WATER POLLUTION CONTROL RESEARCH 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-
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Inquiries pertaining to the Water Pollution Control Research
Reports should be directed to the Head, Project Reports
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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 of Documents, U.S. Government Printing Office, Washington, D.C., 20102 - 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
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high back pressures that will be encountered with dry-
type cooling 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 for economically
optimized dry-type cooling systems are estimated to be
as follows:
Fossil-Fueled Plant
Type of tower:
Capital cost, $/kw
Initial temperature difference
% loss of capability during
hot weather
Penalty for loss of capability
at $100/kw replacement* .
Capital cost of dry tower
system plus replacement
peaking capacity
Mechanical
Draft
$17
60° F
7.6%
$ 8
$25
Natural
Draft
$20
56° F
6.4%
$ 6
$26
Nuclear-Fueled Plant
Type of tower:
Capital cost, $/kw
Initial temperature difference
% loss of capability during
hot weather
Penalty for loss of capability
at $100/kw replacement* .
Capital cost of dry tower
system plus replacement
peaking capacity
Mechanical
Draft
$23
65°F
13.6%
$14
$37
Natural
Draft
$27
62° F
12.5%
$13
$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).
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The foregoing analysis is on the basis of 5,250,000 mwh
annual production (75 percent plant factor).
8. The use of dry-type pooling systems with steam-electric
generating plants .wilf eliminate the need for a large
supply of water as a basic site requirement and will re-
sult in greater freedom of plant siting than has been
possible.
9- There are large deposits of coal and lignite in the United
States 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 i
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 6
Presently Used Methods of Rejecting Heat from Gener-
ating Stations 6
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
I ndi rect 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 air 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
VIM
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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
Genera I 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 Fouling 106
Power Plant Operation 106
Effect of Corrosion on Performance of Coils 107
V AUXILIARY EQUIPMENT 108
General 108
Condensers 108
Air Removal Equipment 110
Pumps 112
Recovery Turbi nes 112
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 Pqge
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
I ntroduction 124
Dr. Ldszlo Heller and Hoterv 124
M.A.N. (Maschinenfabrik Augsburg-Nurnberg) 126
GEA — Gesellschaft Fur Luftkondensation 126
English Electric Company 129
Brown Boveri Corporation 130
United States Turbine Manufacturers 130
Hudson Products Corporation 131
XI
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TABLE OF CONTENTS
Section Description Pa9g
The Marley Company ^ 31
Ingersoll-Rand Company 1 33
GKN Birwelco Limited 133
IX OPTIMIZATION PROGRAMS 134
I ntroduction ' 34
Method of Analysis and Description of Tower
Optimization Program 1 35
Factors Affecting the Economic Optimization of
Dry-Type Cooling Towers 1 39
Performance related to ITD 1 39
Capital cost of the dry cooling system 1 39
Elevation 141
Fixed-charge rate 141
Ambient air temperatures 141
Fuel costs 141
Turbine performance 141
Auxiliary power requirements 142
Replacement of capacity losses 142
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 CON TENTS
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 shelI 223
Condenser 223
Sector valves 225
Auxi Iiary 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
Conclusion 231
IBBENBUREN PLANT 232
Introduction 232
Description of Plant 232
Water Circuit 234
Design Parameters 236
Capital Costs 239
Manpower Requirements of the Tower 239
Winter Operation 242
Auxiliary 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
Manpower Requirements 276
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
Concl usion 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
1
2
3
4
5
6
7
8
9
10
11
12
13
14
15
16
17
National Power Survey Regions
Water Consumption Versus Wet- Bulb Temperature
Evaporative Cooling Tower Condensing System
Indirect Dry-Type Cooling Tower Condensing System With
Natural-Draft Tower
Indirect Dry-Type Cooling Tower Condensing System With
Mechanical-Draft Tower
Direct-Type Cooling Tower Condensing System
Horizontal Air-Cooled Heat Exchanger
Heat Exchanger — Fin Tube Types
Air- Evaporative Cooled Heat Exchanger Systems
Temperature Diagrams of Direct and Indirect Dry Cooling Tower
Heat Transfer Systems
Heat Transfer Effectiveness as a Function of Number of Transfer
Units (Nj-y) Crossflow Exchanger With Air Mixed
Carnot Cycle Plotted on Temperature-Entropy Diagram
Diagram of Rankine Cycle
Typical Flow Diagram for Regenerative Reheat Cycle
Coil Performance Versus Air and Water Flow
Coil Performance Versus Water Flow, Tower Height and Initial
Temperature Difference (ITD)
Cooling Units Required for Mechanical-Draft Dry-Type
Cooling System
Page
4
9
14
16
17
19
25
26
29
32
44
45
47
51
54
56
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 HI
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
On — 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 ($/Kw) for the Range of Economically Opti-
mum values of ITD Shown on Figure 43— Fossil-Fueled Generat-
ing Unit — Mechanical-Draft Tower 167
XXI
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LIST OF FIGURES
Page
56 Capital Cost of the Dry Cooling System ($/Kw) Plus Capital Cost
of Peaking Capacity ($/Kw) for the Range of Economically
Optimum Values of ITD Shown on Figure 44— Nuclear-Fueled
Generating Unit — Natural-Draft Tower 1 68
57 Capital Cost of the Dry Cooling System ($/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 1 69
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
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 196
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 Generati ng Uni t 193
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
Generati ng Uni t ] 99
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, Nature I-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 ($/Kw) for the Economically
Optimum Values of Initial Temperature Difference Shown in
Table 13, Nuclear-Fueled Generating Unit, Natural-
Draft Tower 180
xxv
-------
LIST OF TABLES
Page
22 Capital Cost of the Dry Cooling System ($/Kw) Plus Capital
Cost of Peaking Capacity ($/Kw) for the Economically
Optimum Values of Initial Temperature Difference Shown in
Table 14, Nuclear-Fueled Generating Unit, Mechanical-
Draft Tower 181
23 Optimized Total Annual Costs (in Mills per Kwh) Influenced
by the Cooling System — 800-Mw, Fossil-Fueled Generating
Unit, Natural-Draft, Dry-Type Cooling Tower System 183
24 Optimized Total Annual Costs (in Mills per Kwh) Influenced
by the Cooling System — 800-Mw, Fossil-Fueled Generating
Unit, Mechanical-Draft, Dry-Type Cooling Tower System 184
25 Optimized Total Annual Costs (in Mills per Kwh) Influenced
by the Cooling System — 800-Mw, Nuclear-Fueled Generat-
ing Unit, Natural-Draft, Dry-Type Cooling Tower System 185
26 Optimized Total Annual Costs (in Mills per Kwh) Influenced
by the Cooling System — 800-Mw, Nuclear-Fueled Generat-
ing Unit, Mechanical-Draft, Dry-Type Cooling Tower System 186
27 Auxiliary Capacity Required (in Mw) for Cooling System
Pumps at the Optimum ITD for an 800-Mw, Fossil-Fueled Unit,
Natural-Draft, Dry-Type Cooling Tower System 187
28 Auxiliary Capacity Required (in Mw) for Cooling System Pumps
and Fans at the Optimum ITD for an 800-Mw, Fossil-Fueled
Generating Unit, Mechanical-Draft, Dry-Type Cooling Tower
System 188
29 Auxiliary Capacity Required (in Mw) for Cooling System Pumps
at the Optimum ITD for an 800-Mw, Nuclear-Fueled Generat-
ing Unit, Natural-Draft, Dry-Type Cooling Tower System 189
30 Auxiliary Capacity Required (in Mw) for Cooling System Pumps
and Fans at the Optimum ITD for an 800-Mw, Nuclear-Fueled
Generating Unit, Mechanical-Draft, Dry-Type Cooling Tower
System 190
XXVI
-------
LIST OF TABLES
Page
31 Initial Temperature Differences of Dry Cooling Systems,
Existing Installations Visited 192
32 Effect of Fuel Cost on Optimum ITD (Chicago Fossil-Fueled
Plant, 15% Fixed-Charge Rate) 194
33 Effect of Peaking Capacity Cost on Optimum ITD (Fossil-
Fueled Plant, Chicago, Fuel Cost = 35$ 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
-------
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.
-------
By far the greatest heat rejection is from the main steam condenser. Other
minor heat rejections are from the generator 1 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.
-------
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 anywhich
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.
-------
LSOUTH CENTRAL
\
FEDERAL POWER COMMISSION
POWER SUPPLY AREA
REGIONS SELECTED FOR UPDATING
THE NATIONAL POWER SURVEY
FIGURE I— NATIONAL POWER SURVEY REGIONS (2)
-------
TABLE 2
Estimated Number of Thermal Generating Plant Sites
500-Mw Capacity and Above for Year 1990
Fossil-Fueled Plants by Mw Capacities
Northeast:
Total Sites
New Sites ....
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
14
4
10
129
30
38
1,000
to
2,000
15
1
5
12
3
1
26
9
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
34
6
4
62
16
17
92
49
27
25
6
6
38
9
21
292
91
83
Nuclear-Fueled Plants by Mw Capacities
"500TTOOO27000
to to to Over
1,000 2,000 4,000 4,000 Total
10
8
7
28
23
15
19
18
4
22
18
14
10
8
4
73
63
33
17
14
3
21
14
7
73
58
19
13
12
2
26
20
45
38
9
60
45
32
21
17
6
22
22
5
19
11
33
31
15
200
164
75
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Water Requirements
Almost 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.
