WATER POLLUTION CONTROL RESEARCH SERIES •16130 ONE 03/71
  Advanced  Nonthermally  Polluting
           Gas Turbines in
          Utility Applications
ENVIRONMENTAL PROTECTION AGENCY • WATER QUALITY OFFICE

-------
     WATER POLLUTION CONTROL RESEARCH SERIES
The Water Pollution Control Research Series describes
the results and progress in the control and abatement
of pollution in our Nation's waters.  They provide a
central source of information on the research , develop-
ment, and demonstration activities in the Water Quality
Office, Environmental Protection Agency, through inhouse
research and grants and contracts with Federal, State,
and local agencies, research institutions, and industrial
organizations.

Inquiries pertaining to Water Pollution Control Research
Reports should be directed to the Head, Project Reports
System, Office of Research and Development, Water Quality
Office, Environmental Protection Agency, Room 1108,
Washington, D. C.  20242.

-------
        ADVANCED NONTHERMALLY POLLUTING GAS TURBINES
                   IN UTILITY APPLICATIONS
            United Aircraft  Research Laboratories
              of the United  Aircraft Corporation
              East Hartford, Connecticut  06108
                            for the
               Environmental Protection Agency
                   Water Quality Office
                      Project #16130  ONE
                    Contract Ho. llj-12-593

                           March 1971
For sale by the Superintendent of Documents, U.S. Government Printing Office, Washington, D.C. 20402 - Price $2.00
                         Stock Number 5501-0121

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

-------
                                  ABSTRACT
     Detailed performance, size, and cost estimates were made for advanced
simple-, regenerative-, and compound-cycle gas turbine engines for turbine
inlet temperatures of 2000° F and above as anticipated to be commercially
available in the next two decades.  Conceptual designs for 1000-Mw central
power station utilizing gas turbines and comparisons of complete gas turbine
and steam turbine power station installed costs and total busbar power costs
were made for the various regions of the US.

     It is shown that the gas turbines in the 1970 decade could produce electric
power at lower costs than steam turbines in the South Central region of the
US where natural gas is readily available.  Elsewhere in the US the gas turbines
would be economically competitive if moderately priced clean fuels are available.
Advanced gas turbines will become more competitive in the 1980 decade as anticipated
increases in turbine inlet temperature, component efficiences and larger engine
designs lead to more efficient and lower-cost engines and power stations.

     Although the development costs for large, advanced gas turbines would
approach from 100 to 200 million dollars, the total amount that utilities
are expected to expend for cooling devices to combat thermal pollution over
the next two decades will exceed more than ten times this amount.  Thus
advanced gas turbines should be given serious consideration for increased
research and development support.

     This report was submitted in fulfillment of Contract 14-12-593 under
the sponsorship of the Environmental Protection Agency, Water Quality Office.
                                     iii

-------
                                   CONTENTS
Section	Page
             CONCLUSIONS
 II          RECOMMENDATIONS
III          INTRODUCTION.
 IV          SCOPE OF THE STUDY.
  V          SYNOPSIS OF STUDY RESULTS 	       11

 VI          DESIGN REQUIREMENTS OF FUTURE FOSSIL-FUELED
             THERMALLY NONPOLLUTING POWER STATIONS 	       15

             SUMMARY	       15
             REVIEW OF NATIONAL ELECTRICAL LOAD GROWTH AND FUEL
             USAGE PATTERNS	       16
             ESTIMATES OF REGIONAL FUEL AVAILABILITY AND COST
             PATTERNS	       17
                  Fuel Usage	       18
                       Natural Gas	       19
                       Oil	       21
                       Coal	       23
             REVIEW OF REGIONAL COOLING WATER AVAILABILITY AND
             THERMAL POLLUTION RESTRICTIONS	       26

-------
                              CONTENTS  (CONT.)
Section _ _ _ Page

            ESTIMATES OF PRESENT-DAY AND FUTURE CONVENTIONAL STEAM  ....
            POWER PLANT PERFORMANCE AND COST CHARACTERISTICS                   29

                 Unit Capacities .....................     29
                 Steam Conditions .....................     30
                 Performance .......................     32
                 Station Costs ......................     33

            ESTIMATE OF PRESENT AND FUTURE PERFORMANCE AND COST
            CHARACTERISTICS OF ALTERNATIVE METHODS FOR COOLING
            CONDENSER WATER DISCHARGES ..................     3k

                 Description of Alternative Systems ............     3^
                      Once-Through Cooling ................     3^
                      Cooling Ponds or Reservoirs .............     35
                      Spray Ponds  ....................     36
                      Spray Cooling Canals ................     36
                      Wet Cooling Tovers .................     37
                      Dry Cooling Tovers .................     38
                 Performance Penalty with Alternative Cooling Systems. .  .     39
                 Total Cost Penalties ...................     Ul
                 Other Considerations ...................     1*3
                 Advanced Cooling Systems .................     UU

 VII        TECHNICAL AND ECONOMIC CHARACTERISTICS OF ADVANCED
            GAS TURBINE POWER GENERATING SYSTEMS .............     1*5

            SUMMARY ............................     1*5

            DESCRIPTION OF BASIC THERMODYNAMIC CYCLES ...........     h6

                 Simple Cycle .......................     U6
                 Regenerative Cycle ....................     kj
                 Intercooled Cycle ....................     US
                 Reheat Cycle  ......................     U8
                 Compound Cycle ......................     1*8
            GAS  TURBINE DESIGN  CONSIDERATIONS
                                         VI

-------
                            CONTENTS (CONT.)


Section	         Page

            PROJECTED ADVANCES IN GAS TURBINE COMPONENT TECHNOLOGY  	    50

                 Compressors	    51
                      Performance Parameters 	    51
                      Construction Design Features 	    52
                      Materials	    53
                 Combustors	    5!;
                      Performance Parameters 	    5H
                      Construction Design Features 	    55
                      Materials	    55
                 Turbine	    56
                      Performance Parameters 	    57
                      Materials.	    57
                      Coatings	    58
                      Disks	    59
                      Turbine Cooling Techniques	    59
                 Regenerators	    60
                 Recuperator Materials 	    62

            BASIS FOR SELECTING DESIGN PARAMETERS	    Sh

            PERFORMANCE ESTIMATES	    65

                 Simple-Cycle Engines	    65
                 Regenerative-Cycle Engines	    67
                 Compound-Cycle Designs  	    69

            SELECTION OF GAS TURBINE PARAMETERS FOR MINIMUM-COST POWER  ...    70

                 Simple-Cycle Engine Designs 	    70
                      Engine Size	    70
                      Engine Pressure Ratio and Turbine Inlet Temperature.  .    72
                           C_omp£n_ent_ _Cp_st_Breakdowns_	    73
                      Single- vs  Twin-Spool Designs  	    73
                      Pover Turbine	    7U
                           Exit_Vel£city_	    75
                           Mater_ials_ Chang_es_	    75
                      Coating Life	    76
                 Regenerative-Cycle Engine Designs 	    76
                      Recuperator Surface Characteristics	    77
                                        vii

-------
                               CONTENTS (CONT.)


Section	  Page

                      Effectiveness 	    78
                      Total Pressure Loss	    79
                      Pressure Loss Split	    80
                      Flov Arrangement	    80
                      Compressor Pressure Ratio 	    80
                 Compound-Cycle Designs 	    82

            ADVANCED GAS TURBINE STATION CHARACTERISTICS	    83

  VIII      POWER GENERATION COSTS FOR SYSTEMS DESIGNED TO
            ELIMINATE THERMAL POLLUTION 	    87

            SUMMARY	    8?

            CAPITAL INVESTMENT AND OPERATING COSTS FOR
            ADVANCED POWER GENERATING SYSTEMS 	    88

                 Steam System Costs	    88
                      Station Investment and Total Installation Costs ...    88
                           Rje£iojiaJL_S^eam-Electric_ Station^ Cjosts_	    89
                      Annual Owning and Operating Costs 	    90
                 Gas Turbine System Costs	    90
                      Capital Costs 	    91
                      Annual Owning and Operating Costs 	    92

            COMPARISON OF POWER GENERATION COSTS   	    92

                 South Central 1970-Decade Stations  	    93
                 South Central Early 1980-Decade  Stations  	    93
                      Sensitivity to Economic Factors  	    9^
                 South Central Late 1980-Decade Stations	    95
                 Other Regions	    95
                 Use of Dry Cooling Towers	    96

            POTENTIAL SITING,  TRANSMISSION,  AND RESERVE MARGIN
            ADVANTAGES OF GAS  TURBINES	    97

                 General Transmission and Distribution Considerations ...    97
                      Effect  of Unit Output  Capacity on System  Reliability.    99
                      Effect  of Degree  of Mix and  Forced Outage
                         Rate  on System Reliability	    99
                                         viii

-------
                             CONTENTS  (CONT.)


Section   ___ Page

                      Mixed-System Cost Credits ..............    100
                      Installed Capacity Expansions ............    101
                 Effect of Expansion to Meet Future Demands ........    101
                 Effect of Unit Size and Location on Transmission
                   and Distribution System Costs ..............    102
                 Concluding Remarks ....................    103

            GAS TURBINE FUELS .......................    10U

                 Gas Turbine Fuel Specifications ..............    10U
                 Coal Gasification Technology ...............    106
                      Autothermal Gasifiers ..... ...........    107
                 External Heating Processes ................    108
                 Coal Gasification Costs ..................    109

            DEVELOPMENT TIME AND COST FOR ADVANCED GAS TURBINES ......    109

            ESTIMATE OF ADDITIONAL CAPITAL COSTS FOR
            COOLING TOWERS AND COOLING PONDS  ...............    Ill

  IX        ACKNOWLEDGMENTS ........................    113

  X         REFERENCES ...... . .....................    115

  XI        TABLES I THROUGH XXXI .....................    127

  XII       FIGURES 1 THROUGH 82  .....................    l6l

  XIII      PUBLICATIONS  .........................    2^3

  XIV       APPENDICES
            APPENDIX A - OXIDES OF NITROGEN EMISSIONS FROM GAS
            TURBINE-TYPE POWER SYSTEMS
            APPENDIX B - DESCRIPTION OF GAS TURBINE DESIGN PROGRAM  ....

                 Gas Turbine Design Computer Program ............
                 Basic Assumptions .....................    250
                                       ix

-------
                              CONTENTS  (CONT.)


Section	Page

            APPENDIX C - DESCRIPTION OF GAS TURBINE COST MODEL	   255

                 Economic Analysis 	   255
                 Component Cost Information	   257

            APPENDIX D - G/S TURBINE OFF-DESIGN CHARACTERISTICS	   263

                 Part-Load Characteristics 	   263
                 Effect of Ambient Conditions	   26U

-------
                                 LIST OF FIGURES
FIG. NO.
   1       YEARLY ADDITIONS TO GENERATING CAPACITY AND YEAR-END MARGINS IN
           ELECTRIC UTILITY INDUSTRY

   2       PREDICTED GROWTH OF ELECTRIC UTILITY GENERATION CAPACITY

   3       REGIONAL FORECAST OF ELECTRICAL GENERATION IN THERMAL PLANTS

   1*       LOCATION OF NATURAL GAS RESERVES

   5       COAL FIELDS OF THE UNITED STATES

   6       PROJECTIONS OF REGIONAL FRESH WATER SUPPLIES FOR ONCE-THROUGH
           CONDENSER COOLING

   7       TYPICAL TEMPERATURE VARIATIONS ALONG MONONGAHELA RIVER DUE TO HEAT
           REJECTION FROM VARIOUS SOURCES

   8       DISTRIBUTION OF UNIT SIZE FOR 1968-1971 NUCLEAR AND FOSSIL STEAM
           INSTALLATIONS

   9       ELEVATION VIEW OF TYPICAL STEAM POWER STATION

   10      TYPICAL INSTALLED COSTS OF STEAM POWER PLANTS

   11      SCHEMATIC DIAGRAMS OF ALTERNATIVE CONDENSER COOLING METHODS

   12      TYPES OF WET COOLING TOWERS

   13      GEOGRAPHICAL AREAS OF COOLING WATER SUFFICIENCY

   1^      TYPICAL MECHANICAL-DRAFT DRY COOLING TOWER SYSTEM

   15      ESTIMATED EFFECT OF CONDENSER BACK PRESSURE ON STEAM PLANT PERFORMANCE

   16      DIAGRAMS FOR SELECTED GAS TURBINE CYCLES

   17      SCHEMATIC DRAWINGS OF TYPICAL GAS TURBINE ENGINES

   18      THEORETICAL PERFORMANCE FOR MODIFIED GAS TURBINE CYCLES
                                         xi

-------
                          LIST OF FIGURES (Continued)

FIG. NO.

   19      PROGRESSION OF AIRCRAFT COMPRESSOR TECHNOLOGY

   20      ADVANCES IN COMPRESSOR PERFORMANCE PARAMETERS

   21      COMPRESSOR CONSTRUCTION TECHNIQUES

   22      ESTIMATED TURBINE INLET TEMPERATURE PROGRESSION

   23      ADVANCES IN TURBINE BLADE MATERIALS

   2U      SUMMARY OF PROJECTED CREEP STRENGTH PROPERTIES FOR ADVANCED
           TURBINE BLADE MATERIALS

   25      TURBINE COOLING SCHEMES

   26      ADVANCED BLADE COOLING CONFIGURATIONS FOR AIRCRAFT POWERPLANTS

   27      TURBINE BLADE COOLING BLADE IMPROVEMENTS

   28      ESTIMATED 1970-DECADE SIMPLE-CYCLE BASE-LOAD GAS TURBINE PERFORMANCE

   29      ESTIMATED 1970-DECADE SIMPLE-CYCLE BASE-LOAD GAS TURBINE PERFORMANCE
           WITH SUPPLEMENTARY COOLING

   30      ESTIMATED 1980-DECADE SIMPLE-CYCLE BASE-LOAD GAS TURBINE PERFORMANCE

   31      GAS TURBINE FLOW DIAGRAMS

   32      ESTIMATED 1970-DECADE REGENERATIVE-CYCLE BASE-LOAD GAS  TURBINE
           PERFORMANCE

   33      REGENERATOR GAS  TEMPERATURES FOR  1970-DECADE DESIGNS

   3U      ESTIMATES OF MATERIAL TYPES REQUIRED IN  REGENERATIVE-CYCLE  ENGINE

   35      ESTIMATED 1970-  AND EARLY 1980'S-DECADE  REGENERATIVE-CYCLE  BASE-LOAD
           GAS TURBINE PERFORMANCE

   36      ESTIMATED LATE 1980'S-DECADE REGENERATIVE-CYCLE  GAS TURBINE PERFORMANCE

   37      ESTIMATED 1980-DECADE  COMPOUND-CYCLE BASE-LOAD GAS TURBINE  PERFORMANCE
                                        xii

-------
                           LIST OF FIGURES (Continued)

FIG. NO.

   38      EFFECT OF LOW-PRESSURE COMPRESSOR PRESSURE RATIO ON COMPOUND-
           CYCLE GAS TURBINE PERFORMANCE

   39      EFFECT OF WORK SPLIT ON COMPOUND-CYCLE PERFORMANCE

   1+0      EFFECT OF GAS TURBINE UNIT CAPACITY ON SELLING PRICE

   1+1      EFFECT OF COMPRESSOR PRESSURE RATIO AND TURBINE INLET TEMPERATURE ON
           GAS TURBINE SELLING PRICE

   1+2      COMPONENT COST DISTRIBUTION OF ADVANCED GAS TURBINE ENGINES

   1+3      EFFECT OF COMPRESSOR DESIGN PARAMETERS ON ENGINE SELLING PRICE

   1+1+      EFFECT OF POWER TURBINE DESIGN PARAMETERS ON ENGINE PERFORMANCE AND COST

   1+5      EFFECT OF POWER TURBINE MATERIALS AND DESIGN PARAMETERS ON
           SELLING PRICE

   1+6      EFFECT OF VANE COOLING REQUIREMENTS ON ENGINE PERFORMANCE AND
           COATING LIFE

   1+7      VARIATION OF REGENERATOR SIZE WITH EFFECTIVENESS

   1+8      INFLUENCE OF REGENERATOR EFFECTIVENESS ON POWER COSTS

   1+9      EFFECT OF PRESSURE LOSS PARAMETERS ON REGENERATOR SIZE CHARACTERISTICS

   50      EFFECT OF COMPRESSOR PRESSURE RATIO ON RECUPERATOR SIZE

   51      EFFECT OF TEMPERATURE AND COMPRESSOR PRESSURE RATIO ON RECUPERATOR COST

   52      EFFECT OF DESIGN PARAMETERS ON REGENERATIVE-CYCLE GAS TURBINE ENGINE
           SELLING PRICE

   53      CONCEPTUAL DESIGN OF 200-MW BASE-LOAD GAS TURBINE ENGINE

   5l+      HEAT BALANCE FOR SIMPLE-CYCLE GAS TURBINE

   55      1000-MW GAS TURBINE POWER PLANT, ELEVATION
                                         xiii

-------
                           LIST OF FIGURES (Continued)

FIG. HO.

   56      1000-MW GAS TURBINE POWER PLANT, PLAN VIEW

   57      HEAT BALANCE FOR REGENERATIVE-CYCLE GAS TURBINE

   58      REGENERATIVE-CYCLE ENGINE FLOW PATH

   59      EFFECT OF CAPITAL CHARGES AND GAS COSTS ON STATION POWER COSTS

   60      EFFECT OF GAS TURBINE PERFORMANCE AND COST CHARACTERISTICS ON
           BUSBAR POWER COSTS

   6l      COMPARISON OF BUSBAR POWER GENERATION COSTS IN SELECTED REGIONS

   62      COMPARISON OF BUSBAR POWER GENERATION COSTS IN SELECTED REGIONS

   63      EFFECT OF UNIT SIZE ON LOSS-OF-LOAD

   6k      EFFECT OF SYSTEM UNIT CAPACITY COMPOSITION ON RELIABILITY AND FUEL COST

   65      EFFECT OF EXPANSION AND SYSTEM COMPOSITION ON RELIABILITY AND FUEL COST

   66      EFFECT OF EXPANSION AND SYSTEM UNIT SIZE ON RELIABILITY AND FUEL COST

   6?      POWER TRANSMISSION SYSTEMS

   68      EFFECT OF FORCED OUTAGE RATE ON LOSS-OF-LOAD

   69      ALLOWABLE FUEL COST INCREMENT RESULTING FROM THE REDUCTION IN
           TRANSMISSION REQUIREMENTS

   70      SIMPLIFIED SCHEMATIC DIAGRAMS FOR COAL GASIFICATION  PROCESSES

   71      ESTIMATED DEVELOPMENT COSTS  OF ADVANCED GAS TURBINES

   72      TYPICAL MULTISTAGE COMPRESSOR EFFICIENCY

   73      TYPICAL HIGH-PRESSURE TURBINE PERFORMANCE

   7^      COOLING EFFECTIVENESS  CORRELATION  FOR  ADVANCED  IMPINGEMENT-CONVECTION
           COOLED  BLADES
                                       xlv

-------
                           LIST OF FIGURES (Continued)

FIG. HO.

   75      SCHEMATIC DIAGRAM OF MODEL USED IN GAS TURBINE COST ANALYSIS

   76      MANUFACTURERS' PRICES FOR SOLID COMPRESSOR AND TURBINE BLADES

   77      PRICE ESTIMATES FOR FORGED COMPRESSOR DISKS

   78      ILLUSTRATION OF TYPICAL IMPINGEMENT-COOLED TURBINE BLADE DRAWINGS
           SENT TO BLADE MANUFACTURERS

   79      MANUFACTURERS' PRICES FOR IMPINGEMENT-COOLED TURBINE BLADES

   80      MANUFACTURERS' PRICES FOR IMPINGEMENT-COOLED TURBINE VANES

   8l      PRICE ESTIMATES FOR FORGED TURBINE DISKS

   82      GAS TURBINE OFF-DESIGN CHARACTERISTICS
                                           xv

-------
                                   LIST OF TABLES


 Ho.

 I         UNITED STATES CONSUMPTION OF ENERGY RESOURCES BY ELECTRIC UTILITIES

 II        REGIONAL ELECTRIC GENERATION BY FUEL TYPE AND HYDROELECTRIC POWER

 III        REGIONAL FOSSIL FUEL COSTS FOR ELECTRIC ENERGY GENERATION

 IV        SULPHUR CONTENT AND DISTRIBUTION OF COAL RESERVES

 V         DISTRIBUTION OF COAL WITH SULFUR CONTENT OF ONE PERCENT OR LESS

 VI        SUMMARY OF PROJECTED FUEL COSTS IN SELECTED REGIONS  OF THE US

 VII        SUMMARY OF EXISTING AND EMERGING REGIONAL WATER MANAGEMENT PROBLEMS

 VIII       LIMITING TEMPERATURE CRITERIA IN WATER  QUALITY STANDARDS FOR
           SOUTH CENTRAL POWER REGION

 IX        SCHEDULED ADDITIONS OF  STEAM POWER ELECTRIC GENERATING CAPACITY  BY  YEARS

 X         SCHEDULED ADDITIONS OF  STEAM POWER ELECTRIC GENERATING CAPACITY  BY  REGIONS

 XI         PERFORMANCE OF STEAM POWER STATIONS

 XII        ESTIMATED CAPITAL COST  SUMMARY  FOR COAL-FIRED STEAM  STATIONS

 XIII       INVESTMENT COSTS  FOR ALTERNATE  METHODS  OF COOLING CONDENSER
           WATER DISCHARGES

 XIV        ADDITIONAL COST FACTORS  FOR ALTERNATIVE COOLING SYSTEMS

 XV         GAS TURBINE  COMBUSTOR MATERIALS

 XVI        PROJECTED  TECHNOLOGY FOR BASE-LOAD GAS  TURBINE ENGINES

 XVII       INFLUENCE  OF  COMPRESSOR PRESSURE RATIO  ON  POWER COST

 XVIII      APPROXIMATE  CHARACTERISTICS  OF  UjO-MW COMPOUND-CYCLE GAS TURBINE  DESIGN

XIX        CHARACTERISTICS OF ADVANCED  GAS TURBINE POWER  STATIONS

XX        POWER PLANT CHARACTERISTICS  - EARLY 1980'S  DESIGN TECHNOLOGY

                                          xvl-

-------
                             LIST  C? TABLES  (Contihued)

No.

XXI       1000-MW STEAM-ELECTRIC STATION COSTS - 1980-DECADE DESIGNS

XXII      1000-MW STEAM-ELECTRIC STATION CAPITAL COSTS - 1970-DECADE DESIGNS

XXIII     1000-MW STEAM-ELECTRIC STATION CAPITAL' COSTS - 1980-DECADE DESIGNS

XXIV      BREAKDOWN OF CAPITAL INVESTMENT COSTS

XXV       DETAILED COST BREAKDOWN FOR 1000-MW SIMPLE-CYCLE AND REGENERATIVE-CYCLE
          GAS TURBINE STATIONS

XXVI      1000-MW GAS TURBINE STATION COSTS

XXVII     POWER GENERATION COSTS FOR ADVANCED POWER SYSTEMS LOCATED IN
          SOUTH CENTRAL REGION - 1970-DECADE DESIGNS

XXVIII    POWER GENERATION COSTS FOR ADVANCED POWER SYSTEMS LOCATED IN
          SOUTH CENTRAL REGION - EARLY 198o'S DESIGNS

XXIX      POWER GENERATION COSTS FOR ADVANCED POWER SYSTEMS LOCATED IN
          SOUTH CENTRAL REGION - LATE 1980'S DESIGNS

XXX       PROPOSED ASTM SPECIFICATIONS FOR GAS TURBINE FUELS

XXXI      REPRESENTATIVE COAL GASIFICATION PROCESSES SURVEYED
                                         xvii

-------
                                SECTION I
                               CONCLUSIONS
High-output-capacity open-cycle gas turbines  incorporating the  desj^n  advance-
ments projected to become available during the next  two  decades  could  eliminate
river and lake thermal pollution while producing power at lower  busbar costs
than steam power generation systems in those  regions of  the  country where
natural gas or other suitable gas turbine fuel is  available  at a price level
comparable to competing fossil fuels.

Costs to develop nonthermally polluting gas turbine  power  systems  are  estimated
to be at least an order of magnitude less than the anticipated investment
costs by utilities for supplementary cooling equipment such  as  cooling towers,
ponds, etc., needed for steam plants to reduce thermal pollution over  the
next two decades.

Gas turbines which will become available in the 1970 decade  could begin to
penetrate the swing-load and base-load electric utility  market  in the  South
Central Region of the US where natural gas is available.  In other regions of
the US the penetration of gas turbines for base-load operation could be delayed
until the early 1980's unless low-cost clean fuels become  available sooner than
anticipated.

It is anticipated that domestic natural gas in selected southern regions of
the US, and synthetic pipeline gas or imports of LNG in most remaining regions,
will be available at price levels within the limits  needed to insure competitive
busbar electric power costs from gas turbine stations.

Improvements in high-temperature turbine materials,  turbine cooling techniques,
and aerodynamic design derived from current aircraft engine development
programs could permit progressively higher maximum operating temperatures  and
cycle pressure ratios in base-load gas turbines for electric utility applica-
tions by the 1980's if pursued vigorously.  These advances could result in
system efficiencies approaching and exceeding the levels available with modern
steam power plants and unit output capabilities of 200  Mw and above for open-
cycle fossil-fueled gas turbine plants.

The projected growth rate of the utility industry and the emergence of nuclear
power will heighten the present cooling water shortages, and together  with
federal and local regulations, will require a broader evaluation of electric
power generation methods and cooling devices.  Although  cooling towers and
ponds will be used with increasing frequency as short-term solutions to avoid
thermal pollution, they will add from 1% to 10$ to busbar power costs  and
occupy much-needed space.
                                    1. _

-------
?.  Independence from cooling vater,  smaller  plant  layouts, potential site
    advantages, quicker delivery schedules, and substantially lower installed
    power plant costs relative to steam stations will make gas turbines especially
    attractive for the future mid-range and base-load needs of the electric
    utility industry.

8.  The performance of conventional steam power systems  is projected to remain
    essentially constant over the next  two decades.  Slight improvements in
    component efficiency can be expected, but increases  in cycle operating condi-
    tions to give better performance  cannot be economically justified unless
    very expensive fuels are  utilized.   The  total  installed costs are expected
    to remain relatively constant (in 1970 dollars) since the anticipated economies
    of scale will be offset  somewhat  by the need for cooling towers or ponds
    required to reduce thermal pollution and  the continual pressure for higher
    construction labor rates.

-------
                                   SECTION  II
                                 RECOMMENDATIONS
1.  Programs aimed at transferring the technology  developed for  aircraft gas
    turbine engine applications to base-load electric utility use should be
    promoted and sponsored by the federal  government as  a means  of eliminating
    thermal pollution, improving the utilization of our  vater resources, and
    providing low-cost electric power.

2.  Additional investigations should be undertaken to determine  the potential of
    advanced open-cycle gas turbine power  generation systems which can operate
    independently of a source of cooling water  as  a solution to  the siting of
    power generation stations and transmission  lines while allowing beneficial
    utilization of our dwindling land resources.

3.  Improved processes and techniques leading to the development of adequate
    supplies of low-cost clean fuels, i.e.,  LNG, synthetic gas from coal gasifi-
    cation, and domestic natural gas from  previously untapped sources, should be
    encouraged and supported by the federal  government as well as the utility
    industry as a means of producing fuels resulting in  lower air pollution and
    suitable for use in gas turbines so as to eliminate  thermal  pollution.

U.  The encouraging results of this program  suggest that a study of the application
    of gas turbine technology for nuclear-fueled power generation systems is
    warranted since nuclear power is projected  to  have an ever-expanding role in
    the electric power generation industry.

-------
                                    SECTION III
                                   INTRODUCTION
     The increasing problem of temperature elevation or thermal pollution  of river
and lake vaters used to cool electric utility- pover  generating  plants  is becoming
a major concern to federal and state governments  as  well as  to  electric utilities
and conservation groups.  Some evidence already exists  that  the effects of heated
water from power stations can be harmful to aquatic  life and can adversely change
biochemical reaction rates, thus limiting the capability of  these waters to
assimilate other wastes.  Furthermore, the value  of  water for drinking, recrea-
tional, and industrial use usually decreases with higher water  temperatures.

     If present projections of the electric power industry growth rates are
correct, the power generation capacity and output in the United States by  1990
will have expanded to approximately four times the present-day  level,  and  by
the year 2000, the generating capacity will be approximately ten times the present-
day levels.  The cooling of steam condensers in the  electric generating plants
presently operating in the United States requires over  100,000  million gallons
per day of cooling water.  If unrestricted use of once-through  condenser cooling
continues to be permitted, by the year 2000 as much  as  600,000  million gallons per
day, or the equivalent of one-half of the average daily runoff  of all  rivers in  the  US,
would be needed for cooling power generating systems.   The cooling  water shortage
will accelerate as nuclear-fueled stations provide a larger  portion of electric
power demand in the future, since most modern nuclear plants discharge about 50/5
more waste heat to cooling water than do fossil-fueled  plants of the same  output.

     Technological solutions to waste heat disposal  have not kept pace with  the
increased power production, and concentrated efforts are under  way  by  the  electric
power industry and government agencies to find solutions through a  broad range of
approaches.  For example, numerous studies (reported in government  and trade
publications) have been initiated to determine means of minimizing  the effects of
discharge heat on the aquatic environment, to develop beneficial uses  for  waste
heat, to reduce the waste heat produced from power plants, to utilize  cooling
schemes that produce no harmful effects, and to devise  new and  nonpolluting  methods
of power generation.

     Modeling techniques, experiments, and analytical programs  are  being pursued
to minimize the effects of waste heat on the aquatic environment through increased
turbulence, greater dilution, and faster dispersion  of  the cooling  water.
Unfortunately, these solutions are usually not widely applicable to other  locations,
and thus power plant site selection where adequate cooling water is available can
become a costly procedure.  Substantial performance  improvements of steam-electric

-------
 generating systems  resulting in reduced vaste heat emissions appear unlikely.
 Careful projections indicate that the maximum operating temperatures of steam
 power  plants vill be limited to approximately 1000 F by the excessive costs for
 materials  capable of operating at higher temperatures.  Several potentially
 beneficial uses  for the  enormous quantities of hot water are being explored,
 including  sea  farming and  irrigation.  The widespread economic utilization of waste
 heat for such  purposes is  uncertain at this time.  The most promising near-term
 solutions  appear to be increased use of cooling towers, cooling reservoirs, and
 spray  canals.  Dry or nonevaporative towers consume almost no water and discharge
 heat directly  into  the atmosphere, but these towers, like wet towers, are costly
 to build,  require substantial space, and result in higher fuel costs due to low
 power  plant efficiencies.

     Long-term solutions require new methods of generating power which reject
 their  waste heat directly  to atmospheric air and hence require no cooling water.
 An example of  such  an open-cycle system, i.e., one which utilizes ambient air as
 the working fluid of the thermodynamic cycle, is the gas turbine engine.  Other
 power-generating methods,  such as magnetohydrodynamic generators, thermionic power
 generators, or other unconventional power generation systems, are being investigated
 and could  reduce thermal pollution, but these methods will require enormous
 financial  support and substantial technological advances before they can be reduced
 to commercial  practice.  The gas turbine, however, which requires no cooling water
 and is  already used extensively for peak-power applications as well as in numerous
 other  industrial and military systems, has the potential of eliminating thermal
 pollution  based  upon the numerous development programs in progress to date.

     Presently,  the  utilization of gas turbine engines for stationary electric
 power  generation is  limited to peaking power applications because of their
 relatively low thermal efficiency in comparison to fossil-fueled steam plants.
 However, recent  engineering advances achieved during extensive research and develop-
 ment efforts on  military and commercial aircraft applications have provided the
 basis  for  substantially improved large-capacity base-load gas turbine power systems
 with significantly higher thermal efficiencies than are attainable with present
 systems.   Because of the higher compressor pressure ratios and higher turbine inlet
 temperatures which will be attainable within the next two decades, it is possible
 that base-load gas turbine power plants capable of producing 150 to 350 Mw per
 unit will become commercially feasible in the foreseeable future.

     As a  result of these technological advances in gas turbine design and the
necessary  compromises  in steam-electric power plant design (to adhere to recently
 imposed water temperature standards),  it appeared that future gas  turbine power
systems might be capable of generating base-load electric  power at costs compe-
titive with fossil-fueled steam-electric systems.   Therefore,  the  primary objectives
of this study were:    (l)  to identify the design  requirements for future fossil-
fueled thermally nonpolluting power stations;  (2)  to define and select  advanced

-------
fossil-fueled open-cycle "base-load gas  turbine  systems  that have  the potential
for generating lowest-cost electric pover while eliminating thermal pollution;
and (3) to estimate and compare the costs of producing  electric power with
advanced open-cycle base-load gas turbine stations  and  advanced fossil-fueled
steam stations designed to reduce or eliminate  thermal  pollution  during the  1970
and 1980 decades.

-------
                                   SECTION IV
                                SCOPE OF THE STUDY
     To achieve the objectives of this study,  conceptual design  and  cost programs
developed at UARL vere utilized as a means of determining the  approximate per-
formance, cost, and size characteristics of advanced gas turbine engines and to
incorporate the design advancements, materials,  and other features which are
already in use in aircraft engines or projected to become available  in the next
two decades in base-load gas turbines.  To cover the wide range  of operating
conditions and design parameters such as turbine inlet  temperature and compressor
pressure ratio, the cost and design analyses were heavily dependent  on a number
of simplifying assumptions.  Thus the results  are not intended to reflect compre-
hensive design aspects of advanced gas turbines  which would require  much more
extensive and costly efforts but rather to show general features and levels of
performance and cost that might be attained for utility applications.

     To provide a realistic appraisal of the potential  of gas  turbines as a
means of eliminating river and lake thermal pollution,  extensive review of the
available literature and discussions with electric power industry representatives
were held to estimate (l) the availability and range of prices for suitable gas
turbine and steam turbine system fuels, (2) the extent  and severity  of cooling
water shortages and the implications of thermal pollution restrictions, (3) the
operating limitations, performance, and cost characteristics of  present-day and
projected future steam power stations, and (k) the operating and cost charac-
teristics of cooling towers and reservoirs suitable for use with steam power plants.

-------
                                     SECTION V
                            SYNOPSIS OF STUDY RESULTS
     The national demand for electric power will double every ten years  during
the next two decades, and nuclear fuel will become  a. significant source  of  energy
after 1980.  The demand for power from thermal sources  will  increase  at  the fastest
rate in the West, West Central, South Central, and  Southeast Regions  of  the US.
Except for the Southeast Region, these same areas will  generally lack sufficient
natural sources of cooling water to utilize once-through cooling systems in nuclear-
and fossil-fueled steam power plants, and the majority  of new stations will employ
some alternative cooling system such as cooling ponds or wet cooling  towers.  State
and federal restrictions on thermal discharges will further  stimulate widespread
utilization of cooling ponds and towers except for  isolated  ocean power  plant
installations.

     Long-term supplies of cheap fossil fuels capable of complying  with  present
and anticipated pollution-control laws are not adequate to meet the demands of
the utility industry.  Substantially higher prices  will be needed to  stimulate
the development of low-sulfur coal and natural gas  supplies. However, natural
gas should be available at price levels of 26<£ to ^Oi£/million Btu near the  sources
of supply (South Central and Pacific Regions) during the next two  decades and
at price levels of about Uotf to 60<£/million Btu from coal gasification or in the
form of LHG imports in other coastal and midwestern areas.

     The price of coal and residual oil for utility application will  increase to
the 30
-------
 use of dry rather than wet  cooling towers would alleviate some siting difficulties
 but at a substantial 9 to 10$  increase in the busbar power costs.  However, the
 widespread use of dry towers is not anticipated in the next decade and perhaps
 longer.

      It  is estimated that gas  turbine engines capable of operating at turbine
 inlet temperatures  as high  as  2200 to 2hOO F could be in operation in utility
 power generation systems before the end of the 1970's if turbine materials and
 blade cooling  techniques presently under investigation for aircraft and other
 applications were utilized.  By the early 1980's turbine inlet temperatures 200 to
 UOO F higher can be anticipated, and these advances together with similar improve-
 ments in compressor and combustor technology will form the basis for gas turbine
 engines  capable of  providing 200 to 250 Mw of electric power in a single unit
 while achieving overall plant  thermal efficiency levels of 36 to 3Q% in simple-
 and regenerative—cycle configurations.  The higher turbine inlet temperatures will
 also permit substantial improvements in engine specific power levels which are
 projected to result  in 20 to 30% reductions in future engine and power station
 selling  prices  relative to  present-day prices for gas turbine systems.  The total
 site area requirements for  gas turbine stations would be on the order of 10% of
 those for conventional steam power stations.  Together with elimination of the
 need for cooling water, the reduction in area requirements could tremendously
 simplify utility planning.

      Precooling the  compressor bleed air to approximately 200 F in external heat
 exchangers prior to  its use in the turbine section enhances the performance and
 cost  characteristics  of gas turbine systems; hence, precooled air will be used
with  increasing frequency in the next two decades.

      The compound-cycle gas turbine engine offers attractive levels  of performance
and  cost  for central power stations, and further study to confirm this preliminary
result is recommended.

      Advanced open-cycle gas turbines utilizing technology derived from aircraft
engine programs offer a means of eliminating thermal pollution while generating
electric power at busbar costs substantially below those which will  be attainable
with  future conventional steam systems in the natural-gas-rich South Central region
of the US.  The estimated busbar costs of the simple-cycle gas turbine station
vary  from approximately 0.5 mills/kwhr to 1.0 mill/kwhr below those  projected
for steam stations in the South Central Region during the 1970 and 1980  decades,
respectively.   The regenerative-cycle gas turbine system would generate  power at
costs lower than those for the steam system but  at a somewhat higher level than
the simple-cycle gas turbine system.   The conclusions are relatively insensitive
to the capital and interest  charges,  as  well as  to the fuel cost  used in the
comparisons.
                                        12

-------
     Simple-cycle gas turbine designs which are projected to "be commercially
available by the early 1980's could produce power at busbar costs competitive
with residual-oil- or coal-burning steam stations in the remaining regional
locations for load factors up to approximately 70$, even when burning a fuel
costing as much as 20<£/million Btu more than for the steam system.

     The reduced transmission-distribution network requirements associated with
relatively small gas turbine power generating units, which may be located close
to the load centers due to their independence from cooling water supplies, can
result in an appreciable savings as compared to networks required with large steam
stations.  Increased reliability can be achieved for a given power system through
a reduction in power generating unit size and a diversification of unit size.
The combined effects can result in equal electric power costs with dispersed gas
turbines, in comparison with power costs from large steam stations, notwithstanding
a -cost increment of up to several ^/million Btu for the gas turbine fuel.

