power Test Code for Atmospheric Water-Cooling Equipment, PTC-23-1958,
will serve as a useful guide  in establishing much of the required test procedure.

       The main purpose of testing the dry tower installation would be to determine
whether the guaranteed heat rejection can be achieved with the design 1TD.  In
addition, information would be obtained during the test concerning the water-pres-
sure drop through the cooling coils of the tower, the air-pressure drop across the
cooling coils, the fan brake horsepower requirements, the effect of wind upon per-
formance and the noise level of the fans.

       It is important that performance  tests be conducted during periods when the
heat rejection load  and the atmospheric conditions are stable.  After reaching
steady-state conditions, the duration of each test run should be at least 1 hour.

       Since one of the most important  aspects of the testing method is to obtain
accurate data, a program  to assure accuracy of measurements should be agreed upon
and undertaken before starting  the testing. Where the tests involve contractual ob-
ligations, the parties of the test should reach definite agreements covering the
following:

       1.    Object of the tests.

       2 .    The number  of test runs.

       3.    Allowance for measurements and errors.

       4.    The method  which will be used  to operate the equipment.

       5.    The test apparatus to be used.

       Provisions should be made to accurately measure the following:

       1 .    The flow rate of circulating water.

       2.    The temperature of circulating water to and from
             the tower.
                                      315

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       3.    The ambient air temperature.

       4.    Power input to fans (for mechanical-draft towers).

       The use of Pitot tubes for measurement of water flow is considered to be the
most accurate measurement. Calibrated orifice plates or venturi  tubes could also be
used for water-flow measurement.  The accuracy of temperature measurement should
be within  1°F.

       The following are suggested limitations for dry-type tower tests:

             Flow rate of circulating water:        - 10 percent of design
             Heat rejection:                      - 20 percent of design
             Cooling range of circulating water:   - 20 percent of design
             Ambient air temperature:             - 10°F of design

       The wind velocity during acceptance tests involving contractual  obligations
would generally be less than 10 mph.

       During the test run, the variations from maximum to minimum would be held
within the following  limits of variation:

             Circulating water flow:               5 percent
             Heat rejection:                      5 percent
             Cooling range of circulating water:   5 percent

       A time should be chosen for the test when the rate of change of the air tem-
perature does not  exceed 2°F per  hour.

       Readings should be taken at regular intervals not exceeding the following:

                                                 No./hr.

             Ambient air  temperature:                 6
             Cold water temperature:                 6
             Warm water temperature:                 6
             Circulating water flow:                  3
             Wind direction and velocity:            6
                                      316

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                                  Appendix  E

                         Cooling System Cost Structure
       Construction cost figures were developed for the various components of a dry-
type cooling system from the turbine steam exhaust flanges through the cooling towers
themselves.  The requirements of various components were determined for initial tem-
perature differences from 30°F to 80°F at 10° intervals.  These values were then used
in the computer program to determine construction costs at all  intermediate points
used in the analysis.

       The steel towers used for natural-draft towers (Figure A28) were analyzed for
12 different sizes of towers. Cost estimates were prepared for cooling range from
0.4 to 0.6 of ITD which varied the assumed water flow rates and other parameters.
Smooth curves were developed from the matrix  of estimates representing optimum
conditions so that the computer program could determine construction costs for all
intermediate sizes.

       An example of the cost structure used in our computer program for a dry-type,
natural-draft cooling system for use with an 800-mw, fossil-fueled generating plant
at sea-level elevation with a cooling system initial temperature difference of 60^F7
a range of 30°F, a cooling water circulating requirement of 266,550 gpm, a tower
height of 450 feet, a top diameter  of 350 feet and a bottom diameter of 450 feet,
based on 1970 price data,  is as follows:

Cost Estimate of Natural-Draft Cooling Tower

       A.   Steel Tower with Aluminum Siding and Heat Exchangers:

             1 .  Central tower stack to include  galvanized
                 structural steel,  aluminum siding, reinforced
                 concrete footings, steel piles and excavation:       $ 1,633,000