-------
In the evaporative-type cooling tower, the circulating water is pumped
through the surface condenser where it picks up the heat of condensatipn 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 water falls 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, 11 6 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.
-------
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 Wafer 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.
-------
UJ
<
e>
3. Mechanical-draft evaporative cooling tower.
4. Spray pond.
5. Natural-draft evaporative cooling tower.
6. Natural lake or river.
2 ACRES/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.
-------
en-
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
vironmental questions associated with the utility industry as reported in (7).
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
Ibbenbiiren, Germany* 150 mw Heller 1967
Wolfsburg, Germany* 3-50 mw GEA Direct 1961-67
11
-------
TABLE 3 (continued)
Location
Grootvlei, South Africa
Gybngybs, 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
11&15 mw
5 mw
3.6 mw
38 mw
5.3 mw
1 6 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
1960-61
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
-------
\
AIR FLOW
FAN
-COOLING TOWER
/ * < » 4 I \
X—''—' • ' i ' i "\
y « 4 4 & V.
/ ' ' ' ' ' ' ' ' N>FILL OR PACKING
/ 4 4 i & 4 v
/\ I L I I I I I I l\
AIR FLOW
I i « & A 6 |\
SURFACE
CONDENSER
CIRCULATING WATER
PUMP
CONDENSATE PUMP f,
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
-------
AIR
FLOW
NATURAL-
DRAFT TOWER
.nr COOLING COILS
STEAM
TURBINE
EXHAUST
STEAM
DIRECT-CONTACT
CONDENSER
WATER RECOVERY
TURBINE
CIRCULATING PUMP
MOTOR
CIRCULATING
WATER PUMP
TO BOILER
FEEDWATER
CIRCUIT
FIGURE 4 — INDIRECT, DRY-TYPE COOLING TOWER
CONDENSING SYSTEM WITH NATURAL-DRAFT TOWER
-------
MECHANICAL-
DRAFT TOWER
- COOLING COILS
STEAM
TURBINE
EXHAUST
STEAM
DIRECT CONTACT
CONDENSER
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 diagram 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
-------
CONDENSATE
HEADER
TO BOILER
FEEDWATER
CIRCUIT
STEAM
HEADER
EXHAUST STEAM
CONDENSATE
HEADER
STEAM
TURBINE
CONDENSATE
RECEIVER
CONDENSATE
PUMP
EXHAUST-
STEAM
TRUNK
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
-------
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
-------
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
-------
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
-------
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 VIM 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
-------
SEAL DISC
TIP SEAL
PLENUM
AN RING
MOTOR-
1
oc
I
o
_J
u_
i
OOO OO
OOO TUBE OOO
-------
r*A
(a) FIN TUBES IN TRIANGULAR
PITCH ARRANGEMENT
l\\\\\\\
(b) L-SHAPE
FOOTED FIN
(c)EMBEDDED
FIN
(d) EXTRUDED
FIN
(e) OVERLAPPED
FOOTED FIN
FROM (14)
(f ) HELLER- FORGO
SLOTTED PLATE FINS
FROM (15)
FIGURE 8—HEAT EXCHANGER- FIN TUBE TYPES
26
-------
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 Figure 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
-------
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
-------
AIR COOLER
PROCESS
t \ t
FINNED
COIL
(a) DIRECT COOLING
AIR
AIR COOLER ^ f LOUVERS
tit/iff ttt ft -
I I
——
FINNED
V/'i
k __ N ___ '
I I
LOUVERS
PROCESS
EXCHANGER
| AIR
| COOLER
AIR
(d) INDIRECT COOLING WITH
SPRAY CONDENSER
AIR COOLER
LOUVERS
WARM AIR
RECIRCULATION
COOLING
TOWER
PROCESS
J V
AIR f
r^ I MPROCE
I i
PROCESS
WATER
COOLED
EQUIPMENT
(b) DIRECT COOLING
WARM AIR RECIRCULATION
(e) COMBINATION AIR COOLER-
COOLING TOWER UNIT
PROCESS
EXCHANGER
AIR COOLER
PROCESS
(c) INDIRECT COOLING WITH LIQUID
COOLED EXCHANGER
TUBE
BUNDLE
\ tAIR
SPRAYS
LOUVERS
(f ) DIRECT COOLING-
"V"TYPE AIR COOLED
EXCHANGER WITH SPRAYS
FIGURE 9—EVAPORATIVE-TYPE
HEAT EXCHANGER SYSTEMS (14)
29
-------
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
-------
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 Tsi . 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 Tsi and Tw] 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 Tw] 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 Ta] comes into contact first
31
-------
t
UJ
a:
cr
UJ
a.
2
UJ
1
UJ
ct
IT
UJ
a.
•5.
UJ
DIRECT-CONTACT
CONDENSER
TURBINE EXHAUST __ / ,.^,^., , ,,
STEAM TSI / TO TOWER
COOLING COILS
/TRANSFER OF HOT CIRCULATING
/WATER FROM CONDENSER
TRANSFER OF COLD CIRCULATING WATER
FROM TOWER TO CONDENSER
(I)
(2)
(3)
u/
cc
£
(T
UJ
o
z
UJ
OC
UJ
NOTE: THE ABOVE SKETCH DOES NOT IMPLY
(a) INDIRECT SYSTEM
AMBIENT AIR
FLOW RELATIONSHIPS
(I)
(2)
(3)
(4)
TURBINE EXHAUST
STEAM
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 1
TRANSFER OF
EXHAUST STEAM
FROM TURBINE TO
CONDENSING COILS
AMBIENT AIR
I
(5)
(6)
I
(7)
(b)
DIRECT SYSTEM
STEAM LEAVING TURBINE
STEAM ENTERING CONDENSING COILS AND AIR
LEAVING CONDENSING COILS
AIR ENTERING CONDENSING COILS AND HOT
WATER(CONDENSATE) LEAVING COILS
(5)
(6)
(7)
FIGURE 10 — TEMPERATURE DIAGRAMS OF
DIRECT AND INDIRECT DRY COOLING TOWER
HEAT-TRANSFER SYSTEMS
32
-------
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 Ts| 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 Fg [2]
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
Fq = dimensionless 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
-------
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:
A0
R = r« + -r^
or
1 = 1 +^ Tr +r +Ji+ll [4]
u ha A; L g f kt k J L 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.2 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
9
rr = fouling resistance (°F hr.)/Btu
The apparent coefficient of heat transfer of the external surface, h , 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:
= h
f
[5]
where: Ar = surface area of fins, ft.
total area of fi
and collar, ft.
A. = total area of fin surface
r
= 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 (17)]. However, the only truly accurate method of determining
hQ 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
-------
The fin efficiency, ^r, is given by the following formula:
"f =
[6]
where: f = g [ —
t = fin thickness, ft.
r = radius of curvature of fin
tip, ft.
r = radius of curvature of fin
* root, ft.
h. = mean surface coefficient of
heat transfer of a finned
surface, Btu/(hr. ft.2 °F)
k = thermal conductivity of fin,
f Btu/(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.
k2nVw
37
-------
f\
where: Ar = 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.
P = air density, Ibs./ft.3
p = number of water passes
H = height of cooler, ft.
n = number of tube rows
V = water velocity in cooler, ft ./sec.
DC = over-all crossflow heat transfer
coefficient, Btu/(hr. °Fft.2)
k, , k«, k« = 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:
Heat rejected to circulating water:
GHRW = WwCw(Twl-Tw2) [9]
38
-------
Heat rejected by finned heat exchanger:
GHRC = UALMTDFg [10]
where: C = specific heat of air at constant
a
pressure, Btu/(lb. °F)
C = specific heat of water,
W Btu/(lb. °F)
GHR = gross heat rejected to air,
a
Btu/hr.
C = gross heat rejected by coil,
Btu/hr.
GHR = gross heat rejected to circulat-
w
ing water, Btu/hr.
T i = temperature of air entering
coil, °F
T « = temperature of air leaving
coil, °F
T , = temperature of water entering
^ • I Or
COll, 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:
GHRQ = GHRw = GHRc Ql]
Solving simultaneous equations yields the following expression for gross heat
rejected at the balance point, GHR, :
ITD(ez-l)
GHR, =
,
b z
we we
v*a ^a "w Sv
where: z = F U A | ^-
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
-------
Gates expresses U as:
U = -L + ^J.. +± D4
h1 o hj 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 r- 1
' ,0.2 LIDJ
a
where: V = water velocity, ft./sec.
d = inside diameter of tube, inches
t = temperature of water, F
Equation [15J is a simplification of the more general equation for liquids
in fully developed turbulent flow given by McAdams (22) as:
C \.33 K,
Pl 4 06]
where: d = inside diameter, ft.
2
G = mass velocity, lbs./(hr. ft. )
M- = absolute viscosity, lbs./(hr. ft.)
41
-------
K, = 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 (17) 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 =
C (T -T
a<~°Uw 'a
where: E = exchanger heat transfer effective
ness, nondimensional
W = water flow rate, Ibs./hr.
W - air flow rate, Ibs./hr.
42
-------
C = specific heat of water, Ibs./hr.
C = specific heat of air, Btu/(lb.°F)
TW = temperature of water, F
T = temperature of air, F
The number of heat transfer units, Nj.U7 is a nondimensional expression of
the heat transfer size of the exchanger.