     Coal gasification technology is becoming available as the result of various
incentives, so that both pipeline-quality high-Btu/ft3 gas and low-Btu/ft3 producer-
type gas are anticipated to become available in the next decade at 20 to Uo<£/million
Btu above the price of the coal or residual oil used as feedstock.

     Utilities will spend approximately $2 to $H billion in each of the next two
decades for cooling towers, ponds, and other devices in an attempt to reduce or
eliminate thermal pollution of the nation's rivers and lakes.  Advanced open-cycle
gas turbines capable of generating low-cost electric power could be developed for
perhaps one-tenth of that earmarked for low-pollution cooling systems for steam
power plants.
                                        13

-------
                                   SECTION VI
              DESIGN REQUIREMENTS OF FUTURE FOSSIL-FUELED THERMALLY
                          NONPOLLUTING POWER STATIONS
                                     SUMMARY
     An investigation vas undertaken to determine the  design  requirements  for
future fossil-fueled thermally nonpolluting steam power  stations  to  provide  a
realistic reference for subsequent comparative evaluation of  advanced-design gas
turbines.  A review of available literature was made to  determine those  geographical
areas of the country which are experiencing or are expected to  experience  cooling
water shortages and/or thermal pollution restrictions.   Estimates are  presented of
the national and regional electric power load growth rates, and the  availability
and range of prices for suitable fuels which meet pollution regulations.   Operating
limitations, performance, and cost characteristics of  present-day and  projected
future steam power stations were established from a survey of available  literature
and from discussions with representatives of the public  utility industry.  Evalua-
tions of the future need for condenser heat discharge  methods other  than once-
through cooling such as cooling ponds, wet and dry cooling towers were made, and
estimates are presented of the present and potential future operating  and  cost
characteristics of towers and cooling reservoirs suitable for use with steam power
plants.

     The estimates were made for both the 1970 and 1980  decades;  reliable  predic-
tions further in the future often are not based upon realistic  assumptions.  To
conform with projections of advanced-design gas turbines, steam power  plant  techno-
logy levels have been defined for the 1970 decade, the early  1980's, and the late
1980's.

     For the purposes of this study, comparisons among the competing power systems
were made on a regional basis rather than on a national, statewide,  or even
utility level.   Comparisons on a statewide or utility  level would provide  additional
insight as to the potential for open-cycle gas turbines  as a  means of  eliminating
thermal pollution, but at a substantial increase in the  level of  effort.   However,
sufficient similarity exists within a regional area relative  to the  dominant type
of utility fuel and its availability and price, load profiles,  supplies  of cooling
water, and other factors considered by utilities in selecting a power  system so
that a realistic competitive analysis can be of benefit  on this level.  Therefore,
estimates are provided for the average utility plant size, fuel cost,  types  of
condenser cooling system, and steam plant characteristics in  each of the six FPC-
(Federal Power Commission) designated power regions of the US.
                                         15-

-------
                   REVIEW OF NATIONAL ELECTRICAL  LOAD GROWTH AND
                                FUEL USAGE  PATTERNS
      Over the past twenty-five years the  installed  capacity of the electric
 utility industry in the United States has doubled every decade to the present
 level of approximately 3^3 million kw.  Numerous surveys (Refs. 1 through h) have
 indicated that the growth  rate for this industry will accelerate slightly at
 least through the 1970 decade  and should  continue to produce a doubling of the
 installed capacity every 10 years through the  last  decade of this century.  Thus
 by the end of 1975, the installed capacity of  the utility industry is expected to
 reach 530 million kw (Ref.  2)-,  as of early 1970, some 200 million kw of new
 generating capacity were already  on order and  scheduled for operation.  Interestingly,
 the new capacity to be added exceeds the  electric utility capacity in operation as
 recently as the beginning  of 1963.

      Projections of the yearly generating additions, based on data compiled in
 1969 and. 1970 by the National  Electrical  Manufacturers' Association (NEMA), are
 shown in Fig. la.   If the  capacity in service  by 1978 does reach 625 million kw,
 as forecast, the average yearly growth rate will have been 8.0$ over the decade
 from 1968 to 1978.   The actual yearly additions are not constant but exhibit the
 historical trend of substantial year-to-year variations.  One of the reasons for
 the cyclical behavior in orders for generating additions is the desire by the
 utilities to have additional protection in the event of possible delays (which
 have indeed occurred)  in some  of  the very large advanced-design units which will
 be coming into operation in that  period.   An approximate indication of the reserve
 margins available is shown  by  the ratio of the year-end capacity to the summer
 peak load as shown in Pig.  Ib.  The ratio was  extremely high in the early 1960's
 but is expected to  remain at the  1.20 to  1.21*  level during the 1970's.

      Other forecasts of the growth of the electric utility industry in the US have
 been made,  and the  pertinent results of the most recent surveys are summarized in
 Fig.  2 along with the  NEMA  data through 1978.  The NEMA and EEI (Edison Electric
 Institute)  data shown  in Fig.  2 include the hydroelectric capacity in the US as
 well as the thermal capacity (both fossil-  and nuclear-fueled).  A similar break-
 down according to type of generation is available from the AEG (Atomic Energy
 Commission)  data contained  in  a 1967 report to the President (Ref.  5)  and Ref.  6.
 The Fig.  2  data illustrate  (l)  the proportion  of the total industry capacity in
 hydroelectric and thermal plants through 1990  and 2000,  respectively,  (2) the
 expanding portion of the thermal  capacity which will be provided by nuclear systems
 in  the next  30  years,  and (3) the large discrepancy which already exists  between
 the nuclear  forecasts  presented in the 1967 AEC supplement and the 1970 FPC pre-
 liminary  data for the  forthcoming national power survey.   Based on data in Ref.  U,
which  indicate  that  32.5/5 of the 200 million kw of new additions already on order
will be nuclear units, the 1970 FPC data appear to provide a more accurate picture
                                         16

-------
of the extent of the nuclear penetration into the utility industry.   Furthermore,
since 60.U$ of the additions are in other thermal units and only 1.1% in hydro-
electric units, which in the past have provided 16 to 20$ of the capacity,  the
diminishing role of hydroelectric plants due to the reduction in the number of
desirable sites is also indicated.  Reference U also provides an indication of the
types of generation equipment which are forecast for the 1969-1978 period (see
Fig. l).  The data confirm the trends in certain types of capacity forecast,  at
least for the next decade.  In Ref. 6, it is forecast that conventional and
pumped-storage hydroplants will account for only 12.% of the 575 million kw  of
generating capacity in 1980, while fossil-fueled and nuclear-fueled steam plants
will comprise 62 and 21%, respectively, and gas turbines and diesel plants  the
remaining 5%.  Recent orders for gas turbines, however, have been running con-
siderably ahead of this prediction and many sources now indicate that gas turbines
will provide from 15 to 25$ of utility installed capacity by the 1980-decade.

     Most of the surveys also agree that the electricity generated by those power
plants which will be in operation in 1990 will be more than four times the  1970
level.  Thus, the number of kilowatt hours generated in thermal power plants  is
expected to increase from 1300 billion in 1970 to over 5500 billion in 1990.
Furthermore, it is stated in Ref. 7 that, "Since the nuclear plants  that will be
in operation during the next two decades will be base-loaded and will operate at
75 to 80$ capacity for most of their life, the generation of power by nuclear plants
will grow from a predicted level of 68 billion kwhr in 1970 to 1290  billion kwhr
in 1980, and to a level of nearly UOOO billion kwhr by 1990."   Thus,  by the 1980's,
various references predict that nuclear power generation will  account for from 30$
and 70%,  respectively,  of the total power generated in thermal plants.   Data
from Refs.  8 through 10 tend to confirm these estimates,  while a Bureau of  Mines
projection (Ref.  11) indicates  that nuclear  power will provide only  20$ of  the
total utility energy requirements in 1980 and only  60% by the  year 2000.  The
consumption of natural gas in the utility industry  is  projected by all five surveys
to increase in spite of the diminishing reserves.   The role of oil and coal
as utility fossil fuels appears to vary among the various surveys.   A summary of
the role predicted for various  energy sources in the utility industry from  selec-
ted studies is presented in Table I.   Additional data from other surveys and  a
discussion of the methodologies used in various studies is presented in Ref.  12.
                           ESTIMATES OP REGIONAL FUEL
                         AVAILABILITY AND COST  PATTERNS
     Although the national picture with respect  to  installed  generating  capacity,
electrical generation, and raw energy sources  is of overall interest,  significant
changes will be occurring on a regional basis  as well.   The increase  in  electrical
generation from thermal plants for each of the six  regions in the  National  Power
                                        17

-------
 Survey is shovn in Fig.  3.   Over the twenty-year period from 1970 to 1990, the
 West Region,  which contains  one-third  of the  contiguous United States, is predicted
 to experience an annual  increase in thermal generation capacity of almost 10/S,
 while the South Central, Southeast, and West  Central Regions will experience
 annual growth rates of 1.2%  or higher.  Electrical generation growth from thermal
 plants will be the slowest in  the more populous Northeast and East Central Regions.
                                    Fuel Usage

      Due to the  emergence  of nuclear energy, the growing concern and associated
 legislation for  preserving the  environment, and the presently predicted shortage
 of fossil fuels,  the  utilization  of fuels for electric power generation within
 each  region is expected to undergo dramatic changes during the remainder of the
 twentieth century.  Traditionally, natural gas has been the dominant fuel source
 for power generation  in the South Central Region, supplying over 95% of the energy
 requirements.  Natural  gas has  also been the main fossil fuel in the West Region,
 supplying almost  75$  of the energy for  thermal generation, but since hydrogenera-
 tion  has supplied about 50^ of  the total power generated, gas accounts for only
 about 31% of the  total  electrical generation in this region.  In the remaining
 regions  of the US,  coal has been  the principal fuel, supplying as much as 95%
 of the raw energy in  the East Central Region during 1966.  During this same year,
 coal  was used to  provide about  60%, J2%, and Jk% of the raw energy for electric
 utility  power generation in the Northeast, West Central, and Southeast Regions,
 respectively.  Oil  is used predominantly only along the east and west coasts and
 provides no more  than 20%  of any  regional energy resource.

      However, during the next twenty years, nuclear fuel is expected to carve out
 a  substantial portion of the energy market in almost every region of the country.
 The FPC  estimates of the nuclear penetration in each region, summarized in Table II,
 provide  an indication of the shift likely to be experienced in the energy market
 during the next two decades.  For example, in the East Central Region, the FPC
 estimates  (Ref. l) that  nuclear fuel will share the raw energy market with coal by
 1990.  The estimates of one of the largest architect-engineering firms (Ebasco
 Services  Incorporated)  (Ref. 10) appear to provide essentially the same conclusions.
 Nuclear  fuel will dominate in the Northeast, Southeast, West, and West Central
 Regions  as well.   Only  in the natural gas-rich South Central Region will a single
 fossil fuel, natural gas, provide the bulk of the regional energy requirements.
 Of  course, the ultimate  utilization of each fuel source will depend upon the
 availability, deliverability, and the final relative prices of the fuels and
 associated power systems in each region.  Thus a brief review of the extent and
 location of the different fossil fuel resources and the costs which may be incurred
to bring them to power generation sites is appropriate.
                                        18

-------
Natural Gas

     Despite the fact that the use of natural gas to generate electric power has
been consistently criticized as an inferior use of the best fuel available to
mankind, the consumption of gas by electric utilities has  risen steadily.   At
present, gas accounts for about 26% of the total fossil fuel used in steam-electric
power generation, and this gas comprises about 16% of the  total gas used in the
country.  In fact, due to the recent growing concern over  sulfur dioxide emissions
and other types of air pollutants, the consumption of natural gas for utility
operations and other manufacturing processes has accelerated even faster than
anticipated.  For example, the consumption of natural gas  in 1968 in the Southeast
Region was 60% higher than the 1966 level and had already  exceeded the amount
forecast in late 1965 ani^ early 1966 by the FPC for use in 1980.   The burgeoning
utilization of natural gas has served to highlight the decreasing reserves of
natural gas and has sparked appeals for additional exploration and production of
natural gas (Refs. 13 and I**).  Between 195^ and 1968, gas reserves were growing
at a rate of 2.1% per year, against a consumption rise of  5-3% per year.  As a
result, the reserve fell from 29 years of gas supplies in  195^ to 14.6 years in
1968.  Today, the recoverable proven gas reserves are down to about 11 years of
gas supplies.  Reserves are decreasing because their development  has been dis-
couraged by low well-head prices set by the FPC for gas that will be used interstate
and not because the US or the world is running out of gas.  On the contrary, if
the estimated potential gas supplies as of 1968 of approximately 1227 trillion
cubic feet (Ref. 15) were added to the proven gas reserves of 287 trillion cubic
feet, there would be over 62 years of gas at the current annual level of consumption.
These estimates may even be pessimistic, since the US Geological Survey (Ref. l6)
estimates total proven and unproven gas reserves at 1700 trillion cu ft, while in
Ref. 17 the gas reserves are placed at 2300 trillion cu ft.  However, the develop-
ment of these reserves may involve increased costs.   Although the annual consump-
tion of natural gas is expected to double in the next twenty years (Ref. 18), new
techniques are being investigated to increase natural gas  supplies.  For example,.
the current work under the AEC Plowshare program could also result in substantial
additions to the US gas reserves.  The first Plowshare nuclear shot for gas stimu-
lation was Gas Buggy in New Mexico in December 1967, while the second was Rulison
in Colorado.  Gas Buggy resulted in the production of 280  million cu ft of gas in
17 months or about three and one-half times the output of  the nearest conventional
gas well in a 10-year period.  These results have prompted the AEC to promote a
3-year, $75 million program to solve the gas shortage.  Potential output from
stimulated fields is placed at a trillion cu ft within 10  years from now and
ultimately 317 trillion cu ft.  The quantity and distribution of proven and poten-
tial reserves of natural gas in various parts of the US, shown in Fig. k, highlight
the vast reserves that would be available for use in the South Central Region
from those areas denoted as D, E, F, G, and J in the Potential Gas Committee Survey.
Furthermore, there are proven reserves of 52 trillion cubic feet in Canada, and some
west coast utilities are even exploring the possibility of bringing in gas from
                                        19

-------
South America   (Ref.  19)-  Liquified natural gas will be available for use on
the  east  coast,  and studies are under vay to determine the costs of bringing
Alaskan gas  via  pipeline to the midwest or as LNG to the west coast.

     In the  early  days of natural gas utilization after World War II, gas was sold
to pipeline  customers at prices as low as 3 to 5# per Mcf.  (Gas prices are often
quoted in cents  per thousand cubic feet (Mcf) and gas from the well will generally
average 1075 Btu/Mcf.  However, after processing, the heat content of the gas
available for sale is usually reduced to about 1000 Btu/Mcf, and this value has
been used throughout this report to avoid confusion.  Consequently, prices in <£/Mcf
are  equal to prices in ^/million Btu, another pricing quantity often encountered.)
Recently,  there  have been examples where gas has been bought in the field in the
South Central Region  for more than 20$ per Mcf, and one pipeline reportedly paid
28#  per Mcf  or 12£ per Mcf above the in-line price set by FPC (Ref. 20).   Although
it is generally  conceded that an increase in the well-head price of gas would
end  the reserves decline (Ref. 21), the price increase needed to stimulate the
development  of additional natural gas production is, at best, unclear at this time;
estimates  range  from 2$ per Mcf up to 10<# per Mcf and above (Refs. 21, 22, and 23).
An FPC staff study (Ref. 2k) indicates that the prices of natural gas which would
stimulate  the production of supplies adequate to meet demand would range from
22 to 250/million  Btu in the major gas-producing areas of Texas, Louisiana,
and  the Rocky Mountains.  .A large utility serving the South Central Region presently
pays 20<£/million Btu for gas in Louisiana and 20§^ in Texas and has contracts through
1977 which will  provide for a slow escalation for Texas gas to 23tf in 1980 through
198U.  The same  utility estimates that Louisiana gas, if purchased today, might
cost 280/million Btu.   However, they indicate that the break-even point for fossil
fuel as compared with nuclear is in the neighborhood of 37^ to l*0<£/million Btu.
Thus, on the basis of these estimates, the price of natural gas in the South
Central Region can be expected to increase by about 5 to lOtf/million Btu over the
next twenty  years, and the price projections made by the FPC of about 30$/million
Btu with  extremes  of 21 to 38^/million Btu by 1990 appear reasonable (Ref. l).
If the average figure is accepted, prices in various parts of the US can be esti-
mated using  a figure of about 1.1^/million Btu as the cost of transporting the gas
each 100 mi  via  the established network of pipelines.  As a result, the average
city-gate  price  of gas in the Chicago area would be about 1*0<2/million Btu as
compared with estimated prices of about h3$ for imported Canadian gas (Ref. 21)
and  about  50<£ in the New York market.  This figure appears to fall within the
lower range  of prices (52 to 58^/million Btu) for LUG imported from Africa reported
in Ref.  25.  Thus, most sections of the US would be accessible to some supplies
of natural gas although at prices above today's unrealistically low level.  Average
gas  costs  reported in 1965 by US utilities on a regional basis are presented in
Table III.  The  California electric utilities,  which account for about Qd% of
the  gas used for electric power generation in the eleven western states,  are
estimating a price increase at an average annual rate of 0.5% compounded through
1990 to about 36$/million Btu or about 1
-------
Thus, even though there may "be other less costly fuels available,  utilities  in
certain areas where stringent air pollution regulations are in forcu,  such as
southern California, vill be compelled to use higher-priced low-sulfur,  low-ash
fuels such as natural gas.

Oil

     Although petroleum products have never been a dominant energy source  in the
generation of electricity in this country, supplying only 6% of the energy used,
its role in some geographical regions may be changing.  Typically, residual  oil,
that fraction of the crude barrel which remains  after the light products are
distilled, is the petroleum product used in power generation plants.   US refineries
attempt to minimize the production of residual oil to meet their market  demands
for gasoline, jet fuels, and other high-priced products, and thus  only about 1%
of the crude oil processed in the US ends up as  residual oil.  In  South  America
and Europe, however, residual oil comprises about hl% and 30%, respectively, of
the crude barrel because of the different petroleum product market in  these  areas.
As a result, over 85% of the residual oil burned by utilities in the US  originates
from Caribbean crude oils (Ref. 25), and 63% of  the total residual oil is  burned
between Maine and Florida.  Since the oil is delivered via tanker, the transporta-
tion costs are an important factor.  Most of the residual oil burned elsewhere is
of US origin.  The sulfur content of Caribbean residual oil is typically 2.5%.
Mid-continent residual oils have sulfur contents between 0.5 and 1%, while West
Texas and California residuals will usually contain more than 1% sulfur  (Ref.  25).
Presently there is some excess supply of oil in  the Middle East and South  America,
and in spite of the high tanker freight rates, this oversupply has no  doubt  led
to the long-term contracts for high-sulfur oil at $1.60 per barrel (bbl)*  or 2U<£/
million Btu delivered to utilities on the US east coast (Ref. 7)-   Venezuelan
residual oil containing 2.% sulfur is being offered to midwest utilities  at $2.15
per bbl or 32^/million Btu.  However, such fuel  oil will not meet  most air
pollution limits which at present require not more than 1% sulfur  and  ultimately
will require as low as 0.3% sulfur.  The cost of processing typical Caribbean
residuals to reduce the sulfur content to 1% and 0.3% has been estimated at  about
$0.30 and $1.00 per bbl, respectively (Ref. 25).  Substantial desulfurizing  capacity
has already been added or is under construction  in the Caribbean to produce  1%-sulfur
residual oil.  However, in the northeast, there  is a reported shortage of  low-
price residual oil, and prices of 50 to 55^/million Btu have been  reported for
low-sulfur residual oil in New England.

     Since the Caribbean residuals contain roughly 900 ppm of metals  (of which 85%
is vanadium) and thus require an abnormally high catalyst replacement  rate,  the
costs for treatment are somewhat higher than if  typical Middle East residual
fuels were used as feedstocks.  The US reserves  prior to the Alaskan North Slope
discovery were approximately Uo billion barrels, and the ratio of  reserves to
* bbl will be used as the abbreviation for barrel in this  report,  based on
  gallon capacity.
                                       21  _

-------
 annual production  rate was less than ten years.  However, recent estimates indi-
 cate that  the Alaskan North Slope reserves may alone reach Uo billion barrels,
 and this  discovery has dulled the oil industry's interest in tar sands, shale
 oil, and  coal as supplementary sources of oil supplies.  Even more important,
 however,  on  a worldwide basis there are proven reserves of almost 1*00 billion
 barrels which would be sufficient to meet more than thirty years' consumption.
 Furthermore, the recent discovery of large oil fields in the North Sea is expected
 to  offset  some of  the traditional imports from the Middle East.  This discovery
 will produce additional surpluses of oil, but as the petroleum product consumption
 patterns  in Europe and the rest of the world approach those of the US, the amount
 of  residual oil will decline.

     Imports of residual oil from Europe, the Middle East, Africa, and other
 sources,  for utility generation will depend to a large extent on the tanker
 freight rates.  Due in part to the Suez Canal closing and other factors, a shor-
 tage of tanker capacity has driven the present freight rates up to $3.05 per bbl
 for movements from the Persian Gulf to the US east coast.  Such a freight rate is
 almost  an  order of magnitude higher than that existing under normal conditions
 (Ref.  26).  As tanker capacity is added, though, the rates should return to near-
 normal  rates, and  imports of residual oil should increase.  The interrelationship
between fuel availability and fuel price can be clearly seen with respect to the
present residual oil shortage in the east.  The shortage of low-priced residual
oil produced by the removal of import quotas and the deeper distillation processes
together with the  tanker shortage have driven the price of residual oil to levels
twice that of last year.   These prices are now attractive to the oil companies,
and five major producers have indicated that they would make available about
U00,000 barrels more per day.  In Ref. 27. it is estimated that the price of
residual fuel oil  in the world market, including delivered cost in the US (based
on  1968 dollars),  is expected to trend downward moderately for residual fuel oil
with no sulfur guarantee.   The future cost of low-sulfur residual (containing less
than 0.5$) is more uncertain, but assuming a current premium price on the order
of  60<£ per barrel, its price is also expected to trend downward, based on 1968
dollars.  Reference 7 estimates that the consumption of oil by US electric
utilities will continue to grow from the current level of 250 million bbl/year
to  6kb million bbl/year by 1990.   Due to the projected availability ,of low-sulfur
crude from Alaska  on the west coast, midwest, and possibly in the future on the
east coast, adequate supplies of relatively low-cost residual oil are also predic-
ted in selected locations.

     Recent residual oil fuel costs  for electric utility generation in selected
areas of the country are shown in Table III from Ref.  1 and provide a basis for
future projections.  A utility in the South Central Region studying the use of
residual oil estimates a late-1970-decade price of 38<# to ko^/million Btu.
                                        22

-------
Coal

     According to Refs. IT and 28, about 83% of the knovn economically  recoverable
energy reserves in this country are in the form of coal.   These  sources estimate
that about 220 billion tons of coal are recoverable, sufficient  to  meet the  coal
needs of the country for more than kOO years at the present  rate of consumption.
Hovever, many states and localities concerned with the  potential harmful effects
of certain air pollutants have passed sulfur oxide regulatory  laws  which currently
restrict the sulfur content of coal and oil to be burned  in  selected industries to
less than 1.0% and, in the future, to as low as 0.3%.   Since the electric utility
industry consumes about 60% of the total coal used in the country each  year, the
sulfur content of the coal reserves, the production and transportation  costs for
coal, and the economic feasibility of removing sulfur either prior  to combustion
or from the combustion products are of importance in assessing the  future
availability and prices for this energy resource.

     Unfortunately, more than one-third of the total coal reserves  in the US are
high in sulfur content (> 1$).  Much of the low-sulfur  coal  is lignite  or sub-
bituminous coal, with a heat content lower than that of bituminous  coal which now
represents over 95% of the present production.  A summary of the sulfur content
of US coal reserves according to tonnage and heat content is presented  in Table IV.
Data on reserves, however, can be misleading because much of the readily available
coal, especially that located near the major eastern markets,  is of high sulfur
content.  Figure 5 shows the major coal fields in the US, and  Table V summarizes
the distribution of low-sulfur coal, by type and by state.   Data in Ref. 25  indi-
cate that virtually all the low-sulfur coal west of the Mississippi is  located in
the Rocky Mountain states.  Thus its use would require  mine-mouth generation* or
long-distance rail movements to generate the power near large  load  centers.  East
of the Mississippi, the largest reserve of 1%-or-less-sulfur bituminous coal is
in West Virginia, but about one-fifth of this coal is contained  in  narrow seams
and/or excessively deep mines which would substantially increase the cost of its
recovery.  In addition, a large fraction of the coal has  chemical characteristics
which make its use for steam generation unattractive, without  extensive modifica-
tion, due to different slagging characteristics.  Furthermore, the  bulk of the low-
sulfur coal reserves in the Appalachian states is of metallurgical-grade coking
quality and thus commands premium prices from such users  as  steel companies.  For
example, it is reported that the Japanese are paying $12/ton or  about 50<£/million
Btu for southern Appalachian coking coal.  A number of  long-term contracts
to provide this high-quality coal for export have been  signed  recently, further
*  The availability of low-sulfur, low-cost coal for mine-mouth  steam power
   generation stations in many locations in the arid western  states  where  cooling
   water shortages exist has been responsible,  in part,  for the  increased  interest
   in the use of large, dry cooling towers by utilities  and federal  officials.
                                        23

-------
 reducing the available  supplies.  Although  such quality coal has desirable proper-
 ties,  it is suggested in Ref.. 25 that  a demand for similar-quality power plant
 coal vould result  in  coal prices about  $2 to $3/ton higher than high-sulfur
 conventional utility  bituminous coal.   Since -the average value of coal at the
 mine is  about $^.65/ton, the  use of low-sulfur coal would increase the fuel cost
 about  *K) to 65%  or would add  about 8 to 12^/million Btu to the present price of
 high-sulfur fuel.  These figures correspond to recent fuel cost prices, presented
 in  Ref.  7,  which indicate that low-sulfur coal, when available, costs about kQ
-------
comprise as much as 50$ of the delivered cost.   Rail transportation costs  rose
sharply after World War II, and only the development and use of "unit  trains"  has
produced important transportation cost savings.   Further development of this
concept into "high-speed shuttle trains" and the use of coal slurries  via  pipelines
may result in some transportation efficiencies.   Rates for these shipments have
been increasing lately, and the shortage of rolling stock is expected  to accentuate
the trend.  In the western states, rail costs to transport coal are about  3.6<£/
million Btu/100 mi.  In Ref. 1 it is stated that the transportation costs  and  the
overall cost of coal are expected to rise in the Southeast Region,  hut in  Ref. 27
the transportation component of coal costs is expected to decrease  slightly.

     Sulfur Removal from Coal
     Several mechanical and chemical processes are available or have been  suggested
for the partial removal of some forms of sulfur from coal prior to  its  use in
combustion devices.  These methods of cleaning coal will, at best,  result  in only
partial removal of the pyritic sulfur which is only a fraction of the total sulfur
in coal, and therefore would not be applicable unless it would provide  coal of
acceptable quality through a simple means.

     Gasification and liquefaction of coal  to produce high-quality  low-sulfur  fuels
suitable for use in utilities are being studied intensively  for a number of reasons
and ultimate market uses.  Although a number of studies have been made  and pilot
plants are under consideration, it appears  that gasification processes  will
produce a fuel whose costs are about 20 to  35^/million Btu higher than  that of the
basic feedstock.  Its ultimate use as an electric utility fuel will depend upon
the economics and availability of alternative fuels, including nuclear  power,  at
the particular locations.  Descriptions and preliminary cost projections for
processes suitable for the production of high-quality low- and high-Btu gas are
described in Section VIII of this report.

     Opinions differ widely as to the technical and economic feasibility of removing
sulfur oxides from the flue gases of oil- and coal-burning power plants.   A large
number of processes have been advanced, and some are undergoing tests in power
stations or experimental facilities.  Preliminary results, however, are encouraging
for a number of processes as it is estimated in Ref. 25 that, "The  first generation
of sulfur dioxide removal plants will operate with additional costs of  only $0.75
to $1.00 per ton of coal fired."  Furthermore, Ref. 25 states that, "As more becomes
known about the technology of the various processes, second- and third-generation
systems will incur added costs in the range of 20 to 25<£ per ton of coal fired."
Contrary opinions concerning the economic and technical feasibility of  stack gas
processes are presented in Ref. 31 and elsewhere.  It appears that  it will be
several more years before the final results on stack gas processes  are  available,
but it is clear that these processes probably represent the  pivotal factor in
determining the future widespread utilization of coal in the utility industry.
                                        25

-------
 If sulfur oxide  stack gas  cleanup systems  or fuel pretreatment schemes become
 available at  the cost levels predicted, then the long-term utilization of coal
 would be assured.   If not,  the  fuel patterns in the utility industry may be changed
 drastically,  since  nuclear power would have almost an unchallenged position in the
 industry.

      In summary,  it may safely  be concluded that the price of fossil fuels, with
 only minor exceptions, i.e., coal burned in the Rocky Mountains at mine-mouth
 plants where  dry cooling tower? would permit economic utilization of this coal,
 will be higher by from 5 so 15^/million Btu in the-next two decades in all geo-
 graphic regions  of  the US.  Substantially  higher spot prices-for low-sulfur coal
 will.also be  experienced until  competitive forces tend to stabilize the market.
 These higher  prices have contributed to higher overall busbar energy costs, and
 prompted most utility companies to apply to regulatory agencies for rate increases
 (Ref.  32).  Recently in some areas, i.e.;  the Northeast,utilities have been
 permitted to  pass along higher  fuel costs  to the customer without applying for
 continual increases.   As a  basis for comparison and for use in later phases of the
 study,  prices projected .for the various fuel sources, when applicable, are presented
 in Table VI.
                REVIEW OF REGIONAL COOLING WATER AVAILABILITY AND
                         THERMAL POLLUTION RESTRICTIONS
     During 1965, the cooling of steam condensers in electric generating plants
accounted for almost 60% of the total of 110,000 million gallons per day of water
used in the US for industrial cooling, and for nearly one-third of the total water
used for all purposes (Ref. 33).  If present projections of the electric power
industry growth rates are correct, and steam power plants remain the dominant
type of power generation system, the once-through cooling requirements of this
industry alone would reach a point, possibly by the year 2000, where one-half the
average daily runoff of all rivers in the US would be needed.  The cooling water
shortage will accelerate as nuclear-fueled stations provide a larger portion of
the electric demand, and with continuing growth in population and industrial
productivity.  Regional redistribution of population and economic activity toward
the west will further aggravate the local water shortages and degradation in
qualify of water resources in many parts of the country.  The availability of
cooling water for future power plant sites as well as for additions to existing
plants poses a major problem to the electric utilities.

     The first national assessment (completed in 1968) of the adequacy of supplies
of water necessary to meet all water requirements (domestic,  industrial,  agricultu-
ral, electric power, etc.) in each of IT water resource regions in the US (see
                                        26

-------
Fig. 6) showed that shortages of natural runoff and ground water supplies  have
become serious problems in nearly half of the regions (Ref. 3^).   The pertinent
results of this assessment are summarized in Table VII in which the relative
severity of existing and emerging water management problems for each region is
identified by an assigned rating from 1 to H.  Further, 11 of the IT water resource
regions surveyed (see Fig. 6) presently lack sufficient natural runoff to  satisfy
the year-round power plant condenser requirements with once-through cooling; by
1980 only coastal states will possess adequate supplies of cooling water.  In
Ref. 35, it was estimated that in 1972 about k5% of the total US power generation
capacity will require some type of supplemental cooling apparatus,  such as cooling
towers, to alleviate the demand for condenser cooling water.   By the 1980's, the
same reference estimates that almost 10% of the installed capacity will use
supplementary cooling devices, and only those power stations  convenient to the
oceans will be able to reject heat in once-through cooling systems.

     However, even in those local areas where adequate cooling water exists,
unlimited use of rivers, lakes, and estuaries for cooling will not  be permitted
since effective action was taken through the Water Quality Act of 1965 to  control
waste heat discharges as well as other types of water pollutants.   As a result of
this legislation, water quality standards are being set and implemented for all
coastal and interstate waters.

     All 50 states have submitted water quality standards containing temperature
criteria to protect designated water uses, particularly aquatic life propagation.
Standards for temperature changes and maximum temperature limits  vary from state
to state.  Table VIII summarizes the temperature criteria proposed by the  individual
states in the South Central Region, as of the end of 1968, for interstate  and
coastal waters.  As of April 1970, 20 states did not yet have their water  temperature
standards approved in entirety (Ref. 36).  Most states have established 68 F as
the maximum allowable temperature and from 0 to 5 F as the maximum allowable change
in temperature for streams with cold-water fisheries.  For warm-water fisheries,
the maximum allowable temperatures are generally in the range of 83 to 93  F, and
the maximum allowable rise is in the range of k to 5 F (Ref.  33).

     Although the importance of the mixing zone in determining the amount  of heat
that may be discharged to a water body is recognized, allowable limits for this
parameter have not been clearly defined in all cases, and as  a result a number of
utilities have delayed complying with the specified temperature standards  (Ref. 36).
Furthermore, the heat-accepting capability of a lake, reservoir,  or stream is diffi-
cult to estimate because of the many variables involved, and  some utilities may
be anticipating upgrading revisions in the standards.  Several studies relating
to the ability of water bodies to dissipate heat to the atmosphere have been
completed; however, the apparent results of these studies vary considerably among
water bodies and geographicl locations.  Since the cooling water discharged from
power generating plants using once-through cooling is often heated 10 to 25 F
above the intake water, both substantial local heating and high temperature levels
                                         27

-------
 are  sometimes  produced  in  cooling streams, especially during lov water conditions.
 However, in  other situations, stratification may occur without substantial mixing
 and  may extend for long distances.  In another instance, the temperature of
 the  Monongahela River in August along a 4o-mi stretch upriver from its confluence
 with the Allegheny averages approximately 85 F, with local temperatures as high
 as 95 F, as  shown in Fig.  7 (Ref. 37).  Although much of the heat input is from
 a concentration of industrial plants located near the larger cities along the
 river, measurements and even model simulations have indicated that 15 to 20 mi can
 often be required for a river to return to normal temperatures after waste heat
 discharges from a large power-generation facility (Ref. 38).

     Attempts  have been made in a number of studies to predict the number of power
 stations that  will be required to meet the electric power demands of the United
 States to 1990 and the  number of these stations that will need auxiliary cooling
 systems.  In one such study (Ref. 33), it is predicted that by 1990 158 stations
 of a total of  ^92, of 500-Mw capacity and above, will require cooling towers.  In
 making these projections it was assumed that future stations in the coastal areas
 and  in the vicinity of  large lakes and streams would use once-through reservoirs
 or cooling ponds.  Another source (Ref. 39) indicates that these assumptions are
 optimistic and that the number of stations requiring cooling towers will be even
 greater than those estimated in Ref. 33.   A major manufacturer of steam-electric
 power plants predicts that by 1990 the typical station in some locations might be
 required to  utilize nonevaporative cooling towers due to the unavailability of
 cooling water.  Another manufacturer concedes that dry towers may be more frequently
 used in specific locations but they would not be typical even by 1990.  It is the
 opinion of this manufacturer that it would be generally cheaper to transmit power
 over a longer  distance  to the load center if availability of cooling water at the
 first station  site is a problem.

     A recent  survey of the proposed plans by utilities to meet thermal standards
 indicates substantial increases in the use of supplementary cooling devices (Ref. 36)
 The  69 companies which  replied to the survey reported that l6U stations presently
 use  some form  of tower  or ponds.  All but two of the larger utilities (with capa-
 cities greater than 200  Mw) responding to the survey were located in the south-
 west or arid plains states.

     However,  in Ref.  kOt estimates are made of the potential utilization of
 supplementary  cooling devices under three different  assumptions of thermal quality
 standards.   This analysis indicates that less than 8$ of the approximately 200
 million kw of  installed major thermal generation (500-Mw station capacity and
 above) in the  US would  use cooling ponds  and about 13$ would use cooling towers.
 Under more stringent thermal quality assumptions, almost 90% of the new major
 capacity added in the US during the 1970's and 1980's would utilize towers or
 ponds.  The  Northeast and East Central Regions would utilize cooling towers for
 from ko to 6ofa of the new plant capacity in the next twenty years but there would
be relatively minor utilization of cooling ponds for about 20 to 30% of the new
                                        28

-------
capacity.  Cooling ponds and cooling towers would "be used in at least 50% of the
nev capacity in the other four regions of the country.   Thus, it may be concluded
that only those utilities close to ocean water or large rivers such as the
Mississippi will be permitted to use once-through cooling systems.   Most areas  in
the South Central Region, as noted in Fig. 6, would require cooling ponds or
towers (Ref. kl) and are planning to use these alternatives.
             ESTIMATES OF PRESENT-DAY AND FUTURE CONVENTIONAL STEAM
                POWER PLANT PERFORMANCE AND COST CHARACTERISTICS
     For many years, steam power systems have maintained an overwhelming position
in the field of electric power generation.  Consequently, this type of power
system is and will continue to be the standard against which the feasibility of
alternative methods of generating electric power must be compared.   This section
includes a description and discussion of present-day steam power plant charac-
teristics and limitations.  Also included are estimates of the performance and cost
characteristics of advanced steam power plants which might be built with technology
potentially applicable in commercial configurations during both the 1970 and 1980
decades.
                                 Unit Capacities

     The unit capacity of a base-load or cycler steam power plant purchased by an
electric utility is selected after detailed analyses which include the effects of
power generation, transmission, and distribution costs, as well as reliability and
availability, on the total cost of providing power.  In recent years,  the expansion
of intertie systems has permitted utilities to take advantage of the economies of
scale in purchasing steam power units, and the average-size unit is rapidly
increasing.  As a result, a rough rule of thumb which has been applied to past
base-load capacity additions is that the capacity of new units should not exceed
10$ of the total utility system generating capacity.  For a large electric
utility system, such as TVA or American Electric Power, this 10% rule  would dictate
the selection of 1000-Mw units today.  However, these large systems are not repre-
sentative of the US electric utility industry as only three utility systems have
capacities close to 10,000 Mw.  It may be noted that a steam power station may
consist of one or more units and the units may be of different sizes.