             2.  Bottom shed to accommodate heat exchangers
                 and to include galvanized structural steel,
                 aluminum roofing, reinforced concrete foot-
                 ings and excavation:                                 496,000

                      Total Tower Structural Costs:                $ 2,129,000

             3.   Heat exchangers:                                $ 4,408,000

                      Total Tower with Heat Exchangers:            $ 6,537,000
                                       317

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OUTER SHELL AND
ROOF MADE OF
CORRUGATED
ALUMINUM PLATES-
COOLING DELTAS
ARRANGED ON TWO
LEVELS 67' HIGH,
STIFFENING
RINGS
PREFABRICATED
STEEL FRAMEWORK
WELDED TOGETHER
                                                    o
                                                    in
                                          COOLING
                                          DELTAS
 FIGURE A28—OUTLINE OF NATURAL-DRAFT TOWER (FOR A
 60° ITD DRY-TYPE COOLING SYSTEM FOR USE WITH  AN
800 MW FOSSIL-FUELED GENERATING PLANT AT SEA LEVEL
 ELEVATION) USING STEEL AND ALUMINUM CONSTRUCTION
                         318

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       B.    Direct-Contact Condensers:                          $    832,000

       C.    Piping,  Valves, Pumps and Storage Tanks:             $ 3,381,000

       D.    Controls and Automation applicable to
             Cooling System:                                     $    500,000

                     Complete Cooling System
                     Construction Costs:                          $11,250,000

             Plus 25% for Engineering, Contingencies,
             Interest  and Taxes:                                    2,813,000

                     Total Estimated Natural-Draft
                     Cooling System Cost:                        $14,063,000

                 Cost in $Aw of Plant Capacity:  $17.60

             Note: Above numbers rounded to nearest $1,000
                   and are based  upon 1969-1970 cost levels.

       For elevations other than sea level, cost curves were also developed for
3,000 feet and 6,000  feet,  and  the computer program was arranged to interpolate
for elevations between these limits.  For  instance, a  natural-draft cooling system
for an elevation of 3,000 feet but otherwise the same as in the cost estimate above
was priced at $14,705,000  (an approximate 4-1/2 percent increase over sea-level
costs) and a system for an elevation of 6,000 feet was priced at $15,740,000 (an
approximate 10 percent increase  over sea-level costs).

       Variations in the natural-draft tower size accounted for most of the cost
difference at the various elevations used.

       A comparison of tower dimensions for the same  design conditions as for the
detailed  estimate shown above is shown below.

                  Dimensions of 800-Mw, Natural-Draft Tower
                                 for Fossil Fuel

        Elevation     Tower Height    Top Diameter     Bottom Diameter
       Above MSL         (ft.)             (ft.)               (ft.)

                                                              450
                                                              450
                                                              450
0
3,000
6,000
450
540
655
350
350
350
                                     319

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        Costs for mechanical-draft towers were developed in a similar manner with
 the basic cooling unit consisting of heat exchangers,  fans, gear boxes, motors, motor
 couplings, fan stacks, fan decks, supporting steel structure and concrete footings.
 Costs for this basic cooling unit were supplied by the  Hudson Products Corporation.
 The number of cooling units were determined and all other accessories were added
 similar to that done for the natural-draft system.  The cost figure used for a mechan-
 ical-draft, dry-type cooling system with a 60 F initial temperature difference and
 all other parameters the same as used for the sea-level unit for the natural-draft sys-
 tem is $13,281,000 and consists of the following components:

 Cost Estimate of Mechanical-Draft Cooling Tower

             Piping, Valves, Flanges and Tanks:                   $  1,427,000

             Pumps and Recovery Turbines:                           1,280,000

             Controls and Automation applicable to
             Cooling System:                                         500,000