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
Ntu 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 N^ for a crossflow cooling situation with
the air assumed to be mixed is shown in Figure 11, 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
V)
V)
UJ
z
UJ
o
UJ
u.
u.
UJ
012345
NO. OF TRANSFER UNITS, NjU 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
-------
a:
UJ
a.
2
UJ
I-
HEAT AVAILABLE
FOR
WORK
HEAT UNAVAILABLE
FOR
WORK
3 ( MAXIMUM TEMPERATURE )
4 ( MINIMUM TEMPERATURE )
0 S, S4 S
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:
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] + KE1 + WH] + Q = PE2 + KE2 4- 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(H]-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)
Qr HEAT
(HEAT REJECTED
FROM CYCLE)
(PUMP WORK)
FIGURE 13—DIAGRAM OF RANKINE CYCLE
-------
Referring to Figure 13,
a. Heat added to the cycle by the boiler:
Qs = WO^-H^ [22]
b. Heat rejected from the cycle by the condenser:
Qr = W(H2-H3) [23]
c. Work done by turbine:
Wkt = WfHj-Hj) = Wr?t(H1-H2,) [24]
where: 1. = over-all turbine efficiency
H, — Hr,, = total isentropic available
energy, Btu/lb.
d . Generator output:
Wkg = Wr,g(H1-H2) = ngWkt [25]
where: »j = over-all generator
efficiency
e. Boiler feed pump work:
f. Thermal efficiency based on gross generator output:
Wk
Gross thermal efficiency = _ i_ [27]
s
g. Thermal efficiency based on net generator output:
Wk -Wk
Net thermal efficiency = a _ btp
48
-------
h. Gross turbine cycle heat rate:
r L * * 3,413
Gross heat rate = -= '—.—r?7-
gross thermal efficiency
3,413 Qs _ Qs
Wkg KWg~
i. Net turbine cycle heat rate:
3 413
Net heat rate = ——j- ', ,,. .
net thermal efficiency
3,413 Qs Qs
Wk - Wk, t KW - KW,f
g bfp g bfp
Plant net heat rate, taking into account boiler
efficiency and auxiliary power requirements:
Net plan, hea,ra,e -
(KWg _ . KWa)
where: \ = boiler efficiency
KW = generator output at generator
terminals, kw
KWi r = boiler feed pump power, kw
• bfp
KWQ = auxiliary power (excluding
boiler feed pump), kw
, r
btp
= generator work, Btu/hr.
= boiler feed pump work, Btu/hr.
[29]
[30]
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
-------
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.-hr ) + W, . (H, -H )
Gross heat rate = f f fw rhtr hr ¥- [32]
generator output
where: Wf = throttle flow, Ibs./hr.
^rhtr = rer|eater flow, Ibs./hr.
Ht = throttle enthalpy, Btu/lb.
hr = final feedwater enthalpy, Btu/lb.
H, = enthalpy leaving reheater, Btu/lb.
Hcr = enthalpy entering reheater, Btu/lb.
50
-------
Ol
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 cooling 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.'2) per row.
An analysis of the heat transfer effectiveness "E" versus heat transfer size
"Ntu" 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 mo/ing 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
CO
1.5
NOTE:
UNITS FOR ABOVE CURVES
ARE SHOWN IN RELATIVE
MAGNITUDE ONLY TO
PROTECT PROPRIETARY
INFORMATION
1.0 1.5 2.0 2.5
RELATIVE AIR FLOW
Q.
O
or
o
u
o
UJ O
o:
<8
UJ x
> K
Ul
o:
3.0
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 = .256hp'(~-^M [33]
where: D = draft, inches H«O
h = effective stack height, ft.
p1 = atmospheric pressure, inches Hg
T = ambient air temperature, R
T = inside tower air temperature, °R
y
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
-------
4.0
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 (ITO)
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 ITD'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
-------
loo- -
u.
o
I
UJ
CE
a:
UJ
Q.
2
UJ
UJ
m
TURBINE
EXHAUST
PRESSURE
6O
RELATIVE NUMBER OF COOLING UNITS
FIGURE 17—COOLING UNITS REQUIRED FOR
MECHANICAL- DRAFT, DRY-TYPE COOLING SYSTEM
58
-------
Cn
Z 30
UJ
2
UJ
CL
13
O 25
UJ
CE
cr
UJ
ao
Q.
O.
Q
2
15
10
30C
800 MW, FOSSIL FUEL PLANT
40° 50° 60° 70°
INITIAL TEMPERATURE DIFFERENCE ( I TD ), ° F
80°
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:
ITD = 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 Ibbenburen and Rugeley plants, which have been built and are in
operation:
Ibbenburen (150 mw) 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 = 0.410Q'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
-------
= DES GN POINT
2. THE PERFORMANCE
ON THIS CHART ARE IN NO WAY INTENDED
TO REFLECT SUPERIORITY OF ONE TYPE
OF COIL OVER ANOTHER
10
0
0 200 400 600 800 1000 1200
HEAT REJECTION ( I06 BTU / HQUR )
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 sfope 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 [l 9] was
established at 0.75 for natural-draft 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
-------
guaraniti of ml
Ethouil Slta/n Roll
EnthatfHf of £>Aoustf Sltam
si Prttstjrr
Intrt Ttmpfralurt at Cooling Air
Borernrltr
total Pew Ctms
at tfm Pert Shafts
calculated operating characteristic of the air-cooled steam condensing plant
Black Hills Power S Light Company - Wyodak
Cooling air t»mp*rotur*
FIGURE 20-GRAPH OF CALCULATED OPERATING CHARACTERISTICS
(PREDICTED PERFORMANCE) FOR DIRECT, AIR-COOLED CONDENSING
SYSTEM.NEIL SIMPSON PLANT, WYODAK.WYOMING (FROM GEA)
OCA
"*• llOliah
IOCHUM
-------
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 10 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
-------
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COOLING TOWER SYSTEM OPERATING CHARACTERISTIC CURVE' ITO=AQb
_l I i I I I I i L_
34567
HEAT REJECTION ( I09 BTU/HOUR)
FIGURE 21 —NATURAL-DRAFT, DRY-TYPE COOLING TOWER
OPERATING CHARACTERISTICS
(ITD = 50°F AT 4.0 x I09 BTU / HR 8 ITD=67.8°F AT 6.0 x I09 BTU/ HR )
-------
40° 50° 60° 70° 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 Hj. 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 Hnr 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 Ho— He (with
appropriate correction made for the exhaust losses of the exhaust steam which are
unavailable as work). If the exhaust pressure is raised from 2 inches Hg to 5 inches
Hg, less work is accomplished by each pound of steam by the amount H4 - H^ Btu
per pound (also corrected for exhaust losses for 5 inches Hg back pressure) because,
at 5 inches Hg back pressure, expansion of steam to "B" and corresponding work
ceases at a higher enthalpy with the result that the increment between point 4 to
point 5 is lost to the cycle.
The vertical lines 3-4' and 3-5" represent the isentropic expansion line
for a turbine with 100 percent mechanical efficiency, and the ratio Hnr - H4
compared to Hnr — H^ , or Hnr - H^ compared to H^r — H£I is a measure of the
mechanical efficiency of the turbine.
The increase in exhaust pressure will result in an increase in the turbine
cycle heat rate, since less work is accomplished for the same flow of steam through
the turbine. The increase in exhaust pressure will also result in a decrease in the
amount of work accomplished by the turbine, assuming constant steam flow.
67
-------
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REHEATER
PRESSURE
HOT REHEAT
TEMP.
INITIAL
PRESSURE
Ht INITIAL ENTHALPY
HIGH PRESSURE TURBINE '
EXPANSION LINE
Hcr ENTHALPY, STEAM
ENTERING REHEATER
Hhr ENTHALPY, STEAM
LEAVING REHEATER
THROTTLE
TEMP
INTERMEDIATE AND
LOW PRESSURE
TURBINE EXPANSION
LINE
EXPANSION LINE
END POINT(ELEP)B
EXPANSION LINE
END POINT(ELEP)"A"
COLD REHEAT /
TEMP.
H4 ENTHALPY ELEP AT 5
H5 ENTHALPY ELEP AT 2
EXHAUST PRESSURE "B"
EXHAUST PRESSURE "A"
ENTROPY, S
FIGURE 23 —DIAGRAM OF STEAM EXPANSION LINE
68
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The turbine cycle heat rate may be expressed as a function of heat input
and generation.
Heat rate = *«** 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-nc)
where: Qf = heat rejection by exhaust steam,
Btu/hr.
W = flow of exhaust steam to con-
denser, Ibs./hr.
H = 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
-------
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HEAT REJECTION X I09 BTU / HOUR
(a) EFFECT OF INCREASED TURBINE BACK PRESSURE ON TOTAL HEAT REJECTION TO
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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, 1.0 inch 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
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
-------
GO
JAN FEB MAR APR MAY JUNE JULY AUG SEPT OCT NOV DEC
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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URSINE BACK PRESSURE — INCHEJ
10 30 50 70 90 0
AMBIENT AIR TEMPERATURE — °F
50% 100%
TURBINE OUTPUT
(a) COMPARISON OF TURBINE
BACK PRESSURE AND EXHAUST
TEMPERATURE AT CONSTANT LOAD
AS A FUNCTION OF AMBIENT AIR
TEMPERATURE
(b) COMPARISON OF TURBINE
BACK PiRESSURE 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 76°F
Water cooling range 24°F
Approach to wet-bulb temperature 18 F
Terminal temperture difference 6°F
Condensing temperature: 124 F
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 88).