     Data from Ref. 2, on scheduled additions of steam power generating capacity
'by years, is presented in Table IX.  These data are based on scheduled dates of
commercial operation as of October 1, 1969.  A general trend of increasing unit
size with time may be observed for both conventional (fossil) and nuclear steam
                                         29

-------
 power  plants.   Conventional units scheduled  for operation this year and for the
 early  1970"s have  average  capacities of approximately 300 Mw and 500 Mw, respec-
 tively.   This  trend is  expected to continue, and by the 1980's average steam power
 plant  units may reach output  capacities of approximately 1000 Mw.  Data in Ref. 1
 appear to substantiate  the present average unit size and the trend to larger units.
 Thus,  unit sizes of 500 and 1000 Mw may be considered representative for the 1970
 and 1980  decades,  respectively.  It may be noted that the average size of nuclear
 units  is  running ahead  of  conventional units.  Average nuclear unit sizes would
 be  expected to reach the 1000-Mw level by the mid-1970's.  The distribution of
 unit size for  1968-to-1971.fossil- and nuclear-steam power plant installations
 (Ref.  1*2) is shown in Fig. 8.  However, the availability of large (800 to 1300-Mw
 fossil-fueled  steam plants has been discouragingly low and until there are units
 which  reach the previous levels of availability, there will be a pause in the
 trend  to  increasing unit sizes.  Although there are some differences between the
 data in Fig. 8 and the  data in Table IX, both sets of data highlight the trend
 toward larger-capacity  units.  The scheduled additions of steam power plant
 generating capacity (based on the same data included in Table IX) are presented
 in  Table  X for the six  power regions of the US.

     The  capacity  of conventional fossil-fueled units scheduled for operation in
 late 1969 for  almost all regions will average less than 500 Mw, whereas nuclear
 unit size additions will average approximately 900 Mw or more in all regions.
                                Steam Conditions

     Historically, improvements in steam power system technology have permitted
increases in steam temperature and pressure, thereby resulting in increases in
station efficiency and lower net fuel charges.  Generally, however, as technology
advanced, more expensive equipment was required to contain the steam so that capital
cost increased significantly. Until recently, increases in unit size together with
higher specific power levels achieved from the higher operating conditions and the
resulting reduced fuel charges always outweighed the higher capital costs  due to
the advanced steam conditions.  At the present time, the highest practical steam
temperature in new plants appears to have reached a plateau of approximately 1000 F
with a single reheat to the same approximate temperature.  A second reheat is
not justified in present-day units because the capital costs would outweigh the
marginal saving in fuel charges at present levels of fuel costs (Ref.  U3).
Presently, 2^00 psig is the most common steam inlet pressure to the high-pressure
turbines in medium-size units, although several large units incorporating  super-
critical boilers operating at 3500 psig are now operating, under construction,  or
in the planning stages.  A tabulation of units under construction and scheduled
for operation by 1973 in Ref. hk further illustrates the diversity of steam condi-
tions in present plants.  Data on the characteristics of medium- and large-size
units installed by TVA show that 2^00-psig steam pressure was used on unit sizes
up to 700 Mw, and 3500-psig steam pressure was used on units of 950 Mw and larger
                                        30

-------
(Ref. UU).  Operating problems vith supercritical units,  hcrwever,  have not  been
completely overcome, so that the 2UOO psig/1000 F/1000 F  steam cycle vould  be
considered representative for present-day and 1970-decade steam power  plants
(Ref. U5).

     Equipment manufacturers as well as architect-engineering  firms  were  questioned
specifically regarding the long-term (10 to 20 years)  trends in steam  conditions
(see Ref. 29).  Surprisingly, there was unanimous agreement that there would be no
increases in steam temperature beyond approximately 1000  F, nor increases in steam
pressures beyond approximately 3500 psig.  This general conclusion is  also  stated
in Ref. U6.  Previous experience with units rated at 1100 F to 1200  F  (e.g., the
Eddystone Station of the Philadelphia Electric Co.  and the Bergen  Station of the
Public Service Co. of New Jersey) has not been encouraging, and these  stations
have been downrated to approximately 1000 F to improve their availability.
Although these units were installed primarily to determine the reliability  and
performance obtainable at higher steam conditions,  their  continued operation at
high temperature could not be Justified economically.   The basic cost  problem
with high-temperature operation arises because austenitic-type stainless  steels
must be used above 1000 F.  Since austenitic steels are considerably more expen-
sive than ferritic steels, high-temperature boilers would be very  expensive and
the incremental cost of these boilers over boilers  constructed from  ferritic
steels generally would not be offset by the incremental fuel saving.  For example,
several boiler manufacturers (Refs. bl and U8) estimated  that  a boiler designed
to generate 1100 F steam would cost approximately Q% to 10% more than  a 1000 F
boiler which generally costs about $35/kw«  Similarly, steam turbine manufacturers
indicated that steam turbines designed for 1100 F and 1200 F would cost approxi-
mately 10 to 15$ and 20 to 25/5, more respectively,  than steam  turbines designed
for 1000 F steam.  Several equipment manufacturers, as reported in Refs.  U3 and  U9,
conducted analytical tradeoff studies which indicated that these increased  equip-
ment costs -rould not oe Justified unless the fuel cost were to exceed  approximately
^5 to 50<£/million Btu.  This result can be substantiated  by comparing  the fuel
economics of a 2UOO psig/1000 F/1000 F cycle with a UOOO  psig/1200 F/1200 F cycle.
The total incremental cost for the high-temperature system was estimated  to be
approximately $17.3/kw more than the cost of a 1000 F system,  whereas  the differen-
tial efficiency between the high- and low-temperature systems  was  estimated to be
approximately 3.U percentage points.  Using a lk% fixed charge, 10%  load  factor,
and the figures stated above, it was calculated that the  1200  F system could be
Justified only at fuel charges greater than 53<£/million Btu.

     Consequently, 1000 F would be considered representative as the  maximum cycle
temperature for the 1980-decade as well as the 1970-decade conventional steam
stations.  Since the 3500-psig supercritical cycles do show a  slight increase in
erriciency relative to cycles operating at 2^00 psig,  and since operational
problems experienced with these supercritical pressures should be  overcome  by
the 1980 time period, the 3500-psig pressure level would  be considered representa-
tive for the 1980-decade steam stations.
                                        31

-------
     It may be noted that "boiler manufacturers have indicated (see Ref.  U9) that
 the  fuel  type or its cleanliness has relatively little effect on the maximum
 allowable steam temperature, although it does have a significant effect on boiler
 cost.

     Thus, during the 1970 decade, representative steam systems would be 500-Mv
 units operating with 2^00 psig/1000 F/1000 F steam conditions.  The steam turbines
 would be  3600-rpm, tandem-compound, l;-flow machines with 30-in. last-stage blades.
 There would be provisions for 7 stages of extraction for feedwater heating.  In
 addition,  steam would be extracted from the low-pressure crossover pipe at approxi-
 mately 185 psia to supply steam to the two boiler feed pump turbine drives.  The
 boiler and boiler auxiliaries would be completely enclosed in a. building in most
 areas of  the country.  Each boiler would be pressurized by two half-capacity
 forced-draft fans, and would be equipped with an air preheater, soot blowers,
 economizer, and a 100/5-capacity, condensate polishing demineralizer to prevent
 carry-over of dissolved solids.  The steam would be condensed in a single-pressure,
 two-pass ,  twin-shell condenser wherein the tubes would be perpendicular to the
 turbine centerline.  Wet cooling towers would be used rather than once-through
 cooling.   Steam from the boiler feed pump auxiliary turbine drives would also be
 condensed in the main condensers.

     The  design of the 1980-decade units, averaging 1900 Mw and operating with
 3500 psig/1000 F/1000 F steam conditions, would be similar to that described
 above except for the size of the last-stage turbine blades.  A detailed  arrangement
 of equipment of a representative large coal-fired station is presented in Fig. 9-
                                   Performance

     Although the steam temperature is projected to remain essentially constant
throughout the entire 20-year time period under investigation, slight performance
improvements are anticipated because of the increase in pressure level between
1970-decade and 1980-decade systems and also because of slight improvements in stear
turbine and boiler efficiencies during these time periods.  Projected performance
characteristics for 1970- and 1980-decade steam power stations are presented in
Table XI.  Station efficiencies are presented for design-point operation and for
operation at 70$ load factor, and include allowances for all auxiliaries.  Diffe-
rences in boiler efficiency and power station auxiliary power requirements  account
for the differences in the net station efficiencies among the coal-,  oil-,  and
natural gas-fueled power systems.  The oil-fired power systems exhibit slightly
higher efficiencies than coal-fired systems, primarily because of the lower
auxiliary power requirements.  Power systems fueled with natural gas  exhibit lower
efficiencies primarily because of the greater moisture losses in the  boiler.  In
actual practice the level of sophistication in the plant design and the ultimate
efficiency and cost of the station is related to the cost of fuel,  and where low-
cost fuel is available plant efficiency would be decreased accordingly.
                                        32

-------
                                  Station Costs

     The cost of "building a modern conventional steam power system is  dependent
upon a great many factors including unit size, type of fuel burned, sophistication •
of system design, location, type of air pollution controls (if any), type of
construction (indoor or outdoor), and the type of heat rejection system.   Repre-
sentative total station costs given in Ref.  50, based on discussions with an
architect-engineering firm, for indoor present-day large-capacity base-load
coal-fired stations range from about $l65 to $175/kw,  and for gas-fired stations
from $1^0 to $150/kw.  Details describing the cost breakdown for each  type of
station are presented in Section VIII of this report.   Detailed estimates per-
formed for this study based on Q% interest rates and 15% fixed charges indicate
that the cost will remain about the same through the 198d's.  A method of
generalizing these costs for each region will be described in Section  VIII.
These costs include all indirect items such as engineering, design, and escalation
and interest during construction as well as direct component cost.   It has been
estimated that complete outdoor construction would result in costs  $5  to  $10/kw
lower than those quoted above (Ref. 29).

     A summary of costs for the TVA 950-Mw Bull Run 1 coal-fired steam plant taken
from Ref. kh is presented in Table XII and the total of $15T-6U essentially
verifies the values quoted above.  Although the Bull Run cost figures  include all
direct and indirect costs such as interest and escalation, it should be noted that
TVA building costs, in general, are often not representative of the electric
utility industry because of their ability to borrow money at low rates and to buy
equipment at low prices.  Also included in Table XII is a summary of costs from
Ref. U5 for a 1000-Mw nominal station containing two 500-Mw coal-fired units.
The summation of the direct costs amounts to $12U.72/kw, whereas when  the indirect
costs are included the total cost is $176.95-  It should be noted that power plant
costs quoted in the literature often are summations of the direct costs only.
It is obvious that costs significantly lower than those quoted above,  on  the order
of $80 to $120/kw, do not include the indirect costs and therefore do  not reflect
the true cost of building a steam power plant.

     As previously mentioned, steam power plant specific cost depends  upon unit
capacity.  A typical variation of specific cost ($/kw) with unit capacity (taken
from Refs. 27, 51, and 52) is depicted in Fig. 10 for conventional (fossil-fueled)
steam power systems together with that for the nuclear plants.  It may be noted in
Fig. 10 that significant cost savings can still be achieved by building steam
power plants in sizes greater than 1000 Mw.   Data from a local utility (Ref. 52)
for oil-fired base-load, oil-fired cycler, and nuclear units are also  shown in
Fig. 10, and for the most part verify the level of costs for fossil fuels taken
from Ref. 27-  It may be seen that cycler-type plants  cost approximately  $20 to
$2l*/kw less than base-load type plants.
                                        33

-------
      Although the costs quoted above pertain primarily to present-day systems,
 they are expected to apply to  future systems as veil.  Costs will tend to in-
 crease due to the addition of  equipment  for control or elimination of air pollution
 and vater thermal pollution but vill also  tend to  decrease due to the increased
 use of larger-capacity units.
      ESTIMATE OF PRESENT AND FUTURE PERFORMANCE AND COST CHARACTERISTICS OF
            ALTERNATIVE METHODS  FOR COOLING  CONDENSER WATER DISCHARGES
      The previous  discussion  indicates that the availability of naturally occurring
 vaters  suitable  for large  condensing water systems used in steam-powered electric
 generating plants  will be  limited  in many areas of the country.  Where sufficient
 quantities of cooling water do  exist, their use may be restricted because of the
 temperature effect on these waters  from a once-through condensing system.  As a
 result,  alternative methods to  once-through systems using river or sea water are
 being given greater consideration  in many areas of the country for condenser
 cooling in steam-powered generating plants.  The types of cooling systems which
 are in  use at the  present  time  or  considered feasible in the near future include
 once-through systems using river or sea water, cooling ponds and reservoirs,
 spray ponds, spray cooling canals,  and wet and dry cooling towers.  A brief
 description of the operating  principles, performance limitations, area of mainte-
 nance requirements,  cost characteristics, and effect on the performance of steam
 systems  when using these alternative systems are presented in the following para-
 graphs .
                      Description of Alternative Systems

Once-Through Cooling

     The most common method used for the removal of heat from steam power plant
condensers in water-rich sites in the United States is once-through cooling (see
Fig. lla).  Cool water is diverted from a river, natural lake, or estuary and
pumped through the power plant condenser tubes, thus condensing the steam working
fluid on the outside of the tubes.  The river or lake cooling water is generally
heated from 10 to 25 F in the condenser, depending upon the number of passes of
the cooling water.  Single-pass condensers are normally used in once-through system
to minimize their size and surface area.  However, to reduce the temperature
rise of the water, large flows of cooling water are required.  For example, to
cool a typical 1000-Mw fossil-fueled plant and limit the temperature rise of the
cooling water to 10 F, about 2100 cu-ft/sec of water would be required.   Where the
natural flow of water available for cooling over the entire year may not be

-------
adequate to meet the requirements, two-pass condensers can be used,  resulting in
higher cooling vater temperature rises, higher turbine back pressures,  and reduced
plant efficiency.  Although once-through cooling is usually- the  most economical
to install and operate, the intake and discharge channels  must be carefully lo-
cated to prevent recirculation of the warm discharge water.  Sometimes  skimmer
walls, diffuser systems, or long intake and/or discharge lines are used,  and often
costly models are needed to predict the hydraulic and thermal flow patterns
(Refs. 53 and 5^)-  All these items add to the costs of the once-through  cooling
system.

     Once-through cooling using sea water is perhaps the second  most common
type of condenser system and much the same factors must be considered in  its
design.  However, higher-quality corrosion-resistant materials are required in
sea water systems, as well as long intake and discharge lines to maintain an
adequate supply of water during tidal movements.  These factors  can  raise the
cost of a sea water installation considerably.   According  to data in Refs.  55 and
56, the cost of the condenser alone can be 25$ more than conventional once-through
river units.  Discharge conduit costing $500 to $1000 per  ft or  capital costs from
$500,000 to $1 million per 1000 ft are indicated in Ref. 55.  Discussions with an
architect-engineering firm confirms that once-through sea  water  systems could add
as much as $5 to $10/kw of installed capacity (Ref. 29) in extreme cases.

Cooling Ponds or Reservoirs

     A cooling pond or reservoir is a man-made body of water into which the warm
condenser discharge is pumped so that it may be cooled and eventually circulated
through the condenser.  Although it is stated in Ref. 57 that the natural rolling
topography of much of the nation is favorable for the formation  of man-made lakes
to retain the water during high runoff periods, many cooling ponds are found in
the hilly Southeast and lower portions of the East Central Regions of the US.
A lake may be constructed by placing an earth dam at the junction of one  or more
small streams and allowing the runoff to fill up the low area.  In some instances,
it may be necessary to construct the lake and retain the vater using earth dikes.
With water supply to the lake from runoff, the drainage area necessary is depen-
dent on the natural and forced evaporation rates, rainfall, and  expected  periods
of drought.  In general, the drainage area required is about ten times  the lake
surface area and two or three years may be required for initial  filling of the
cooling pond.  The pond surface area required will depend  upon the temperature of
the condenser discharge and other factors, but in general  1 to 2 acres  of water
surface area are used per megawatt of generating capacity.  However, the  diffi-
culty of installing a cooling pond in the minimum area very often requires that
the utility must buy from 2 to 3 times the pond acreage in the surrounding area
for an adequate site.  Kolflat, in Ref. 57, estimates the  cost of a  cooling pond
at about $2.50/kw for a 1000-Mw plant with the cost of land accounting for about
^0% of this total, hO% to clear the land, and the remaining 20$  for  the darn and
spillway.   These costs appear reasonable if the land for the entire  site  costs
only several hundred dollars per acre.  However, it is suggested in  Refs. 56 and
                                        35

-------
 58 that  the cooling pond costs may be somewhat higher, especially if a dual-pressure
 condenser vere  used to minimize the pond area required.  The accessory equipment
 and a detailed  listing of the costs associated vith the construction of a cooling
 pond are presented in Ref. 58.  Cooling ponds are often classified according to
 the circulation pattern and temperature distribution as completely mixed, flow-
 through, or internally circulating, as described in detail in Ref. 56.

 Spray Ponds

      The area requirements of a cooling pond can be reduced by at least an order
 of magnitude if the water used for cooling is sprayed into the air.  The evapora-
 tion rates  are  enhanced and cooling occurs more rapidly, since the water droplets
 remain in intimate contact with the air for longer periods of time.  The spray
 nozzles  are usually located up to 10 ft above the surface of the pond and must be
 carefully placed with respect to each other and according to the prevailing winds
 if the spray pond is to be effective.  Extensive data on spray pond design for
 utility  application is not available although some tests have been under way to
 verify their performance.  The costs of spray ponds, including the piping, nozzles,
 pumps, and  installation are estimated to be about $2.50/kw in Ref. 56.

 Spray Cooling Canals

      A potentially low-cost variation of the spray pond for condenser cooling is
 the  use  of  spray cooling canals.  In this system the condenser cooling water is
 directed through a canal where it passes through a series of floating spray nozzles.
 The  water is cooled by evaporation and convection through each pass of the nozzles
 until the desired approach to the ambient wet bulb temperature is achieved.  The
basic  spray nozzle unit is comprised of a pump and four spray heads with inter-
 connecting  straight-line piping.  The entire unit floats in the water and is moored
 in place.   It is claimed that flotation eliminates the need for special and expen-
 sive basins, foundations, and complex pump and piping distribution systems such
 as those required for spray ponds (Ref. 59).  In addition, a much coarser droplet
 size of  3/8 to  1/2 in. dia is produced which eliminates clogging of the spray
nozzles  and reduces carry-over loss.   Designs using a multiple of spray units
arranged in rows across a channel l60 ft long have been made that will accommodate
the waste heat  from a typical 1100-Mwe steam plant.   Such a system is being
 installed in a  utility in the northeast and pilot tests are in progress at other
utilities in the east central, south central, and south atlantic regions.   The
 costs of the powered spray modules only are estimated at approximately $2/kw
 (Ref.  59)-  Added costs for channel construction, field assembly, moving materials,
etc., are included in Ref.  60 and indicate only another $0.50 to $1.00/kw for
these items.  The cost of installing a spray cooling canal system is apparently
a strong function of the cost of constructing the cooling canal in many installa-
tions.
                                        36

-------
Wet Cooling Tovers

     In a vet tower, vater is cooled largely by evaporation  of  a portion of the
circulated flow.  The evaporated water is  absorbed by  the  air flowing  through the
cooling tower which is in direct contact with  the  circulated water.  During
summertime operation, when the air temperature is  high,  evaporation provides
for the larger fraction of the heat transfer to the  air.   For example, the waste
heat from a 1000-Mw fossil-fueled plant would  require  the  evaporation
of approximately 6750 gallons of water each minute of  operation (Ref.  6l).  A
nuclear-fueled plant, due to the larger quantities of  waste  heat that  must be
rejected, would require the evaporation of about 10,125  gallons of water each
minute of operation.  During the cooler months of  the  year,  the evaporation quanti-
ty would be reduced to about 5^00 gallons  per  minute and about  8100 gallons per
minute for the fossil- and nuclear-fueled  plants,  respectively.  The remaining
waste heat is removed by sensible heat transfer within the water cooling tower
due to the temperature difference between  the  water  and  air.  Of course, the
portion of the circulated cooling water that leaves  the  water cooling  tower system
as evaporation must be replaced and represents the largest portion of  the water
demand for siting considerations in a power plant  with wet towers.

     The most often-used cooling tower arrangement is  the  total recirculation
system.  With this system, the tower to be installed must  satisfy the  total waste
heat dissipation requirements of the generating plant  regardless of the season of
the year or the generating load.  Such systems are usually termed "closed-circuit"
recirculation systems since the water used for cooling remains  within  the system
and only evaporation, drift, and blowdown  losses must  be replaced from the water
source (see Fig. lib).  Open-circuit, once-through systems are  used where suffi-
cient water is available to supply the plant's requirements  but water  discharge
temperatures must be limited (see Fig. lie) (Ref.  6l).   Temperature reduction of
the discharge flow from the plant condenser is accomplished  in  the cooling tower
and the flow is then returned to the water source.   Sometimes the entire flow is
pumped through the cooling tower or the tower  is used  to cool only a portion of
the flow which is then mixed with warm water from  the  condenser discharge to meet
the temperature standards.  Other configurations are also  available, as described
in Ref. 6l, to match the power plant cooling requirements  and remain within
regulated temperature limits over the entire year.

     Wet towers can be further classified  as either  mechanical- (forced or induced)
draft or natural-draft towers referring to the means of  providing the  air circula-
tion through the tower.  In the mechanical-draft design, large-diameter fans
driven by electric motors induce the air through the circulating water which
flows over splash surfaces that are provided to interrupt  the flow of  water and
increase the contact period between the air and water  (see Fig. 12a).  A number
of different arrangements of the airflow and cooling water streams are possible
(i.e., cross-flow, counterflow) as indicated in Ref. 56.   Natural-draft wet cooling
                                        37

-------
 towers (Fig.  12b)  utilize concrete chimneys  to  induce  air through similarly-
 arranged heat transfer surfaces.

      The total water requirements  necessary  for the installation of a wet cooling
 tower include that portion needed  for  evaporation, plus an amount for physical
 water losses  known as "drift"  due  to droplets entrained in the leaving airstream,
 and an amount which must  be bled from  the  circulating water system to limit the
 concentration of dissolved solids  in the circulating water.  Estimates of the
 total water demand for 1000-Mw fossil- and nuclear-fueled plants are presented
 in Ref.  6l and indicate that from  7700 to  20,250 gallons per minute could be
 required depending on the design temperature range of the tower and allowable
 chemical concentrations of the makeup  water.

      Several  factors are  important in  determining the size of a wet cooling tower:
 (l)  heat load,  (2)  range, (3)  approach, and  (1+) wet bulb temperature.  Normal
 approach temperatures of  10 F  are  used in  mechanical-draft towers with 15 to 20 F
 typical in natural-draft  towers.   Due  to the higher drafts obtainable with
 mechanical-draft towers,  higher packing water loads are possible relative to those
 in natural units and the  ground area requirements are usually only one-half to
 one-third of  equivalent-capacity natural-draft  units.  However, mechanical-draft
 units must be located further  away from the  plant than natural-draft towers and
 thus require   considerably more connecting piping.  Typical dimensions are
 given in Ref.  56 for a 1000-Mw nuclear plant and indicate base areas of 2.67 x
 105  ft2  and 1.33 x 105 ft2 for the natural-draft and mechanical-draft units,
 respectively.   Thus,  two  natural-draft units each Ul2 ft in diameter could be used,
 while 10 square cells arranged in  a configuration 115 ft wide by 1150 ft long would
 be used  for the mechanical-draft units.  Typical heights would be 60 ft for the
 mechanical-draft tower and over UQQ ft  for the natural-draft unit.  In addition
 to the area required for  the tower itself, land is also required for reservoirs,
 blowdown ponds,  and  pumping and storage areas.  The towers must also be located
 with proper spacing between each other to  avoid damage and possible destruction
 during high winds such as  that which occurred in England several years ago.   The
 location of some of the larger mechanical- and natural-draft tower installations
 in the US  are shown  in Fig. 13.

 Dry  Cooling Towers

      Dry or nonevaporative  cooling towers  reject the waste heat in the warm conden-
 ser  discharge flow entirely through convective heat transfer,  depending upon the
 temperature difference between the heated water and ambient air.   Dry cooling
 towers have been used extensively  in the chemical processing and petroleum
 industry but, except  for small installations, are not used in the US electric
 utility  industry.  However, dry cooling towers have been used in steam genera-
ting plants with capacities of 200 Mw in Europe and South Africa (see Ref.  62).
There  are two basic types  of air-cooled condensing systems — the indirect  system
                                        38

-------
and the direct system.  With the indirect dry-type cooling tower  system (see
Fig. 1*0» a spray condenser is used at the turbine exhaust vith circulating
vater sprayed into the unit to condense the steam.  The heated water  is then  circu-
lated from the spray condenser to a tubed cooling coil in the  dry cooling  tower
over vhich the cooling air is passed and the warmed air is discharged to atmosphere.
For each pound of condensate leaving the condenser hotwell and going  to the power
plant cycle, about 25 to kQ pounds leaves the hotwell and is pumped through the  air-
cooled heat exchanger coils.  The pressure in the piping and cooling  coil  tubes
is maintained above atmospheric pressure by pumps to prevent air  leakage into the
system.  In some instances, power recovery is proposed by letting the pressure
down through a turbine to the pressure level used in the spray condenser.  At
the present time, dry cooling towers using the indirect condensing cycles  (some-
times referred to as the Heller system) are being studied extensively for  potential
use in power plants (Refs. 6k and 65).  The use of a direct air-cooled combining
system in which the exhaust steam is piped to tubed condensing coils  for air
cooling is not considered feasible for large power generating  plants  beyond 200  Mw
because of the large piping and tubing requirements when forced to operate under
low absolute pressure conditions.

     Dry towers are possible using either mechanical draft or  natural draft to
provide for air circulation over the cooling coils.  Area requirements for each
type are estimated in Refs. 62, 63, and 6k and depend upon a number of design
factors such as the condenser operating pressure, ambient air  temperature, etc.
Typical areas are about 300 sq ft and 900 sq ft for each Mw of fossil-fuel plant
capacity for the mechanical- and natural-draft types, respectively.   Nuclear-
fueled plants would require about 60% more area than fossil-fueled plants  per Mw
of output for the same design conditions.  Reference 62 contains  an outstanding
summary of dry-cooling tower technology as well as the estimated  costs and design
parameters for cooling towers suitable for conventional nuclear-  and  fossil-
fueled steam plants in 27 US locations.
              Performance Penalty with Alternative Cooling Systems

     One of the principal parameters which governs the thermal efficiency  of a
steam power plant is the back pressure at the turbine exhaust.   This  pressure and
thus the temperature at which the steam condenses is  a strong function of  the
temperature of the condenser cooling medium (usually  circulating water).   Although
the condensing temperature of the steam could theoretically approach  that  of the
cooling water in typical systems, the condensing temperature is usually some 25
to 35 F higher than that of entering cooling water to provide the nroper economic
balance between condenser cost and plant efficiency.   For example, with cooling
water available at 55 F, a 90 F condensing temperature (equivalent to a l.U2-in.
Hg abs back pressure) might be maintained in the system.   Thus  with higher cooling
water temperatures as would be experienced in the summer  months, the  steam con-
                                       39

-------
 dens ing temperature would have  to  increase  accordingly and the plant efficiency
 would be reduced.   Cooling water temperatures  for power plant use in the US
 range from 32 F to as  high as 95 F.   It  appears that average cooling water and
 condenser temperatures would be about 65 F  and 100 F, respectively.  The use of
 wet or dry cooling towers would therefore require somewhat higher condenser design
 temperatures if the heat  is  to  be  rejected  to  ambient air.  In Ref. 56, condenser
 temperatures 10, 15, and  20  F higher  than those of once-through cooling systems
 are suggested for mechanical-draft wet towers, natural-draft wet towers, and dry-
 towers, respectively.   Higher condenser  temperatures are indicated in Refs. 62
 through 66 for the dry tower systems.  Optimum economic condenser design pressures
 were  found to vary from 8.5  to  15-6 in.  Hg  abs (corresponding to about 155 to
 l80 F)  in a study to determine  the applicability of mechanical-draft dry cooling
 towers  in steam power  plants suitable for various climatological conditions around
 the US  (Ref.  6h).   In  a study of dry  cooling systems for location all over the US,
 the optimum initial temperature difference  was found to be from about 50 to 60 F.
 Thus  the condenser temperature  would be  about  130 to lUO F for a design dry bulb
 temperature of 80  F.   Generally, large central power station steam turbine genera-
 tors  are limited to operation at back  pressures below 5 in. Hg.   Present designs
 of  large turbines  are  based  on  achieving maximum guaranteed kilowatt output at
 back  pressures of  3-5  in.  Hg, with reduced  capability for back pressures above 3.5
 in. Hg.   If prolonged  operation were attempted at back pressures above 5 in.  Hg,
 some  problems  would be anticipated due to bucket heating and vibration, thermal
 distortion of  the  exhaust  hood  and diaphragms, and abnormal stress due to thermal
 cycling (Refs.  62  and  6U).

      A  turbine which operates satisfactorily at back pressures above 5 in.  Hg
 could possibly be  achieved by several  alternative methods:   (l)  eliminating the
 last  row of blades  in  the  low pressure turbines in present turbine generators,
 (2) designing  a  large  turbine to operate at high back pressure by using somewhat
 shorter blade  lengths  in the last  stages than present stages but opening up the
 flow  passages  to permit higher  steam flows, and (3)  modifying present turbine
 designs  by  using blades only 25 to 30  inches in length,  increasing the blade
 structural  strength, and using  smaller hood structure and shorter bearing span.
 The effect  of  variations in  the turbine exhaust pressure on the fuel consumption
 and power  output of typical  nuclear- and fossil-fueled steam plants is shown in
 Fig.  15  based  on GE data presented in Refs.  62 and 6U for a present turbine
 design modified to  operate over high backpressures.   The data indicate that the
 steam power plant efficiency would be reduced by 8%  to lk% for fossil-fueled plants
 operating with turbine exhaust pressures of 8 in.  Hg abs  to 15 in.  Hg abs,
 values which appear typical  for dry cooling tower  designs.   Thus,  fossil-fueled
 plants could achieve a thermal efficiency of 38% (HR = 8980 Btu/kwh)  using  a  once-
 through  cooling system, whereas  if a mechanical-draft dry-cooling tower system
were  used, a thermal efficiency from 32.7 to 35.0$ would be achieved.   Efficiency
 penalties for nuclear plants would be almost 60% higher  than in  fossil-fueled
 stations but low fuel costs tend to offset  this effect when determining the total

-------
cost of producing power.  The efficiency penalties would be about  3 to k%  with
most wet-type cooling tower installations in either fossil- or nuclear-fueled
plants.  The performance of a new turbine design would be flatter  over the
range of exhaust pressures with poorer performance at low pressures and
improved performance at high pressures (see dashed line  in Fig. 15a).

     The power output from a steam plant also decreases with higher turbine
exhaust pressure as shown in Fig.l5b.  Thus, during summer months  when ambient
dry- and wet-bulb temperatures are above the yearly average, the power output from
a given plant would be reduced and additional plant capacity would be  needed  to
meet the nameplate rating of the station.  This factor could be especially costly
and serious in some regions of the country such as the Northeast,  East Central,
Southeast and South Central Regions where annual peak demands  are  experienced
during the summer months.
                               Total Cost Penalties

     A complete economic comparison of the alternative methods  of  rejecting
condenser waste heat from steam power systems requires specification  of a number
of design and economic factors including plant geographic  location, fuel costs
and capital charges, utility system load characteristics,  etc.   Several studies
have been devoted to such comparisons (Refs.  35,  56,  56,  and 60 through 62)  and
provide valuable sources of owning and operating  cost data.   A  selected summary
of these data is presented below for each alternative cooling system.

     At least four major factors must be evaluated in a cost comparison for
alternative.condenser cooling systems.  They  include:   (l) the  total  capital cost
to purchase and install equipment such as the condenser,  pumps, motors, piping,
cooling towers and/or cooling ponds, and accessories,  (2)  the cost of the auxiliary
power needed for circulating pumps and fans to operate this  equipment,  (3) the
maintenance costs for chemical treatment, makeup  water, and  repairs where applica-
ble, and (it) the increased annual fuel consumption and loss  of  power  output  as
the result of- operating at higher condenser pressures than would be required
with a once-through river system.

     A comparison of the total capital costs  for  the  alternative cooling systems
is presented in Table XIII based on data presented in selected  references.   Slightly
different assumptions were used to obtain the values  in the  various studies  but
the data appear fairly consistent for most systems.   However, several minor
discrepancies are apparent.  For example, the Ref. 63 data,  unlike those presented
in Refs.  56 and 35, indicate that once-through cooling using ocean water is  less
expensive than using river water.  Apparently the Ref.  63  estimates do not include
costs for discharge temperature regulation such as skimmer walls or long discharge
lines, especially for the ocean installations. The Ref.  56  data also indicate

-------
 only about a $l/kw additional cost  for  a  cooling pond system when compared to a
 once-through river installation.  Data  in Ref.  57, "by Kolflat, as well as those
 presented in Refs.  63 and 65, would appear to substantiate a cost differential of
 about $2.50 to $3.00/kw.   The Ref.  56 cost data for the mechanical-draft dry
 cooling towers also appear to be  somewhat low,  especially in light of the extensive
 optimization analyses performed to  arrive at the data presented in Ref. 6U.
 The cost estimates presented for  the natural-draft dry cooling towers show a
 wider range of differences,  but detailed  estimates presented in Ref. 62 and those
 in Ref. 66 indicate that  $20 and  $26/kw are realistic.

      The costs for the auxiliary  power  to drive the circulating pumps, fans, and
 appropriate accessories in alternative  cooling  systems must also be included in
 economic analyses.   Reference 63  estimates that O.U25$ of the generator output in
 a fossil-fueled plant is  required for a once-through river system and cooling lake,
 whereas a once-through ocean system will  need only 0.375$ because of the generally
 lower circulating  rates (see Table  XIV).   The same reference estimates 0.875 and
 1.075$ of the generator output is needed  for auxiliary power in natural-draft
 and mechanical-draft dry  towers.  As a  rule of  thumb, these values can be in-
 creased by 60$ for nuclear plants.   Data  in Ref. 67 state that the cooling fans in
 a mechanical-draft  wet tower would  require as much as 0.50$ of the power plant
 normal generating  capacity while  typical  values of 0.65$ (in Ref. 56) and 0.8$
 for a nuclear unit  are given by Kolflat.   The auxiliary power requirements for
 mechanical-draft dry towers  are given in  Ref. 6U and range from 1.60 to 3-50$ of
 the generator output,  depending upon the  condenser pressure level.  A value of
 3.05$ appears  only  slightly  above the average.  Although the added fuel cost to
 operate the auxiliaries produces  cost penalties of less than 0.1 mill/kwhr for
 an  80$ load factor,  the added capital cost  for the auxiliary power equipment
 can also add a comparable  cost  penalty.    For example, the capital cost to provide
 an  added 3$ for the  auxiliary power  in mechanical-draft dry-tower systems can
 add 0.065  mill/kwhr  to the cost of  generating power if the auxiliary power cost
 is  $100/kw (see Table  XIV).   Maintenance  and water makeup costs for the various
 systems  were  assumed to be one-half of the  costs for the added fuel to operate
 the  auxiliaries, based in part  on the limited data presented in Ref.  56,  except
 for  the  dry towers which,  according to Refs. 62 and 6k,  should require only
minimal  maintenance.

      The final  factor in evaluating the alternative methods is the possible
 increased  annual fuel  consumption and loss  of power output due to operation
 at slightly higher turbine back pressure  in comparison to most once-through systems
For  the  purposes of this comparison, it can be assumed that the use of cooling
ponds, spray ponds, and wet towers will result in the use of a condenser  tempera-
ture  oi  115 F  (or 15 F above  once-through cooling systems) and, for dry cooling
towers,  160 F.  Thus, the increased fuel consumption as  read from Fig.  15 would
amount to  about O.U$ and 10,0$, and the loss in turbine  capability about  0.6$
and 9.1$ for the wet-tower and dry-tower systems,  respectively.   Estimates of the

-------
added costs for both of these factors, when converted to mills/kwhr, are shown in
Table XIV based upon fuel costs of 30<£/million Btu and $100/kw as the value of
the incremental capability that would "be needed to make up the turbine output
deficiency.  The added costs for mechanical-draft wet cooling towers in excess
of that for a once-through river or ocean system would amount to about 0.07
mills/kwhr which is in general agreement with the extensive cost data for various
cooling systems presented in Ref. 60.

     In summary, it does not appear that the added installed costs or performance
penalties are particularly severe when cooling ponds, spray canals or wet towers
are selected as cooling systems for fossil-fueled plants, i.e.,  they add only
1 to 3% to the overall busbar cost of power production.  However, the costs of
these alternative systems tend to be substantially higher for nuclear plants, and
the use of dry cooling towers in either draft configuration would add at least
to the cost of power production and would result in serious consideration of
alternative power systems.
                            Other Considerations

     Other factors in addition to the general economic ones  previously described
are usually considered by a utility before selecting a particular method for
cooling condenser water.  These factors include the availability and cost of
water and environmental effects such as fog potential, consumptive water loss
by evaporation, drift, blowdown and aesthetic distractions.   Dry cooling towers,
unlike wet cooling methods, should have no adverse effects on the environment
(Ref. 60) and other than the large units which result with natural-draft types
would be completely satisfactory.  Their higher operating and capital costs in
many instances could be offset in areas where low-cost fuels and makeup water is
costly.  Cooling ponds are often used by utilities because they can be carefully
integrated into the surrounding area and provide beneficial recreation sites open-
to the entire community.

     Although the fog-producing potential of wet cooling towers is often mentioned,
visual observations and studies have shown that this is not the case.  An excellent
review and estimate of the fog potential of wet cooling devices is presented in
Ref. 60.  Wet cooling towers can sometimes produce localized vapor plumes which
together with their physical size can be somewhat objectionable in terms of
overall station appearance, proper site location.  Careful planning can essen-
tially eliminate all problems.  Concern over evaporation losses from wet cooling
devices is  considered in Ref. 60 and the results indicate that cooling ponds
would result in about 25% higher water losses relative to once-through cooling
based on average normal conditions in the Lake Michigan area.  The water losses
for other wet cooling devices would be only about 15% higher than for once-through
cooling systems.  Drift is usually encountered only in areas in the immediate
vicinity of the tower according to Ref. 60 and can be almost completely eliminated
"by control of air velocity and design of drift eliminators.
                                       1*3

-------
      Proper  maintenance  of vet  cooling towers to insure extended life requires
 coating and  chemical  treatment  of the system elements subject to corrosion and
 deterioration.   These inhibitors and the blow-down vater rejected to maintain a
 given vater  concentration must  "be carefully controlled to avoid discharging
 pollutants into  the cooling vater source.  Hovever, in most cases cooling vater
 can "be controlled to  avoid objectionable waste vater discharges.  The use of
 salt  vater in  cooling tower installations has been considered, and manufacturers
 have  apparently  stated it vould be practical although the tover costs vould be
 about 25% higher than for fresh vater use (Ref. 63).

      The availability of makeup vater to replace evaporation and other sources of
 vater losses from a vet  cooling tover and cooling pond is essential for these
 cooling methods.  Sufficient vater should be available in all areas generally east
 of the Mississippi and along the vest coast.  In the plains states and other vestera
 locations, the cost for  makeup  vater vould have to be carefully considered
 before a selection is  made.