             Direct-Contact Condenser:                               832,000

             Cooling Units:                                         6,586,000

                      Total  of Above:                             $10,625,000

             Add 25% for Engineering, Contingencies,
             Interest and Taxes:                                     2,656,000

                      Total  Estimated Mechanical-Draft
                      Cooling System Cost:                        $13,281,000

                  Cost in  $/kw of Plant Capacity: $16.60

        The cost of the mechanical-draft equipment is affected to a  lesser degree by
altitude than that for  the natural-draft system.  The total system cost for the mechan-
ical-draft, dry-type cooling system with parameters as above at an elevation of
3,000 feet is approximately $13,580,000 (an increase of approximately 2-1/4 per-
cent over sea-level costs)  and at an elevation of 6,000 feet is approximately
$13,945,000 (an  increase  of approximately 5 percent over sea-level costs).
                                      320

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                           APPENDIX REFERENCES
1A     Christopher, P. J. and Forster, V. T.  "Rugeley Dry Cooling Tower System",
       The Institution of Mechanical  Engineers — Steam Plant  Group, October,
       1969.

2A     Christopher, P.J.  "The Dry  Cooling Tower System at the Rugeley  Power
       Station of the Central  Electricity Generating Board",  English Electric
       Journal,  February, 1965.

3A     "Rugeley Power Station",  Central Electricity Generating Board, Midlands
       Region Public Relations Branch.

4A     Goecke, Ernst; Gerz, Hans-Bernd; Schwarze, Wlnfried; and Scherf, Ottokar.
       "Die  Kondensarionsanlage des 150-Mw-Blocks im  Kraftwerk Ibbenburen
       der Preussag AG", V.I.K. Berichte -  Nr. 176 - Mai, 1969.

5A     Scherf, O.  "Air Cooled Condensation Installation fora 150-Mw set  in the
       Ibbenburen Power Station", E.I.S. Translation Number 18150 from Bronnat-
       Whrmo - Kraft 20 (1968) No. 2 February.

6A     Durr, Rolf Dietrich; Von Cleve,  Hans  Henning; Kirchhubel, Erich.   "Die
       Kondensationsanlagen des Kraftwerks der Volkswagenwerk Aktieng-
       essellschaft Wolfsburg".

7A     From data furnished by Black Hills Power and Light Company.
                                     321

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                        SECTION XV
                     ACKNOWLEDGMENTS
 Acknowledgment is due to Paul R. Cunningham, Kenneth C.
O'Brien, Clarence J. Steiert, and Jack R. Lundberg for
their over-all assistance; to Donald W. Bird, Paul J.
Bride and Rodger Young for computer programming; to
Winston E. Knechtel, Jr. and T. V. Stradley for structural
design analysis and cost estimates for natural draft towers;
to Guido Chibas for special assistance; and to Dr. Francis J.
Badgeley of the University of Washington and Naydene N.
Maykut for analysis of meteorological aspects of the problem.
                            322

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BIBLIOGRAPHIC:  R. W. Beck and Associates, "Research on Dry-
Type Cooling Towers for Thermal Electric Generation."  FWQA
Publication No. 16130EES11/70.
ABSTRACT:  An economic analysis is made for the use of dry
cooling towers in thermal power plants in the United States.
Twenty-seven sites were examined providing in each case
capital and operating cost for natural and mechanical draft
systems both for fossil and nuclear plants.
System optimization was based on capital cost, auxiliary power
cost, cost due to loss of capacity, and fuel cost.

Comparison was made with wet cooling tower systems.  It was
found that with all factors considered, dry towers would be
economically competitive with wet cooling tower systems.