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
-------
67°
XI
25
II5C
100
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
Low-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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HIGH BACK PRESSURE-
E:SIGN TURBI
NE
INTERMEDIATE
.DESIGN TURBINE
CONVENTIONAL TURBINE MODIFIED
"FOR OPERATION AT HIGH EXHAUST
PRESSURE
NOTE
STEAM CONDITIONS
2400 PSIG, IOOO°/IOOO°F
BASIC CONVENTIONAL UNIT-
3.5 INCHES Hg EXHAUST
PRESSURE DESIGN ~~
SOURCE
GENERAL ELECTRIC CO.
468
EXHAUST PRESSURE (INCHES Hg)
10
12
14
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 annul us 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
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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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CONDENSER
^
f
< Jt i
"X
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s *
^~^e — '- — ^=r
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1
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WATER
TURBINE
• D •• ]
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(5)
/
I
, M V A H , SUPPLY PIPE
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I
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2
13
Q.
—
A
i
i
\Jti
.
Hi
H, RETURN
PIPE
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T.(®A T\
9 AH.couJ \
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00
5
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POSITION A
t^\ .1.
A H, COIL
2
i
ATMC
xPRESSl
U
^ M.
^^ V PRESSURE
*" AH, CONDENSER ABOVE ATMC
"^ A NOZZLES AT TOP OF
1 r.n\\
FIGURE 29 — PRESSURE HEAD DIAGRAM FOR
CIRCULATING WATER SYSTEM OF INDIRECT
DRY TOWER EQUIPPED WITH WATER TURBINE
83
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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).
MuIti-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 multi-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 T i 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/multi-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
STEAM TO
REHEATER
INTERMEDIATE
PRESSURE
TURBINE
CIRCULATING WATER IN
j
SURFACE
CONDENSER
f
1
SURFACE
CONDENSER
CIRCULATING WATER OUT
(a) DIAGRAM OF PARALLEL CONNECTION OF
CIRCULATING WATER FOR SURFACE CONDENSERS
STEAM FROM REHEATER
»•
STEAM FROM
SUPERHEATER
-»
STEAM TO
REHEATER
INTERMEDIATE
PRESSURE
TURBINE
a ri
GENERATOR
SURFACE
CONDENSER
CIRCULATING WATER IN
CIRCULATING
WATER OUT
(b) DIAGRAM OF SERIES CONNECTION OF
CIRCULATING WATER FOR SURFACE CONDENSERS
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)
TC 1
Tc2 = Tw5
t _^x*
^H^- Tw4X^^
^^
TW3
CONDENSER CONDENSER
SHELL SHELL
NO. 1 NO. 2
1
TEMP OF AMBIEN1
7
)
t
r AIR
i i
I
V)
o
h-
M
i
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. Heller'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-generator 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°FITD.
87
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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.
88
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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.
89
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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 nature I-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-
90
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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 natural-draft, dry-type cooling tower than the performance obtained under con-
91
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editions 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 Ibbenbijren 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 natural-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.
92
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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 Wyoming 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.
93
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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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UJ
UUJ
NOTE i HEAT REJECTION IS
CONSTANT
-10
0 10 20 30
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 lOOfeet,
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 Tower
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.
97
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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 ofeither 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
tp 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 Dritower 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 uncoated 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 11QQ-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
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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 1100-7-24
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.
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"There is little 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:
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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.
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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.
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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 parricu-
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.
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SECTION V
AUXILIARY EQUIPMENT
General
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) .
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20' dia bore atmospheric
exhaust branch pipe
12' bore air suction
pipe to air ejectors
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 RUGELEY STATION (31)
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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 horwell. 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 the
steam or water to the cycle, to handle the task.
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8
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 the 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
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condenser. Approximately 20 to 40 percent of the total 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
a 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 95 F cooling water during hot weather.
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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 95r 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
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Ol
STEAM TURBINE
EXTRACTED STEAM
DRY
COOLING
TOWER
DIRECT CONTACT
CONDENSER
STEAM JET EJECTOR
CIRCULATING
WATER PUMP
AUXILIARY
COOLING LOAD
AUXILIARY COOLING
WATER PUMP
FIGURE 35—AUXILIARY COOLING BY STEAM-JET REFRIGERATION
-------
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.
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SECTION VI
DRY-TYPE COOLING TOWER USE
WITH BINARY CYCLES
General
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 Ti to To , 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 1* where it is condensed and flows to the heat exchanger 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
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ENTROPY S
BINARY CYCLE TEMPERATURE-ENTROPY DIAGRAM
DRY COOLING
TOWER
CONDENSER
HEAT EXCHANGER"
FIGURE 36 — FLOW DIAGRAM OF BINARY CYCLE
WITH DRY COOLING TOWER
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by Slusorek), 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.
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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 1 8 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.
Dry-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 main trans-
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 will cause a major pollution problem. The rateof oxidation
of SC>2 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-ty,pe 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 region1 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 natural-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 author1 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 government'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 45feet
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 natural-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. (Maschinenfabrik 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 ^at load, TO6 Btu/hr.
(ITD, °F)K25
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:
()-oi/-v nnn heat Ioad, 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 heat load, IP6 Btu/hr. ff>
ITD, °F
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to
CXI
<&
FIGURE 37—DIRECT CONDENSING SYSTEM
UTRILLAS POWER STATION, SPAIN (GEA PHOTO)
-------
Installed fan power is:
heat load, 106Btu/hr.
250 x
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 Boveri 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. Miihlhauser, 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 world'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. Cates, 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 Marley1 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 in eval-
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 Section XII 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+05
AMBIENT AIR TEMP = 50
TOWER SIZING
ELEVATION = 3000
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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LEGEND: UPPER SET OF CURVES FOR FOSSIL-FUELED UNIT(4 X I09 BTU PER HOUR HEAT REJECTION)
LOWER SET OF CURVES FOR NUCLEAR-FUELED UNIT(6 X I09 BTU PER HOUR HEAT REJECTION)
SEE APPENDIX 'E' FOR OUTLINE OF STEEL TOWER
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FIGURE 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
-------
IB-
ID
in
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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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o
Q 0
V)
O
o
<40
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
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
40
50
60 70 80 30 40
INITIAL TEMPERATURE DIFFERENCE (°F)
50
60
70
80
FIGURE 40-RELATIONSHIP OF DRY COOLING SYSTEM CAPITAL COST
TO ITD AND ELEVATION-800 MW GENERATING PLANT
(1970 COST LEVEL)
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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 owndemands,
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 performqnce 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 1 0 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 1 0 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.
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TABLE 8
Heat Rejection Versus Back Pressure
for an 800-Mw Generating Unit
(Full Throttle Flow Performance)
Turbine Fossil Nuclear
Back Pressure Output Heat Rejection Output Heat Rejection
(in. Hg) (mw) (IP9 Btu/hr.) (mw) (TO9 Btu/hr.)
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 the 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 100 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
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00
TABLE 9—COMPUTER PRINT-OUT, ECONOMIC OPTIMIZATION.800 MW,FOSSIL-FUELED
GENERATING UNIT, NATURAL-DRAFT TOWER, BURLINGTON .VERMONT
CAPITAL COST FACTORS: PLANT - IS 0/0
PLANT FUEL COST - 25 cENTS/io*«6 STU
PEAKING CAPITAL COST - 100 J/KW
PEAKING CAPACITY - 15 0/0 AUXILIARIES - 15 0/0
PEAKING FUEL COST - 40 CENTS/10«6 BTU
AUXILIARY CAPITAL COST - 150 I/KW
INIT.
TEMP.
OIFF.