      Finally,  climatic conditions such as high vinds and high vet- or dry-bulb
 temperatures vill affect the performance of cooling tovers and other systems to
 varying degrees.  Wet  tovers, vhich depend on lov vet-bulb ambient air temperatures
 for efficient operation, vill be favored less in hot, humid areas.  Conversely,
 dry cooling  tovers vould be favored in areas of lov dry-bulb temperature.  Hurricane-
 force  vinds  can  cause  damage to natural draft tovers, high vinds can cause uneven
 operation, and structural damage due to freezing can occur if the tovers are
 not properly designed.
                             Advanced Cooling Systems

     With increased emphasis on the use of cooling tovers and ponds for large
central pover stations in the next tvo decades, larger-capacity designs, improved
materials, and greater utilization of remote operation of these systems are anti-
cipated.  Most of the advances vill occur as nev materials such as plastics and
fiber glass are used for piping and coatings to reduce costs. Maintenance
problems associated vith heat exchanger corrosion and fouling are expected to
be reduced by the utilization of tougher coatings.  A number of advanced cooling
tover concepts have been studied in an effort to achieve improved performance or
lover installed costs (see Refs. 56, 63, and 65).  These programs are in the
preliminary stage and specific cost data are generally not available.

     There is also a need to explore the use of nev exchanger surfaces and
tover designs rather than continue vith designs vhich have evolved vithout exten-
sive analysis over the past half century.

-------
                                    SECTION VII
               TECHNICAL AND ECONOMIC CHARACTERISTICS OF ADVANCED GAS
                         TURBINE POWER GENERATING SYSTEMS
                                      SUMMARY

     An investigation was undertaken to define and select advanced fossil-fueled
open-cycle base-load gas turbine systems that have the potential for generating
lowest cost electric power while eliminating thermal pollution.   A review of
aircraft and industrial gas turbine research and development programs was made to
provide the basis for the selection of pertinent base-load engine design parameters.
Detailed estimates of the performance, size, and cost characteristics are presented
for advanced simple-, regenerative- and compound-cycle gas turbine engines.
Conceptual designs of selected engine configurations which are Judged to have the
greatest technical and economic potential for providing minimum power costs  and
an engineering layout of a possible future large central power station utilizing
open-cycle gas turbines are included.

     The performance, cost, and size estimates of advanced gas turbines were made
for engines which could be in commercial operation during three time periods —
the 1970-decade and the early and late periods of the 1980 decade.  The estimates
are based on the assumption that the substantial gas turbine technology developed
for aircraft and aerospace application will be transferred unhindered to various
industrial applications, including those in the electric utility industry.
Natural gas has been selected as the basic gas turbine fuel for convenience  in the
study; however, a number of other gaseous and liquid distillate fuels are also
suitable for use in gas turbines.

     The advanced gas turbine power systems considered in this study would not
contribute to thermal pollution of rivers and lakes.  The total heat load that
must te rejected from gas turbine systems includes: (l) the waste heat in the
exhaust gases, which accounts for over 90? of the total heat rejection load;
(2) heat generated in the bearings, seals, etc.; and (3) heat of compression in
the compressor discharge air bled for purposes of turbine cooling; removal of this
heat, as shown later in this report, improves power plant performance considerably.
The waste heat in the exhaust gases is rapidly dispersed into the atmosphere.  The
bearing heat load is rejected to a circulating oil cooling system and the compressor
bleed air heat load may be rejected to a circulating water cooling system; in both
cases final heat rejection to the atmosphere is achieved via coolant-bo-air  heat
exchangers.  The elimination of large supplies of cooling water as a siting

-------
 criterion  and  design  factor for gas turbine power systems can also provide
 additional transmission and reserve margin savings, the details of which will be
 presented  in Section  VIII of this report.
                     DESCRIPTION OF BASIC THEEMODYNAMIC CYCLES

     A brief review of the thermodynamic cycles used in gas turbine power
 generation  systems is presented in this section to provide background concerning
 the major system components and arrangements as well as the primary operating
 parameters  and to establish a framework for the discussion and presentation of
 results which follow.
                                   Simple Cycle

     The Brayton cycle is the basis of gas turbine engine power generation systems.
Ideally it consists of isentropic compression, constant pressure heating,  isen-
tropic expansion, and constant pressure cooling.  Thus only a compressor,  combus-
tion chamber and turbine are needed to achieve the various processes and this
arrangement is commonly called the simple cycle.  The processes for the ideal
simple cycle are depicted on a temperature-entropy (t-s) diagram in Fig. l6a.
A flow diagram for a simple-cycle engine is shown in Fig. ITa and representative
performance given in Fig. l8a.  The difference between the shaded area in  Fig. l6a
(heat input in combustion chamber) and the cross-hatched area (heat rejected in
the turbine exhaust) represents the net work output.   Of the total shaft work
developed by the turbine, approximately one-half to two-thirds is used to  drive
the compressor and the remainder to drive the load, i.e., the electric generator.
The electrical generator may be mechanically coupled to the same shaft as  the com-
pressor turbine (single-shaft version) or it may be driven by a free power turbine
on a separate shaft"(two-shaft version) aerodynamically coupled to the compressor
turbine as shown in Fig.  ITa.  The combination of the compressor, combustor, and
compressor turbine, in a free power turbine configuration, is commonly called the
gas generator.

     In the basic open-cycle configuration, the working fluid (air and combustion
products) passes through the gas generator only once; modifications to this cycle
include the closed-cycle and semi-closed-cycle configurations.  In the closed-
cycle configuration the working fluid, which is usually selected to minimize high-
temperature oxidation problems, is continuously recycled.  Heat addition is
accomplished by heat transfer through the walls of a heat exchanger, to which heat
from an external source (such as a furnace or nuclear reactor) is supplied.  In

-------
the  semiclosed-cycle configuration, approximately two-thirds  of the working  fluid
is recirculated.  This type of gas turbine requires a precooler for the recirculated
gas and a "charging" compressor to provide the necessary air  for combustion.   The
closed-cycle and  semiclosed-cycle configurations  were not  investigated in this
study since these configurations are generally not considered competitive  for
fossil-fuel plants and thus are not being actively pursued  in the US at the present
time.

     To achieve high thermal efficiency and minimum size with the simple open-
cycle gas turbine, high values of compressor pressure ratio and turbine inlet gas
temperature are required together with efficient components,  i.e., compressor,
turbine, and combustor.  Because of the importance of these parameters  to  the
success of base-load gas turbines for power generation applications, progress
made during the last two decades in achieving high compressor ratios and turbine
inlet temperatures, together with that projected for the next two decades, is
considered in detail in subsequent sections of this report.
                               Regenerative Cycle

     Because materials capable of withstanding high operating  temperatures  were
not available up until the early 1950's,  and since  the  attainment  of high component
efficiencies required basic research programs beyond the financial capability of
private industry, designers of industrial gas turbines  began incorporating
changes to the simple cycle which would increase thermal efficiency.   One of the
changes often selected is the use of regeneration in which a portion of the heat
in the hot exhaust gases leaving the turbine is transferred through an exchanger to
raise the temperature of the compressor exit airflow prior to  combustion.  The
processes in a regenerative cycle are shown on a t-s diagram in Fig.  l6b.  The
heat addition from the turbine exhaust gases (see Fig.  IJb) results in an increase
in the average temperature of heat addition to the  cycle and the net effect is a
marked improvement in thermal efficiency as illustrated in Fig. l8b.*

     The Fig. l8b results also indicate that for a  given turbine inlet temperature,
the maximum value of thermal efficiency (nth) is generally reached at a lower
compressor pressure ratio in a regenerative-cycle engine when  compared to t-he
simple-cycle engine.  However, if turbine inlet temperature is increased, the
maximum value of r\ ,  will occur at higher compressor pressure  ratios.
                  "Czl
   The performance data shown in Fig.18 are based upon sinplying assumptions and
   are intended to illustrate broad trends rather than specific  levels.

-------
                                Interceded Cycle

     In the intercooled cycle, the gases in the compression stage axe cooled
between one or more stages of the compressor and thus the work of compression is
reduced.  As shovn on the t-s diagram in Fig. l6c, the intercooling step tends to
lover the average exhaust temperature.  An arrangement of components in an inter-
cooled cycle is shovn in Fig. ITc, and the effect on the thermal efficiency for
engines designed vith relatively high compressor pressure ratios is shovn in
Fig. l8c.  Hovever, the main effect is to increase the specific pover of the engine
and thus reduce the physical size and cost of basic equipment, but these benefits
are partly offset by the added cost of the intercooler.
                                  Reheat Cycle

     In the reheat cycle, combustion gases vhich are partially expanded in a
turbine to an intermediate pressure, are then reheated in a secondary cozabustor,
and finally expanded in a second turbine to atmospheric pressure.  Reheat raises
the average temperature of heat addition to the cycle, as shovn in Fig. l6d,
thereby resulting in a higher output per Ib of air.  A typical arrangement of
components in a reheat cycle is shovn in Fig. 17d.  Reheat alone does not improve
the thermal efficiency of gas turbines; in fact, as shovn in Fig. l8d, the use of
reheat can result in a slight reduction in r\.,  of lov-to-moderate pressure ratio
engines.  A significant improvement in r\.,  can be achieved, hovever, vhen reheat
is used in conjunction vith regeneration and/or intercooling in a compound cycle
(see Fig. l8e).
                                 Compound Cycle

     The compound cycle, shovn schematically in Fig. I7e, can incorporate all of
the foregoing features to achieve maximum efficiency as depicted in Fig.  l8e.  The
t-s diagram for this cycle is shovn in Fig. l6e.  The shaded area under c-3-d-3
represents fuel-energy input, the hatched area under a-b represents heat  rejected
in the intercooler, and the hatched area under f-1 represents heat rejected in the
regenerator exhaust gas.

     The study presented herein has focused on the simple-, regenerative- and
compound-cycles; the compound cycle utilizes only reheat and intercooling.

-------
                         GAS TURBINE DESIGN CONSIDERATIONS

     Although the same basic components, namely,  a compressor,  combustion chamber,
and turbine, are found in both the aircraft and industrial gas  turbines,  their
engineering development has been based on distinctly different  design philosophies
for each of the respective applications.

     Basic criteria for Jet engines are compact size,  light weight  and low fuel
consumption for a given thrust output.  Also required are such  features as wide
operating range, fast acceleration, low thermal stresses, and above all,  high
reliability.  Consequently, the development of aircraft  gas turbines has  generally
been centered about the simple open-cycle engine, and has involved  highly
sophisticated and costly programs to develop engines with the (l) highest flow
rates and highest work transfer capabilities in each rotating stage, (2)  lightest
weight materials capable of operating at the highest possible stresses and
temperatures without compromising reliability, and (3) highest  component
efficiencies.

     Some effort has been expanded by the aircraft industry,  in particular the
Allison Division of General Motors, Pratt & Whitney Aircraft, and the Garrett
Corp.. on the development of regenerative-cycle aircraft engines for long-
duration missions where minimum fuel consumption  was of  particular  importance;
however, many problems such as coinbustion-product-deposit accumulation on the heat
exchanger surfaces, excessive pressure losses, high costs, and  relatively heavy
regenerator weights have hindered these programs.

     A review of industrial gas turbine designs built over the  last 20 years  shows
that they tend to oscillate between complex reheat, intercooled, and/or regenerative
cycles offering relatively high thermal efficiencies,  and simple designs  which
stress reliability.  The continued re-emergence of the simple-cycle gas turbine
indicates that, despite the effort expended, the  industry has failed to provide a
completely satisfactory high-efficiency prime mover.  Much of the trouble
encountered in the development of the more complicated cycles for relatively  large
gas turbines has been due to miscalculations of pressure losses, particularly in
ducting and elbows.  The effect of these losses  on performance in  future high-
pressure-ratio (50  to 100:l) designs could be less pronounced  due  to the high
absolute levels of operating pressures.

     However, articles such as those appearing in Refs.  68 through  71 continue to
promote the use of compound cycles in various applications ranging  from moderately
large central power stations to land transportation.  A  37-Mw compound open-cycle
gas turbine power station was constructed by the  Fiat Company of Italy; industrial
operation of the unit started in September 1962.   The system  utilizes two
compressors with interceding and two turbines with reheat (Ref. 69).  In order
to allow for the combustion of cheap residual fuels, both combustion chambers

-------
consist of single-can combustors arranged remotely from the compressor turbine
assembly.  A similar unit, rated at 100-Mv output, was built and rig-tested in
Russia in 1967  (Ref. 68).  The principal difference between this unit and the
Fiat design is  the use of multiple-can straight-through combustors instead of the
remotely located single-can configuration.  The 100-Mw unit was reportedly designed
to achieve a pressure ratio of 26:1 even though previous Russian studies has
indicated that  a pressure ratio over 100:1 would be advantageous.

     Most industrial gas turbines follow a set design pattern; a compressor is
followed by either a single-shaft or two-shaft turbine, with the simplest bearing
arrangement.  Single-shaft machines are more often used for constant-speed
applications such as electric power generation.  Two-shaft machines are more
suitable for variable-speed service such as gas or air compressors and pumps.
A single large  combustor, favored in European designs, or multiple, symmetrical,
can combustion  chambers, favored in the United States designs, provide heat
addition to the engine.  Axial-flow turbomachinery is used nearly exclusively in
all large gas turbine designs.  Centrifugal turbcmachinery has been used on a few
small units (with ratings up to about 3000 hp), although not extensively.

     It should be noted that increased turbine inlet temperatures above the 1500-
to 1600-F level presently used in industrial gas turbines represent one of the
design advancements available from materials technology in aircraft-type gas
turbine programs that will lead to lower fuel consumption per unit of shaft work
output.
              PROJECTED ADVANCES IN GAS TURBINE COMPONENT TECHNOLOGY

     The application of gas turbine engines to provide the mid-range and base-
load requirements of the electric utility industry has been limited because of
the modest thermal efficiency levels attainable in comparison with conventional
steam power systems.  The present limited output capability per engine and the
requirement for relatively clean-burning and, therefore, moderately expensive
fuels (Refs. 68 through 77) have also severely limited nonaircraft gas turbine
application.  Further significant improvements in specific power (hp per Ib/sec
of engine airflow) and thermal efficiency for gas turbine power plants can be
achieved only by increasing turbine inlet temperature and compressor pressure
ratio, since turbomachinery component efficiencies have reached a relatively high
level due to over 30 years of extensive research and development on compressors
and turbines for aircraft propulsion, and only small further gains may be anticipate
                                       50

-------
     Recent advances in turbomachinery,  materials technology,  aerodynamic design,
heat transfer, and fabrication techniques for gas turbines  have resulted from
the requirement for nev military and commercial aircraft (Refs. 78 through 85)
as veil as various industrial applications (Refs. 86 and 87).   These technological
developments provide the basis for substantially improved gas  turbines  with
significantly lower fuel consumption rates than presently attainable.   Pertinent
aspects of these technological advances, including their effect on gas  turbines
designed for electric power generation,  and projections  of  further advances
anticipated during the 1980's are discussed in the following paragraphs.
                                    Compressors

Performance Parameters

     Technological advances in aircraft gas turbine compressor design have resulted
in substantial improvements in cycle pressure ratios,  increasing from about U:l
in the early English (National Gas Turbine Establishment)  centrifugal compressors
of the 19^0's to 2U:1 in present twin-spool axial-flow compressors  (see Fig.  19).
A spool is commonly considered to be a compressor and  connected turbine.   The
compressor or turbine may have one or more stages.   Thus,  twin-spool compressors
consist of two tandem compressors with their respective turbines that form two
rotor systems which are mechanically independent but related aerodynamically.
These configurations facilitate proper compressor-stage matching for relatively
high pressure ratio designs as an alternative solution to  single-spool compressors
that would have to incorporate variable geometry stator vanes.  The technology
is becoming available that will permit the introduction of machines with pressure
ratios exceeding 30:1, a prerequisite for high-efficiency  simple-cycle base-load
gas turbine engines.  Proportionate increases in the number of axial compressor
stages to meet these pressure ratio requirements have  been avoided through
intensive research and development efforts to increase stage pressure ratios.
For example, it is presently possible to achieve single-stage pressure ratios of
1.2 to l.k while still maintaining stage efficiencies  of 90% or more as shown in
Fig. 20.  These stage performance levels have been extended to multi-stage aircraft
compressor designs and permit the attainment of average stage pressure ratios on
the order of 1.2 to 1.3 with associated polytropic efficiencies of approximately
90$ or greater (Ref. 70).  These trends are clearly indicated by the statistical
data shown in Fig. 20 for compressors which have been  built or are in the study or
development stage.  Further improvements in polytropic efficiency to peak values
of 91 to 93$ within the next decade appear feasible (Refs. 70, 88,  and 89).

     Principal factors which influence stage pressure  ratio are rotor tip speed
and aerodynamic loading.  Rotor tip speeds of 1100 to  1200 ft/sec are attainable
with current advanced lightweight compressor designs (Ref. 70) and values as high
as ll+OO ft/sec are forecast for the early 1980's (see  Fig. 20).  Although light-
weight fan stages are in operation at tip speeds of 1500 ft/sec and above, high
                                       51

-------
 stress  limits  and  supersonic Mach number operation will limit compressors to the
 aforementioned levels.  The aerodynamic loading, at a given tip speed, establishes
 stage pressure ratio.   A measure of aerodynamic loading is the airfoil diffusion
 factor  (Df), a parameter vhich  includes a  factor reflecting the overall change in
 relative  velocity  across the blade  row and a term proportional to the conventional
 lift coefficient.   High values  of Df  are desirable since they result in a reduction
 in  the  number  of compressor stages; values  of O.Uo and 0.^5 are currently obtainable
 The present limit  in Df is due  primarily to excessive secondary flov or endwall
 losses  and to  compressor surge  margins that are dangerously low.  Further improve-
 ments in  tip speed and  diffusion factor will result from aircraft compressor
 research  and development programs .

     The  specific  flow  (iVsec  per unit flow area) of axial aircraft compressors
 varies  from approximately 33 to ^0 lb/sec/ft2 of frontal area.  Increases in this
 parameter result in more compact designs (smaller cross-sectional area); however
 for a given wheel  speed (Um) and work coefficient, higher specific flows and,
 alternatively, higher axial velocities (Cx) reflect high flow coefficients
        and correspondingly lower compressor efficiency.
     Adaptation of these developments to industrial units has been slow because
of the lack of incentives to undertake parallel costly development programs for
long-life stationary gas turbine power plants.  Consequently, present industrial
gas turbine axial compressors incorporate moderate blade loadings and tip speeds
of only 650 to 750 ft/sec, and achieve average stage pressure ratios only on the
order of 1.1 to 1.15 (Refs. 71*, 86, and 90 ).  As a result, as many as 15 to 1?
compressor stages are required to produce cycle pressure ratios above 10:1.  Thus,
high-thermal-efficiency performance cannot be  easily achieved with these designs.
European industrial units appear to favor the use of large, massive, multi-unit
compressors incorporating stage intercoolers (Refs. 68, 69, 75, and 91 ) to achieve
cycle pressure ratios on the order of 15:1 to 25:1'

Construction Design Features

     Future large industrial compressor units will have to incorporate low-cost
materials and inexpensive construction techniques without sacrificing reliability,
life, and performance.  Although typical lightweight aircraft compressors are
designed as shown in Fig. 21a, with the rotors formed by rows of 'blades rooted in
structural disks, other configurations may have to be examined in detail (see
Figs. 21b and 22c).  In aircraft designs,, the compressor stages are jointed and
positioned axially by cylindrical inner spacers and conical spacers.  The inner
wall of the flow path is formed partly by the disk, rims, and blade root platform,
and partly by the inner shroud of the stators.  The outer wall is formed by the
outer shroud of the stators.  Present industrial compressor designs incorporate
disk-drum rotor assembly (Fig. 21b), or a drum rotor assembly (Fig. 21c) construct:
In the disk-drum arrangement, the rotor for each stage consists of a solid disk.
                                        52

-------
Through-bolts connect the rotor disks to the forward and aft  subshafts  to form
the disk-drum assembly.  The bolt circle is placed at the maximum diameter
possible within the confines of the rotor design to ensure a  strong,  stiff rotor
assembly.  This rotor configuration eliminates the need for labyrinth seals at
the blade root section between compressor stages since a close-fitting  hub is
formed by the adjacent riins of the stacked disk assembly.

     The drum rotor configuration affords perhaps the simplest  type of structure.
Individual blades are attached to the outer surface of a cylindrical  cast drum
structure thus avoiding the necessity of using individually forged compressor
disks.  A constant hub diameter design also offers an estimated 15-to-20/£
reduction in blade cost, relative to a tapered hub configuration, as  a result  of
the simplified blade root pedestal machining procedure.   Maximum rotational
speeds with the drum rotor construction are limited, however, because high rim
speeds must be avoided in order not to exceed maximum allowable stress  limits.

Materials

     Continuing efforts to eliminate excessive weight in aircraft engines have
pushed the limits of materials properties.   Until recently, the operating
conditions in the compressor section were such that titanium  alloys (Ti-6Al-Uv)
and low-alloy and stainless steels (Types 17-22A, Greek Ascoloy stainless AMS56l6,
and AISlUlO) satisfactorily fulfilled all design requirements.   However,  with  the
development of high-pressure ratio, high-airflow handling turbofans,  exhibiting
rim temperatures in excess of 800 F, new superalloys such as  A-286, Incoloy 901,
and Inconel 718 were developed for disks as well as blades in the high-pressure
sections of these advanced designs (Ref. 9l)> and most of these materials could
be  easily used in advanced industrial-type engines.  A complete listing  of materials
used for compressor parts in existing and new designs is presented in Ref. 91.

     High-pressure ratio advanced-design industrial compressors should, therefore,
continue to utilize the martensitic corrosion-resistant AISI  Type UlO steel and
low-alloy (AISI^S^O) steel extensively for airfoils and disks,  respectively,
in the low-temperature regions (rim temperatures below 800 F).   However,  aircraft-
proven nickel-base alloys such as Incoloy 901, Inconel 7l8, and A-286 will be
necessary in the high-temperature sections (800-to-1200 F).  The low-alloy-steels
are usually protected from rusting by a 0.5-mil coating of nickel cadmium (Ref.  92).

     The next generation of aircraft compressors will see an  increase in  the use
of composite parts (Refs. 91 and 92); among the most promising  are silicon
carbide-coated boron fiber and carbon fiber.  The composite materials have not
yet been proven ready to meet commercial airline reliability  requirements as far
as erosion resistance and foreign-object damage are concerned.   However,  these
composites may permit the inexpensive fabrication of large parts without  the need
for new and expensive machine tools; consequently, they show  the potential of
significant future cost savings, particularly in the manufacture of very  large-
capacity gas turbines.
                                        53

-------
                                     Combustors

      The primary characteristics  of combustors which affect gas turbine performance
 are combustion efficiency,  pressure drop,  and uniformity of outlet gas temperature.
 Whereas  combustion efficiency affects  only fuel  consumption, pressure drop  affects
 both fuel consumption and power output.

 Performance Parameters

      Combustion efficiencies  of 100$ are currently  achieved in aircraft gas
 turbine  combustors and similar performance can also be  achieved in advanced
 industrial combustors burning clean gaseous  fuels.   The outstanding achievement
 in  combustion technology has  been in the reduction  in pressure drop, defined as
 the difference in total pressure  between the compressor exit and turbine inlet.
 Pressure drops in aircraft  combustion  designs have  been reduced from about 10$
 in  19^5  to under 5% in current advanced designs.  Of particular importance is the
 fact that these reductions  in pressure drop  have been accompanied by improvements
 in  combustor outlet temperature distribution (Ref.  89).   Combustor pressure drop
 is  also  related to reference  velocity, defined as the theoretical velocity for
 flow of  combustor inlet air thorugh an area  equal to the maximum cross section of
 the combustor casing.   Reference  velocities  of 80 to 150 ft/sec, with corresponding
 combustor liner pressure losses of 2 to 5$>  are common  in present aircraft
 designs;  20 to 80 ft/sec with correspondingly lower pressure losses will be found
 in  the industrial external  combustor and multiple-can designs.

      The heat release  rate  within the  combustor also affects engine size and is
 especially important  in aircraft  gas turbines wherein heat release rates of from
 2 to  k million Btu/hr-ft3-atm are common and rates  as high as 8.5 million Btu/hr-ft
 atm have  been used (Ref.  93).   Heat  release  rates for industrial gas turbines are
 substantially lower"since these units  are  commonly  designed to burn a wide variety
 of  fuel  types  (Ref. 93).

      The  combustor must be  designed to avoid exit gas temperature peaks since a
 large temperature  gradient  reduces  the allowable average gas temperature into the
 turbine  and thus  limits  engine  output  and  efficiency.   In addition, turbine vanes
 are exposed directly to  local  gas  temperatures, and high peak gas temperature will
 reduce vane life.  Continuous  efforts  to reduce the ratio of the difference betweer
 the peak  local-to-average gas  temperature  and the average combustion temperature
 rise have  resulted in  typical values for this parameter  of between 0.05 and 0.15
 for vanes.   These  techniques  can  be  utilized in gas  turbines designed for electric
power generation  systems  and will enhance  their performance and operating
 characteristics.

-------
Construction Design Features

     Combustor design configuration and shape depend largely upon the intended
applications.  Aircraft gas turbines, for instance, specify compactness and
lightveight as the primary design criteria.  These designs, therefore, will
utilize either a single annular combustor, a number of tubular combustors (multiple-
cans), or a can-annular arrangement vith a number of tubular liners vithin an
annular casing.  The combustors may be placed for straight-through flov between
the compressor and the turbine, or designed for reverse flow and wrapped around the
compressor olr turbine.  Heavy-duty residual-oil burning industrial turbines are
usually designed for ready access and removal of combustor parts, and may use
either multiple-can combustors discharging directly into the turbine vanes (Refs.
68, 7!*, 76, and 90) or one or two external combustors firing into a large turbine
nozzle box (Refs. 69, 71, and 77).  However, as turbine inlet gas temperatures
in advanced-design industrial gas turbines are raised to higher levels to achieve
improved thermal efficiency, existing differences between the design philosophies
of industrial- and aircraft-type combustors will diminish considerably.  For
instance, the external combustors and multiple-can combustor construction currently
used in industrial designs operating with turbine inlet temperatures of approximately
1300 F to 1TOO F will have to be replaced with annular configurations in power
plants designed to operate at turbine inlet temperatures much above 2000 F.  These
changes will be necessary to avoid excessive liner cooling flow rates, high pressure
losses, and high local peak temperatures on first-stage turbine vanes.

     Conventional louvered liners, used in industrial and production aircraft
coabustors, will be permissible for use at turbine inlet gas temperatures below
approximately 2^00 F; at 2^00 F and above, advanced-design liners with high heat
transfer surfaces developed for aircraft combustor designs will also be specified
for the industrial counterpart.  Length-to-height ratio for combustors has ranged
from 2.7 to 5.0, but generally varies between 3 and 3.5.  In the future, however,
short burners will be favored since recent experiments indicate that in such
designs the formation of nitrogen oxides seems to be inhibited.  Discussions of the
nitrogen oxide emissions from gas turbine power plants and research efforts under
way to inhibit the formation of this pollutant are presented in Appendix A.  The
ability to shorten burners is dependent upon the fuel distribution system as well
as the linear hole pattern and liner cooling scheme.

Materials

     Alloys used for aircraft combustors and liners and for the transition ducts
connecting the combustors to the turbine inlet nozzle must be  formable, weldable
sheet alloys capable of at least 8000 to 10,000 hr of operation in an oxidizing
environment at average metal temperatures of 1650 F and above  (Ref. 9l).  These
sheet alloys must be stable for this long-time, high-temperature service and have
optimum erosion-corrosion, thermal fatigue, and distortion resistance.  Liners are
always air-film cooled and are often coated for additional protection.  Hastelloy X
                                       55

-------
 is by  far the most  commonly used alloy for aircraft combustion liner applications,
 vhereas  AISI Type 321  and  310, oxidation-resistant austenitic stainless steels,
 have been -used  in the  relatively low-temperature industrial units  (Eef. 91).
 Haynes Development  Alloy No. 188,  superior to Hastelloy X in high-strength ductility
 and oxidation resistance,  is nov videly used by aircraft engine builders for
 advanced-design combustor  applications.  Materials presently under development for
 the next generation of aircraft engine combustor designs include coated TD-Ni,
 TD-NiCr, and dispersion-strengthened materials (thoria dispersed in nickel).
 These  materials will have  substantially higher service temperatures than present
 materials (see  Table XV),  and it is foreseeable that many of these nev materials
 would  permit long-life operation of 2U,000 hr and above,
                                      Turbine

     Increases in turbine inlet operating temperatures through the use of improved
materials  and cooling techniques represent the principal means of improving the
performance  and increasing the output potential of future gas turbine designs.
Although historically most of this effort was directed primarily toward aircraft
propulsion systems, some of the advances were eventually used to improve industrial
designs.   This trend is evident by an inspection of Fig. 22 which shows the pro-
gression of  maximum operating cycle temperatures with time for gas turbines designed
for various  aircraft, industrial, and electric power generation applications.

     The data indicate that turbine inlet temperatures up to 1800 F are being used
in gas turbine electric power plants, and several hundred degrees higher in commercii
and military aircraft.  Prior to the mid-1960's, advances in materials technology
traditionally accounted for a respectable 20~to-l*0 F increase per year in turbine
inlet temperature (Refs. 78 and 89).  Recently, however, significant increases in
turbine inlet temperature (Refs. 70, 75, 78, 79, 83, and 8U) approaching 70 to 80 F
per year,  have been achieved through substantial improvements in turbine cooling
techniques in combination with newer materials.  With the expected continuation of
current trends, turbine inlet temperatures up to 2^00-to-260Q F may be common in
military aircraft engines before the end of the 1970's and should be available in
industrial engines early in the 1980 decade.  Turbine inlet gas temperatures as
high as 1900 F for continuous cruise operation, and 2100 F for take-off, are now
experienced by the engines utilized in today's 7^7 and C5A Jumbo jets.  AiResearch
Division of  Garrett Corp. has made extensive studies of simple- and regenerative-
cycle gas turbines capable of operating at a turbine inlet temperature of 2300 F.
Turbine inlet gas temperatures as high as 2^*00 to 3000 F in base-load industrial
gas turbines have been projected for the next decade by manufacturers of industrial
units (Refs.  9^ and 95).
                                        56

-------
Performance Parameters,

     Improvements in turbine performance resulting from technological advances
in aircraft turbine designs have paralleled those for the compressor, with turbine
isentropic efficiency increasing from about &5% in early-1950 designs to over 90%
in current designs; efficiencies of 92 to 93% are feasible in the near future,
particularly in high-output designs (Refs. 70 and 89).  The design philosophy of
turbines for industrial engines vill also approach that of aircraft engine turbines.
For instance, future industrial turbine designs will incorporate a higher degree
of impulse staging than exists in most present designs to achieve a relatively
large gas temperature drop in the nozzle, thus avoiding high gas temperatures with
corresponding high cooling flows in the rotor (Ref. 75).  The stage work for aircraft
turbines varies from a low of 20 Btu/lb in the last stage of low-pressure ratio
units to a maximum of about 120 Btu/lb in the high-pressure ratio stages of current
production units; technology presently exists to increase this value to about
155 Btu/lb without sacrificing turbine efficiency.

     High stage work in itself does not cause aerodynamic problems, but when
associated with high efficiency it requires high wheel speeds which become
limited by structural considerations.  Turbine tip speeds in current twin-spool
aircraft designs vary from 900 to 1000 ft/sec in the low-pressure ratio stages
to as high as 1500 to 1600 ft/sec in the high-pressure ratio stages.  These
tip speeds can be achieved in future industrial-type designs that utilize the
improved turbine materials and cooling techniques described in the following
paragraphs.

Materials

     The improvements in the temperature capability of turbine materials, with
time, for aircraft gas turbines is shown in Fig. 23a.  The gains have been most
significant for the nickel-base alloys, with an improvement of more than 300 F
in temperature capability over the past decade (Eef. 78).  Presently, nickel-
base cast blade alloys, such as Inconel 7l8, B1900, and IN-100, have in some
aircraft turbine designs replaced forged Udimet 700 (u-700) for high-temperature
applications and, at the same time, decreased thermal fatigue failures by a factor
of five (Ref. 78).

     Figure 23b shows the stress-to-rupture strength of the nickel-base alloy
33-100 for a range of material temperatures.  Curves for 1000-, 10,000-, and
100,000-hr lifetimes are shown and illustrate the sharp reduction in allowable
stress as the desired operating time is increased.  The 1000-hr-life curve is
based on experimental data while the 10,000- and 100,000-hr curves are extra-
polated data from short-duration test results.  Although these materials were
developed initially for relatively short-time (1000-hr) high-strength aircraft
                                        57

-------
 turbine applications, modifications to the heat treatment cycle have permitted
 their  use for long-time  operation  (approaching 100,000-hr) as required in base-
 load pover generation.   It is  anticipated that similar modifications for the
 aforementioned high-temperature nickel-base alloys vlll allow the eventual
 replacement of such  alloys as  wrought Udimet  500 and M252 presently used in
 industrial turbines  and  thus permit high operating temperature.

     Turbine blade materials for industrial designs anticipated by the 1980's
 will be the superalloys  currently  under development for advanced aircraft gas
 turbines;  these include  unidirectionally solidified eutectic alloys such as
 Ni3Al-Ni3Cb (Ref. 85) and particle or fiber dispersion-strengthened metals (Refs.
 91  and 96).   On the  basis of efforts currently underway it is reasonable to
 expect a continuation of at least  a 20 F per year improvement in material
 temperature.   Chromium-base alloys, for instance, with potential firing temperature
 to  2000 F,  are being investigated.  Recent breakthroughs in the ability to coat
 columbiurn-base alloys also offer a possibility for the use of these alloys,
 currently being investigated for advanced aircraft propulsion systems.  Ceramic
 materials  such as silicon-nitride  and silicon-carbide composites are under
 intensive  study for  automotive gas turbine applications and offer firing
 temperatures  on the  order of 2500  and 2600 F  (Refs. 97 and 98).

     A brief summary of  the projected creep strength properties of present-day
 and advanced turbine blade materials which could be used for future designs is
 presented in  Fig. 2k.  The band in Fig. 2k labeled "present materials" represents
 properties  of cobalt-base and  nickel-base alloys that are now in use in industrial
 and aircraft  gas turbines.  The other bands represent estimated properties of
 advanced nickel-, chromium-, and columbiurn-base alloys and combinations currently
 in  advanced  stages of development  for aircraft engines, which will eventually be
 adapted to base-load applications.  The projected creep strength characteristics
 of  these materials are based on applying Larson-Miller parameter extrapolations
 from available  short-term property data for these materials.

     The first-stage nozzle vanes are the hottest parts in the gas turbine and
 are also subjected to high thermal shock stresses.  These vanes are usually coated
 for improved  oxidation resistance.  Since they are stationary, they are not highly
 stressed, and cast nickel- and cobalt-base alloys such as Inconel 7l8 and 738 and
 WI-52, which have been used as vane materials in aircraft gas turbines, would be
 satisfactory  for industrial applications.

 Coatings

     Coatings have been developed which provide adequate oxidation-corrosion
resistance for blades and vanes in aircraft turbines.   These coatings are
aTuminide types, applied by pack or slurry techniques.   These coatings have
eliminated intergranular oxidation attack (which could cause a thermal fatigue

-------
failure) and have provided some protection (Ref,  92)  against sulfidation attack
(corrosion due to the presence of sodium sulfate  compounds in the combustion
products of distillate fuels).  However, coating  life decreases  rapidly with
increases in allowable metal temperatures.  For instance, at a metal temperature
of approximately 1^50 F, a coating life of 50,000 hr  may "be achieved.   A 200 F
increase in metal temperature reduces coating life to less than  10,000 hr.
Consequently, coating life rather than creep strength properties as determined
by peak local temperatures in the first-stage turbine vanes will be an important
criterion used to specify turbine inlet gas temperature for different  power plant
operating modes.  Adequate coating life could be  especially critical when the
fuel used in the engine contains quantities of ash, metals, or sulfur  that could
cause erosion and corrosion problems.  Natural gas, in this respect, is an ideal
fuel for high-performance gas turbines.  Meanwhile, new coating  materials and
application techniques including nondiffusion-type coatings are  being  developed
which, when applied in sufficient thicknesses, should provide uninterrupted service
for periods exceeding 25,000 hr.

Disks

     Turbine disks in new, industrial-type gas turbines operating at 1^00 to 1700 F
are maintained at metal temperatures of 600 to 750 F  (Ref. 86) by the  utilization
of cooling air extracted from the compressor (see Fig. 25a).  As a result of these
moderate disk temperatures, fairly inexpensive materials such as austenitic
stainless steels (A-286) are used.  In production aircraft gas turbines, disks
are fabricated from nickel-based alloys such as Inconel 7l8, Incoloy 901, and
Astroloy, and are capable of withstanding metal temperatures of  900 to 1UOO F
(Refs. 91 and 92).  It is anticipated that these  materials would be used also in
advanced gas turbine power systems.

Turbine Cooling Techniques

     Only first-stage vanes and disks of current  advanced-design industrial gas
turbines, operating at 1600- to 1700-F turbine inlet  gas temperatures, are cooled.
For long-life, base-load operation at turbine inlet temperatures of 1800 F and
above, successive stages of turbine blades will also  require cooling (Fig. 25b).
Turbine cooling can be accomplished with coolants such as air, water,  or liquid
metals, but because of the complex cooling system designs and mechanical problems
associated with liquid systems, air has been used exclusively as the coolant in
all aircraft propulsion systems and for most stationary applications.

     Turbine cooling systems in current aircraft  engines have progressed from
simple convective cooling configurations incorporating cast, round, radial
passages, for vanes, and single cavities  for blades  to advanced convective-heat
                                        59

-------
 transfer designs utilizing impingement cooling of the inside surface of the leading
 edge  (Fig. 26).  In film-cooled designs, a layer of coolant is injected in hollow
 "blades through radial slots to form an insulating air blanket for the outside
 blade surface.  These designs are in the advanced stages of development for the
 next generation of aircraft engines.  In transpiration-^cooled blades, coolant is
 bled through a porous material vhich may be formed by a series of small drilled
 holes along the airfoil surface.  These designs are also in the early stages of
 development and, when applied to 1980-time period aircraft turbines, will offer
 operation at turbine inlet temperatures approaching 3000 F.  Figure 27 summarizes
 the progress that has been made in the different aforementioned turbine blade
 cooling system for commercial aircraft engines.  The temperature values recorded
 on the figure for each type of cooling scheme reflect levels that have been
 demonstrated by actual tests and which have been accepted or appear feasible for
 commercial aircraft propulsion system applications.