This report was submitted in fulfillment of Contract No.
14-12-823 under the sponsorship of the Federal Water
Quality Administration.
ACCESSION NO.
KEY WORDS:

Dry Cooling Towers
Cooling
Thermal Power Plant
Economic Evaluation
BIBLIOGRAPHIC:   R. W. Beck and Associates, "Research on Dry-
Type  Cooling Towers for Thermal Electric Generation."  FWQA
Publication No.  16130EES11/70.
ABSTRACT:  An economic analysis is made for the use of dry
cooTing  towers in thermal power plants in the United States.
Twenty-seven sites were examined providing in each case
capital  and operating cost for natural and mechanical draft
systems  for both fossil and nuclear plants.

System optimization was based on capital cost, auxiliary  power
cost, cost due  to loss of capacity, and fuel cost.

Comparison was  made with wet cooling tower systems.   It was
found that with all factors considered, dry towers would  be
economically competitive with wet cooling tower systems.
This  report was submitted in fulfillment of Contract  No.
14-12-823 under the sponsorship of the Federal Water
Quality  Administration.
ACCESSION NO.
 KEY WORDS:

 Dry Cooling Towers
 Cooling
 Thermal  Power  Plant
 Economic Evaluation
BIBLIOGRAPHIC:   R.  W.  Beck and  Associates,  "Research  on  Dry-
Type Cooling Towers for  Thermal  Electric  Generation." FWQA
publication No.  16130EES11/70.
ABSTRACT;   An economic analysis is  made for the  use of dry
pooling towers  in thermal  power plants  in the  United  States.
Twenty-seven sites were  examined providing in  each case
capital and operating  cost for  natural  and mechanical draft
systems for both fossil  and nuclear plants.
System optimization was  based on capital  cost, auxiliary power
cost,  cost due to loss  of capacity, and  fuel  cost.
Comparison was  made with wet cooling tower systems.  It  was
found that with all factors considered, dry towers would be
economically competive with wet cooling tower systems.
This report was submitted in fulfillment  of Contract  No.
 14-12-823 under the sponsorship of the  Federal Water
Quality Administration.
 ACCESSION NO.
 KEY WORDS:

 Dry Cooling Towers
 Cooling
 Thermal Power Plant
 Economic Evaluation

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    Accession Number
                        Subject Field &. Group
                                013D
                                           SELECTED WATER RESOURCES ABSTRACTS
                                                  INPUT TRANSACTION  FORM
    Organization

     R. W.  Beck  and Associates
    Title
      "Research  on  Dry-Type Cooling Towers for Thermal  Electric Generation
1Q Authors)
.Inhn PT Rossis and
Edward A. Cecil
16

21
Project Designation
FWQA Contract 14-12-823;
EES .
Note
 22
     Citation
     FWQA  R  &  D Report #16130EES11/70
 23
     Descriptors (Starred First)
     *Cooling,  *Thermal  Power Plant, *Economic Evaluation,  Water Pollution, Heat
     Exchanger,  Waste Treatment
 25
Identifiers (Starred First)

 *Dry Cooling Towers
 27
    Abstract
         An economic  analysis is made for the use of dry cooling  towers in thermal power
plants in the United  States.   Twenty-seven sites were examined  providing in each case
capital and operating cost  for natural  and mechanical draft  systems  both for fossil
and nuclear plants.                                    ,

System optimization was  based on capital cost, auxiliary power  cost, cost due to loss
of capacity, and  fuel  cost.

Comparison was made with wet  cooling ,tower systems.  It was  found that with all
factors considered, dry  towerswould be economically competitive with wet cooling
tower systems. (Shirazi, EPA)

This report was  submitted in  fulfillment of Contract No. 14-12-823 under the
sponsorship of the Federal  Water Quality Administration.
Abstractor
Mostafa  A.  Shirazi
                          Institution
                           EPA/FWOA/National Thermal
Pollution Research
 WR:IOa (REV. JULY 1969)
 WRSIC
                                           SEND TO: WATER RESOURCES SCIENTIFIC IN
                                                  U.S. DEPARTMENT OF THE INTERIOR
                                                  WASHINGTON. D. C. 20240
                                                                                     N CENTER
                                                                                * GPO: 1969-359-339

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