IOEG)
I F )
30
31
32
33
34
35
36
37
38
39
40
41
42
43
44
45
46
47
48
49
50
51
52
53
54
55
56
OPTIMUM
57
58
59
60
61
62
63
64
65
66
67
68
69
70
71
72
73
74
75
76
77
78
79
80
GROSS
ENERGY
800 MW
UNIT
(MWH)
5285897
5285716
5285511
5285289
5285049
5284789
5284495
5284161
5283799
5283413
5282994
5282538
5282023
5281468
5280876
5280236
S2795SO
5278787
5277968
5277095
5276159
5275160
5274069
5272901
5271663
5270351
5268947
5267444
5265842
5264153
5262376
5260488
52S8491
5256376
5254158
5251839
5249400
5246839
5244144
5241331
5238413
5235365
5232185
5228856
5225404
5221834
5218126
5214269
5210265
5206124
5201851
EXCESS
ENERGY
DUE TO
EXTRA
CAPACITY
(MWH)
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
CAPACITY
PENALTY
ENERGY
(MWH)
3
7
11
IS
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
49592
53109
56718
60417
AUXILIARY
ENERGY
(MWH)
93375
90624
87966
85401
82929
80550
78264
76071
73971
71964
70050
68229
66501
64866
63324
61875
60519
59256
58086
57009
56025
54878
53778
52724
51717
50756
49842
48974
48153
47378
46650
45861
45099
44364
43656
42975
42321
41694
41094
40521
39975
39402
38847
38309
37788
37284
36798
36329
35877
35442
35025
MAXIMUM
LOSS OF AUXILIARY
CAPACITY
(KW)
249
1077
1942
2846
3791
4819
5944
7144
8426
9791
11242
12786
14429
16160
17939
19797
21737
23690
25669
27694
29717
J1711
33691
35673
37664
39648
41724
43878
46089
48334
50623
52939
55279
57635
59995
62387
64850
67357
69907
72509
75182
77916
80676
83528
86455
89420
92425
95445
98306
101180
104133
POWER
IKW)
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
1566193
ANNUAL
FUEL
COST
OF
800 MW
UNIT
(I)
11969738
11969991
11970270
U970573
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 CAPACITY
EXCESS
ENERGY
(1)
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
4S25B
43997
43015
41899
40685
39385
38084
37140
36049
34859
33580
32212
31254
30158
28960
27669
PENALTY AUXILIARY
COST
IS)
4055
17490
31525
46193
61527
78244
96610
116199
137124
159400
183149
208757
236000
264701
294213
325053
358052
391350
425158
459800
494510
529939
565508
601297
637400
673574
712506
753445
795601
838571
882539
928120
975204
1022939
1071092
1120144
H71322
1224678
1279177
1334972
1392323
1451437
1512705
1575968
1640924
1706971
1774260
1843763
1911211
1979421
2049456
COST
II)
500907
486156
471919
458171
444914
432169
419915
408175
396925
386165
375918
366162
356920
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
189B46
TOTAL ANNUAL
COST OF COOLING
SYSTEM AND TOTAL
PLANT FUEL
===========3
======
(SI (MILL/KWH)
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
Ib214321
15227669
15233700
15242990
15255004
15269493
15286906
15308311
15333467
15361889
15393670
15429066
15461258
15495989
15533481
15573467
15615356
15659299
15705841
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
2.9331
2.9268
2.9218
2.9154
2.9097
2.9050
2.9012
2.8982
3. 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
2.9827
2.9916
3.0002
3.0092
3.0186
-------
SECTION X
RESULTS OF THE ECONOMIC OPTIMIZATION
Economic opHmization 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
Cn
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 (cj/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.
-------
o
o
<
3.\J
4.5
4.0
C f\,
3AJ
4 ^v
/in.
NO
N.
\
X
V
\
\
\
\
\
^
V
\
\
"**
TE: DOES NOT INCLUDE:
a. PLANT CAPITAL COST EX
COOLING SYSTEM FROM T
EXHAUST FLANGE OUT
b. OPERATION AND MAINTEN/
OF PLANT OTHER THAN
COOLING SYSTEM AND TO
1
SUMMER PEAK
LOAT
•
IAI 1 kl T* C B
JINb
—^
r^r A tS
WINTER rc.M*
LOAD
^
CEPT ^-^-^
URBINE
\NCE COST 1
APPLICABLE TO
TAL PLANT FUE
IN
-------
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 -5/ 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
-------
Ol
co
OPTIMUM ITD-FOSSIL FUEL- NATURAL-DRAFT
LEGEND:
I. BASED ON THE SITE DATA AND STUDY ASSUMPTIONS SUMMARIZED IN TABLE 9 \
WITH SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT * 100/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 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 42— ECONOMICALLY OPTIMUM VALUES OF INITIAL
TEMPERATURE DIFFERENCE (°F)—FOSSIL-FUELED
GENERATING UNIT—NATURAL-DRAFT TOWER
-------
Ol
OPTIMUM ITD-FOSSIL FUEL-MECHANICAL-DRAFT
LEGEND:
I. BASED ON THE SITE DATA AND STUDY ASSUMPTIONS SUMMARIZED IN TABLE 9
WITH SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT 9 100/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 MILLS/KWH
OF THE COST AT THE OPTIMUM POINT
(I) NOT OPTIMIZED. THE LOWEST COST WAS FOUND AT THE HIGHEST ITD VALUE INVESTIGATED, 80 «p'
FIGURE 43—ECONOMICALLY OPTIMUM VALUES OF INITIAL
TEMPERATURE DIFFERENCE (°F)— FOSSIL-FUELED
GENERATING UNIT- MECHANICAL-DRAFT TOWER
-------
OPTIMUM ITD-NUCLEAR FUEL-NATURAL-DRAFT
Oi
LEGEND:
I. BASED ON THE SITE DATA AND STUDY ASSUMPTIONS SUMMARIZED IN TABLE 9 '
WITH SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT 8 100/KW ASsJ
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
i THE LOWER FIGURES, IN PARENTHESES, INDICATE THE RANGE OF ITD VALUES
FOR WHICH THE TOTAL ANNUAL COST OF PLANT OPERATION IS WITHIN QOI MILLS/KWH
OF THE COST AT THE OPTI 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
-------
Oi
O
OPTIMUM ITD-NUCLEAR FUEL-MECHANICAL-DRAFT
LEGEND:
I BASED ON THE SITE DATA AND STUDY ASSUMPTIONS SUMMARIZED IN TABLE 9 ^
WITH SUMMER PEAKSIEXCEPT ANCHORAGE) AND PEAKING CAPACITY AT 8 100/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
1 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
-------
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
-------
Ui
CO
% 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
(I) NOT OPTIMIZED. THE LOWEST COST WAS FOUND AT THE HIGHEST ITD VALUE
INVESTIGATED, 80 °F
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
-------
Oi
% CAPACITY LOSS- FOSSIL FUEL- MECHANICAL-DRAFT
._-• "NASHVILLE ^ CA*O<-INA
• K n- 10.1 s son' n
LEGEND:
I. THE LOSSES SHOWN ARE THOSE WHICH WOULD OCCUR AT THE AMBIENT
AIR TEMPERATURE EQUALLED OR EXCEEDED 10 HOURS PER YEAR
Z SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY AT 1 IOO/ KW ASSUMED
(I) NOT OPTIMIZED. THE LOWEST COST WAS FOUND AT THE HIGHEST ITD 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
-------
% CAPACITY LOSS-NUCLEAR 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
(I) NOT OPTIMIZED. THE LOWEST COST WAS FOUND AT THE HIGHEST ITD VALUE
INVESTIGATED, 80 °F
FIGURE 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
-------
% CAPACITY LOSS-NUCLEAR FUEL-MECHANICAL-DRAFT
._.-.- . NASHVILLE ^NOBTI( cM,oU««
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
(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
-------
&
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 * 100/KW ASSUMED
(I) NOT OPTIMIZED. THE LOWEST COST WAS FOUND AT THE
HIGHEST ITD VALUE INVESTIGATED, 80° F
CAPITAL COST- FOSSIL FUEL- NATURAL-DRAFT
FIGURE 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
-------
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
-------
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
2. SUMMER PEAKS! EXCEPT ANCHORAGE) AND PEAKING CAPACITY
AT 8 100/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(8/KW)
FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 44
NUCLEAR-FUELED GENERATING UNIT— NATURAL-DRAFT TOWER
-------
CAPITAL COST-NUCLEAR 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 8 100/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
-------
COOLING AND PEAKING CAPITAL COST-FOSSIL FUEL-NATURAL-DRAFT
PHOENIX ,
3"-33.6
LEGEND:
I. SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY
AT 8 100/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 (B/KW)
PLUS CAPITAL COST OF PEAKING CAPACITY (8/KW)
FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 42
FOSSIL-FUELED GENERATING UNIT— NATURAL-DRAFT TOWER
-------
DOLING AND PEAKING CAPITAL COST-FOSSIL FUEL-MECHANICAL-DRAFT,
LEGEND^
I SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY
AT SlOO/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 (8/KW)
PLUS CAPITAL COST OF PEAKING CAPACITY ($/KW)
FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 43
FOSSIL-FUELED GENERATING UNIT—MECHANICAL-DRAFT TOWER
-------
C^^^COOLING AND PEAKING CAPITAL COST-NUCLEAR FUEL-NATURAL-DRAFT
LEGEND:
I SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY
AT BIOO/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 SYSTEM (8/KW)
PLUS CAPITAL COST OF PEAKING CAPACITY (ft/KW)
FOR THE RANGE OF ECONOMICALLY OPTIMUM VALUES OF ITD SHOWN ON FIGURE 44
NUCLEAR-FUELED GENERATING UNIT—NATURAL-DRAFT TOWER
-------
COOLING AND PEAKING CAPITAL COST-NUCLEAR FUEL-MECHANICAL-DRAFT
LEGEND:
I SUMMER PEAKS (EXCEPT ANCHORAGE) AND PEAKING CAPACITY
AT SlOO/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
-------
TABLE 11
Economically Optimum Values of Initial Temperature Difference (r)
Fossil-Fueled Generating Unit, Natural-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, 111 .
Nashville, Tenn.
Burlington, Vt.
Philadelphia, Penna.
Charleston, W. Va.
Atlanta, Ga.
Miami, Fla.