     In aircraft engines, the air used to cool the turbine blades and vanes is
 at an elevated temperature, 800 to 1200 F, even before it is introduced into, the
 turbine, since it is bled from the compressor discharge airstream.  Precooling
 the compressor bleed air to fairly low levels before it is used to cool the
 turbine has not been a general practice in these aircraft applications because
 the added cooling system weight detracts from the potential gains in performance
 that might otherwise be achieved by reducing the turbine cooling flow.  However,
 as turbine inlet temperatures and compressor pressure ratios continue to rise and
 approach the limits established for proven turbine blade cooling techniques,
 lightweight low-power, compressor bleed air cooling systems will have to be devised.

     For stationary industrial power plants, air-, water-, or even possibly fuel-
 cooled heat exchangers could be used to reduce the temperature of compressor
 discharge cooling air to fairly low levels (perhaps to 100 to 200 F) since in
 these stationary applications weight is not an important criteria.  Therefore,
with this technique it should be possible to achieve the gas temperature levels
 indicated in Fig. 27 on a continuous basis in base-load plants with presently
 available combinations of impingement-conversion cooling techniques, provided
 the anticipated extended coating lifetimes with the previously mentioned coating
 improvements are achieved.
                                   Regenerators

     Research and development programs on regenerative heat exchangers have
concentrated on three different types of regenerator designs, the stationary
regenerator (commonly referred to as a recuperator), the rotary regenerator, and
the  liquid-coupled indirect-transfer regenerator.  The liquid-coupled indirect-
transfer type has been considered primarily for military aircraft applications.
However, several current military programs have been concerned with the recuperator
                                     60

-------
type.  Most rotary regenerator programs have been limited to automotive applications.
The basic characteristics of the three types of gas turbine regenerators and the
reasons for selecting the recuperator type for the intended base-load electric
pover application are described in the following paragraphs.

     In the liquid-coupled indirect-transfer regenerator concept, tvo separate
gas-to-liquid heat exchangers and a circulating liquid loop, coupling the two
heat exchangers, are employed.  In one heat exchanger, hot engine exhaust gases
heat the liquid which is circulated and subsequently cooled in the other exchanger
during the process of heating the compressor discharge air.  This design has the
advantage that the heat exchangers may be conveniently located, thereby requiring
the fewest changes in the normal gas turbine flow path.  Since the external
ducting handles only a high-density liquid, the frontal area is minimized, an
important factor in an aircraft application.  However, the liquid coupling loop
does add complexity and cost.  Furthermore, liquid metals which have the best
heat transfer characteristics require elaborate handling techniques and special
materials because of their corrosiveness and other undesirable characteristics.
Some liquid metals are toxic or combustible in air and therefore would create
serious problems if a leak occurred. Thus, this type of regenerative gas turbine
has not become operational even for military aircraft applications.  Even if some
of the technical problems of this type of regenerator were solved, it is unlikely
that its limited advantages would warrant its use in an industrial gas turbine
and thus was not considered further in this study.

     Rotary regenerators involve the direct heat transfer to and from a rotating
matrix which acts as a heat sink and source, as opposed to heat transfer from one
fluid to another through heat exchanger walls, and thus need not withstand large
pressure differences across tube walls.  Although many matrix designs and
materials have been considered in past research and development programs, the
porous ceramic core built in the form of a disk such  as the Cercor* regenerator,
is the only one used to any extent to date.  These ceramic rotary regenerators
are currently employed in automotive-type gas turbines.  The ceramic core is
relatively inexpensive and can operate at very high temperatures relative to
other materials; furthermore, it can be designed with a high effectiveness (90$
or higher) without serious penalties to cost, size, and gas turbine pressure loss.
However, the automotive applications of these regenerators have been limited to
engines of 300 to kOQ hp.  In discussions with the leading manufacturer of
ceramic regenerator cores (Ref. 99)»  it was learned that a UOO-hp engine requires
a. 28-in. diameter core and that 36 in. appears to be the largest size that can be
built with present technology.  This limitation is due to structural considerations
which also limits their use to gas turbines with compressore pressure ratios of
6.5 or less.  One of the problem areas associated with the rotary regenerator is
that of sealing the rotating matrix to prevent the higher-pressure compressor
discharge air from leaking directly into the turbine exhaust, thereby resulting in
a degradation of performance.  Obviously this problem would be worse at the higher
                                        61

-------
 compressor pressure  ratios  and  is  another  reason for the pressure limit previously
 mentioned.  Furthermore,  advances  in  technology in the next tventy years are not
 expected to significantly affect these limitations.  Thus it appears that ceramic
 rotary regenerators,  even in modular  form, vould not prove practical for the large
 industrial-type  regenerative gas turbines  considered in this study.  This conclusion
 is  also mentioned  in  Ref.100, vhich states that as the pover of the gas turbine
 exceeds 300 hp,  large pressure  loads  inherent with the regenerator introduce
 structural complexity that  tends to offset the low-cost advantages of the ceramic
 matrix.

      In the recuperator  (stationary regenerator), compressor discharge air and
 turbine exhaust  gases are ducted to a single heat exchanger in which heat is
 transferred directly  from one fluid to the other through the exchanger wall.
 Extensive work is  currently being  done to  design and develop lightweight regenerative
 aircraft gas turbines utilizing the recuperator type of regenerator (Ref. 88)
 and Harrison Radiator Division  of  General Motors Corporation has built 75
 recuperators since 1957 for use in industrial gas turbines.  Since recuperators
 tend  to be large (a recuperator for a current industrial gas turbine of
 approximately 20,000  hp output  capacity weighs approximately 50 tons installed)
 it  is  a practical  necessity to  construct several small modules and then manifold
 them  together into one unit.    Most of the industrial gas turbine recuperators in
 use are constructed with  plate-fin type surfaces made of mild steel and operate
 with  a maximum gas temperature  less than 1000 F.  However, aircraft gas turbine
 recuperators have  been built and operated successfully at significantly higher
 temperatures.  A number of  high-temperature recuperator materials are available,
 and others which are  under  investigation will provide the desired characteristics
 for the advanced regenerative-cycle gas turbines which, in general, operate at
 recuperator inlet  temperatures  above  1000 F.  A brief description of these materials
 follows.
                               Recuperator Materials

     The primary factors considered in the selection of recuperator base materials
are their mechanical properties, hot-corrosion resitance, fabricability,
compatibility with brazing alloys, metallurgical stability, and cost,

     A recuperator material must have adequate mechanical properties during its
design lifetime to withstand the stresses due to thermal transients and fluctuating
and steady-state pressure differentials.  Therefore, before selecting a material
the degradation of its mechanical properties due to environmental attack and by
metallurgical changes, such as aging reactions and carbide precipitation, must be
considered.  Fatigue, ultimate strength, and stress-to-rupture properties are all
                                        62

-------
important in designing a recuperator.  Erosion resistance is  also  important  because
the gas turbine environment usually contains  some solid or liquid  particles  (dust,
sand, or carbon).

     Hot corrosion, which is the attack on metal alloy components  caused  directly
or indirectly by contact with the gas turbine combustion products, includes  all
the effects that contribute to corrosion such as sulfidation,  oxidation,  erosion,
and stress corrosion.  This type of attack, even on stainless  steels  and  super-
alloys , occurs at temperatures likely to be encountered in advanced gas turbine
recuperators.

     Since brazing is the most economical means of producing  recuperator  structures,
compatibility of the base or structural materials with available brazing  filler
alloys is extremely important.  Base metal-filler metal combinations  must have
compatible brazing temperatures to prevent metallurgical changes in the base
metal as a result of the brazing temperature.cycle.  Adequate  vetting, brazed
joint strength, and hot corrosion resistance  are also required characteristics.
Another requirement is that there should be little or no embrittlement of the
base metal by diffusion of the brazing filler metal constituents into the base
metal.

     Materials selected for formed parts, such as fins and pans, should have
adequate formability at room temperature and must be amenable  to brazing, welding,
and other contemplated manufacturing operations to minimize fabricating costs.

     As previously noted, mild (carbon) steel has been used extensively in
industrial gas turbine recuperators and possesses adequate properties when
operated at temperatures up to approximately 1000 F (maximum  gas temperature).
This material is relatively inexpensive and is amenable to most low-cost  fabrication
techniques.  Above this temperature level, alloy steels and superalloys must be  used.

     Type 3^-7 stainless steel, which has been used in many recuperators,  provides
excellent performance at lower temperatures.   It is relatively inexpensive compared
to other high-temperature alloys.  It has good oxidation resistance at moderate
temperatures, relatively good corrosion resistance, and is easily  brazed.  This
alloy has proven successful in recuperator applications (nonindustrial) operating
below 1300 F (Eef. 101).  Type 3^7 stainless steel was specified for  the  recuperator
by a recuperator manufacturer (Ref. 102) for a typical IpSO-decade engine design
(turbine exhaust temperature of 1286 F), whereas a combination of  3^7 and carbon
steel was specified for the 1970-decade engine (turbine exhaust temperature of
1128 F).  Because of the relatively low cost of this alloy, another manufacturer
even specified 3^7 steel for the core of its industrial gas turbine recuperator
which would operate at turbine exhaust temperatures below 1000 F (Ref.  103).
                                         63

-------
      Type  1*30 stainless  steel  is  a lov-carbon, high-chromium ferritic steel
 possessing good resistance to  oxidation  and  corrosion at elevated temperatures.
 It is lover in cost  than Type  3^7 and has good room-temperature strength and
 ductility.  However, it  has one-half the yield strength of Type 3^7 at L200 F,
 and its  stress-to-rupture properties are also lov.

      Incoloy 800,  an iron-based superalloy,  has been used in some high-temperature
 gas turbine recuperators.  Its mechanical properties at 1000 to 1500 F are
 comparable to Type 3^7 stainless  steel,  as is its cost.  Incoloy 800 is sometimes
 specified  instead  of 3^7 steel because of its superior resistance to stress
 corrosion  cracking in a  chloride  ion environment, a situation which might be
 encountered in a marine  gas turbine application.  Incoloy 800 is only slightly
 more expensive than  3^7  stainless steel.  The brazing characteristics of Incoloy
 800 are  also very  good.

      Other superalloys which have been considered for high-temperature recuperator
 applications are Hastelloy X and Inconel 625 (Ref. 101).  Although these alloys
 also have  good brazing properties, they  are  significantly more expensive than
 3^7 stainless  steel  or Incoloy 800.

      The characteristics of the brazing  alloy are also important in specifying a
 recuperator for high-temperature application.  The selection of the brazing alloy
 is  made  in conjunction with the selection of base material because the brazing
 temperature must be  compatible with the  base material properties.   In Ref.  101,
 Palniro  7>  a silver  alloy, was selected  as the first choice for use with both
 3^7 stainless  steel  and  Incoloy 800 because of its excellent brazing characteristics
 and its  hot corrosion resistance;Nicrobraz 135. a nickel-base  alloy, was selected
 as  the second  choice.  Although Palniro  7 is probably more expensive than Nicrobraz
 135 on a per-pound basis, the  important  consideration is how the choice of brazing
 affects  the overall  cost of the recuperator which must operate to a specific
 requirement.

      Discussions with the Hamilton Standard Division of United Aircraft have
 indicated  that  several new brazing materials are under development which would be
 applicable  to the  regenerators of interest.  Some of these brazing materials tend
 to  completely coat the base material during the brazing process and, since they
 are  corrosion-resistant, they may even permit the use of low-cost  base materials
 at  relatively high operating temperatures.
                        BASIS FOR SELECTING DESIGN PARAMETERS

     The anticipated advances in design technology and materials improvement
projections discussed in the preceding sections of this report formed the basis
for parametric studies of future natural gas-fueled gas turbine power plant designs
                                        61*

-------
Simple-cycle pover plant designs, incorporating single-spool and twin-spool
turbomachinery configurations, regenerative-cycle power plant designs  with single-
spool machinery, and compound-cycle power plant designs incorporating  one stage of
intercooling and one reheat were investigated.

     Ranges of design parameters for base-load gas turbines, reflecting component
and materials technology currently available and that projected to become available
during the 1970 and 1980 decades, are presented in Table XVI.  These design
parameters reflect projected advances in turbomachinery flow-path aerodynamic
performance and mechanical design features,  as  well as projected advances in
combustor and turbine materials and turbine  cooling techniques.  The data shown in
Table XVI indicate that if compressor bleed  air is precooled to temperature levels
of 125 to 250 F before being used for turbine vane and blade cooling,  increases
in turbine inlet operating temperatures of as much as UOO deg can be achieved.
The turbomachinery component efficiency values  projected in Table XVI  for each  of
the respective time periods indicated could  be achieved if anticipated aircraft
development programs are pursued.  A wide ran^e of compressor pressure ratios was
selected, including those levels which are already commonly used, so that the
determination of optimum cycle conditions for maximum thermal efficiency simple-
cycle, regenerative-cycle, and compound-cycle,  gas turbines could be made.  In  the
compound-cycle power plant studies, overall  cycle pressure ratios as high as 100:1
were investigated, although the individual unit pressure ratio did not exceed
23:1.  Parametric performance studies for the compound-cycle configurations were
not as extensive as those for the simple-cycle and regenerative-cycle  designs
and were restricted primarily to the design  technology projected to become
available during the early 1980's.  The ranges  of regenerator effectiveness and
pressure drops in Table XVI were selected on the basis of previous experience.
A detailed description of the methods and assumptions used to perform  these
parametric performance and power plant design studies is presented in  Appendix  B.
                               PERFORMANCE ESTIMATES

                               Simple-Cycle Engines

     Thermal efficiency and specific output (shp per unit airflow) for simple-
cycle gas turbine power plants incorporating the three levels of design technology
defined in Table XVI are depicted in Figs. 28 through 30.  The performance presented
in these figures and in those to follow for the regenerative- and compound-cycle
engines are based on the use of methane (HHV = 1000 Btu/ft3)  as the fuel at an
ambient state defined in accordance with NEMA standards (80 F and 1000 ft altitude).
In addition, these performance values are for the gas turbine only, and although
they have been corrected for intake and exhaust stack pressure drops, they would
have to be adjusted for auxiliary power loads and generator efficiencies to reflect
                                         65

-------
 net station performance.   Since  these  corrections usually represent a constant
 percentage of the output  and thus  would  not  affect the  comparions, they have
 been omitted until a final design  point  has  been selected for each system under
 investigation.

      Present-day gas turbines which  operate  at turbine  inlet temperatures up to
 1700 F can achieve thermal efficiencies  on the order of 20 to 21% (Refs. 71, 7U,
 76, to 78, and 86) and specific  outputs  of 80 to ikO shp per Ib/sec airflow.  During
 the 1970 decade it is  estimated  that simple-cycle gas turbines incorporating the
 best presently available  materials,  turbine  cooling schemes, and component
 performance will be capable of substantial improvements in both thermal efficiency
 and specific output.   The specific output is an important parameter since it is
 directly related to engine size  and hence cost; thus the relative increases in
 this parameter provide an approximation  to the potential reduction in power plant
 size for a given output.   The estimated  performance of  simple-cycle engines (shown
 schematically in Fig.  31)  which  could  be in  commercial  operation during the 1970
 decade is shown in Fig. 28.   Performance data are presented for engines designed
 to  operate at turbine  inlet gas  temperatures from 1600  to 2200 F without cooling
 of  the compressor bleed air in a separate exchanger, for engines operating at
 2200 F turbine inlet temperatures  with compressor bleed air precooled to 125 F,
 and with no compressor bleed air for turbine cooling at all.  The Fig. 28 results
 indicate that if the present practice  of using uncooled compressor bleed air is
 continued, thermal efficiencies  and specific outputs as high as 31$ and 200 shp per
 Ib/sec airflow,  respectively, could be achieved.  A comparison of the results
 obtained at turbine  inlet  temperatures of 2200 F indicates that preceding the
 compressor bleed air prior to its  use  for turbine cooling would account for about
 a two-percentage point increase  in thermal efficiency.  The use of precooled
 compressor bleed air could lower the actual  blade operating temperatures by several
 hundred degrees  when compared to the use of  uncooled bleed air or alternatively,
 could reduce  the total amount of bleed air necessary to maintain a given blade
 temperature.   The  latter  approach  was  used in this study, and for typical conditions
 the  use of preceding  reduced the  quantity of bleed air required by about 50/5.
 For  example,  at  a  turbine  inlet  temperature  of 2200 F and pressure ratio of l6:l,
 the  engine with  the uncooled bleed air would require about 1.1% of the compressor
 air  for turbine  cooling while in the precooled design only 6% would be needed.
 The  reduction  in bleed flow  is largely responsible for the higher efficiency in
 the  precooled  designs.  The  Fig. 28 data also indicate that substantially higher
 performance could  be achieved if no compressor bleed air were required for turbine
 cooling.   This curve implies the use of  turbine materials capable of continuous
 operation at a temperature of 2200 F,  and the prospects for achieving such
materials  in the 1970  decade are not encouraging.  However, the results do provide
 an indication  of the performance that might be achieved with future uncooled
 ceramic  composite  turbine materials.
                                        66

-------
     Brief studier of the effects of precooled compressor bleed air temperature
levels on power plant performance indicated that the cooling air might be provided
to the turbine section at temperatures between 125 F and 250 F without serious
compromises in performance, thus affording less stringent design conditions  on  the
precooler heat rejection system.  Consequently, 200 F was selected as  the tempera-
ture level for the precooled compressor bleed air, and the performance for simple-
cycle gas turbines incorporating 1970-decade design technology is depicted in
Fig. 29 as a function of compressor pressure ratio.  The Fig.  29 data  are based
on the use of compressor bleed air precooled to 200 F for turbine cooling and
turbine inlet temperatures up to 2^00 F.  The Fig. 29 results  indicate a dimini-
shing rate of improvement in thermal efficiency as turbine inlet temperature is
increased above approximately 2200 F.  However, significant improvements in  specific
power continue to be realized above a gas temperature of 2200  F.

     The performance of simple-cycle gas turbine systems which could be in commer-
cial operation during the early and late parts of the 1980 decade are  summarized
in Fig. 30.  Thermal efficiency and specific.power levels as high as 3,8% and 270
shp per Ib/sec airflow, respectively, should be achievable in  the early 1980's
with simple-cycle, high-pressure-ratio engines operating with  2^00 F turbine
inlet gas temperatures and utilizing precooled compressor bleed air turbine
cooling techniques.  The improved component efficiencies and materials technology
projected for the early 1980's (Table XVI) relative to the projections for the
1970 decade result in the superior performance shown in Fig. 30, as compared to
the Figs. 28 and 29 performance estimates for a given turbine  inlet temperature.
The Fig. 30 data also indicate that increases in turbine inlet temperature above
2i*00 F up to the anticipated level of 2800 F in the early part of the  1980's
result in specific power increases of about 20 hp per Ib/sec of airflow for  each
100 F rise in turbine inlet temperature, with essentially no improvement in
thermal efficiency.  However, thermal efficiencies of nearly kl.% and specific
power approaching ^00 hp per Ib/sec of airflow could be achieved with  a 3000-F
turbine inlet temperature, a level which could be reached in late-1980 decade
engines.
                             Regenerative-Cycle Engines

     If a portion of the waste heat available in the exhaust gases of a gas
turbine is used in a regenerator to heat compressor discharge air prior to
combustion, a reduction in required fuel flow rates and,  hence,  significant
improvements in thermal efficiency may be realized.  The  thermal efficiency  and
specific output performance of regenerative-cycle base-load gas  turbines (shown
schematically in Fig. 31) operating at a turbine inlet gas temperature of 2000 F
and based on 1970-decade technology is shown in Fig. 32.   This performance reflects
two levels of regenerator total pressure drop, h% and 8$, and is based upon  the
                                        67

-------
 assumption of uncooled compressor "bleed air being  used to  cool  the  turbine section.
 These engines offer up to seven percentage  points  improvement in  thermal  efficiency,
 depending upon the  assumed values of regenerator effectiveness  and  pressure drop,
 when  compared with  a simple-cycle gas  turbine  designed for 2000 F.  Thermal
 efficiencies  approaching 39%  are possible with high-effectiveness,  low-total
 pressure  drop regenerator designs (Fig.  32).   Figure  32 also shows  that the peak
 thermal efficiencies for the  regenerative-cycle  design occur at relatively low
 compressor pressure ratios  (from about 5:1  up  to 8:1);  thus, single-spool
 compressor gas turbine configurations  would be possible.   The variation of regenerat:
 hot-  and  cold-side  temperatures with compressor  pressure ratio  for  a regenerator
 air-side  effectiveness of Bo% is shown in Fig. 33.  Eelatively  high hot-side gas
 temperatures  will exist with  compressor pressure ratios of about  U:lt but at a
 pressure  ratio of 10:1 or higher,  temperatures below  1000  F would be encountered
 and mild  steel construction could be used.   Estimates  of the materials required
 for regenerators at various levels of  turbine  inlet gas temperature and compressor
 pressure  ratio are  shown in Fig.  3^.   This  figure  indicates that mild steel could
 be used for regenerator construction at  turbine  inlet  gas  temperatures of about
 2200  F if the engine design is  based upon a compressor pressure ratio of  10:1 or
 above.  Stainless steels wouJd be  required  at  higher  turbine inlet  gas temperatures
 (and  the  same pressure ratios), and ultimately,  nickel-base alloys  such as Incoloy
 800 would be  needed.

      The  performance of regenerative-cycle, base-load gas  turbines  utilizing com-
 pressor bleed air precooled to  200 F is  shown  in Fig.  35 for systems which could
 be in commercial operation during the  1970  decade  and early in  the  1980's.  The
 utilization of precooled compressor bleed air  enables  the  attainment of 200-F
 turbine inlet temperatures with 1970-decade design technology and the achievement
 of thermal efficiencies on the  order of 31% with practical regenerator effective-
 nesses  (80 to &5%)  as  is  shown  in Fig.  35-  This performance reflects about a four-
 percentage point improvement  relative  to that of the  simple-cycle designs (based
 on the  same design  technology)  described previously.  Furthermore,  as previously
mentioned,  the  relatively high  thermal  efficiency associated with the regenerative-
 cycle designs  is achieved at  low compressor pressure ratios (approximately 8:l),
by comparison with  optimum thermal efficiency simple-cycle designs which  require
 compressor pressure  ratios on the order  of 16:1 to 20:1.  A comparison of Figs.
 35 and  30  also  shows that the high thermal efficiency levels for the regenerative
 cycles  can be achieved without  a compromise in specific output.

      Regenerative-cycle  designs, based  on early-1980's  technology, should be
 capable of achieving thermal  efficiencies of about Ul#  at a turbine inlet temperat1.
of 2^00 F,  a  regenerator  effectiveness of 80JJ, and moderate engine pressure ratios
 (see  Fig.  35).  Increases in  turbine inlet temperature  to 2800  F result in only
minimal increases in thermal  efficiency, but a 25$ increase in  specific power.
However, with the projected improvements in component efficiencies, and materials
                                        68

-------
anticipated by the late 1980's, additional improvements  in both  thermal efficiency
and specific output can be realized by further increases  in turbine  inlet gas
temperatures to 3000 F.  As shown in Fig.  36,  thermal  efficiencies approaching  k5%
will be achieved with regenerative-cycle designs,  based  on regenerator  effective-
ness values of 80 to
                              Compound-Cycle Designs

     The addition of reheat and intercooling to the simple-cycle gas  turbines
(see Fig. 3l) offers the potential of achieving further improvements  in power
plant performance.  In the cycle configuration envisioned herein,  air enters a
low-pressure ratio compressor and is partially compressed.  The air is then cooled
to a temperature level of 125 F within the intercooler.   The  cooled air is then
further compressed in a high-pressure ratio compressor to the final overall cycle
pressure ratios of 50 to 100 atm and heated in the combustor  by the combustion of
fuel.  These high cycle pressure ratios are possible  through  the use  of the inter-
cooling which serves to reduce the total compression  work required for a  given
pressure ratio.  Partial expansion of the combustion  gases occurs  in  the  gasifier
turbine, which drives the high-pressure ratio compressor.   Then further heating
back to the initial turbine inlet gas temperature occurs in the reheater.  Since
turbine work is proportional to temperature, the use  of reheating  increases the
total work obtained for a given expansion.  Final expansion to atmospheric pressure
then occurs in the low-pressure turbine, which drives both the low-pressure ratio
compressor and the generator.

     A summary of the calculated compound-cycle gas turbine performance is shown
in Fig. 37» based on component efficiencies and materials projected for the early
I960's.  The results indicate thermal efficiency levels as high as h2% at turbine
inlet temperatures of 2200 to 2^00 F.  Figure 37 also shows that turbine  inlet
gas temperatures beyond 21+00 F would not produce higher thermal efficiencies
but would provide practically a linear increase in the specific power capability
beyond UOO shp per Ib/sec.  Specific powers of UOO shp per Ib/sec  represent a
significant 33% increase in specific power beyond that achievable  with simple-
cycle gas turbines at the same turbine inlet temperatures and technology  base
(compare Figs. 37 and 30).

     Prior to computing the engine performance data depicted  in Fig.  37 a pre-
liminary analysis was conducted to determine the effect of varying the amount  of
compression between the previously defined low-pressure ratio and  high-pressure
ratio compressors (see Fig. 31).  Typical results for total cycle  pressure ratios
of 50:1 and 100:1, respectively, are shown in Fig. 38.  The performance depicted  in
Fig. 38 indicates that if the low-pressure ratio compressor provides  more than 50
percent    of the total cycle pressure ratio, the specific power increases while
                                        69

-------
 thermal efficiency falls  off gradually.   Since "base-load operation emphasizes the
 achievement  of high thermal  efficiencies, the distribution of  compressor work was
 selected to  provide maximum  efficiency.   Thus pressure ratios  of 3, U, and 5 vere
 selected for the lov-pressure ratio  compressor to  correspond with total cycle
 pressure ratios of 50,  75, and 100,  respectively.

      Reheat  for these designs occurs after expansion to a pressure ratio
 corresponding to a work split of  from kQ% of the total turbine work in the
 relatively low-cycle pressure ratio  (50:l)  designs to about 65$ of the total
 turbine work in the high-cycle pressure ratio (lOO:l) designs.  If reheat of the
 combustion gases was provided following complete expansion through the gasifier
 turbine and  prior to the  low-pressure turbine, further, improvement in specific
 power would  be achieved but  at a  loss in  thermal efficiency.   This trend is apparent
 from  an inspection of Fig. 39 which  depicts the effects on thermal efficiency and
 specific output performance  of reheat at  different levels of expansion through
 the gasifier turbine.
                      SELECTION OF GAS TURBINE PARAMETERS FOR
                                MINIMUM-COST POWER

     Parametric investigations of the interrelationships among gas turbine per-
formance, design  configurations, and engine cost were conducted for the simple-,
regenerative-, and compound-cycle gas turbines to determine those system designs
which would have  the potential for generating lowest-cost electric power within
the next two decades.  These investigations utilized the gas turbine engine design
and costing procedures, developed under Corporate sponsorship, which are described
in detail in Appendices B and C, respectively.  The selling prices (based on 1970
dollars) were estimated based on a projected market of UOOO Mw per year (see SECTION
VIII for further  details).  Appropriate factors for development, assembly, test costs,
general and administrative expenses, and profit are included in the estimated selling
price of each unit.

                            Simple-Cycle Engine Designs

Engine Size

     The effect of engine size (output capacity) on the estimated gas turbine
specific selling prices is depicted in Fig. kQ for combinations of turbine inlet
temperatures and compressor pressure ratios chosen to reflect high-thermal effi-
ciency designs (see Figs. 29 and 30).  The results are presented for power turbine
output speeds of 3600 and 1800 rpm to match the synchronous rotationals speeds of
large electric generators.  The trends illustrated by the curves shown in Fig. Uo
clearly indicate that specific selling price decreases with unit power capacity
up to approximately 100 Mw which is the approximate upper limit for the power which
                                        70

-------
can be developed by a 3600-rpm power turbine.   To produce more power than  100 Mw,
an l800-rpm power turbine must be used which initially  results in  a penalty  in
specific price relative to 3600-rpm designs.   However,  as unit power capacities
increase, specific price is seen to decrease to apparent minimum attainable
levels for capacities in the range of 250 to 300 Mw.  Gas turbines of  this latter
power capacity must be considered to be very large machines  compared with  those
presently in commercial operation.  An alternative means of  attaining  high unit
power capacities would be to utilize multiple exhaust ends on the  power  turbine.
This is common practice in steam turbine design as a means of achieving  unit
capacities of 500 to 1000 Mw.  The Fig. Uo data illustrate that the maximum  output
capacity of gas turbine engines designed with two exhaust ends at  3600 rpm could
achieve somewhat lower specific prices and extend the unit capability  to about
150 Mw.  However, the added complexity of multiple exhaust ends is considered
impractical.

     The results also illustrate that the selling price of these advanced, large-
capacity gas turbines could approach levels as much as  30 to 50$ lower than  those
of present-day gas turbines used in power generation application.  Smaller reductions
in engine specific selling price ($/kw) are also projected as turbine  inlet
temperatures are increased toward 2600 to 2800 F, a level representative of
early-lpSO's technology.  However, to achieve these cost reductions, blade
centrifugal stresses approaching ^5,000 psi and above,  as indicated in Fig.  Uo,
will be experienced in the last stage of the power turbine,  particularly when
extending the single-exhaust-end design configurations  to output capacities  on
the order of 75 Mw for 3600-rpm turbines and 250 Mw for l8oO-rpm designs.

     Beyond output capacities of about 100 Mw, the power turbines  must be  designed
for rotational speeds of l800 rpm to avoid prohibitive  blade stresses.  However,
for a given turbine inlet temperature and compressor pressure ratio, a step  in-
crease in specific selling price occurs, as mentioned previously,  when the power
turbines are designed for output speeds of 1800 rpm instead  of 3600 rpm.  This
increase is due to the larger power turbine components  associated  with the l800-
rpm designs.  The utilization of an l800-rpm power turbine enables gas turbines
to be designed with a single exhaust end at unit output capacities approaching
250 Mw without exceeding allowable stress limits (see Fig. hO).  For l800-rpm
designs, miniaum engine selling prices are achieved at  about 250 Mw for  2600 F
to 2800 F turbine inlet temperature designs and remain  relatively  constant above
the 200 Mw level for the 200 F to 2^00 F turbine inlet  temperature level
representative of 1970 technology designs.  The Fig. ^0 results provided the
basis for concentrating further cost studies on the 200- to  250-Mw size  engines
for base-load, power generation applications.  The high power plant capacities
associated with the minimum specific selling price designs also provide  further
power station cost savings since fewer units and hence  less  ancilliary equipment
and smaller floor spaces would be required in a typical large, i.e., 750-Mw  to
1000-Mw power station.
                                       71

-------
 Engine  Pressure  Ratio  and  Turbine Inlet Temperature

      The effects of variations in turbine inlet temperature and compressor pressure
 ratio on the  gas turbine engine selling price are shown in Fig. hi for engines
 (based  on 1970-, and 1980-decade design technologies) designed to operate at
 l800  rpm output  speed  and  to provide a nominal 200-Mw electrical output.  All
 of the  engine designs  included in Fig. hi except the 1970-decade, 2000-F turbine
 inlet temperature engines  can employ pover turbines with single exhaust ends
 without exceeding practical blade stress limits of about 3^,000 psi in the last
 stage.   Power turbines with two exhaust ends are required in the 2000-F engine
 designs in order to remain within those specified stress limits.  Nearly a 10$
 reduction in  engine selling price could be achieved, as indicated by the circled
 point in Fig.  Ul, for the  2000-F engine designs by relaxing the constraint on
 blade stress,  and thereby  eliminating an exhaust end, if the blade stress in the
 last  stage of the power turbine is allowed to approach U6,000 psi at the expense
 of reduced power turbine life.  However, even with the relaxation of this design
 constraint, the  specific costs ($/kw) for the 2000-F engine designs remain relative!}
 high.

      The  estimates  in Fig. Ul also indicate that increases in turbine inlet
 temperature up to 2600 to  2800 F provide substantial reductions in selling price.
 However,  above turbine inlet temperatures of 2800 F, the engine selling prices
 begin to  increase again.   A partial explanation of these trends is evident by
 reviewing the  effect of increases in turbine inlet temperatures on the specific
 power level of the  gas turbine (see Fig. 30).  As turbine inlet temperature is
 increased from 2^00 ? to 2800 F  and above, the specific power increases by as much
 as  50$;  thus  a lower airflow rate is required to achieve a given power level,
 and hence smaller and lower-cost components are needed.  However, above a turbine
 inlet temperature of approximately 2800 F, the costs of the hot-section components,
 i.e., the compressor turbine and the power turbine, begin to increase sharply due
 to  the  need for more sophisticated construction materials in each component to
withstand high gas  temperatures.   As a result, the cost reductions for the compressor
 and burner components, accruing from technology improvements (higher stage loadings,
 component  efficiencies, etc.) as  well as higher specific power levels, are offset
by  the  higher  costs of the compressor turbine and power turbine sections.

      To narrow the  range of engine design parameters considered in this study, the
influence  of compressor pressure  ratio on power cost at turbine inlet temperatures
corresponding  to  the various time periods (1970-decade and early and late 1980's)
was estimated.  The results, based on 80% load factor, 15$ capital charges, and 30
and 50<£/million Btu fuel costs,  are summarized in Table XVII for selected 200-Mw
plants.    The effect of variations in the engine specific selling price on the
overall cost of the power generating station were also included in the analysis.
The results indicate that the cost differentials among the various pressure ratios
are generally  small.  The optimum compressor pressure ratios range from 20:1 at
                                         72

-------
a 2000-F turbine inlet temperature to 36:1 at 2900 F.   The improved performance
achievable at high pressure ratios is the primary factor in determination of the
minimum-power cost system.  Hovever, a reduction in the load factor to 10% would
tend to favor the lover compression ratio designs which are somewhat less expensive
to build.  For very low load factors of 20%, comparable to utility peaking applica-
tions, compressor pressure ratios of about 15 to 20:1 would be selected for the
various time periods.  Based on the Table XVII data, attention was focused on
simple-cycle engine designs representative of the 1970 decade, operating at
compressor pressure ratios of about 20:1.  For the early- and late-1980's engine
designs somewhat higher compressor pressure ratios, between 2U and 30:1, were
selected for the systems which would provide minimum-cost power.
     Cpmponent_
     Examination of Fig. 1*1 also reveals that increases in compressor pressure
ratio at any given turbine inlet temperature generally tend to result in higher
unit selling prices.  As pressure ratio is increased, the number of compressor
stages as well as the number of stages in the compressor turbine must increase
accordingly.  In addition, since an increase in compressor pressure ratio could
also mean a decrease in specific output (higher airflow rates for a given output)
the sizes of these components increase.  As a result, the manufacturing costs of
these components are increased.  The curves in Fig.  k2 provide an indication of
the typical distribution of costs among the various  major components in an early-
1980 's gas turbine engine design as pressure ratio is increased.  The component
costs shown in these figures account for approximately 70$ of the total engine
cost.  Much of the remaining cost is distributed among such components as the casings
and the bearings, seals, and shafting.  The costs for the assembly and testing
of each unit, which account for approximately 10% of the total manufacturing costs,
are also not included in the Fig. **2 estimates.

Single- vs Twin-Spool Designs

     The estimates presented in Fig. ^1 were based upon the use of twin-spool
compressor designs and indicate that engine specific selling price tends to
minimize at compressor pressure ratios below 15:1.  This result suggests that a
further reduction in selling price would be possible through the use of a single-
spool rather than a twin-spool design.  In a single-spool engine design, the
desired engine cycle pressure ratio is achieved by a series of compressor stages
vhich are mechanically constrained to operate at the same rotational speed.  If
the design pressure exceeds about 10:1, an excessive number of compressor stages are
required as well as variable-pitch stator blades on the first few stages to avoid
compromising the startup and part-load performance characteristics of the engine.
&i a twin-spool design, the compression process is provided in two separate
compressors, each operating at its own optimum rotational speed to provide maximum
Performance.  Thus, the forward, or low-pressure, compressor section is on a
                                       73

-------
 common shaft with the lov-pressure stages of the compressor turbine.   The aft,
 or high-pressure, compressor section is on. a common shaft with the initial (high-
 pressure) stages of the compressor turbine.   The lov-pressure spool shaft is
 concentric with and extends through that of the high-pressure spool.   The twin-
 spool design is a more complicated configuration and requires more bearings,
 supports, controls, and shafts  and thus often more maintenance than a single-
 spool design.

      The variation in specific  selling price with compressor pressure ratio for
 a single-spool engine designed  to provide approximately 200 Mw when operating at
 a turbine inlet turbine of 2200 F is shown in Fig.  l+3a.   The results, shown as  a
 dashed line in Fig. l+3a, were obtained using a compressor corrected tip  speed of
 1000 ft/sec and a diffusion factor of 0.1+0 (as were the estimates  shown  in Fig.
 The specific engine selling prices for the single-spool designs vary  between  $32
 and $37/kw, substantially higher than the costs for the twin-spool designs of
 Fig.  1+1.   If the compressor corrected tip speed were raised to 1150 ft/sec and  a
 diffusion factor of 0.1+5 were used,  a level  which is representative of more
 advanced aircraft engine technology, the selling prices  of the single-spool
 engine designs would be reduced about $5/kw  as shown by the solid  line in Fig.  l+3a.
 These cost levels are still not lower than those shown  in Fig.  1+1  for the twin-
 spool designs.   Furthermore, compressor tip  speeds  above 1000 ft/sec  could also
 be  utilized on the twin-spool designs to achieve modest reductions in engine
 selling price.   The Fig.  l+3b data indicate that operation at a compressor tip
 speed of 1150  ft/sec would eliminate 5 stages from  the  compressor  as  well as  one
 stage from the compressor turbine (not shown).   The variation in engine  selling
 price with compressors  designed for  tip speeds  above 1000 ft/sec are  shown in
 Fig.  l+3c.   However, unlimited increases in compressor tip speed above about
 1200  ft/sec are restricted by a rapid increase in the disk stresses and  hence the
 requirement for more advanced materials.