Honolulu, Hawaii
Anchorage, Alas. (1)
8%
10%
12%
15%
18%
57
57
55
57-58
56
58-59
57
56
47-48
56
55
54
49-52
52-54
51-53
57
55-56
55-56
55-56
52
56
54
55
52
48-49
49
80
57-58
57-58
55-56
58
56-57
58-62
57
57
48
56
56
55
51-53
53-55
52-54
57
56
56
56
52-53
56
54-55
55
53
49-51
51-52
80
58
57-58
56
58
56-57
58-62
58
57
49
57
56
55
51-53
54-55
53-55
57
56
56
56
53
57-58
55-56
55-56
54
51
53-54
80
58
58
56
59
57
59-62
58
58
51-52
57
57
55-56
52-54
54-56
53-55
57-58
57
57
57
54
57
55-56
56
54-55
52
53
80
58
58
56-57
59-61
58
59-62
58
58
52-53
57
57
56
53-54
55-56
54-55
58
57
57
57
54-55
57
56
56-57
55
53
53-54
80
Note: Based upon the site data and study assumptions summarized in Table 10
with summer peaks (except Anchorage), peaking capacity at $100/1
-------
TABLE 12
Economically Optimum Values of Initial Temperature Difference (°F)
Fossil-Fueled Generating Unit, Mechanical-Draft Tower
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. (1)
Initial Temperature Difference (°F)
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
-------
TABLE 13
Economically Optimum Values of Initial Temperature Difference (°F)
Nuclear-Fueled Generating Unit, Natural-Draft Tower
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. (1)
Initial Temperature Difference (°F)
8%
10%
12%
15%
18%
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
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.
172
-------
TABLE 14
Economically Optimum Values of Initial Temperature Difference (°F)
Nuclear-Fueled Generating Unit, Mechanical-Draft Tower
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. (1)
Initial Temperature Difference (°F)
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 $100Aw;
fan and pump replacement capacity at $225/kw.
(1) Not optimized — winter peak assumed.
173
-------
TABLE 15
Capital Cost of the Dry Cooling System ($/Kw)
for the Economically Optimum Values
of Initial Temperature Difference Shown in Table 11
Fossil-Fueled Generating Unit, Nature I-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% T2~%"T5%" T8%~
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 $100Aw;
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% 18%
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, 111. 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 $225/kw.
(1) Not optimized — winter peak assumed.
177
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TABLE 19
Capital Cost of the Dry Cooling System ($/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: ~H% 10% T2% 13*%" T8%
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 $150/kw.
(1) Not optimized— winter peak assumed.
178
-------
TABLE 20
Capital Cost of the Dry Cooling System f$/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
$/Kw
Fixed-C^rge Rate: ~8% T0% T2%T5% T
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, Ga. 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
$/K
w
Fixed-Charge Rate- 8% 10% 12% 15% 18%
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, Ariz^ 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
$/Kw
Fixed-Charge Rate: 8% 10% 12% 15% 18%
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
Natural-Draft, Dry-Type Cooling Tower System
00
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.
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, switchgear equipment, and associated structures and foundations)
-------
TABLE 24
Optimized Total Annual Costs (in Mills per Kwh) Influenced by the Cooling System
800-Mw, Fossi I - Fue I ed 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
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
Medium Fuel Cost Range
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)
d) the total annual plo
(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 equipment associated w,th the bo.ler and turbine-generator, step-up transformer, switchgear equipment, and associated structures and foundations).
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 auxihary power and energy required for the cooling system; c) the annual cost of replacing capacity and energy lost at high turbine back pressures- and
a; the total annual plant tuel cost. r /
-------
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
00
Oi
Low Fuel Cost Range
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
1
3%
.72
.73
.75
.77
.83
.75
.84
.80
.97
.80
.78
.83
.80
.84
.79
.74
.78
.80
.80
.80
.72
.80
.79
.81
.86
.86
.62
10%
.83
.85
.88
.89
.96
.88
.98
.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
1
1
1
2
2
2
2
2
2
2
2
2
2
2
2
2
1
2
2
2
2
1
2
2
2
2
2
1
2%
.95
.97
.00
.01
.09
.00
.11
.05
.27
.05
.03
.10
.06
.10
.05
.98
.03
.06
.06
.06
.95
.06
.05
.07
.14
.13
.81
1
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
1
5%
.13
.15
.19
.20
.29
.18-
.31
.24
.50
.25
.22
.30
.25
.30
.23
.16
.22
.26
.25
.25
.13
.26
.24
.27
.34
.33
.95
1
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
2
8%
.30
.33
.37
.38
.49
.36
.51
.43
.72
.44
.41
.50
.45
.50
.42
.33
.40
.45
.44
.44
.30
.45
.43
.46
.54
.53
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
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.
Nashvilie, Tenn.
Burlington, Vt.
Philadelphia, Penna .
Charleston, W. Va.
Atlanta, Ga.
Miami, Fla .
Honolulu, Hawaii
Anchorage, Alas.
(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. ^lo^uies, ana
(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).
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
-------
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
Low Fuel Cost Range
Medium Fuel Cost Range
8%
1.77
1 .78
1 .81
1.81
1 .87
1 .79
1 .88
1 .83
2.03
1.84
1 .83
1 .88
1 .86
1 .89
1.85
1 .79
1 .82
1.85
.84
.86
.77
.86
.85
.87
.91
.91
1 .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
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
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.
00 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 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).
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
3.0V
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
-------
TABLE 27
Auxiliary Capacity Required (in Mw) for Cooling System Pumps
at the Optimum ITD for an 800-Mw, Fossil-Fueled Generating Unit
Nature I -Draft, Dry-Type Cooling Tower System
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 .
Nashvi Me, Tenn.
Burlington, Vt.
Philadelphia, Penna.
Charleston, W. Va.
Atlanta, Ga.
Miami, Fla.
Honolulu, Hawaii
Anchorage, Alas.
Low Fuel Cost Range
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
Medium Fuel Cost Range
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
(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.
-------
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
CO
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
Low Fuel Cost Range
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.
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
Medium Fuel Cost Range
High Fuel Cost Range
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
10.0
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 head
and to control the water pressure in the condenser spray nozzles.
-------
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, Ga.
Miami, Fla.
Honolulu, Hawaii
Anchorage, Alas.
Low Fuel Cost Range
Medium 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
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 control 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 whicKare 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)
Power Plant
Dry-Cooled Generating Units
No. Capacity per Unit
1
120 mw
Cooling System
Type
Natural draft
Indirect cooling system
Design ITD
35(1)
-o
Ibbenburen
(Germany)
150 mw
Natural draft
Indirect cooling system
50
Volkswagen
(Germany)
50 mw
Mechanical draft
Direct condensing system
51
Gyongybs
(Hungary)
2
2
100 mw
200 mw
Natural draft
Indirect cooling system
46
47
Neil Simpson
(Wyoming, U.S.)
20 mw
Mechanical draft
Direct condensing system
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 $28 per kw . The corresponding values for
nuclear-fueled units are shown on Figures 56 and 57.
The effects of the various parameters on the economic optimization analyses
have been studied and are discussed below.
Effect of Fixed-Charge Rate
The effect of increasing the fixed-charge rate is to increase the value of
the economically optimum ITD and, therefore, to reduce the cooling system invest-
ment. This effect is clearly shown in Tables 11, 12, 13 and 14. In these tables ,
the optimum ITD at a fixed-charge rate of 18 percent is a few degrees higher than
the optimum ITD at a fixed-charge rate of 8 percent.
Effect of Fuel Cost
The effect of increasing the plant fuel cost is to reduce the optimum ITD.
This reduction in the optimum ITD increases the cooling system investment and im-
proves the plant efficiency. The range of fuel costs investigated in our analyses
had only a minor effect on the optimum ITD. In many cases, as shown in Tables 11
through 14, the optimum ITD was not sufficiently affected by the fuel cost to change
the ITD by a full degree F. In other cases, the fuel cost difference caused the op-
timum ITD value to vary over a range of 1°F to 4°F for a given fixed-charge rate.
In order to test the effect of varying the fuel cost over a somewhat wider
range, some supplemental runs were made for the Chicago site assuming fossil-fuel
costs ranging from 1 0£ to 80$ per million Btu. The results of these analyses are
summarized in Table 32.
193
-------
TABLE 32
Effect of Fuel Cost on Optimum ITD
(Chicago, fossil-fueled plant, 15% fixed-charge rate)
Fuel Cost Optimum ITD (°F)
(C/106 Btu) Natural Draft Mechanical Draft
10 57 61
25 57 61
35 57 61
40 57 61
80 55 59
/
Effect of Air Temperatures
The results of the economic optimization analyses indicated that the optimum
ITD was affected both by the general level of ambient air temperatures and by the
summer high temperatures. The results have been analyzed by plotting optimum ITD
values as a function of both the median air temperature and the air temperature
which is equalled or exceeded only 10 hours each year. Figure 58 shows such a plot
for fossil-fueled units with natural-draft towers assuming a fixed-charge rate of 15
percent and a fuel cost of 25£ per million Btu. Figures 59 through 61 show this type
of plot for mechanical-draft fossil, natural-draft nuclear, and mechanical-draft
nuclear generating units. The grouping of values on these figures indicates the
general relationship between these parameters, but there is some scatter of the data.
Apparently, this scatter is due primarily to the differences in site elevations and in
capital cost multipliers which were reflected in the results of the basic analyses.
Therefore, supplemental analyses were made using the actual air temperature data
from all the sites and assuming uniform elevations, capital cost multipliers and fuel
costs. The results of these analyses are shown in Figures 62 through 65. Isolinesof
optimum ITD values have been drawn on these figures to indicate the nature of the
relationship between the optimum ITD and air temperatures based on the results of
the studies reported herein. The isolines were drawn as straight lines which appar-
ently fit the data, but it is recognized that additional sites should be investigated
to better define this relationship between the optimum ITD and'ambient air tempera-
tures.
Effect of Assumptions as to Lost Capacity
As discussed in Section IX, the assumptions as to the replacement of lost
generating capacity at high ambient air temperatures and the cost of that capacity
have a significant effect on the economic optimization . The results reported here-
in are based on the assumption that capacity lost at high air temperatures must be
194
-------
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58
OREEN BAY
OPTIMUM ITD VALUES
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APIDS
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55
MIDLAND
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52
LEGEND: i. FIXED CHARGE RATE = is %
2. FUEL COST=250/IO« BTU
3. ACTUAL SITE ELEVATION
4. CAPITAL COST MULTIPLIER
APPLICABLE TO SITE
o
UJ
"80 85 90 95 100 105 110 115
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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64
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64
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62
GREEN BAY
OPTIMUM ITD VALUES
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60
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61 *"60
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60* GRAND R
60
* MINNEAI
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62
CASPER
BISM/
LITTLE ROCK
• • 55
96 NASHV
ILADELPHIA
OMAHA
58
BOISE
60
APIDS
*RENO
OLIS
RCK
*60
57
MIDLAND
ILLE
LEGEND: 1. FIXED CHARGE RATE = 15 %
PHOEMX
51
2. FUEL COST=25 0/I06 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 FOR A FOSSIL-FUELED 800 MW GENERATING UNIT
196
-------
80
70
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•53
LOS ANGELES
61
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67
N6TON 67
• • G
65
66
•
GREEN BAY
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ATLANTA
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•
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63
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66 65
"••GRAND RA
66
* MINNEAI
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,69
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A ITD VA
LITTLE ROCK
• • 59
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LA DELPHI A
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64
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65
PIDS
• 66
RENO
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BISMARCK
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LUES
MIDLAND
ILLE
PHOENIX
59
JD: 1. FIXED CHARGE RATE - 15%
Z. FUEL COST= 15 0/10 « BTU
3. ACTUAL SITE ELEVATION
4. CAPITAL COST MULTIPLIER
APPLICABLE TO SITE
80 85 90 95 IOO 105 110 115
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
ECONOMICALLY OPTIMUM ITD VALUES
LJ
tr
cr
LJ
Q.
2
Lu
o
LJ
LEGEND: i. FIXED CHARGE RATE = 15%
2 FUELCOST=25#/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 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
75
70
ECONOMICALLY OPTIMUM ITD VALUES
54
51 '
65
60
.55
LJ
OC.
UJ
0.
2
UJ
\
•57
-------
80
75
40
35
ECONOMICALLY OPTIMUM ITD VALUES
LEGEND' I. FIXED CHARGE RATE = 15%
2. FUEL COST= I 507 10 6 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
70
t65
LJ
£ 60
LU
Q.
UJ
I-
tr. 55
111
50
45
40
35
ECONOMICALLY OPTIMUM ITD VALUES
54
i 60
64\
62*
62*
62-64
• 53
• 56
> 60
60 •
60
•0
56
• 57 • 57
58
..60
• 60
60
55
55-57
51
LEGEND: i. FIXED CHARGE RATE = 15 %
2. F U EL COST: 15^/10 6 BTU
3. ELEVATION = 0
4. CAPITAL COST MULTIPLIERS: 1.0
1 1 1
80 85 90 95 100 105 110 115
AIR TEMPERATURE EQUALLED OR EXCEEDED TEN HOURS PER YEAR(°F)
FIGURE 65-RELATIONSHIPOF 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 $150 per kw—was studied. The results of these
analyses are indicated in Table 33.
TABLE 33
Effect of Peaking Capacity Cost on Optimum 1TD
(fossil-fueled plant, Chicago,
fuel cost - 35$ 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°-57 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 for a 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
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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
percent 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: 7478"
(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 —$11 .OOAw.
(4) The cost figures in this table are based upon a 15% fixed-charge rate, 35(J/
10" Btu steam turbine fuel cost and 40£/10° Btu gas turbine fuel cost.
(5) Weather data used is listed in Table A-VI of Appendix B.
(6) The evaporative-type system analyzed is based upon a range of 24 F, an
approach of 18°F and a terminal temperature difference of 6°F.
(7) The dry-type system initial temperature difference is 61°F.
205
-------
As shown above, the cost penalty of the dry-type cooling tower system is a
small percentage of the total cost of power and energy delivered to the customer,
even if no offsetting cost savings are assumed. It should be pointed out, however,
that, since the dry cooling system requires very little water, an electric utility
would have much more flexibility in locating its generating plant and would have
the opportunity for certain cost savings. For example, a saving in fuel cost of
approximately 5$ 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 £ 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-
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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-
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5. Mauser, 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 mit
luftgekuhltem Kuhlwasserkrelslauf und die Moglichkeiten der
Weiterentwicklung", World Power Conference, Vienna, 1956.
207
-------
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
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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).
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(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 Krafrwerk 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
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36. Aynsley, E. "Cooling Tower Effects: Studies Abound", Electric World,
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37. Appelman, H. S. and Coons, F. G. "The Use of Jet Aircraft Engines to
Dissipate Warm Fog", Journal of Applied Meteorology, June, 1970.
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209
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39. Waselkow, C. "Design and Operation of Cooling Towers", Federal Water
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40. "Hydroelectric Power Evaluation", Federal Power Commission, 1968.
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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 Gyongyos 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 cooling 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 cool ing 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 Nature I-Draft Tower Performance 244
Al 6 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 Ait—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 Gybngybs 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 Gybngyos Power Station 267
A22 Water Circuit for Heller Dry Tower, Gybngyos 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-ll Operating Data — Preussag-Kraftwerk, 150-Mw Turbine
Generator — Ibbenburen, 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-Draft, 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 coolingtower 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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FIGURE A
RUGELEY GENERATING STATION
SITE LAYOUT
(3A)
-------
K>
00
FIGURE A2—RUGELEY POWER STATION
-------
STEAM TURBINE (120 MW)
SPRAY VALVES
WATER
TURBINE-*/
SPRAY CONDENSER
MOTOR
^^
NATURAL DRAUGHT1
COOLING TOWER
AUXILIARIES
^CIRCULATING WATER
EXTRACTION PUMPS
QUADRANTS
SECTOR
VALVES
-TO BOILER WATER
EXTRACTION PUMPS
TRANSFER.
VALVE —'
BYPASS
VALVE
] ! EMERGENCY DRAIN 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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120
IOO
0 10 20 30 40 50 60 70 80 90
AMBIENT AIR TEMPERATURE (°F)
ONE QUADRANT OPERATION
0 10 20 30 40 50 60 70 80 90
AMBIENT AIR TEMPERATURE (°F)
TWO QUADRANT OPERATION
0 10 20 30 40 50 60 70 OO 90
AMBIENT AIR TEMPERATURE (°F)
THREE QUADRANT OPERATION
0 10 20 30 40 50 60 70 80 9O
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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Tests 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 unitwas
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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ro
FIGURE A5—INSIDE VIEW, NATURAL-DRAFT, DRY-TYPE COOLING TOWER — RUGELEY STATION
(ENGLISH ELECTRIC PHOTO)
-------
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 1°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 8.9 mw 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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N5
O
FIGURE A6—SECTOR VALVES, VALVE HOUSE LOCATED INSIDE NATURAL-DRAFT,
DRY-TYPE COOLING TOWER RUGELEY STATION (ENGLISH ELECTRIC PHOTO)
-------
TABLE A-l
hO
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
41.0
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.
4 S.W.
8S.E.
8 N.
25 N.E. /N.W.
10E.
20 E.
10 W. /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
-------
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
deposits.
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>
X
I
Z
o
o
o
111
IT
0.4
0.3
0.2
0 10 20 30 40
AVERAGE WIND SPEED-MPH
FIGURE A7—WIND EFFECT UPON PER FORMANCE (I A)
0 JO" 60° SO* I2O° ISCP WO" 210° 2«O° 27O° 3OO" 33O° 360"
POSITION AROUND TOWER
FIGURE A8-AIR MASS VELOCITY VARIATION
THROUGH THE COOLERS WITH A 20 M.PH. 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 Rrm 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.
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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
IBBENBU'REN (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 33t|: 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 Preussag's water-treatment plant, and 65tJ 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
-------
FIGURE AIO— COOLING WATER CIRCUIT DIAGRAM —150 MW
IBBENBU'REN GENERATING STATION (5A)
235
-------
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./ 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. CONDENSATE INLET TO DEAERATOR
FIGURE All — DIRECT CONTACT CONDENSER (5A)
237
-------
CO
00
NOTE = CALCULATED VALUES
(5)
50 60 70
AMBIENT AIR TEMPERATURE (°F)
FIGURE A 12 — IBBENBUREN PLANT—NATURAL-DRAFT, DRY-TYPE COOLING TOWER
TURBINE BACK PRESSURE VARIATION WITH AMBIENT AIR TEMPERATURE
-------
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
-------
FIGURE AI3— HYPERBOLIC CONCRETE DRY-TYPE COOLING
TOWER INSTALLATION AT IBBENBUREN- 150 MW
GENERATING PLANT
(GEA PHOTO)
240
-------
+20AA
120
MW
+10
50
-10
(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 NSTRUMENT(5A)
241
-------
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
-------
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 l,640kw,
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
-------
6
UJ uj 6
Q >
Q
UJ Z c
ft- — -J
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-------
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 milligrams per 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
20
10
WARM WATER TEMPERATURE
COLD WATER TEMPERATURE
JDEAL VALUE OF THE COLD
WATER TEMPERATURE
AMBIENT AIR TEMPERATURE
10 12 14 16 18 20 22 24 2
DEVIATION OF MEASURED
COLD WATER TEMPERATURE
FROM IDEAL VALUE -,_|
(m/s)
3.