 Power Turbine

      Since  power turbine  component costs  are a significant fraction of the total
 engine costs  (Fig.  1+2),  efforts were made  to achieve further engine cost reductions
 from  cost  optimization  studies  of the power  turbine design configurations.
 Figure  1+1+a  shows  the  effect  on  engine specific  selling price of variations in the
 power  turbine  last-stage  hub/tip  ratio  for a 200-Mw,  l800-rpm gas  turbine
 operating at a 21+00-F turbine inlet  temperature and a 20:1 compressor pressure
 ratio.  As  indicated in Fig.  1+1+a,  an increase in hub/tip  ratio  from 0.50 to 0.70
 results in  the  elimination  of two  stages  from the power  turbine (not shown) but  the
 result  is less  than  a 10$ reduction  in  engine selling price.  Despite the  decrease
 in blade and vane  costs accruing  from a reduction in the  number  of stages  as
well as from the decrease  in blade height, these  cost reductions are  offset by
higher  disk costs  resulting  from the increase in  disk diameter.  Hence,  although

-------
two stages have been eliminated by increasing the last-stage hub/tip ratios from
0.50 to 0.70, disk costs have also increased by a factor of almost tvo.  Further
cost studies were confined to the range of hub-to-tip ratios between 0.50 and
0.70.
     A more pronounced effect on engine selling price as well as on engine per-
formance appears to be due to the limits placed on the exit velocity of the gases
leaving the last stage in the power turbine.  A large annulus area reflects the
use of long, highly stressed, high-cost blades and vanes and low exit velocities.
The low exit velocities permit more efficient  extraction of the energy in the
gas (lower leaving losses) and thus more efficient power plant performance.  This
trend is shown in Fig. ItVb which depicts the effects of power turbine exit
velocity on engine specific selling price and thermal efficiency performance for
a 200-Mw, l800-rpm, gas turbine designed to operate at a 2^00-F turbine inlet
temperature and a 20:1 compressor pressure ratio.  It can be seen from Fig. hkli
that an increase in exit velocity from 500 to 800 ft/sec affords approximately a
2Q% decrease in specific selling price, but this is accompanied by approximately
an 0.8$ loss in thermal efficiency.  However, a brief analysis indicated that
the reduction in engine selling price of 5 to 7 $/kw would offset the higher fuel
costs due to the lower power plant efficiencies; hence, the remaining engines
were designed for exit velocities of 600 ft/sec or more.  The loss in engine
performance is minimized by installing a suitable diffuser having a conservative
diffuser energy recovery coefficient downstream of the power turbine.

     Materi als_ Changj2s_

     The effects on the engine selling price of substituting lower-cost materials
and of relaxing the first-stage hub-to-tip ratio design constraints in the power
turbine were also studied.  Attempts to maintain the hub-to-tip ratio at 0.85,
a value which generally ensures low turbine losses, had resulted in relatively long
last-stage turbine blades for the engines which could be commercially available
by the late 1980's.  As a result, these engines were found to be somewhat more
costly than engines investigated for the earlier time periods (see Fig. Ul).
However, by relaxing the hub-to-tip ratio constraints to a. level of 0.875, a
substantial reduction in engine selling price could be achieved, especially as
indicated in Fig. U5, at the highest turbine inlet temperature of 3100 F.

     The Fig. 1*5 results also indicate that, because of increased cooling require-
ments, the substitution of advanced nickel alloy blades and vanes for the more
costly columbium alloys tends to increase the engine selling price by about 10
to 13$.  At turbine inlet temperatures of 2900 and 3100 F as many as three
stages of the power turbine must be cooled if nickel-based alloys are used, and
although the nickel-based alloys are less expensive than columbium alloys the
costs associated with providing cooling increases the engine selling price.
                                        75

-------
 Coating Life

     During the  course of the study, additional data were received indicating that
 the  effects of erosion and corrosion on the compressor turbine initial-stage vanes
 could be  the limiting compressor turbine design parameter rather than vane creep
 strength.  The data  indicate that at the assumed vane operating temperatures
 (i.e., 1700 F, 1900  and 2100 F for the 1970-decade and early and late 1980's,
respectively)  coating life could be relatively short, on the order of approximately
 10,000 hr.  The  coating life vould be considerably shorter if ash bearing fuels
 containing even  a small quantity of sulfur were used.  Although recoating of the
 vanes at  10,000  to 15,000-hr intervals would be possible without incurring high
maintenance costs, a more practical base-load engine design would utilize lower
 vane operating temperatures by using a higher .percentage of compressor bleed air
 for vane  cooling.  The effects of a reduction in allowable vane temperature on
 engine performance were investigated (the selling prices would remain the same)
 and typical results  are presented in Fig. U6.  The Fig. U6 data indicate that an
 eight-fold increase  in coating life could be achieved by cooling the vanes about
 200 F lower than the originally assumed temperatures for a representative early-
1980's technology engine designed to operate at a turbine inlet temperature of
2600 F and a 28:1 compressor pressure ratio.  For example, if the vane temperature
were maintained  at 1700 F rather than 1900 F, a coating life of about 80,000 hr
could be achieved although the thermal efficiency and specific power would be
reduced about  1.5$ and 3.2$, respectively, due to the added compressor bleed
flow required.   A 3-to-l improvement in coating life could be achieved with only
a 125-F decrease in  allowable vane metal temperature while the performance losses
would be reduced. The Fig. k6 results are based on currently available and
projected performance of turbine vane coatings; however advanced coatings under
investigation  are expected to provide protection for time periods on the order of
several years.  The performance penalties estimated for lower vane temperatures
were, nonetheless, included in the final engine efficiency estimates.
                          Regenerative-Cycle Engine Designs

     To define the regenerative-cycle engine designs that have the greatest
potential for generating the lowest-cost electric power, the effects of a number
of engine and heat exchanger design parameters were investigated, including the
type of regenerator surface, gas pressure drop, recuperator* effectiveness and
flow arrangements.
   Recuperator and regenerative are used interchangeably although the type of heat
   exchanger considered is stationary and heat must be transferred across  a surface.
                                        76

-------
Recuperator Surface Characteristics

     The regenerative-cycle gas turbine characteristics presented herein are based
on the use of plate-fin type recuperator surfaces although  bare- and finned-tube
type surfaces vere considered as veil.   Plate-fin type surfaces  have many
advantages in large industrial applications such as  those considered here,  probably
the most important of vhich is the ease in vhich these surfaces  can  be  used in
nodular construction of the recuperator.  Modular construction permits  the  use of
a number of similar small, readily manufactured units  to be built up into a large
installation such as that advocated in Ref. 103.  Compact plate-fin  surfaces vith
high heat transfer surface per unit volume also result in minimal exchanger
dimensions so that the station foundations and building size can be  minimized.
Although tube-type surfaces can be made relatively compact, with surface area
densities approaching those of plate-fin exchanges,  very small tube  diameters
(on the order of 0.125 in.) and thin-tube vail thicknesses  (of approximately
0.00k in.) are required (see Ref. 88).   In the Ref.  88 study, both bare-tube
and plate-fin type surfaces vere considered, but the application vas for a  very
small military gas turbine engine vith an airflov of 5 Ib/sec rather than the
1000 Ib/sec engines considered in base-load plants.   To build a  tube-surface
recuperator for a 200^-lv base load engine considered in this study,  tube header
attachment costs vould be high for the many miles of small-diameter  tubes required.
Exploratory calculations performed at the Hamilton Standard Division of United
Aircraft (HSD) also indicated that the no-flov direction vith tube configurations
vould be several hundred feet long and introduce difficulty in devising a compact
arrangement.

     A number of different plate-fin type surface geometries vere initially
considered, but attention vas focused on the compact configurations. Although
the Harrison industrial gas turbine recuperators have  been  constructed  vith a
rather course matrix (l-in. fin height on the gas side and  a channel vith no fins
on the air side), both Ref.  105 and exploratory calculations performed at  UARL
indicate that significant reductions in core volume could be achieved vith  a finer
matrix (smaller fin height and many fins per inch).   Smaller core volumes result
in smaller foundations easier transportation to the site, and reduced overall plant
size.  The surface geometries employed in the industrial gas turbine recuperators
described in Ref. 103 are very fine, vith 20 fins per inch  and 0.10-in. fin height
on the air side and 16 fins per inch and 0.075-in. fin height on the gas side.

     Discussions vith HSD revealed that the selection of surface fin height and
fin spacing vill be a function of the desired volume and cost criteria  for  the
application.  Furthermore the costs «rre  more sensitive to  the number of pieces
handled in the fabrication processes than the manufacturing techniques. It  vas
evident that, to perform trade-off analyses in hopes of determining  the surfaces
      vould result in the minimum cost pover, a complete breakdovn of material and
                                        77

-------
 fabrication costs  would be  required for  each  candidate surface.  This type of
 generalized cost information was  not available from recuperator manufacturers,
 so a few representative surface geometries were  selected from Ref. 10U which
 contains the basic pressure loss  and heat transfer data needed to perform the other
 trade-off analyses conducted in this study.   As  a guide, surfaces similar -co those
 recommended in Ref.  103 were selected.   Those which appeared to provide the smallest
 core volumes were  plate-fins with 13.1 plain  fins per in. on the gas side and 19.8
 plain fins per in.  on the air  side,  both with fin heights of 0.25 in.  This
 combination of geometrical  characteristics results in a core with heat transfer
 surface  of approximately h6h ft2/ft3 of  core  volume.

 Effectiveness

      Recuperator air-side effectiveness  is one of the most important factors
 affecting engine performance,  recuperator size,  and system cost.  The data in
 Figs.  32 and 35 illustrate  that a significant increase in gas turbine thermal
 efficiency can be  achieved  by  increasing recuperator air-side effectiveness.  For
 example, for a turbine inlet temperature of 2000 F and a compressor pressure ratio
 of 8:1,  representative of a 1970-decade  engine,  the thermal efficiency is increased
 by approximately k points (from approximately 3^$ to 38$) by increasing effective-
 ness  from 10% to 90$.   Therefore, estimates were made of the effects of
 recuperator effectiveness on recuperator size and cost, and the results are shown
 in Fig.  ^7 for a 2000-F turbine inlet temperature engine capable of operating at
 a  pressure ratio of  8:1.  The  data  clearly show  the rapid increase in volume
 requirements,  and  hence heat transfer area, as the effectiveness approaches 90$.
 However,  the  slope of these lines is  a strong function of the type of surface
 selected for  each  side  of the  exchanger  and the  absolute size of the heat exchanger.
 Furthermore,  the UARL  data  in  Fig. U7 are based  on a cross-counter-flow type of
 flow  arrangement and the relative increase in recuperator size with effectiveness
 is  influenced by such  design characteristics as  pressure drop and the coarseness
 of the heat transfer matrix.  Also shown in Fig. ^7 is a manufacturer's estimate
 of the variation in recuperator cost with effectiveness.  This curve is for a
 different  plate-fin type of construction and a smaller size than the other
 regenerators  of Fig. U7.

     In order  to determine  the recuperator effectiveness that would result in
minimum-cost power, analyses showing the trade-off between engine operating and
 capital costs  were made using both UARL and manufacturers' recuperator cost estimate;
 and assuming various capital charges ranging from 12 to 17$ and for fuel costs  of
 30 and 50^/million Btu.  The results in terms of the added total power costs
 above a "base value of regenerator effectiveness  are shown in Fig.  U8.  For example,
using 30i£/million Btu fuel  and 12$ capital charges, the minimum costs would occur
 at a recuperator effectiveness of 80$ based on manufacturers'  estimates of the  cost
variation between 70 and 90$ effectiveness.   The total power costs  using a 90$
recuperator effectiveness would be at least 0.07 mills/kwhr above the minimum,  or
                                        78

-------
base, case.  In general, the results indicate the miniinun power costs  occur at
approximately 80$ effectiveness and therefore this value was  used in conducting
the remaining trade-off analyses made in this study.   Even at very high fuel costs
of 50^/million Btu, a 90$ effectiveness provides  minimum costs only vhen capital
charges are about 12%.  Data from recuperator manufacturers and Ref. 105 also
indicate that 80$ effectiveness is approximately  the  design value for  many
industrial gas turbines regenerators previously built.

Total Pressure Loss
     Another relatively important parameter considered in the  design  of a
regenerative-cycle gas turbine is the recuperator total pressure  loss.   The
effects of total pressure loss on the size characteristics of  recuperators  for
gas turbines designed with 1970 technology are illustrated in  Fig.  1*9 »  and  the
effects on the performance are shown in Fig. 32.   The results  presented in  Fig.  **9a
indicate that, as the total pressure loss is increased from ^.5 to  13.5$, the
decrease in recuperator core volume is most pronounced at the  highest value of
effectiveness (90$).  In designing the recuperators  represented in  Fig.  Upa it
was assumed that ~ of the total pressure loss occurs in the core  and  the remaining
§ in the manifolds and ducting.  The total pressure  loss has even a more pro-
nounced effect on the regenerator no-flow length  i.e., (see sketch  in Fig.  U9),  an
important item when considering the problems of integrating the engine  and
recuperator, which would affect the overall space requirements of the system.

     The effect of varying total recuperator pressure loss between  U  and 8$ on gas
turbine performance is shown in Fig. 32, and the  data indicate only about a one
point reduction in thermal efficiency and about a one-third of a  percent reduction
in specific power at the highest pressure loss.  Since the recuperator costs about
$15/kw of engine output, the savings in recuperator cost for an  8$  total pressure
loss level would offset the loss in efficiency relative to a recuperator designed
for only k% total pressure loss.  Thus, further efforts were confined to
recuperators based on an 8/» pressure loss.

Pressure Loss Split

     The split of the total pressure loss between the air side and  gas  side is
another independent parameter which must be considered in designing a cross-flow
or multipass cross-counterflow type heat exchanger.   The effects  of varying the
pressure loss split are exemplified in Fig. ^9b.   For the conditions represented
in Fig. U9b, the core volume continues to decrease as the percent of total  pressure
loss on the gas side increases, but the no-flow length appears to be a minimum when
tfce split in pressure drop between the air and gas side is equal.  Thus a split
of | of the total pressure loss on the gas side,  the value assumed  for the
recuperator designs in Fig. ^9a, appears to result in a near-optimum recuperator
core configuration.
                                        79

-------
Flow Arrangement

     Rough layouts  of the engine and recuperators utilizing 2-passes (on the air
side) designed for  80$ effectiveness indicated that compact configurations
using the cross-counterflov arrangement vould require no-flov lengths of 100 ft or
more (see Fig. l*9b) and did not appear practical.  Other cross-counterflow
arrangements  suggested by the various aircraft recuperator designs analyzed in
Ref. 88, which involve the use of multipasses on both the gas and air sides, vere
considered briefly.  However, it appeared that these arrangements required
relatively complex  ducting and manifolding and the designs were tending toward
pure counterflow systems.  Thus, counterflow recuperator designs were investigated
and found to  be more desirable for the large industrial engines considered in
this study.   Counterflow designs usually involve more difficult header con-
figurations but provide far more efficient heat transfer and, as shown in Fig. 50,
result in significantly smaller core volumes than the cross-counterflow designs
investigated.  Furthermore, with the counterflow arrangement there is a certain
degree of flexibility which allows for achieving reasonable no-flow lengths for
the core.  With the counterflow design, the flow lengths for the compressor dis-
charge air and turbine exhaust gases must, by definition, be equal, and for given
heat transfer surfaces and heat rejection loads, the split in total pressure is
also fixed, as is the core face area.  Thus, the two no-flow lengths with a counter-
flow design can be  selected to suit the most convenient arrangement of engine and
recuperator,  the only stipulation being that the product of the two no-flow
lengths is equal to the required core face area.

Compressor Pressure Ratio

     Compressor pressure ratio was varied for 1970-decade and 1980-decade engines
to assess its effect on the recuperator size.  The estimates in Fig. 50 show the
reduction of  recuperator core volume with increased compressor pressure ratio for
engines with  200-Mw output capacity.  This trend is due to the higher air densities
and lower airflow requirements (for a given output capacity) associated with the
higher compressor pressure ratios.  The smaller core volumes for the 1980-decade
designs relative to the 1970-decade designs are also due to the lower airflows
required to yield 200 Mw of output power, since the early 1980's design can achieve
substantially higher specific outputs.  However, the 1980-decade designs operate
at higher temperatures and require the use of better materials in the recuperator
and more sophisticated fabrication techniques; these factors are reflected in
the recuperator costs.

     To determine the compressor pressure ratio that would provide minimum power
costs for regenerative-cycle engine power systems, costs were obtained for counter-
flow recuperator designs and combined with engine costs for a range of selected
operating conditions.   Counterflow recuperator costs were estimated from the require
heat transfer surface area, as determined through the use of an existing UARL heat
exchanger computer program, and specific cost factors ($/ft2) obtained from a
                                        80

-------
correlation of manufacturers'data (Refs.  88,  102,  and 106).   These  data were
obtained in response to inquiries and extensive  discussions vith manufacturers'
representatives for recuperators using a variety of different construction materials.
For those selected operating conditions vhere the  temperature capabilities of mild
steel were exceeded, a correction factor was  applied to  the specific  cost for mild
steel recuperators to account for the use of  better materials such  as  Type 3^7
stainless or Inconel 800.  For the recuperator sizes investigated  (between approximately
1 to 3 x 106 ft2) the specific cost of mild steel  recuperators were estimated to
range between $1.00 and $0.8Q/ft2.  Smaller recuperators would have a higher
specific cost.  The recuperator cost factor,  as  a  function of maximum metal
temperature, ia shown in Fig. 51a and illustrates  the sharp increase  in cost as
aaximum metal temperature increases.  The effect of temperature on  cost was
discussed with recuperator manufacturers on a number of  occasions and it  appears
that temperature not only affects the selection  of the base material  but  also the
brazing material and fabrication technique.  Furthermore, in  a very large recuperator
utilizing modular construction, it becomes economically  attractive  to use more than
one material, the choice depending upon the local  temperatures encountered.
However, the cost factor is depicted in Fig.  51a as a smooth  curve  rather than a series
of step increases which would reflect the use of more expensive material.  Mild steel
would be used for metal temperatures below approximately 950  F corresponding to
the maximum gas temperature of approximately  1000  F at the turbine  exhaust.  As
the temperature to which the recuperator materials will  be exposed  is  increased,
^ype i*30 stainless steel might be used.  Type 3^7  stainless or Incoloy 800 would
be used up to metal temperatures of approximately  1300 F. Above this temperature
level, more exotic and more expensive base materials would be required, but the
recuperator temperatures considered in this study  generally did not exceed a
temperature level of 1300 F.  The maximum metal  temperatures  and estimated
recuperator costs for the counterflow designs represented in  Fig. 50  are  shown in
Fig. 51b with the maximum metal temperature taken  to be  the average of recuperator
gas inlet and air outlet temperatures.  Since the  maximum metal temperatures
depicted in Fig. 51b do not exceed the 1300-F level, base materials no better than
Type 3^7 stainless or Incoloy 800 would be required in the 1970-decade and early-
1980 engines.  The decrease in recuperator cost  with engine pressure  ratio (for
a given turbine inlet temperature) is also illustrated in Fig. 51b, and it appears
that the trend is due primarily to the decrease  in maximum metal temperature.
Thus, maintaining metal temperature below approximately  1000  F appears to be an
important factor in determining recuperator cost.

     Representative results showing the effect of  compressor  pressure ratio on the
combined engine and recuperator costs are presented in Fig. 52>-- These costs are
based on the regenerative engine costs as shown  in Fig.  52a,using conservative
levels of compressor tip speeds of 1000 and 1100 ft/sec  for the 1970-decade and
early 1980's.  Also shown in Fig. 52a are the thermal efficiency levels estimated
for the regenerative-cycle engines. Combining the  hardware and fuel costs in
separate calculations indicates that minimum  total power costs would  be achieved
^y selecting compressor ratios of about 9:1 or 10:1 for  turbine inlet temperatures
                                        81

-------
 of 2000 F and about 12:1 for the early-1980 time period when turbine  inlet
 temperatures of 2400 F and above will "be utilized.   Thus*  these  designs were
 utilized in the economic comparisons  with steam plants  that  will be described in
 SECTION VIII.
                                Compound-Cycle  Designs

      The very high specific output  capabilities  associated with  the compound-cycle
 designs  (see Fig.  31b)  provided an  incentive to  estimate  representative compound-
 cycle gas turbine  power plant  costs despite the  fact that large  quantities of
 heat would have to be rejected from this  cycle via the  system intercooler.  Since
 relatively high cycle pressure levels  of  50 to 100 atmosphere are required within
 some of  the cycle  components to make this system attractive, moderate compressor
 design parameters, a maximum cycle  gas temperature of 2200 F in  both the primary
 and reheat oombustors,  and a total  cycle  pressure ratio of 75:1  were selected
 for evaluation. These  values  are representative of projections  of early-1980's
 design technology.  Since  the  low-pressure ratio compressor is driven by the
 l800-rpm power turbine  (see schematic  diagram  in Fig. 31b), system output capacity
 and hence system size was  established  on  the basis of the maximum airflow handling
 capacity of an l800-rpm compressor  with 1100 ft/sec tip speed design characteristics
 and a maximum airflow rate per unit flow  area  of 32 Ib/sec per ft2.  These basic
 assumptions provided the basis for  the design  of a nominal UjO-Mw power plant
 with approximately 2000 Ib/sec airflow handling  capability.

      The design configurations and  pertinent dimensions of the major components
 for the  l*70-Mw compound-cycle  gas turbine pover  plant are presented in Table XVIII.
 The specific outputs  associated with this cycle  configuration (see Fig. 37) are
 approximately 25#  higher than  those associated with simple-cycle configurations
 (Fig.  30)  and offer-the use of more compact turbomachinery units and, hence, the
 realization of attractive  selling prices  approaching $20/kw.  This selling price
 includes  the pricp of the  power plant  plus dry cooling  tower and the recirculating
 water cooliug systems required to dissipate the  heat of compression rejected
 from the  compressor stage  intercooler.

      The  costs  of  the intercoolers  employed in the compound-cycle gas turbines
 and the  air precoolers  employed in  the simple-cycle or  regenerative-cycle power
plants are  based on the  use of a closed cooling  water loop.  The compressor
 air is cooled with water which is heated  and in  turn circulated and cooled in a
 dry cooling tower.  This arrangement permits locating the  relatively large dry
 cooling towers  remotely  from the gas turbine engines with  little cost penalty.
The water-pooled air preheater and  intercooler sizes were  calculated using an
 existing UARL heat exchanger computer  program, and costs were estimated using
manufacturers'  data in  accordance with the heat  transfer  areas required.  The
 dry  cooling tower  costs  were estimated using the  procedure and data contained in
                                       82

-------
Ref.  6^.  These costs include allovances for fans,  motors,  and circulating pumps.
The cost of the water-cooled air precooler or intercooler for  a given system
was found to "be small compared to the dry cooling tower.  Because  even the dry
cooling tower costs were estimated to "be small compared to  the total  installed
engine costs they were checked against the steam power plant dry cooling  tower
system costs given in Ref. 6U.  After adjusting for differences in heat loads  and
temperature levels, the dry cooling tower costs estimated in this  study and those
in Ref. 6^ were found to be in close agreement.

     Although the compound-cycle power plant was estimated  to  be a most attractive
gas turbine system for base-load power generation,  with substantial potential
economic and performance merits, the use of cycle operating pressures of  50
atmospheres and above would introduce significantly higher  blade" bending  stresses
and require more extensive design analysis and cost efforts than that provided in
this  study.  It is recommended that these studies be pursued in additional programs
to assess the application of gas turbines for utility power generation systems.
                    ADVANCED GAS TURBINE STATION CHARACTERISTICS

     The results of the performance and cost studies  described  in the  previous
sections permitted the selection of simple- and regenerative-cycle engine  designs
vhich are judged potentially capable of producing low-cost  electric power  without
river and lake water thermal pollution for the 1970 and 1980  decades.   The design
characteristics, engine and overall station performance levels, and estimated
engine and station selling prices are summarized in Table XIX.   Complete details
describing the basis for the gas turbine station selling price  estimates are
presented in SECTION VIII of this report.  Station thermal  efficiency  levels of
30ft to almost 39% are projected for the simple-cycle  engines , and are  as much as
three percentage points higher for the regenerative-cycle engines. Included in
the station net heat rates are allowances for (l) the power requirements of the
station auxiliaries; (2) mechanical losses in the electric  generator;  and  (3)
reductions in gas turbine engine performance due to operation at  lower vane
temperatures and with higher exit velocities than the values  assumed in the
parametric performance studies.  These losses contributed to  a  k  to 6% reduction
in station heat rate.  No performance penalty has been included in the data for
operation at off-design conditions.  Typically, the performance of steam power
stations is reduced by about 5% to reflect off-design operation.   It is
anticipated that these losses would be negligible in  gas turbine  stations  since
the availability of multiple units in a typical large power generating station
vould provide the flexibility to meet off-design conditions.  However, the effects
°f operating at part-load conditions and at varying ambient conditions are discussed
in Appendix D.  Total station "pacific prices, in $/kw, are about 20 to 30% less
than present-day gas turbine prices.
                                        83

-------
      The engine  selling price  estimates are based on the corporate-sponsored
 computer program developed  to  estimate the manufacturing and selling prices for
 selected simple-cycle  and regenerative-cycle designs.  The devleopment costs for
 advanced gas  turbines  have  been included in arriving at the estimates of engine
 selling  price.   To  assist in these efforts, a conceptual design of a simple-cycle
 base-load gas turbine  engine vhich could be in commercial operation during the
 early part of the 1980 decade  and capable of producing 200-Mw was prepared
 (see  Fig.  53).   The engine  conceptual design is based on a compressor pressure
 ratio of 32:1, a turbine inlet temperature of 2^00 F, and an airflow of 1176 Ib/sec.
 The overall length  of  the engine would be slightly less than 60 ft.  The conceptual
 design shown  in  Fig. 53 differs only slightly with the simple-cycle engine
 incorporating the early 1980 technology advances selected for use in the comparative
 studies.   The pertinent power  plant design characteristics for this engine are
 summarized in Table XX, and the temperatures, pressures and flow rates are shown
 in Fig.  5^.   The primary dimensions of the turbomachinery were obtained from
 supplementary design procedures and from layouts of the aerodynamic flow path.

      The  station cost  estimates for the 1000-Mw simple-cycle gas turbines utilizing
 the early  1980's technology were based on the utilization of four 26d-Mw units
 designed to operate at the  conditions specified in Table XIX and Fig. 5^ for the
 engine size estimates provided in Table XX.  An elevation drawing of such a
 station, illustrating the placement of one of the four engines, is shown in
 Fig.  55.   A plan view of the station, which would require an area only
 approximately 165 x 200 ft, is shown in Fig. 56.  Arrangement of the gas turbines
was made after consideration of the space requirements needed for placement and
maintenance requirements of the engine.  The plan view shows atmospheric air-cooled
heat  exchangers  which would be used to remove heat from the compressor bleed air,
thereby precooling  it prior to its use for turbine cooling.  The turbine air
precoolers, engine  oil coolers, and generator coolers are located below the main
floor level as indicated in Fig. 55> and would be connected to the atmospheric-
air coolers outside the building.  Thus, such a power system could be operated
independently of a  cooling  water source and would require a site area about an
order of magnitude  less than the conventional system.  For example, the total
site requirements for the 1000-Mw gas turbine power plant shown in Fig. 56 might
be only 500 x 500 ft an area about 6 acres square including allowances for a
100-ft exclusion distance,  parking area requirements, etc.  This size would be
about 0.006 acres per Mw of output.  Based on data in Ref. 108, gas-fired steam
plants require about 0.10 acre per Mw of output.

     The pertinent  dimensions of the selected early 1980-decade regenerative-
cycle engine  are given in Table XX.  The temperatures, pressures and flow rates
at the various locations in the engine design are shown in Fig. 57, and a flow-
path diagram  is  shown in Fig. 58.  The gas turbine unit shown in Fig. 58
incorporates  an  80$ effectiveness recuperator, arranged in a counterflow

-------
configuration consisting of multiple nodule?  arranged seven deep by  seven in a
radial array.  The overall dimensions of the  regenerative-cycle engine would be
60 ft  long vith a maximum radial dimension of about  35 ft.  The station  dimensions
for the regenerative-cycle engine would be about  30? greater  than  for the simple-
cycle  gas turbine station, and this added area is reflected,  in part, in the
higher station selling prices.

-------
                                   SECTION  VIII
                  PO¥ER GENERATION COSTS  FOR SYSTEMS DESIGNED TO
                            ELIMINATE THERMAL POLLUTION
                                     SUMMARY
     An investigation was undertaken to estimate and  compare the costs of pro-
ducing electric power with advanced open-cycle base-load gas turbine stations and
advanced fossil-fueled steam stations designed to  reduce or eliminate thermal
pollution during the 1970 and 1980 decades.   A brief  review was made to establish
the capital charges, interest rates during construction, and maintenance and
supervision costs for the competitive systems.  Estimates were made for the total
installed station capital costs of power generating systems based upon the use of
simple and regenerative-cycle gas turbines and conventional steam turbines for
the six major FPC regions.  Detailed technical characteristics are presented for
selected advanced—cycle power stations considered  representative of types which
could be located in the South Central region, and  estimated busbar power costs
for these stations are presented and compared.  The sensitivity of the results to
the basic values used in the study are examined.   Potential advantages that would
accrue to electric utilities due to wider selection and availability of plant
sites, freedom from power plant cooling water requirements with open-cycle gas
turbines, and reduced transmission and lower reserve  margin were identified and
evaluated.  Estimates are provided of the time and approximate cost required to
develop commercial base-load gas turbine power stations.  These development costs
are compared with the incremental capital costs which electric utilities will be
obliged to spend through 1990 for cooling towers and  cooling ponds to eliminate
thermal pollution from conventional steam plants.

     For the purposes of this study, power stations with nominal 1000-Mw capacity
were selected as the basis for the total owning and operating cost comparisons.
These comparisons were focused on stations which could be installed in the
South Central region during the next two decades.  In this region natural gas
is projected to be readily available at moderate price levels, the cooling water
shortage is already acute, and the growth of demand for electric power generation
is expected to continue at or above the present rate.
                                      87

-------
                    CAPITAL INVESTMENT AND OPERATING COSTS  FOR
                        ADVANCED POWER GENERATING SYSTEMS
                                 Steam System Costs

      Capital and operating cost estimates are presented in the  following para-
 graphs for conventional coal-,  residual oil-, and natural gas-fired steam pover
 generating stations vith performance and size characteristics identified in
 Section VI to be consistent with stations available  for commercial operation in
 the 1970 and 1980 decades.   The data from various references summarized in Section
 VI of this report provide the basis  for the steam station costs.  Station costs
 include all p3ant equipment up  to and including the  main station transformers
 as veil as the various  factors  for escalation, interest on capital during construc-
 tion, etc.  The costs are presented  in terms  of 1970 dollars for stations
 capable of providing 1000 Mw of electric power.

 Station Investment and  Total Installation Costs
      Station investment  and total  installed costs were estimated for coal-,
 residual oil-,  and natural gas-fired  steam power generating stations projected
 to be representative  of  those  in commercial operation during the 1970 and 1980
 decades.   The steam stations placed in operation during the 1970 decade would
 consist  of two  500-Mw units, each  operating at 2^00 psig/1000 F/1000 F steam
 conditions.   The basis for these projections and the performance characteristics
 are presented in Section VI  of this report.  The actual investment and total
 installed station  costs  are  based  on  data presented in Refs. 109, M, and 1+5-  For
 example,  data are  presented in Ref. U5 for a typical coal-fired station located
 at a  mine-mouth  site  in  the  East Central region.  The same reference also provides
 extensive station  cost data  for oil-fired and natural gas-fired plants which
 reflect typical  installations  located in the Northeast region.  On the basis of
 these data,  plus that available in Refs. 109 and M» costs were estimated for
 1000-Mw fossil-fueled power  generating plants that could be in commercial opera-
 tion  during  the  1970  and 1980  decades.  The installed costs of stations in-
 corporating  design  advances  projected to become available during the 1980 decade
were  estimated by applying cost scaling factors to the itemized station cost
 estimates  of the comparable  present-day designs, described in Section VI, to
 account for  (l)  differences  in unit size, and (2) for projected improvements in
boiler design and operating  efficiency.  All of the coal-, oil-,  and gas-fired
stations incorporate  design  features that provide full protection against the
weather.  Finally, a period  of four years from date of planning to date of opera-
tion, as determined from actual construction practice, served as  a basis for
station cost estimating purposes.
                                       88

-------
     Installed cost summaries for the 1980-decade station designs are  presented
in Table XXI according to standard FPC account categories vhich include  land
and land rights, structures and improvements,  toiler plant equipment,  turbo-
generator units, accessory electric equipment, miscellaneous  pover plant equip-
ment, and miscellaneous station equipment.  The cost of natural draft  vet cooling
towers (approximately $8/kw) has "been included for each of the  three stations
and is attributed to FPC Account 31 ^ Turbine Generator Units.   The costs of
750-ft and 600-ft stacks, vith appropriate electrostatic precipitators or
mechanical cyclone dust collectors, have also  been included for the coal- and
oil-fired units, respectively (see Item 312).   Also included  in the cost summary
are other expenses associated vith startup, testing, temporary  facilities,
temporary buildings, removal of temporary facilities,  cleanup,  and security
guards.  Indirect construction expenses for engineering, design,  construction
supervision, and contingency as veil as escalation (at 3.5% per annum  compounded),
and interest during construction (at Q% per annum) are shown  in this summary
table.  It can be seen from Table XXI that the indirect costs account  for
approximately 33^ of the total station cost.   The total station costs  vhen all
factors have been included amount to approximately $165.6,  $153.9,  and $13J.9/kw
of installed capacity for the coal-, oil-, and gas-fired plants,  respectively.

                             Stat in Csts
     Station cost estimates for coal-, oil-,  and gas-fired 1000-Mw  steam-electric
station designs that might be installed in each of the  six FPC power  regions  were
Eade by applying appropriate correction factors based on  Handy-Whitman  correla-
tions, furnished by the architect-engineering firm of Burns and  Roe,  Inc.   These
regional installed station costs are presented in Tables  XXII and XXIII  for
1970-decade and 1980-decade designs, respectively.

     Station costs based on outdoor construction as well  as indoor  construction
have been included in Tables XXII and XXIII to provide  realistic bases  of  compari-
son between systems in those regions of the country where indoor construction
features may not be needed.  As indicated in  Tables XXII  and XXIII, the outdoor
plants provide a net $15/kw saving in total station installed cost  over the indoor
designs.  This saving in total installed cost results from a $10 /kw reduction in
building cost (FPC Account 311), substantiated in Refs. 29 and 108  to arise from
the elimination of much of the enclosure around the powerhouse.  The  1980-decade
designs installed costs in Table XXIII are also presented for different rates of
interest during construction to illustrate the effects  of changes in  interest
rate on construction costs.  Whereas 6.25$ was the existing interest  rate  only
two years ago, these rates have currently climbed to  8$ and higher.   A  comparison
°f Tables XXII and XXIII shows that the 1980-decade stations enjoy  approximately
a 10/5 decrease in station cost by comparison  with the 1970-decade systems.  This
cost saving is achieved from the improved steam conditions and from the economies
°? scale.
                                      89

-------
 Annual Owning and Operating  Costs

      The total annual owning and operating costs for the steam-electric stations
 as  well as for the gas turbine power stations described in the following paragraphs
 are expressed in terms of mills/kvhr and are equivalent to the busbar power cost.
 In  order to determine busbar power cost, the annual hours of operation were selected
 on  the basis of an annual load factor of 1Q% (equal to 6l32 hr/yr), consistent with
 present experience in conventional-fueled, base-load, steam-electric systems.
 Although load factors as  high as 80$ are common in nuclear-fueled base-load plant
 operation,  fossil-fueled  plants tend to achieve a lower utilization per year.  The
 busbar power cost consists of three (see Refs. 108 and 2) items as follows:
 capital charges,  operating and maintenance charges, arid fuel charges.  The capital
 charge is the yearly  owning  cost and includes allowances for interest on borrowed
 capital, amortization,  insurance, taxes, annual maintenance, and depreciation.
 A total annual fixed  charge  of 15% (based on the items in Table XXIV), consistent
 with current cost estimating practice was used as the basis for computing capital
 charges.   These charges are  higher than past practice indicates but as noted in
 Ref.  110 are the result of today's high interest rate levels.  Municipalities or
 federally owned utilities capable of borrowing at lower interest rates would use
 capital charges several points lower than 15$.  Supplies and materials were
 assessed at a constant cost  level of 0.200 mills/kwhr; however, operating and
 maintenance labor costs were selected as 0.160 and 0.172 mills/kwhr for the 1970-
 decade oil- and gas-fired stations and the coal-fired stations, respectively.
 These  costs  are consistent with similar data published in the available literature
 and have been substantiated  through discussions with engineering-architect firms
 (Refs.  2, 29,  H5,  52,  108, and 109).  It was assumed that the utilization of a
 single 1000-Mw unit in the 1980-decade designs in place of the two 500-Mw 1970-
 decade units  along with anticipated advances in system designs could afford a
 0.05 mill/kwhr reduction in  the operating and maintenance cost for the 1980-
 decade designs.   Fuel  costs were calculated on the basis of projected fuel prices
 in  Section VT  of this report and net station heat rates that were modified to
 compensate  for part-load operation and station startup.
                             Gas Turbine System Costs

     Detailed capital and operating cost estimates (based on 1970 dollars)  are
presented in the following paragraphs for simple-cycle and regenerative-cycle,
natural gas-fired 1000-Mw gas turbine power stations.  The capital costs  for the
gas turbine systems include all plant equipment up to and including the main
station transformers.
                                         90

-------
Capital Costs

     Detailed cost estimates were made for a number of engine  designs  reflecting
the 1970 and 1980 decades.  The results of these estimates vere  used as  tne  oasis
for the preparation of Table XVIII.  Detailed capital cost breakdowns  are presented
in Table XXV for early 1980's-decade simple-cycle and regenerative-cycle 1000-Mw
gas turbine power systems.  As previously mentioned,  the  gas turbine station designs
(see Figs. 55 and 56) comprise four 260-Mw units for the  simple-cycle  configuration
and five 208-Mw units for the regenerative-cycle configuration.  The detailed cost
breakdowns in Table XXV are presented according to the standard  Federal  Power
Commission (FTC) account categories described in the preceding section of this
report.  The stations were designed and cost estimates made for  full protection
against the weather.  The costs of the outdoor construction designs were then
estimated by applying a 30% correction factor to the cost of structures  corres-
ponding to the indoor designs.  Since the gas turbine designs  are  relatively simple
by comparison with the steam-electric stations and,  in addition, many  of the major
components (gas turbine, generators, etc.) are delivered  to the  site in  modules,
a construction time (as previously defined date of planning to date of operation)
of two years was considered to be adequate for cost estimating purposes.  Present
gas turbine installations using aircraft-type engines are generally completed in
slightly more than one year.

     Summaries of early 1980-decade station costs are presented  in Table XXVI.  It
can be seen from this table that differences in costs between  indoor and outdoor
construction amount to only about $1 to $1.5/kw due to the relatively  compact
construction associated with the gas turbine station designs.  In  addition,
primarily as a result of the shorter construction time for the gas turbine plants,
the sum of the costs for indirect expenses (i.e., engineering  design,  escalation,
and interest during construction) accounts for approximately 21% of the  total
installed gas turbine station costs; this compares with the aforementioned 33^
value for the steam stations.

     Since on-site construction costs are held to a minimum with the gas turbine
plants, due to the fact that most components are assembled before  they arrive
at the plant site, regional differences in installed station costs are not as
significant for these stations as they were for the steam power  stations (see
Table XXIII).   Therefore, gas turbine installed station  costs were assumed  not to
vary among regions.  The total installed costs for the early-1980's design
simple-cycle and regenerative-cycle designs when all factors have  been included
amount to approximately $66 and $92/kw, respectively.  Similar cost estimates were
s^de for the 1970-decade and late-1980 decade   station designs  by applying
appropriate scaling factors to the FPC account categories for  the  early  1980-
decade designs.  Total installed costs of $80 and $100/kw were predicted for 1970-
decade simple-cycle and regenerative-cycle gas turbine power station designs,
respectively, and $71 and $93/kw for the late-19801s simple-cycle  and  regenerative-
cycle designs, respectively (see Table XIX).
                                        91

-------
 Annual Owning and Operating Costs

      Annual owning and operating costs  for  simple-cycle and regenerative-cycle
 1000-Mw gas turbine power  systems  assumed to be placed on line during each of the
 three time periods of interest vere  computed for those regions of the country in
 vhich natural gas now is or will in  the future become a competitive source of fuel.
 The annual capital charges (mills/kwhr) were determined on the basis of an 8%
 interest rate during construction, 15%  per  annum fixed charges, and a fQ% average
 annual load factor.  Fuel  costs (mills/kwhr) were  calculated on the basis of
 projected fuel prices for  different  regions presented in Section VI of this report,
 and from net station heat  rates which include an adjustment of k% to compensate
 for generator losses and other house loads, and appropriate correction factors
 which include the small performance  penalties previously described in Section VII.
 Supplies and materials were estimated to account for 0.200 mills/kwhr.  This is
 the same value estimated for the steam-electric stations.  Operating and maintenance
 costs were estimated at 0.500 mills/kwhr and 0.600 mills/kwhr for the simple-
 and regenerative-cycle stations, respectively, on the basis of discussions with
 the Burns and Roe, Inc., architect-engineering firm that had conducted a survey
 of the operating  and maintenance costs  of gas turbine plants in operation and data
 presented in Refs. 52,  72,  and 111.  The results of this survey indicated that
 selected actual maintenance costs vary  from a minimum of 0.5 mills/kwhr for base-
 load type operations to a  level of about 1.5 mills/kwhr for some peaking plants.
 It is anticipated that  with careful  design  and attention to maintenance-saving
 features in the engine, the maintenance costs would equal the 0.5 mills/kwhr level.
                      COMPARISON OF POWER GENERATION COSTS
     Although the advanced open-cycle gas turbine power systems would provide a
means of eliminating thermal pollution, there are alternative cooling systems
available for use in steam plants which can greatly reduce or eliminate thermal
pollution as well.  Thus the ultimate acceptance of advanced gas turbine
generating systems will depend on the possibility of producing the lowest-cost, power
in competition with steam plants using the alternative cooling systems.  In the
following paragraphs the total busbar power generation costs (including all the
owning and operating cost elements) are presented for the advanced open-cycle
gas turbine systems and steam turbine systems equipped with cooling devices designed
to substantially reduce thermal pollution.  The primary comparisons and analyses
are based on natural gas-fueled stations of 1000-Mw net electrical output located
in the South Central region of the country.  The South Central region was selected
for the comparison since natural gas is abundant and will be available in this
area for the next two decades and water for station cooling is often in short
supply and will be more difficult to obtain.  The comparisons are also generalized
                                         92

-------
to other regions of the country.   No particular  site in the South Central region
was selected but a typical site might be that  chosen by the Houston Plant Lighting
Pover Company for the Cedar Bayou Generator Station (Refs. 112 and 113).  This
plant complex vill ultimately contain 6 units  totaling 5000 Mw and vill utilize
a 2600-acre cooling pond which discharges water  at 95.7 F.  Based on the performance
and cost data presented in Section VI for alternative condenser cooling systems,
the cooling ponds may be considered approximately equivalent to the use of cooling
towers.
                       South Central 1970-Decade Stations

     A tabulation and comparison of the total busbar power generation costs among
1000-Mw conventional steam turbine stations  and advanced simple-cycle and regenerative-
cycle gas turbine stations which are projected to be in commercial operation during
the 1970 decade in the South Central region  are presented in Table XXVII.  The
results are based upon gas costs of 23^/million Btu, the gas level projected in
Section VI for South Central gas in the 1970 decade, 15% capital charges, and
?0$ load factor.  The Table XXVII results indicate approximately a 0.5 mill/kwhr
lower busbar power cost for the simple-cycle gas turbine station in comparison to
the steam turbine station.  The power costs  for the regenerative-cycle gas turbine
station would be approximately 6% higher than the costs for simple-cycle gas turbine
stations but still less than those for the steam station.  The principal advantage
of the gas turbine stations is the substantially lower station capital costs, which
at the present capital charges of 15$ would  result in almost 1.25 mills/kwhr lower
capital costs for the simple-cycle gas turbine station in comparison to the steam
station.
                    South Central Early 1980-Decade Stations

     The comparison of busbar power costs among the various systems projected to
be available for the early 1980-decade is shown in Table XXVIII.  The data indi-
cate that the busbar power costs would be the lowest  for the simple-cycle gas turbine,
in comparison to the steam turbine and regenerative-cycle gas turbine.  The
cost difference would be more than 0.80 mills/kwhr in favor of the gas turbine
station.  Although the installed costs of the steam station were reduced by almost
$12/kw between the 1970-decade and early 1980's designs, the installed costs of
the gas turbine were reduced by approximately the same amount.  In addition, the
net plant efficiency of the gas turbine station increased by more than 5 percentage
Points as described in Section VII, while the efficiency of the steam plant in-
creased by only 2 percentage points.  Thus,  even though gas costs are projected to
increase to the 30<£/million Btu level, the low installed costs of the gas turbine
Cation would offset the relatively high fuel and maintenance costs and still
provide a minimum-power-cost system.
                                       93

-------
 Sensitivity to  Economic  Factors

      Although the values selected herein for capital charges, natural gas costs,
 interest  charges, and  escalation rate are based on careful projections, a number
 of outside  influences  such as the general economic condition could change these
 factors considerably.  Therefore, the sensitivity of the results presented in
 Table XXVIII to variations in several important values vere considered.  The
 variation in busbar pover costs with changes in the capital charges and cost of gas
 are indicated in Figs. 59a and 59b, respectively.  The Fig. 59a results indicate
 that even with  a reduction in capital charges to the 10$ level (a. level which was
 considered  adequate only U or 5 years ago), the simple-cycle gas turbine station
 would have  power costs about O.UOO mills/kwhr lower than either the steam or re-
 generative-cycle gas turbine station.  An increase in the capital charges to 17$
 from 15$  would,  of course, only widen the cost difference between the steam station
 and simple-cycle gas turbine.  The effect of a variation in natural gas costs
 between 20  and  50^/million Btu on the total busbar power costs are shown in Fig. 59b
 and indicate only a minor reduction in the cost advantage of the simple-cycle
 gas turbine stations relative to the other types of power systems at the highest
 levels of gas costs*.

     Only recently has the short-term interest rate climbed to the high 8 to 10%
 levels presently  experienced by the electric utility and other industries.   The
 effect of utilizing a 6% interest rate during construction on the power costs of
 a steam system are shown by the lower line  of the band shown in Fig.  60a.   The
 upper line  shows the busbar power costs with 8% interest during construction as
 used in Table XXVIII.   The dashed line in Fig.  60a depicts the increase in busbar
 power costs as the simple-cycle gas turbine station costs are increased above the
 $66.5/kw level shown in Table XXVIII.  Only at about $93/kw or a level  hQ% above
 that estimated in the study would the total busbar costs of the simple-cycle
 gas turbine station begin to approach that  of the steam plant.   Figure  60b shows
 the effect of changes  in the heat rate of the steam plant on the total  busbar
power costs.  The results indicate that even at  a heat rate of 7500 Btu/kwhr, the
total busbar power costs  of the steam plant would be about 0.1*00 mills/kwhr above
those of the simple-cycle gas turbine plant presented in Table XXVIII,  and about
0.300 mills/kwhr above that of a gas turbine which had a heat rate about 5$
poorer than that level projected in Section VII.
   Fuel costs of 50^/million Btu would stimulate interest  in  using nuclear-fueled
   stations for base-load operation and restrict the use of fossil fuels only for
   the swing-load sector of the load demand (x Uo$  load factor).

-------
                     South Central Late 1980-Decade Stations

     The estimated "busbar pover generation costs  for steam  and gas turbine  systems
that could be commercially available by the late  1960's  for installation  in the
South Central region are shown in Table XXIX.   The cost  of  natural gas  used in
these stations is Uo^/million Btu, the highest  level projected for natural  gas in
the South Central region in Section VI of this  report.   The Table XXIX  results
indicate that the simple-cycle gas turbine station costs vould produce  power at
the lowest busbar costs; approximately 0.1*5 and 0.85 mill/kwhr lower than that of
the regenerative-cycle gas turbine and steam turbine stations, respectively.
                                 Other Regions

     Although the advanced open-cycle gas turbines  are projected to provide  lower
busbar power generation costs than steam systems  in the South  Central  region where
relatively low-cost gas is available, in the other  regions  of  the US the  cost  of
fuels available for use in steam and gas turbine  utility  systems can be considerably
different.  For example, on the Pacific Coast in  the West Power Region of the  US
and especially in Southern California, both residual oil  at 28^/million Btu  and
natural gas at 3^<£/million Btu are projected to be  available for use in the  early
1980's (see Section VT).  In some of the Rocky  Mountain states both coal  and gas
would be available, although the cost of gas would  be about 20<£/million Btu  above
that for coal.  In the Northeast region coal, residual oil, and gas either in  the
form of imported LNG or produced as synthetic from  coal or  via pipeline from the
Southwest are projected to be available for utility use.  The  cost of  the gas
would be in the 50 to 60$/million Btu range, while  low-sulfur  oil might cost
20tf/million Btu less than the gas.  Using the fuel  COST; projections presented  in
Section VI, and the station installed cost and  performance  characteristics
determined in Section VII and VIII, estimates were  made of  the total busbar
power costs for the competing systems located in  different  regions of  the US,   The
results are presented in Figs. 6la and 6lb for  the  Northeast and West  regions  of
the country, respectively, for steam and simple-cycle gas turbine systems that
vould be commercially available in the early 1980's.   The Fig. 6la data indicate
that in the Northeast region the simple-cycle gas turbine station would generate
power at slightly higher costs than the steam systems for load factors greater
than about 60$.   However, at load factors below 60$,  substantially lower  power
costs would be realized with the simple-cycle gas turbine stations in  comparison
to the steam stations, in spite of the 20tf/million  Btu higher  fuel costs  for the
gas turbine station.  For example, at a hQf<, load  factor the cost of producing
power could be 8.3 mills/kwhr for the simple-cycle  gas turbine and more than 9-5
Eills/kwhr for the steam station using residual oil.

     Approximately the same results are illustrated in Fig. 6lb for competing
stations located in the West region.  The steam stations would produce power at
                                        95

-------
 lower costs than the gas turbine stations  only  for load factors of 10% and above,
 in spite of the substantially lover fuel costs  that are projected for coal and oil
 relative to natural gas.

      Similar results were obtained for comparisons of power generating costs in
 the Southeast (not shown), West Central, and East  Central regions of the country
 (see Fig. 62).   These areas are heavily dependent  on coal as a fuel source and
 although limited supplies of natural gas are available,  the higher prices for it
 would make gas  turbines attractive only for  those  utility applications where the
 load factor is  projected to be less than about  60%.   However, this mid-range or
 swing-load market will assume further importance as  nuclear steam power systems
 are more widely used for the high-load-factor,  base-load applications.  But it
 must be cautioned that not all areas of each region  will have access to a suffi-
 cient supply of natural gas for power generation applications.  Furthermore, the
 air pollution regulations must be pursued  vigorously to  prevent reverting to the
 use of high-sulfur-content, low price coal and  oil fuels. If low-cost, high-sulfur
 fuels were used,  their prices would be substantially lower than those for natural
 gas, and it is  likely the gas turbine would  not be competitive.
                             Use  of Dry Cooling Towers

      Although  neither the mechanical- or natural-draft type of dry cooling tower
 is  anticipated to be widely  used during the next two decades in any of the US
 power regions,  except possibly the arid portions of the Western states, the effects
 of  their  use on the  economic comparisons presented herein can be estimated readily
 from the  data  presented in Section VI.  For example, the installed capital costs
 of  a fossil-fueled station equipped with mechanical-draft dry cooling towers would
 be   $8/kw to $12/kw  higher than  that for a similar station utilizing natural-draft
 wet cooling towers (see Table XIII and Ref. 60).  For capital charges of 15/S and
 a load factor  of 70#, this would add almost 0.20 to 0.30 mills/kwhr to the steam
 station busbar costs presented in this section.  The added operating costs for a
 steam station when mechanical-draft dry towers are used rather than wet towers
 are presented  in Table XIV of Section VI.  Depending upon the region considered,
 the added operating  costs could be from 0.10  to as much as 0.3^ mills/kwhr* Thus,
 the added costs  for  a dry cooling tower would be from approximately 0.30 to
 0.65 mills./kwhr.  Even if the lower cost figure were added to the steam station
 busbar costs shown in Figs.  6l and 62, the competitive position for the gas turbine
 would  be  enhanced.   The only advantage possible for the steam station would be the
 utilization of  low-cost coal or  lignite directly in the boiler, whereas this would
be  impossible with the gas turbine.  However, if a suitable gas turbine fuel
were available, the  use of dry cooling towers in a steam system would permit the
 utilization of higher-cost fuel in the gas turbine.  Each 0.1 mills/kwhr in added
busbar cost for the  steam plant would allow the use of fuel costing about 1^/million
Btu more in the gas turbine.
                                        96

-------
               POTENTIAL SITING, TRANSMISSION, AND RESERVE MARGIN
                          ADVANTAGES OF GAS TURBINES
     The successful operation of gas turbines without the use of  cooling water  can
not only provide a possible solution to the thermal pollution problem but also
produce other advantages in a utility system.   For example, since a gas turbine can
operate essentially independent of a supply of  cooling water, considerably more
flexibility is possible in the selection of station sites than is permissible with
conventional steam power stations.   With this extra flexibility the utility
planning engineer vould have the freedom to locate the gas turbine plant to take
advantage of inexpensive land costs, or alternatively to be closer to a. load
center, to a system transmission network, or to a fuel supply.  Furthermore, the
compactness of the gas turbine system together  with the flexibility of site
selection which it allows could provide the basis for a highly reliable integrated
network comprised of moderate-sized units.   Since only 50$ of the cost of providing
power to the consumer is attributable to the cost of power generation and the
remaining 10$ and Uo$, respectively, are attributable to transmission and distribu-
tion costs, these costs must also be considered in an analysis of competing power
systems.

     A brief study was conducted to determine some of the potential advantages
that could accrue to an electric utility through the use of open-cycle gas turbines
for base-load operation as a result of the smaller unit capacities of gas turbine
engines and the wider selection and availability of plant sites made possible
through the elimination of power plant cooling  water.  To assess  the full advantages,
the investigation included:  (l) the effect of  power generating unit size, system
unit mix, i.e., the variations of unit size within a utility system, and unit
forced outage rate on system reliability and reserve requirements; (2) the effect
on system reliability and cost of expanding system capacity in small unit sizes;
and (3) the effect of unit size and location on station-transmission network tie
requirements.  In the following discussion cost credits are presented which can
be attributed to a base-load gas turbine generating plant that has freedom of
location, and whose units are smaller in generating capacity than a conventional
steam-powered generating station.
             General Transmission and Distribution  Considerations

     The transmission and distribution (T&D)  systems  transfer power efficiently
    reliably from the generating plant to the consumer.  The transmission  system
Performs several functions including (l) the  transfer of large blocks of power at
high voltage levels (115 kv to 765 kv) from generating plants to areas  of  high
load density; (2) the intraconnection of generating plants  and substations;  and
(3) the interconnection of neighboring utility systems.  Alternatively  the distri-
bution system distributes the electric power, at a  lower voltage level  (69 kv or
iess) to the consumer (Ref. II1*).
                                          97

-------
     The system  consists largely of circuits and substations.  The circuits may
be  located  either  overhead or underground, and comprise extra-high-voltage (EHV)
transmission-, regular transmission-, subterranean transmission-, primary distribu-
tion-, and  secondary distribution-lines (Ref. 115).  The substations contain the
transformers used  to step down the voltage, buses for interconnecting transformers
and circuits, and  circuit breakers for the automatic disconnect of faulty trans-
formers or  circuits.  In addition to these equipment there are also ancillary
equipment including series and shunt capacitors, shunt reactors, current and
voltage transformers, relays, air-break switches, communication equipment, panel
boards with meters, etc.

     Transmission  and distribution system failures can be caused by natural hazards,
by  man-made hazards, by over-voltages generated within the system as a result of
switching operations, or by maladjustment or failure of system components.  Circuits
and equipment are  designed individually to provide good reliability.  Adequate
system reliability and maintainability are achieved largely by the provision of
redundant circuits, circuit breakers, buses, and transformers.

     Probability methods similar to those outlined in Refs. Il6 through 121 are com-
monly   used by the utilities in an effort to determine the degree of reliability
and thus the reserve requirements* needed for a system.  A loss-of-load value of
0.2 days/yr was used in this study based on Refs. 117 and 122.  This value repre-
sents the probable number of days per year that the demand exceeds the available
generating  capacity.  For example, the probability of load exceeding capacity on
0.2 days/yr may also be stated as:  one day in 5 years or alternatively as a
loss-of-load probability of 0.055$.

     Present T&D systems have, generally speaking, not been designed in the normal
sense of the term, but are the outgrowth of an extensive series of planned addi-
tions.  As  a transmission system grows to meet increasing and expanding loads with
the  addition of substations and generating plants, the paths for the new transmission
circuits (rights-of-ways) and the points at which they terminate are selected so
as  to provide a total integrated generation-transmission system which is flexible,
stable, reliable, and economical.  This system which evolves is a complex inter-
laced network of circuits commonly referred to as a transmission grid.  Most of the
transmission systems in and around large urban areas have reached the grid stage
in their development.   However,  in many urban areas the space needed for trans-
   A proper appreciation of the reserve  requirements  for  an  electric power system
   is quite important and discussed in detail  in  Ref. 123.   Two different "rules
   of thumb" are usually stated as  representing reasonable approximation of system
   installed reserve; they are:   (l)  the installed  capacity  should be at least 15$
   greater  than the annual peak load and (2) the  reserve, i.e., the excess capacity
   above that required to meet  the  annual peak load,  should  be at least as great
   as the sum of the two largest units in the  system  (Ref. 122).
                                        98

-------
mission lines is scarce and is becoming a significant  factor in the selection of
station sites.  In Ref. 29 and in discussions with utility planners, the need for
providing sufficient rights-of-ways for transmission lines has been continually
stressed.

Effect of Unit Output Capacity on System Reliability

     The loss~of-load probability method vas employed  to establish the effect of
unit output capacity on reliability.   The first system considered vas assumed to
be homogeneous, i.e., consisting of only units that are all of the same level of
output capacity and all having the same forced outage  rate.  The unit output
capacities for the homogeneous systems were selected as follows:  20$, 10$,
5$, and 2% of total installed capacity.  Therefore, a  typical 1000-Mw system could
consist of either five units each capable of providing 20% of the system capacity,
or ten units of 10$ capacity each, etc.  The forced outage for all units was
assumed equal to 2.% of the required operating time (Refs. Il6, 117, 120 through
122).  Conversely, the unit would be available 98$ of  the required time.  The
results of the probability analysis that was performed is presented in Fig. 63.
This figure clearly shows the justification for the trend  toward unit sizes of
about 10$ or less than the total system rather than toward the 20$ units.  A
system comprised of units each capable of providing 10$ of the total system capacity
could meet the loss-of-load criteria and still have an annual peak load to
installed capacity ratio of about 0.82.  This result is not totally unexpected in
that given the same forced outage rate it is quite reasonable that the disabling
of one large unit, say a 20$ unit, would more likely occur than the disabling of
ten 2% units. The same arguments would apply whether the units were steam or gas
turbines.  However, in actual practice the cost savings inherent with economies
of scale (see Fig. 10) and the reliability provided by intertie connections is the
major reason for not adopting the use of many small units.  However, if it is
assumed that the competing systems are initially  cost  competitive, that is the
small units can produce power for the same cost as the large units, the advantages
described herein can be considered over and above the  costs.  The competing systems
    be both of the gas turbine or steam turbine type or combinations of the two.
Effect of Degree of Mix and Forced
Outage Rate on System Reliability

     Having established, from a reliability standpoint, the desirability of having
units consisting of 10% of the system capacity, the  effects of forced outage rate
and unit mix on loss of load were explored for the desired reliability of 0.2
3ays/yr.  The forced outage rate equal to 0.02 (fraction of "required" operating
tine that system is experiencing a forced outage) was held constant for these
10/J units.   An additional system was then synthesized having the same total
installed capacity. However, the makeup of this second system is slightly
Afferent and consists of a mixture of eight 10$ units and ten 2% units.  This
                                        99

-------
 system is  called the mixed system since units are included which represent
 different  percentages  of total system capacity.  The 2% units, which could repre-
 sent a reasonable ratio  of gas turbine power to conventional steam power in a
 utility system,  were assumed to vary in forced outage rate from 0.02 to 0.10
 (Ref.  123).

      It is important to  note that in both systems all the installed capacity,
 except that  capacity which is down for either maintenance or as a result of a
 forced outage, is available to supply the system load at the time of peak annual
 load.

      The results of the  comparison between homogeneous systems consisting of units
 having capacities of 10? of the system capacity and the mixed system is shown in
 Fig. 6k&.  The homogeneous system (i.e., having ten units each comprising 10? of
 the total  system output) with a constant forced outage rate of 0.02 meets the loss-
 of-load requirements of  0.2 days/yr if the annual peak load is approximately Ql.1%
 of the system installed  capacity.  Thus the system must have reserves equal to
 18.3?  of the installed capacity.  For example, if the peak load of a utility systea
 were 817 Mw, the installed capacity would have to equal 1000 Mw to provide reserves
 equal  to 183 Mw  or 18.3? of installed system capacity.  The mixed system with a
 forced outage rate of  0.02 for the 2% units achieves the same loss of load at a
 higher annual peak load  percentage (82.7?)-  Therefore, this system would require a
 reserve margin of only 17.3?.  Therefore, the mixed system could meet the peak
 load requirements for  817 Mw (as presented in the example described above) with
 only about 990 Mw of installed capacity or 10 Mw less capacity than the homogeneous
 system.  The advantage for the mixed system decreases as the forced outage rate of
 the 2% units increases and vanishes entirely when the forced outage rate reaches
 about  0.06.   If  the  forced outage rate of the 2% units exceed 0.06, the mixed
 system is  less reliable  than the homogeneous system.  At a forced outage rate of
 0.10,  the  annual peak  load of the mixed system can be only 80.5? of installed
 capacity or  conversely a reserve of 19-5? is needed to maintain the same loss-of-
 load probability.  Figure  6Ua thus illustrates one very important fact:   that a
 mixed  system can have a  greater reliability than a homogeneous system even though
 the reliability  of some  of the units in the mixed system are poorer than those in
 the homogeneous  system.

 Mixed-System Cost Credits

     Although the mixed  and homogeneous  systems can meet the same reliability
 criterion,  the installed capacities  of each system could be different depending on
 the  forced outage rates  of the individual units.   Figure 6^b presents the cost
 credits (or  deficits) that can be assigned to the mixed system as a result of the
 differences  in capacity  for such a system relative to a homogeneous system.   The
 results are presented for forced outage  rates varying from 0.02 to 0.10  for the 21>
units in the mixed system.  For example,  if the 2% units in the mixed system had s.
 forced outage rate of 0.02, approximately 10 Mw less  capacity could be installed
relative to the  homogeneous system without  a loss  in overall- system reliability.
                                        100

-------
If the fixed charges are assumed to be 15* and the  average system thermal  efficiency
is equal to 36/f, the reduced capacity can be converted into a  fuel cost credit
as shovn in Fig. 6Ub.  Figure 6kb illustrates that  there exists a positive allowable
fuel cost increment for the mixed system if the forced outage  rates of the unit  are
less than 0.06.  The magnitude of this positive increment is less than l£/million
Btu, with values of approximately 0.10 to Q.30#/million Btu at a load factor  of  0.8.

Installed Capacity Expansions

     To determine vhat effect the utilization of the homogeneous system or mixed
system vould have on future expansion* a situation  vas synthesized where the
systems would have to expand their original installed capacity by 10%.  Thus  the
expansion could be obtained by either of tvo modes  for the homogeneous system.
The system could be expanded by adding another unit of 10JJ of the original installed
capacity and having a forced outage rate of 0.02, or it could be expanded  by
adding five units, each unit having  2%   of the original installed capacity
and having a slightly higher forced outage rate of  0.03.  In the case of the  mixed
system, the system was expanded by adding five more 2% units.  The results of the
system expansion are shown in Fig. 65a.   Again, the reliability is increased  by
use of a mixed-unit concept.  In fact, a comparison of Fig. 65a with Fig.  6^a shows
that, for a forced outage rate of 0.03,  the mixed system enjoyed a 0.8? reliability
advantage over the homogeneous system before expansion; while after expansion the
advantage has increased to about 1.0% over that of  the homogeneous system.  It is
also seen that the system that was initially homogeneous and utilized a smaller-
unit concept in expansion now enjoys a 0.5^ advantage over the system which
adhered to the homogeneous unit concept.  A cost analysis was performed as described
previously and the results are displayed in Fig.  65b.  The results indicate that
the system which employs a unit mix with large and  small units proves less expen-
sive to own and operate than a system of large units with no mix but of comparable
forced outage rate.
                   Effect of Expansion to Meet  Future Demands

     There is another manner by which system expansion  can be evaluated.  This
method is discussed in Ref.  121* and is utilized here to determine the effect on
electric utility system cost of frequent expansion  in small, low-capital-cost/
kilowatt units such as a gas turbine, as opposed  to less  frequent expansion in large
Mgh-capital-cost units typical of conventional steam systems.  Expansion of the
utility system in smaller units reduces the amount  of capacity in excess of system
demand carried at any one time.  Thus, the excess cost  incurred by producing
additional capacity beyond current system requirements  is eliminated.  It is
clearly a more advantageous  situation for an electric utility to expand its capa-
city by frequently adding units which match load  demand rather than being forced
to add large units simply to gain economies of  scale as would be the situation if
expansion is by means of conventional steam power.
                                         101

-------
      The savings to the utility system which accrue  as  the  result of expanding in
 units which much more closely match tne demand curve is presented in Fig. 66a.
 The results were based on the use of small capacity  units costing only $70/kw
 (i.e., gas turbines) and indicate savings of about 0.25 mills,/kwhr.  The
 potential mills/kwhr savings  shown in Fig.  66a which could  result when expanding
 with small units can be converted into a fuel cost savings  by assuming a load
 factor, fixed charges, and thermal efficiency.   These fuel  cost savings are
 presented in Fig. 66b.  For a system load factor of  0.8, Fig. 66b shows that the
 savings amount to an allowable fuel cost increment on the order of 2.2^/million Btu
 for a system expanded in small (2% of initial system capacity) units, rather than
 in large (10$ of initial system capacity)  units.
                 Effect  of Unit  Size and Location on Transmission
                          and Distribution System Costs

      The last  potential advantage considered involved the determination of the
 effect  of unit size  and location on the system-transmission network  require-
 ments.   To determine this effect, two systems were synthesized; they are presented
 in Fig.  67.  The first  system,  Case I, consists of a single-unit central generating
 plant in close proximity to  the system grid lines.  This single-unit generating
 plant contains 10$ of the system installed capacity.  However, this central station
 is required to transmit  its  power output over some distance to the load center.
 To transmit this electrical  power over the distance, a step-up transformer is
 required at the central station and a step-down transformer at the load center.
 The transmission network tie with the station-load system is at the central
 generating station on the high  voltage side of the transformer.  The tie line costs
 were  assumed negligible  because of the proximity of the central generating plant
 and its  transmission lines to the transmission network.

      The second system,  Case II, involves a multiple-unit central generating plant,
 such  as  that which would occur  if gas turbines were utilized.  This multiple-
 unit  central operating plant also contains 1Q% of the system installed capacity.
 Again, .since the generating  plant utilizes gas turbine units, and since they are
 independent of cooling water, it could be located near the load center.  However,
 this would probably mean  that the transmission network tie would have to travel
 over  some  distance into the  station.   For this comparison, the distance over which
 the energy had to be transmitted in from the transmission network (Case II)  was
 assumed  identical to the transmission distance from the central power station to
 the load center  (Case I).  The only transformer that need be required in the Case
 II system is a  step-down transformer converting the transmission system grid
 voltage to the  distribution voltage.   Finally,  the costs of the equipment required
 for the tie at the transmission network was  assumed to be identical for both Cases
I and II, as a result they would cancel each other out at the transmission network
in an incremental cost approach.
                                        102

-------
     After having synthesized the tvo cases,  it became necessary to  determine  the
required number of circuits from the transmission network to  the station-load
center system.   Two circuits were indicated in Case I; however, since the  trans-
mission distance was assumed negligible,  circuitry cost  considerations  did not
really have to  bfc analyzed for this case.   Reliability considerations were used
for Case II to  determine the number of circuits required.  The results  of  these
considerations  are illustrated in Fig. 68.   This  figure  presents a plot  of loss
of load as a function of the percentage of annual peak load divided  by  system
installed capacity for various unit forced outage rates  and single or double
circuits.  Each circuit carries electric  power equal to  the output of a  single unit.
Again, if it is desirable to  reduce  loss of load to no more than 0.2 days/yr,  it
would appear that a forced outage rate of at  least O.OU  is essential in  the units
if single circuitry is desired.  It should be noted that two  electrical  circuits
will decrease the loss of load by at least two orders of magnitude.  It  is obvious
that either single- or double-circuit transmission lines were acceptable in Case
II, depending on the reliability of the units in  the central  power station.  As a
result for the  remainder of the analysis  both options were considered.

     The results of considering both options  for  a transmission,distance of 25 mi
is shown in Fig. 69a.  This figure presents a. comparison of system allowable fuel
cost increment  as a function of load factor for both circuitry options  and fright-
of-way cost.  It is clearly shown that there is a greater savings associated with
single circuitry as compared to double circuitry.   This  result was expected along
with the decreased saving associated with increased right-of-way cost.   The range
of land costs gives an indication of the  costs that could be  expected in locating
the generation  station near the load center.   Obviously, the  more densely  indus-
trialized or populated the load center is, the greater the premium placed  on the
land in the area of the load center.  Of the three values presented, the figure
of $3000/acre most closely approximates the load  values  within the Northeast
Utilities system (Ref. 125).  The results presented in Fig. 69b indicate the
system allowable fuel cost increment for  the  same parameters  as in Fig.  69a but
for the situation where transmission line distances are  50 mi.  There is a
greater savings in the 50-mi situation as opposed to the 25-mi situation because
°f the increased savings due to lower transmission line  costs.  This reduction in
transmission line costs results from the decreased transmission line load, i.e.,
power equal to  1 of the 5 units in Case II vs power equal to  the single  unit in
Case I.  AS a result, since there is a savings per mile, it follows  that as the
number of miles increases the savings will also increase.
                               Concluding Remarks

     For a given utility system capacity level both  a  reduction  in power  generating
     output capacity and a diversification of unit size  can increase reliability.
      with this increase in reliability there also results, at load factors  of
                                       103

-------
 70 to QQ%> a utility-wide allowable fuel cost  savings  of 0.3<£/million Btu.  There
 is an appreciable system cost credit for expanding an  electric  utility  system more
 frequently and in smaller power generating unit  output capacity that more closely
 approximate the demand curve as opposed to expanding the system over longer time
 periods and in larger unit output capacity.  However,  the costs involved in
 borrowing money have not been included in this comparison.  The system  allowable
 fuel cost saving which can be credited can be  2.2i£/million Btu  for reasonable
 load factors.  The reduced station-transmission  network tie requirements asso-
 ciated with relatively small gas turbine power generating units which can be located
 close to the load center results in an appreciable saving for these units as com-
 pared to the larger steam units located at some  distance from the load  center in
 order to have access to cooling water.   This savings can result in a system allow-
 able fuel cost increment of as great as 1.2#/million Btu.
                                 GAS TURBINE FUELS
      In a previous  section of this report it was noted that natural gas, which
 is  an ideal fuel for use in gas  turbines, will be  in  short supply within
 a few years unless  the  FPC adopts a more realistic policy for establishing well-
 head prices for natural gas.  However, even if the supply of natural gas con-
 tinues to dwindle,  thus  precluding its future use  for electric power generation,
 it  appears  that alternative fuels will become available and that these fuels will
 be  suitable for use in  gas turbine power systems.  The most promising alternative
 fuel appears to be  either  a high- or low-Btu synthesis gas derived from coal.  In
 the United  States,  three incentives exist for development of synthetic coal gasi-
 fication processes:   (l) the United States gas industry, desires to obtain a
 supplementary source of gas to insure its gas supply  in the face of worsening
 reserve-to-production ratio for natural gas in the United States; (2) the federal
 government  wants to  stimulate the utilization of this country's vast coal resources
 to  meet  the country's energy needs; and (3) the federal government desires to
 stimulate the use of nonpolluting, low-sulfur fuels in place of the high-sulfur
 coal and residual fuel  oil in common use today.
                         Gas Turbine Fuel Specifications

     A variety of gaseous, liquid, and solid fuels have been used in gas turbine
engines.  However, with the exception of a few small closed-cycle gas turbine
plants which burn coal or use nuclear power, all present-day aircraft-type and
heavy-duty industrial gas turbines use liquid or gaseous fuels.  Gaseous fuels
such as natural gas, butane, and propane are ideal fuels since they contain no
harmful alkali or sulfur compounds.  Furthermore, they burn readily in small-
volume combustors with no smoke or carbon residue and at low flame luminosity.
                                       101*

-------
Fuels vith high flame luminosity properties  will  exhibit high radiation heat
transfer characteristics.   Consequently,  engines  burning these  fuels would have
to protect against higher liner temperatures than engines burning  gaseous
fuels.   Low flame luminosity has also been found  to be related  to  desirable levels
of smoke emissions and carbon residue.

    Liquid fuels, used by 10% of the gas turbine units in  operation today, are
available in a variety of petroleum distillates,  residual oils,  and blends of
the two.  Liquid fuels for industrial gas turbines traditionally have been petroleum
distillates such as American Society for  Testing  Materials  (ASTM)  Grade No. 1 or
2 diesel fuel oil.  While heavy residual  fuels  such as ASTM Grade  No. 5 or 6  fuel
oil have been used in heavy-duty gas turbines which operate at  relatively  low
turbine inlet temperatures, these heavy fuels must be selected  with care
to minimize maintenance costs, especially in the  hot  section of the engine.   The
raajor problems associated with the utilization  of these heavy residual fuels  in
gas turbines are erosion, ash deposition, vanadium corrosion, and  sulfidation of  the
turbine blades and vanes resulting from the  high  ash, metal, and sulfur contents
of these fuels.

    Gas turbines operating with heavy liquid fuels are generally  restricted  to
operating temperatures which are several  hundred  degrees below  those employed
vhen burning clean gaseous or distillate  fuels.   Ash  deposition due to the
formation of liquid vanadium and alkali metal compounds during  combustion  of
residual fuel oils can result in severe loss of output power, especially under
continuous operating conditions.  At turbine inlet temperatures  below approximately
1200 F, ash deposits are generally loose  and powdery  or can be  made so by  selected
fuel additives.  These deposits can be readily  washed off or spalled off during
frequent shutdown intervals.  Harder bonded  deposits  and subsequent vanadium
corrosion tend to occur at temperatures above 1200 F  with stainless steels,
cobalt-base alloys, and to a lesser extent with nickel-base alloys.  Satisfactory
operation can sometimes be achieved by utilizing  magnesium-base additives  and water
washing of the fuel to reduce the sodium and potassium concentrations.  However,
sulfidation is a formidable corrosion problem at  turbine inlet  temperatures of
1500 F  and above with fuels containing substantial quantities of sodium and sulfur.

    As a result of these operational problems, residual fuels  are generally
unacceptable for use in advanced, high-temperature gas turbines.  The problems
associated with utilizing coal as a gas turbine fuel  would  be far  more severe than
"those encountered with residual oil because of  the higher ash and  sulfur contents
of most coal.  Consequently, in order to utilize  coal as a  gas  turbine fuel it
oust be gasified and purified to a cleanliness  comparable to that  of natural  gas
°r distillate fuel.  No industry-vide specification currently exists for gaseous
fuels,  because most natural gas made available  through gas  distribution networks
exceeds the cleanliness requirements specified  by gas turbine manufacturers.
Recently, the ASTM provided tentative specifications  (see Table XXX) for four
                                       105

-------
 grades of liquid gas turbine fuels.   These specifications provide  a starting point
 for discussions "between gas turbine  manufacturers  and fuel  suppliers.  In addition
 to normal fuel properties,  these specifications provide  for specific limits on
 metals such as vanadium, sodium, potassium,  calcium,  and lead that could cause
 corrosion or ash deposition during turbine operation.  Pratt & Whitney Aircraft
 Division of UA in its heavy distillate fuel  specifications,  further restricts
 the vanadium content to 0.2 ppm by weight, the sodium plus  potassium content to
 0.6 ppm by weight, and the  sulfur content  to 1.3$  by  weight.
                           Coal  Gasification Technology

      The basic  coal gasification process consists of reacting coal with steam
 at elevated temperature  to produce a  synthesis gas consisting of carbon monoxide,
 hydrogen, and a variety  of impurities.  This basic gasification reaction is highly
 endothermic so  that considerable heat must be supplied to sustain the reaction.
 This  heat could be supplied  autothermally, wherein a portion of the coal would be
 combusted with  air or oxygen, or by supplying external heat to the process.
 Simplified schematic diagrams of coal gasification processes utilizing these two
 methods  of heat addition are depicted in Figs. 70a and TOb, respectively.

      In  addition to providing a source of heat for the reaction, all gasification
 processes require equipment  to  scrub  coal dust and condensible hydrocarbon tar
 from  the synthesis gas and gas  purification equipment to remove undesirable sulfur
 compounds and,  in some cases, carbon  dioxide.  This equipment is denoted by the
 solid squares in Fig.  70.  If a low-Btu fuel gas were to be the end product, air
 could be used for partial  combustion  of the coal in autothermal processes and
 no further processing would be  required after the gas purification step.  However,
 if a  high-Btu (900 Btu/ft3 or higher) pipeline quality gas were the desired end
 product,  then autothermal processes would require the use of pure oxygen for par-
 tial  combustion to prevent nitrogen in the air from getting into the synthesis
 gas.   In  addition,  the hydrogen-to-carbon monoxide ratio in the gas must be ad-
 justed to  three parts by volume hydrogen to one part by volume carbon monoxide
 using the  water-gas  shift  conversion  so that these gases could undergo subsequent
 catalytic  methanation to produce the  desired methane-rich, high-Btu pipeline gas.
 This  optional equipment required for pipeline quality gas applications is desig-
 nated by the dashed boxes in Fig. 70.