'
n-
^^^^
r—
j
j
m
WIND VELOCITY
_L
E^^M
8 10 12 14 HOUR
FIGURE AI6— DRY-.TYPE, NATURAL-DRAFT
COOLING TOWER: IBBENBUREN PLANT PERFORMANCE
TEST RESULTS (4A)
246
-------
Ni
TABLE A-ll
Operating Data — Preussag-Kraftwerk
150-Mw Turbine-Generator — Ibbenbiiren
with Natural-Draft Dry Cooling Tower
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
1.0
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
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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 Ibbenburen, 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
-------
K)
Ol
O
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 plant are
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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J TURBINE
SATORLUFTER" OR STANDARD COILS
GMATORLUFTER" OR COUNTER- FLOW COILS
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 Al9), 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-III, which was obtained from actual operating records of the V-W plant.
253
-------
K)
Betriebsdiagramm der GEA-Luftkandensattonsanlage KW-Nord
Block C VW-Wolfsburg ArNr 221/2828
FIGURE AI9-CALCULATED OPERATING CHARACTERISTICS (PREDICTED PERFORMANCE) FOR
THE DIRECT AIR-COOLED CONDENSING SYSTEM-BLOCK 'C1 OF VOLKSWAGEN PLANT (FROM GEA)
GEA
-------
TABLE A-l
NO
Ol
01
Operating Data — Power Station "Wolfsburg"
of the Volkswagenwerk 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
Pressu re
I (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 fans)
6
12
12
9 3 -
6 -
1
2
2
1
Full Speed
Half Speed
Full Speed
Full Speed
Half Speed
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 . Ocj: 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:
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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.
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Freezing Problems
The method 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
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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 Figure 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 .
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The Roman numerals I, II, III, and IV refer to the four zones shown on the
curve in different shades of cross-hatching .
I - AIM2 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 —/ 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 Requirement's
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
107 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.
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Although the exhaust steam from the turbines contains 11 percent moisture,
no erosion of tubes or piping has been reported.
Effect of Wind on Performance
Because the air-cooled condensers are equipped with mechanically 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
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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.
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GYONGYOS STATION
Introduction
On March 24, 1970 John P. Rossie, accompanied by Dr. L. Forgo of Hoterv
and Mr. Istvan Lindner of Transelektro, visited the Gybngyos steam-electric power
station located at Gybngyos, 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
100-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 2.00-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
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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 stator 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.
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FIGURE A20—GYONGYOS POWER STATION
TWO REINFORCED CONCRETE DRY-TYPE COOLING
TOWERS FOR 100 MW GENERATING UNITS (HOTERV PHOTO)
-------
K)
O
FIGURE A2I—REINFORCED CONCRETE TOWER FOR FIRST OF TWO
200 MW GENERATING UNITS IN THE GYONGYOS POWER STATIONtHOTERV PHOTO)
-------
NATURAL DRAFT
TOWER
S3
STEAM
TURBINE
WATER RECOVERY
TURBINE
TO BOILER
FEEDWATER
CIRCUIT
MOTOR DRIVEN
CIRCULATING
WATER PUMPS
I
FIGURE A22—WATER CIRCUIT FOR HELLER DRY TOWER
GYONGYOS STATION
^FILLING
f PUMPS
f==^ 1
h=^===~]
\^====^J
STORAGE
TANK
-------
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 Gybngyos 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
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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.
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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, 467r 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.
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Conclusion
The use of the Heller-type dry tower with the Gyongybs 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 Gybngybs 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.
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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 a special
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.
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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
275
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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
-------
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
277
-------
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 Ijght-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-shaped carbon steel 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 coolingwater
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
16
30
44
KJ CO
00 OZ
4
43
- 8
-22
-18
46
26
74
80
Operating Data -
- Neil Simpson Station
20-Mw Turbine-Generator with Mechanical-Draft,
Gross Output
(lew)
20,687
20,479
19,604
20,400
20,600
17,500
19,667
21,560
21,267
20,600
22,600
22,000
21,600
Direct Air-Cooled
Steam Flow
(Ibs./hr.)
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./hr.)
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
-------
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
Weatherstation 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 'o.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 \ F) Percent of Time
119/115 0 39/35 1.13
114/110 0 34/30 0.11
109/105 0 29/ 25 0
fo 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/ 7Q 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.
-------
TABLE A-VI (continued)
Site: No. 3, Los Angeles, California
Weather Station Location: International Airport
Period: 1951-60
Air Temperature Range (r) 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
53 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.
-------
oo
CO
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.
-------
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.
-------
•o
o
TABLE A-VI (continued)
Site: No. 6, Casper, Wyoming
Weather Station Location: Casper Air Terminal
Period: 1956-60
Air Temperature Range (°F) 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.
-------
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.
-------
-o
to
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
M 104/100 3.69 24/ 20 0
% 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 -31A35 0
44/ 40 4.46 -36/-40 0
Total Percentage = 100.
-------
TABLE A-VI (continued)
Site: No. 10, Bismarck, North Dakota
Weather Station Location: Municipal Airport
Period: 1951-60
Air Temperature Range fr) 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 5.91 -36/-40 0.01
Total Percentage = 100.
-------
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
01 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.
-------
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
,,0 104/100 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
NJ 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 4/ 0 1.83
79/ 75 3.78 - I/- 5 1.08
74/ 70 5.40 - 6/-10 0.48
69/ 65 7.50 -11/-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:
Weather Station Location:
Period:
Air Temperature Range (r) 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
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.
No. 17, Grand Rapids,
Kent County Airport
1951-60
Percent of Time
0
0
0
0
0.06
0.47
1 .56
3.27
5.15
7.23
8.42
8.12
7.37
6.52
6.45
6.33
Michigan
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. 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
g 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.
-------
o
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
114/110 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.
-------
TABLE A-VI (continued)
Site: No. 20, Nashville, Tennessee
Weather Station Location: Berry Field
Period: 1951-60
Air Temperature Range (°F) Percent of Time Air Temperature Range (r) Percent of Time
119/115 0 39/ 35 6.45
114/110 0 34/30 5.29
109/105 0.01 29/25 3.00
co 104/100 0.13 24/ 20 1.51
2 99/ 95 0.75 19/ 15 0.77
94/ 90 2.59 14/ 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.
-------
o
Ol
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 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.
No. 21, Burlington, Vermont
Municipal Airport
1956-60
Percent of Time
0
0
0
0
0.01
0.10
0.60
2.16
4.13
6.53
7.64
8.02
7.91
7.47
6.88
7.28
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. 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.
-------
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.
-------
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.
-------
TABLE A-VI (continued)
Site: No. 25, Miami, Florida
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 0.05
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.
-------
TABLE A-VI (continued)
Site: No. 26, Honolulu^Hawaii
Weather Station Location: International Airport
Period: 1951-60
Air Temperature Range (r) 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.
-------
TABLE A~VI (continued)
Site: No. 27, Anchorage, Alaska
Weather Station Location: International Airport
Period: 1956-60
Air Temperature Range (°F) Percent of Time Air Temperature Range (°F) Percent of Time
119/115 0 39/ 35 8.13
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.
-------
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
-------
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
-------
11 . 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.
16. 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
follovving:
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
-------
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 F,
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,OOP
Total Tower Structural Costs: $ 2,129,000
3. Heat exchangers: $ 4,408,000
Total Tower with Heat Exchangers: $ 6,537,000
317
-------
350' DIA.
OUTER SHELL AND
ROOF MADE OF
CORRUGATED
ALUMINUM PLATES*
COOLING DELTAS
ARRANGED ON TWO
LEVELS 67' HIGH,
STIFFENING
RINGS
xxxxx
PREFABRICATED
STEEL FRAMEWORK
WELDED TOGETHER
COOLING
DELTAS
450' DIA.
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
-------
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
Elevation
Above MSL
0
3,000
6,000
for
Tower Height
(ft.)
450
540
655
Fossil Fuel
Top Diameter
(ft.)
350
350
350
Bottom Diameter
(ft.)
450
450
450
319
-------
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 aJI 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 $Aw 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
-------
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, Winfried; and Scherf, Ottokar.
"Die Kondensationsanlage 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 for a 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
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 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
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 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
-------
1
5
Accession Number-
„ Subject Field & Group
013D
SELECTED WATER RESOURCES ABSTRACTS
INPUT TRANSACTION FORM
Organization
Title
"Research on Dry-Type Cooling Towers for Thermal Electric Generation
10
Authors)
John P. Rossie and
Edward A. Cecil
16
Project Designation
FWQA Contract 14-12-823; EES
21
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 frn 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. (Shirazi, EPA)
This report was submitted in fulfillment of Contract No. 14-12-823 under the
sponsorship of the Federal Water Quality Administration.
Abstractor
JMostafa A. Shirazi
Institution
EPA/FWQA/National Thermal PnllnHnn Research
AR:I02 (REV. JULY 1969)
WRSIC
SEND TO: WATER RESOURCES SCIENTIFIC INF
U.S. DEPARTMENT OF THE INTERIOR
WASHINGTON, D. C. 20240
N CENTER
* GPo: 1969-359-339
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