      Several alternative oxygen separation,  shift-conversion,  and purification
 processes  are commercially available so that development of coal gasification
 processes has concentrated on the basic gasification step and,  in the case of
 pipeline quality gas, the methanation step.   A concise description and evaluation
 of four methanation reactor designs is presented in  Ref.  126 and is not  repeated
herein.  The following discussion concerns  alternative methods  of gasifying coal,
 including two with in-situ sulfur removal.
                                       106

-------
Autothermal Gasifiers

     Countercurrent, cocurrent, and fluid!zed-ted  autothermal gasifiers have been
used commercially (Ref. 127).   Countercurrent  gasifiers with downflow of coal and
upflov of gases have the advantage of high  thermal efficiency and high turndown
ratio.   Their major disadvantages are lov gasification rates, relatively small
capacities, and the formation of tar which  makes gas purification and waste heat
recovery difficult.   Cocurrent gasifiers, with up  or down flow of both coal and gases,
have relatively low thermal efficiencies  unless expensive waste heat recovery
equipment is employed.  Gasification rates  are higher than those of Countercurrent
gasifiers because higher temperatures can be employed.  A major advantage is that
no tar is formed, making gas cleanup and  waste heat recovery easy.  Fluidized-bed
gasifiers have intermediate characteristics. They can be scaled up to relatively
large sizes.  However, turndown ratios for  these units are small due to the neces-
sity of maintaining a minimum fluidizing  velocity  through the bed.

     A summary and comparison of some of  the more  promising commercial coal
gasification processes are given in Table XXX.  Of these processes, only the Lurgi
dry-ash process is carried out at elevated  pressure.  Synthesis, gas must be at
elevated pressure for pipeline transmission and for use in gas turbine power
systems.  In addition, elevated pressure  would be  desired in order to reduce the
physical size and cost of gas purification  equipment.  Also, specific gasification
rates,  i.e., gas produced per cubic foot  of reactor volume, are favored by
increased pressure.   For these reasons, the Lurgi  dry-ash gasifier is the only
commercial gasifier which appears to be suitable (see Ref. 128).  The specific
gasification rate for this gasifier compares favorably with other commercial
gasifiers, and is amenable to scale up in size such that a single gasifier could
provide fuel gas to 80-Mw or larger power stations.  For very large stations
(1000 Mw), the requirement for a relatively large  number of gasifiers could prove
to be a disadvantage for the Lurgi gasifier. This limitation in gasification
rate results because the Lurgi gasifier is  not designed for slagging operation
requiring reaction temperature to be kept below the ash fusion temperature of coal
(approximately 1700 F).

     Advanced autothermal gasifiers could achieve  higher gasification rates by
operating at higher temperatures, in excess of 2200 F.  Under these conditions, the
c^al ash would melt and become slag.  Thus, fixed-bed gasifiers (such as the
Lurgi type) could no longer be used and entrained  or cocurrent gasifiers would
need to be employed.  Various high-temperature, cocurrent flow gasifiers have been
surveyed (see Table XXX).  The Texaco (Ref. 129),  US Bureau of Mines (Ref. 130),
and the Bituminous Coal Research (Ref. 131) gasifiers are representative of ad-
vanced gasifiers.  These advanced gasifier  configurations vary, but the basic
chemistry, gasification rates, and efficiencies appear to be comparable.  In all
cases,  from two- to three-second residence  times are required for 90 to 100$
carbon conversion, and gasification temperatures range from 2200 F to 2500 F.
       or air requirements must be sufficient  to supply heat for preheating feeds,
                                       107

-------
 endothermic reactions, and making up heat  lasses.   Steam requirements must be such
 that the reactant temperature is kept vithin  the 2200 to 2500 F range by the
 endothermic steam-carbon reaction.   Off-gases from these advanced reactors
 generally reach equilibrium with respect to the vater-gas shift conversion, thus
 obviating the need for separate shift conversion facilities.  Unlike the Lurgi
 dry-ash gasifier, vhere the synthesis gas  exits the gasifier at approximately 950
 F and contains less than B% of the  heating value of the coal as sensible heat,
 synthesis gas would exit from advanced cocurrent gasifiers at 2200 F to 2500 F and
 contain 15 to 20% of the heating value of  the coal as sensible heat.  Recovery of
 this heat by preheating feeds to the process  or by generating electricity will be
 necessary in order to realize satisfactory coal-to-gas energy conversion effi-
 ciency.
                            External Heating Processes

      Externally heated gasifiers  differ  from autothermal gasifiers (see Fig. 70)
 in that coal need not  be  oxidized within the gasifier vessel to provide the heat
 necessary to sustain the  endothermic coal-steam reaction.  The advantages of
 external heating are two-fold:  first, no oxygen separation equipment would
 be needed to keep nitrogen  out of the synthesis gas, and second, carbon dioxide
 would not be formed to dilute the synthesis gas.  These characteristics make
 externally heated processes more  adaptable to providing high-Btu pipeline gas
 than low-Btu gas.

      Three processes utilizing external heating that have been studied extensively
 are also listed in Table  XXXI.  In the HYGAS process, which is being developed by
 the Institute of Gas Technology (Ref. 132), gasification would occur in two
 sections.  In the  first or hydrogasification section, coal would be gasified in
 a  series  of  contacting stages by  a mixture of steam and synthesis gas.   The
 synthesis  gas,  consisting primarily of hydrogen and carbon monoxide, would be
 generated in an  electrothermal gasification section.  In this section,  char resi-
 due from the hydrogasifier would be reacted with steam to produce the synthesis
 gas.   Heat for the electrothermal gasifier would be-provided by electrical
 resistance heaters.  The  char residue from the electrothermal gasifier would be
 used to generate the required electrical power in a combined magnetohydrodynamic-
 steam  power  system.  A HYGAS pilot plant capable of processing 80 tons/day of coal
 to  produce 1.5 million cu ft of synthetic high-Btu gas was dedicated in Chicago
 and is expected  to begin  operation in early 1971-   According to an estimate
 recently made by Stearns-Roger Corporation (Ref.  133), a commercial plant based on
 the HYGAS process  could be designed and constructed by about mid-197** if a crash
 program were  inaugurated.   However,  such a program would involve considerable risk
by  the operator.  A more  conservative approach would allow another 2 to 3 years
 for additional pilot plant testing.
                                      108

-------
     The other tvo processes identified in Table XXXI  in which external heating
vould "be used are interesting because a single  recirculated material vould  serve
the dual purpose of conveying heat to sustain the endothermic chemical reactions
while simultaneously providing in-situ removal  of sulfur compounds.  In the carbon
dioxide acceptor process (Ref. 13*0 being developed by Consolidation Coal
Company, hot calcined limestone or dolomite would be used, whereas in the molten
salt process investigated by M. ¥. Kellogg Company (Ref. 126), hot sodium carbonate
would be used.  The off-gases from these gasifiers would have to be scrubbed,
undergo shift conversion, and be methanated as  depicted in Fig. 70.  According
to Ref. 133, the carbon dioxide acceptor process could be commercially available
approximately one year after the HYGAS process.
                             Coal Gasification  Costs

     Numerous estimates for the cost of pipeline quality synthetic coal gas
(900 Btu/ft3 or higher) are available in the literature.  These estimates range
from 30 to 70<£/million Btu, exclusive of transmission  cost,  depending on the
degree of optimism built into the estimates, the cost  of coal, interest rates,
rate of return on investment, and the time period during which the estimates were
prepared.  Most estimates prepared during the late 1960's based on then-current
coal costs, interest rates, and construction costs and using utility accounting
procedures indicate that pipeline gas could be  produced for  ko to 50i£/million Btu,
exclusive of transmission cost (see Refs. 130,  132, and 13U). Other estimates
(Ref. 135) indicate that synthetic pipeline quality gas could be generated in West
Virginia and piped to Philadelphia for just under 50tf/million Btu range.  A recent
analysis of the IGT HYGAS and the carbon dioxide acceptor process (which is also
slated for large pilot plant evaluation) conducted by  Stearns-Roger Corporation
engineers (Ref. 133) indicates the price of pipeline quality gas to be about
^Qtf higher than the price of coal used.  Since  processes designed to produce low-
Btu gas (150 to 250 Btu/ft3) would not require  equipment for oxygen separation,
shift conversion, or methanation, it has been estimated that this type of producer
fuel gas could be manufactured for an incremental cost of about 20
-------
      The costs associated vith the development  of new engines  for aircraft
 applications are staggering.   For example,  the  costs  to  develop the engines  for
 the various Jumbo jets,  i.e.,  the Lockheed  L-1011, the Boeing  7^7 and others are
 projected to reach a level of $200 million  (Ref.  136).   In Ref. 137, estimates
 of the cumulative development  costs for turbojet  engines are presented; and  indi-
 cate that for engine thrust levels of approximately 100,000 Ib (equivalent to a
 power output of approximately  150,000 hp) the costs vould approach $500 million.
 These costs include the  costs  spent for continued product improvement of the engine
 with time.   Product improvement is an important part  of  the engine development
 process and should not be misconstrued as simply  improving reliability or in-
 creasing the number of applications.   The thrust  of the  Pratt  & Whitney J-57, for
 example, was increased from approximately 10,000  Ib to 21,000  Ib over 10 years
 through the product improvement procedure.

      There  are numerous  reasons for the high development costs for these aircraft
 gas turbines.   Each new  engine development  usually involves some advancement in
 the performance and weight characteristics  of the engine above the levels
 available with existing  engines.   Furthermore,  these  advances  must be achieved
 without sacrificing engine reliability or the flexibility to operate over a wide
 range of power settings.   The  requirements  for  aircraft  engine development also
 tend to emphasize the attainment  of higher  and  higher component efficiencies.
 Since each  operating part of the  engine is  so highly  loaded, and minor failures
 in  a local  area might overload other  critical components, the  designer often finds
 his analytical abilities  inadequate and each new  engine  development requires
 repeated designing,  building,  and testing of individual  components as well as
 complete engines to  achieve the desired results.   It  is  not uncommon for as many
 as  two dozen sets of engine parts  to be built and for some engine designs to
 undergo 10,000 hrs  on the test  stand before certification and many times that amount
 during subsequent model  changes.

      Development costs for advanced gas  turbines  for  electrical utility and other
 ground-based applications  would be substantially  less than for aircraft engines of
 comparable  size.  Since the attainment  of high-power-to-weight ratios is not neces-
 sary,  this would provide  an additional  degree of  freedom for the gas turbine
 designer.   In  addition, the technology  of advanced materials and blade cooling
would be available  from aircraft  engine  programs.  Although reliability would still
be  a major  criterion, it  is unlikely that more  than three to five sets of engine
parts would be required   for testing before the design characteristics of a new
engine were  finalized.  Thus, the  extensive component and engine testing such as
that  required  before an aircraft engine  is  certified  could be reduced.   Further-
more, much of  the tooling  and manufacturing facilities used for smaller output
capacity engines  could be  utilized for the  advanced, higher-capacity designs.

     A brief preliminary analysis was made to determine the approximate develop-
ment costs for the advanced engines designed for industrial applications.   The
                                      110

-------
results are shown in Fig. Tla (see line l) and indicate that  from a~bout $100 to
$150 million "ould be required for initial engine development over the first five
years for engines ranging in size from 100 to 250 Mw.   An additional $6U to $100
million would be needed for product improvement over the next 10 years (see line
2 in Fig. Tla).  It is anticipated that the product improvement program could
result in advanced and larger versions of the gas turbine engines.  These estimates
are for a company entering the gas turbine field which would  require the construc-
tion of new manufacturing facilities and tooling capable of handling engines capa-
ble of producing 100 Mw and above.  However, a company which  had existing facilities
and was making large capacity engines, i.e., 50 to 100 Mw,  could develop the
advanced engines at substantially lower costs since manufacturing facilities would
be available.  The dashed line (line 3) shows the estimated costs including the
5-year and product improvement program for a company with existing facilities.
These levels would be substantially lower and would introduce a lower burden on the
selling prices of the engines than  for a new manufacturer.

     The estimated selling prices of the gas turbine engines  discussed in this
report were based on a total engine development cost of $125  million,  or essentially
cnly the costs associated for the first five years of development> and an antici-
pated market penetration of 1*000 Mw/yr.  A review of Figs.  1  and 2 indicates that
from 30,000 to 40,000 Mw of thermal generation equipment will be added in the
electric utility industry during each of the next twenty years.   Therefore,  it
appears reasonable that any one gas turbine manufacturer after allowances for steam
station penetration would be capable of capturing 10$ of this total market (say
1*000 Mw/yr).  However, the effects of variations on market  penetration and engine
development cost on the estimated engine selling price for  a  250-Mw engine are
shown in Fig. 71b.  The results indicate only a modest increase  in engine selling
price of about $5/kw as market penetration falls to about half (2000 Mw/yr)  of
that assumed in the study.
                     ESTIMATE OF ADDITIONAL CAPITAL COSTS FOR
                        COOLING TOWERS AND COOLING PONDS

     It  has  been projected (Ref. 31*) that within the next 50 years  approximately
$20 billion  will have to be invested to provide the cooling water requirements  for
steam-electric  plants and that a large part of the needed investment will  be for
cooling  towers.  This projection is substantiated by a  recent  survey  (Ref.  36)
vhich noted  that twenty-six steam-electric power companies estimated that  they
vill invest  some $170 minion of capital in plants now  under construction  to
comply with  the new or impending water temperature standards.   Specifically, new
^its under  construction for twenty investor-owned companies (representing 1*1,860 Mw)
and six  municipal,  co-op, or public companies  (representing 32hQ Mw) will  require
$156,706,000 and $1**,000,000, respectively, to achieve  this compliance.  It is
                                        111

-------
noted in Ref. 36 that a number of companies contacted in this survey were unable
to provide estimates because of uncertainties existing over the final temperature
standards or permissible mixing zones in their respective localities.  Fourteen
of the companies surveyed claim to have already spent nearly $15 million over the
past five years to bring their stations into compliance vith nev temperature
standards and 6 companies alone plan to spend $U9 million in the near future to
bring existing stations into compliance.

     Projections of future anticipated investments in cooling facilities in the
United States (Ref. Uo) indicate the cost for such investments may amount to from
about $2 billion to over $U billion in the 1970 decade; the actual amount will
depend upon how stringent thermal quality standards become.  It is projected that
an additional $3.5 billion to $5 billion will be invested for this purpose in the
1980-decade.  These costs are about an order of magnitude greater than the develop- |
ment costs of the advanced open-cycle gas turbines which would provide a solution  )
to the thermal pollution of our river and lake waters.   This highlights the        |
need for more intensive development of reliable, high-output-capacity gas          i
turbines operating at temperature levels of 2000 F and above.                      }
                                        112

-------
                                    SECTION  IX
                                 ACKNOWLEDGMENTS
     The work described herein vas performed by the United Aircraft Research
Laboratories (UAEL)  for the United States  Environmental Protection Agency  Water
Quality Office (formerly the Federal Water Quality Administration of the Department
of the  Interior)  under Contract No.  lU-12-593 during the period  from February  16,
1970 to March 15, 1971.

     The support  of the project by the Water Quality Office and  the valuable
guidance and comments provided by Dr. Mostafa Shirazi, Project Officer of  the
contract in the Pacific Northwest Water Laboratory, is acknovledged vith sincere
thanks.

     The assistance provided by the various members of the Energy Conversion
Systems Evaluation Section of UARL, under  the direction of Mr. N. C. Rice  and
various members of the utility industry is gratefully acknowledge.
                                       113

-------
                                     SECTION X
                                    REFERENCES
 1.  Federal Power Commission Regional Reports.   Prepared by Regional Advisory
     Committees, used in Preparation of the 19.70  National Power Survey, various
     1969 dates.

 2.  l*6th Semi-Annual Electric Power Survey.   Edison Electric Publication No.  69-58,
     October 1969.

 3.  Ujth Semi-Annual Electric Power Survey.   Edison Electric Publication No.  70-26,
     April 1970.

 1*.  Second Biennial Survey of Power Equipment Requirements of the US Electric
     Utility Industry 1969-78.  Survey sponsored  by Power Equipment Division,
     National Electric Manufacturers Association, New York, New York, February 1970.

 5.  Civilian Nuclear Power 1967, Supplement  to the 1962  Report to the President.
     USAEC, February 1967.

 6.  McKennitt, D. B.:  The US Electric Power Industry.   Stanford Research Institute
     Report No. 321, May 1967.

 7.  Gambs, G.:  The Electric Utility Industry:   Future Fuel Requirements 1970-1990.
     Mechanical Engineering, April 1970.

 8.  Energy in the United States 1960-1985.   Sartorius &  Co., September 1967.

 9.  Outlook for Energy in the United States.  Energy Division, The Chase Manhattan
     Bank,  October 1968.

10.  Ritchings, F. A.:  Raw Energy Resources  for  Electric Energy Generation.   Paper
     presented at the 1968 American Power  Conference, April 1968, Chicago, Illinois.

11 •  Morrison, W. E.:   Simulated Models of Future Energy  Demand - Probability  and
     Contingencies for 1980 and 2000 A.D.  ASME Paper 68-PWR-U presented at the
     IEEE-ASME Joint Power Conference in San  Francisco, California, September  16-19,
     1968.

12-  A Review and Comparison of Selected United States Energy Forecasts.  Pacific
     Northwest Laboratories of Battelle Memorial  Institute, December 1969.
                                         115

-------
                              REFERENCES (Continued)
 13.   More Natural Gas is Sought for Use in Eastern States.  New York Times,
      August 16, 1970, pp. 1 and 36.

 lU.   Congressional Record - Extension of Remarks,  pp. E6015-E6016, June 26, 1970.

 15.   Potential Supply of Natural Gas in the United States  (as  of December 31, 1968).
      Prepared by Potential Gas Committee, Colorado School  of Mines Foundation,
      Inc., Golden, Colorado.

 16.   US Geological Survey Circular No.  522.

 17.   Cambel,  A. B., et al.:   Energy R&D and National Progress.  US Government
      Printing Office, Washington,  1965.

 18.   Future Natural Gas Requirements of the  United States.  Volume No. 3, Denver
      Research Institute,  University of Denver, Denver, Colorado, September 1969.

 19.   "Big West Coast Utility Eyes  Faraway Gas."  The Oil and Gas Journal, August
      10,  1970, pp.  90-91.

 20.   Pipeline Pays  28^ for Texas Gas to Ease Supply Bind.  The Oil and Gas Journal,
      July 13,  1970,  p.  37.

 21.   Gas  Supply Would Rise with Price Hike.  The Oil and Gas Journal, May U, 1970,
      pp.  9^-95.

 22.   Higher Gas  Prices  Coming, Question is How Much.  The Oil and Gas Journal,
      June 15,  1970,  pp.  33-36.

 23.   Virtually No Uncommitted Gulf Gas Left.  The Oil and Gas Journal, May 18, 1970,
      pp.  U2-U4.

 2k.   FPC  Study Points Up Big Gas-Price Gap.  The Oil and Gas Journal, September lk,
      1970, pp.  62-63.

 25.   Air Pollution and  the Regulated Natural Gas and Electric Utility Industries.
      FPC Report, September 1968.

26.   Soaring Tanker Rates Felt in US.  The Oil and Gas Journal, July 13,  1970, p. ^

27.  Bennett,  R. R.:  Energy  for the Future.  Combustion, April 1970.
                                       116

-------
                             REFERENCES  (Continued)
28.   DeCarlo,  J.  A., E. T.  Sheridan,  and Z.  E. Murphy:  Sulfur Content and United
     States Coal.   Bureau of Mines Information Circular 8312.

29.   Davison,  W.  R.:  Visit to Burns  and Roe, Inc. to Discuss Electric Utility
     Industry  Projects.  United Aircraft Research Laboratories Report UAR-J175,
     July T, 1970,

30.   Sporn, P.:   Developments in Nuclear Pover Economics, January 1968-December 1969.
     Report prepared for the Joint Conmittee on Atomic Energy, Congress of the US.

31.   Congressional Statement - Extension of  Remarks by Hon. J. Randolph (W. Va.),
     pp.  E7963-E7975, September 2, 1970.

32,   To Keep the  Lights Burning.  Forbes Magazine, July 15, 1970, pp. 22-29.

33.   Problems  in  Disposal of Waste Heat from Steam-Electric Plants.  FPC Staff
     Report, 1969.

3^.   The Nation's Water Resources.  The US Water Resources Council, Washington, D.C.,
     November  1968.

35.   Hauser, L. G.:  Cooling Water Requirements for the Growing Thermal Generation
     Additions of the Electric Utility Industry, Paper Presented at the American
     Pover Conference, Chicago, Illinois, April 22-2U, 1969.

36.   Olds, F.  C.:  Thermal Effects:  A Report on Utility Action.  Power Engineering,
     April 1970.

37.   Clark, J. L.:  Thermal Pollution and Aquatic Life.  Scientific American,
     Vol.  220, No. 3, March 1969.

38.   Jaske, R. T.:  The Need for Advance Planning of Thermal Discharges.  Nuclear
     Nevs, September 1969.

39.   Considerations Affecting Steam Poverplant Site Selection.  A report sponsored
     by the Energy Policy Staff, Office of Science and Technology, December 1968.

^0.   Warren, F. H.:  Electric Pover and Thermal Output in the Next Two Decades.
     Stanford  Research Institute Report No.  321, May 1967.

kl-   Presentation Before the New York Society of Security Analysts by Gulf States
     Utilities Company, September 1970.
                                        117

-------
                             REFERENCES  (Continued)
 U2.   Lokay,  H.  E.,  H.  L. Smith, and G. D. Broome:   Changing Patterns in Generation
      Planning Results.  Paper  presented  at American Power Conference, Chicago,
      Illinois,  April 23-25,  1968.

 1+3.   Levis,  G.  P.:   Feasibility Studies  of Advanced Pover Cycles, Progress
      Report  No.  10,  Burns  and  Roe, Inc., Oradell, New Jersey, May 6, 1970.

 kk.   Palo, G. P., et al.:  Units 500 Mw  and Larger  Found to Yield Savings.
      Electrical World, March 31, 1969, pp. 30-31*.

 1+5-   Robson,  F.  L.,  et al.:  Technological and Economic Feasibility of
      Advanced Power  Cycles and Methods of Producing Nonpollution Fuels for Utility
      Power Stations.  United Aircraft Research Laboratories Report J-970855-13,
      November 1970.

 U6.   Woodson, H. H.:  Short  Term Prospects for Improving Efficiency of Power Plants.
      Paper presented at Thermal Considerations in the Production of Electric Power
      (a Joint Meeting of the Atomic Industrial Forum and Electric Power Council
      on Environment), Washington, D. C., June 28-30, 1970.

 U7.   Giramonti,  A. J.:  Discussion of COGAS Systems with Riley Stoker Corporation.
      United Aircraft Research  Laboratories Report UAR-H2U1, September 30, 1969.

 U8.   Giramonti,  A. J.:  Discussion of COGAS Systems with Foster Wheeler Corporation.
      United Aircraft Research  Laboratories Report UAR-H210, September 5, 1969.

 1+9-   Giramonti,  A. J.:  Discussions of Steam and COGAS Systems with Babcock and
     Wilcox Company.  United Aircraft Research Laboratories Report UAR-H2U6,
      September 30, 1969.

50.  Biancardi,  F. R.:   Feasibility Study of Nonthermal Pollution Power Generating
     Systems.  United Aircraft Research Laboratories Report J-970978-U, July 10, 19I;

51.  Nuclear Power for the Under-Developed?  Electrical World, January 12, 1970,
     pp.  21+-26.

52.  Biancardi,  F. R.:   Memorandum of communication with Northeast Utilities,
     "Calculation of Power Generation Costs," May 10,  1970.

53.  "TVA Contrasts Cooling Water Designs," Electrical World,  December 22, 1969,
     pp.  23-26.
                                        118

-------
                             REFERENCES (Continued)
5U.   Neale, L. C.:  The Use of River Models  in Power Plant Heat Effect Studies.
     Paper presented at Meeting on Thermal Considerations in the Production of
     Electric Pover, Washington, D. C.,  June 28-30, 1970 (A Joint Meeting  of Atomic
     Industrial Forum and Electric Pover Council on Environment).

55.   Christiansen, A. G., and B. A. Tichenor:   Economic Aspects of Thermal Pollution
     Control in the Electric Power Industry, No.  67, September 1969.  Federal
     Water Pollution Control Administration Northwest Region, Pacific Northwest
     Water Laboratory, Corvallis, Oregon.

56.   Carey, J. H., J. T. Ganley, and J.  S.  Maulbetsch:  Task I Report; Survey of
     Large-Scale Heat Rejection Equipment Prepared for Federal Water Pollution
     Control Administration.  US Dept.  of the Interior, Corvallis, Oregon.  Under
     Contract No. 1^-12-2177 by Dynatech R/D Company, July 21, 1969.

57-   Kolflat, T.:  Natural Bodies of Water for Cooling.  Paper presented at Thermal
     Considerations in the Production of Electric Power, Washington, D. C.,
     June 28-30, 1970 (Joint Meeting of the Atomic Industrial Forum and Electric
     Power Council on Environment).

58.   A Cooling Pond Proves Cheaper.  Electrical World, November 30, 1953,  pp. 8U-85.

59.   Letter to F. Biancardi from Mr. G.  Crossland of Ceramic Cooling Tower
     Company, Fort Worth, Texas, February 8, 1971.

60.   Feasibility of Alternative Means of Cooling for Thermal Power Plants  Near
     Lake Michigan.  US Department of the Interior, Federal Water Quality  Admini-
     stration Report, August 1970.

6l.   Kadel, J. 0.:   Cooling Towers - A Technological Tool to Increase Plant Site
     Potentials.  Paper presented at American Power Conference, Chicago, Illinois,
     April 23, 1970.

62.   Rossie, J.  P., and E. A. Cecil:  Research on Dry-Type Cooling Towers  for
     Thermal Electric Generation.  Prepared for US Department of the Interior,
     Federal Water Quality Administration,  under Contract No. lU-12-823, November
     1970.

63.   Woodson, R. D.:  Cooling Towers for Large Stean-Electric Generating Units.
     Paper presented at Symposium on Thermal Considerations in the Production of
     Electric Power, Washington, D. C.,  June 1970 (Joint Meeting of the Atomic
     Industrial  Forum and the Electric  Power Council on the Environment).
                                       119

-------
                              REFERENCES (Continued)
 6k.   Smith,  E.  C., and M.  W.  Larinoff:   Pover Plant  Siting.  Performance and
      Economics  with Dry Cooling Tower Systems.   Paper presented at American
      Power Conference, Chicago, Illinois,  April 1970.

 65.   Heeren, H.,  and L.  Holly:   Air Cooling for Condensation and Exhaust Heat
      Rejection  in Large Generating Stations.   Paper  presented at American Power
      Conference,  April 1970.

 66.   Kolflat, T.:  How to  Beat  the Heat  in Cooling Water.  Electrical World,
      October lU,  1968, pp.  31-33.

 67.   Cooling Tower Fundamentals and Application Principles.  Marley Co. publication,
      1967-

 68.   Chervishev,  P.  S.,  et  al.:  Experience with Development Work-and Manufacture
      of 100-Mw  Gas Turbine  Plant at LMZ.   ASME  Paper No. 70-GT-30, 1970.

 69.   Congiu,  A.:   A 37A2 Mw  Gas Turbine for  Power Generation.  ASME Paper No.
      6U-GTP-U,  196U.

 70.   Stewart, W.  L.,  et  al.:  Brayton Cycle Systems.  Selected Technology for the
      Electric Power  Industry.   NASA SP-5057,  September 1968.

 71.   STAL-LAVAL Turbine  Company:   Technical Information Letter - Gas Turbines
      for Peak Load Generation,  196U.

 72.   Baldwin, C.  J.,  et  al.:  Future Role  of  Gas Turbines in Power Generation.
      Proceedings  of the  American Power Conference 27th Annual Meeting, Vol.
      XXVII, 1965,  pp.  U8U-500.

 73.   Bailey, W. D.:   Operating Experience with  a Multiset Gas Turbine-Generator.
      ASME Paper No. 68-GT-57, 1968.

 Ik.   Gatzemeyer, J. B.,  et al.:  Characteristics of a New ^6,000-kw Packaged Gas
      Turbine Power Plant Presented at American Power Conference 30th Annual
     Meeting, 1968.

75.  Starkey, N. E.:  Long-Life Base-Load Service at 1600 F Turbine Inlet
     Temperature.  ASME Paper No. 66-GT-98, 1966.

76.  Gaskins, R. C., and J. M. Stevens:   World's Largest Single-Shaft Gas Turbine
     Installation.  ASME Paper No.  70-GT-12U, 1970.
                                       120

-------
                             REFERENCES (Continued)
77.   Bankoul, V., and J. H. B. Kean:   A New 30-Mw  Packaged Gas Turbine Power  Plant.
     ASME Paper No. 70-GT-ll, 1970.

78.   Martens, W. R. , and W. A. Raabe:   The Materials  Challenge of High-Temperature
     Turbine Vanes and Blades.  ASME  Paper No.  67-GT-17, 1967.

79.   Hare, A., and H. H. Malley:  Cooling Modern Aero Engine Turbine Blades and
     Vanes.  SAE Paper No. 660053, January 1966.

80.   Sharp, W. H. :  High Temperature  Alloys for the Gas Turbine - The State of
     the Art.  SAE Paper No. 650708,  October 1965.

8l.   (Jberg, A. :  Design Features for  Maintainability  in the Pratt and Whitney
     Aircraft JT9D Gas Turbine Engine.  SAE Paper  No. 680337, May 1968.

82.   Keen, J. M. S.:  Design Features of Rolls-Royce  Advanced Technology Engines.
     SAE Paper No. 680338, 1968.

83.   Freche, C. , and R. W. Hall:  NASA Programs for Development of High-Temperature
     Alloys for Advanced Engines.   AIAA Journal of Aircraft, September-October 1969,
     pp.
8U.   Halls, G. A., and S. G. Baker:   Turbine Blade  Cooling - The Global Picture.
     AIAA Technical Information Service.   Presented to the 9th International
     Aeronautical Congress, June 1969-

85.   Thompson, E. R., et al.:  Investigation to Develop a High Strength Eutectic
     Alloy with Controlled Microstructure.   UA Research Laboratories Report
     J-910868-4, July 31, 1970.

86-   Schloesser, V. V.:  A Large Peaking  Gas Turbine.  Proceedings of the American
     Power Conference, Vol. 31, 1969-

67.   Allen, R. P., and R. C. Petitt:   New Gas Turbine Design for Large Power
     Systems.   General Electric Company.   Paper presented at American Power
     Conference 32nd Annual Meeting,  April 22, 1970.

88-   McDonald, C. F.:  Study of a Lightweight Integral Regenerative Gas Turbine
     for High  Performance.  AiResearch Report 70-6179-

89.   Weir,  R.  H.:  Advances in Gas Turbine Technology.  The Chartered Mechanical
     Engineer, March 1962.
                                        121

-------
                               REFERENCES  (Continued)
  90.   Petitt,  R.  C. :   Design  and Development of  a New 11,000 hp Industrial Gas
       Turbine.  ASME  Paper  No.  69-GT-lll, March  1969-

  91.   Battelle Memorial  Institute:  Current and  Future Usage of Materials in
       Aircraft Gas  Turbine  Engines, February 1,  1970.

  92.   Bradley,  E. ,  et al. :  The Pratt & Whitney  Gas Turbine Story.  Metal Progress,
       March  1970.

  93.   Hazard,  H.  R.:   Combustors, Gas Turbine Engineering Handbook, First Edition,
       Gas Turbine Publications,  1966.

  9k.   Burns  and Roe Correspondence vith Westinghouse under Department of Health,
       Education and Welfare Contract CPA-22-69-11H to UA Research Laboratories,
       July 1,  1969.

  95.   Burns  and Roe Correspondence with General  Electric under Department of
       Health,  Education, and Welfare Contract CPA-22-69-11^ to UA Research Labora-
       tories, July 1,  1969.

  96.   Ault,  G.  M. :  Engineering Mechanics and Materials.  Selected Technology
       for the Electric Power Industry.  NASA SP-5057, September 1968.

  97-   Peters, D. , and J. Mortuner:  Ceramic Turbines:  Why Britain is Leading
       the Race. 'The  Engineer, February 26, 1970, pp. 29-33.
 98.  Kraft, E. H.:  An Analysis of V/SI^S^ Composite as a Possible High Strength,
      High Temperature Material.  UA Research Laboratories Report J-110603-1,
      September 9, 1970 (Controlled).

 99.  Landerman, A.:  Discussions vith Corning Glass Works Personnel Concerning
      Cercor* Gas Turbine Regenerators.  UA Research Laboratories File Memorandum,
      June 1, 1970.

100.  Recuperators vs Regenerators.  Discussion by Paul A. Pitt.   Gas Turbine
      Magazine, September-October 1966.

101.  Curbishley, G. , et al.:  Hot Corrosion Resistance of Materials for Small
      Gas Turbine Recuperators.  USAAVLABS Technical Report 69-92, December 1969.

102.  Letter from M. H. McClew, Harrison Radiator Division of General Motors
      Corporation to A. M. Landerman, UA Research Laboratories, June 10, 1970.
                                       122

-------
                            . REFERENCES (Continued)
103.   J«Wboir»ki, S.  T. :   Plate-Fin Recuperator 9000 hp to 2000 hp  Industrial Gas
      TuAine Engines.  AiResearch Manufacturing  Division, Garrett  Corporation,
      Report Ho.  69-5U71, August 27, 1969.

10i*.   K«y», W.  M. , and A. C. London:  Compact  Heat Exchangers,  McGraw-Hill Book
105-  Wolfe,  P.,  and H. F.  May:  Design Experience with Regenerators for  Industrial
     Gas Turbines.   ASME Paper 69-GT-1Q6,  March  1969.

106.  Letter  from M. H. McClew, Harrison Radiator Division, General Motors
     Corporation, to V. R. Davison,  UA Research  Laboratories, January  5, 1970.

107.  Smith,  E.  C.:   Technical Data Relevant  to Direct Use of Mr  for Process
     Cooling,  Hudson Engineering Corporation.

108.  l6th Steam Station Cost Survey.   Electrical World, November  3, 1969, PP-  ^1-56.

109.  Choosing  Your Next Plant?  Interview  vith Ken Hamming, Sargent &  Lundy
     Engineers,  pp. 32-3^.

110.  Svengel,  F. M. :  A New Era of Power Supply  Economics.  Power Engineering,
     March 1970.

111.  Pfersdorff, D. H. :  Electric Utility  Gas Turbines - A Maintenance Report.
     Paper No.  66-GT-100,  presented  at Gas Turtine Conference and Products Show,
     Switzerland, March 1966.

112.  Cooling Pond Planned for Cedar  Bayou  Plant.  Electrical World, June 8, 1970.
     p.  28.

113.  Galveston Bay:  Test  Case of an Estuary in  Crisis.  Science, February 20,  1970,
     pp.  1102-1107-

Hk-  Brovn,  V.  D. :   Twentieth Annual Electric Industry Forecast.  £Lectrical
     World, -Sept ember 15,  1969, pp.  93-98.

115.  Bauer,  H.  E. ,  et al. :  Electric Utility Equipment Requirements, II  - Equip-
     aent Trends.  United Aircraft Research  Laboratories Report E-110303-2,
     August  1966.
                                       123

-------
                              REFERENCES (Continued)
 116.   Baldwin, C.  J.:   Probability Calculation of Generation  Reserves.  Westinghousi
       Engineer, March 1969,  pp.  S^-^O.

 117-   Kirchmayer,  L.  K. :   Application of Probability Methods  to  Generating Capacity
       Problems. AIEE Transactions,  February  1961.

 118.   Calabrese, G.:   Generating Reserve Capacity Determined  by  the Probability
       Method.   AIEE Transactions, Vol.  66,
 119.   Miller,  A.  L.:   Details  of Outage Probability  Calculations.  AIEE Transactions
       August  1958.

 120.   Kist, C.,  and G.  J.  Thomas:   Probability  Calculations  for System Generation
       Reserves.   AIEE  Transactions, August  1958.

 121.   Limmer,  H.  D. :   Determination of Reserve  and Interconnection Requirements.
       AIEE Transactions, August  1958.

 122.   Garver,  L.  L.:   Reserve  Planning for  Interconnected Systems.  Power
       Engineering,  May 1970, pp.  1*0-1*3.

 123.   Parisian, R. W. :  How Reliable are Today's Prime Movers?  Power, January 1970,
       pp. 1*5-1*7.                                                                   j
                                                                                   j
 121*.   Carstens, J. P.:  Economic Advantages of  Power System Expansion with         j
       Dispersed Gas Turbines.  United  Aircraft  Research Laboratories Report        !
       B-110052-6, August 1963.                                                     |

 125.   Lessard, R. D. :  Telephone conversation with Mr. P. Ashton of HELCO.
       Memorandum to Mr. F. R. Biancardi, November k, 1970.

 126.   Skaperdas, G. T.:  Commercial Potential for the Kellogg Coal Gasification
       Process.  M. W. Kellogg Co. Research and  Development Report No. 38,
       Office of Coal Research Contract No. lU-01-0001-380 , 1967.

127-  Bituminous Coal Research:  Gas Generator  Research and Development, Survey
       and Evaluations, Phase 1, Vols.  I and II.  Office of Coal Research Contract
      No. 1U-01-0001-32U, 1965.

128.  Rudolph, P. F. H. :  New Fossil-Fueled Power Plant Process Based on Lurgi
      Pressure Gasification of Coal.  Combined Meeting of the American Chemical
      Society and Chemical Institute of Canada, Toronto, Canada, May 1970.
                                       12U

-------
                             REFERENCES (Continued)
129.   Eastman, duB.:  Gasification and Liquefaction of Coal.   Proceedings of
      the American Institute of Mining Engineers, 1953.

130.   Forney, A. J., et al.:  A Process to Make High-Btu Gas  from Coal.   US
      Department of the Interior, Bureau of Mines Technical Progress  Report 2U,
      April 1970.

131.   Bituminous Coal Research:  Coal Gasification for Combined-Cycle Power
      Generation without Atmospheric Pollution.   BCR Report RPP-127 R2,
      June 23, 1967.

132.   Institute of Gas Technology:  Cost Estimate of a 500 Billion Btu/Day Pipe-
      line Gas Plant Via Hydrogasification and Electrothermal Gasification of
      Lignite.  US Department of the Interior, Office of Coal Research,  Research
      and Development Report No. 22, 1968.

133.   Anon.:  Coal-to-Gas Plant "Possible" by 197^.   The Oil  and Gas  Journal,
      October 26, 1970.

I3h.   Theodore, F. W.:  Low Sulfur Boiler Fuel Using the CONSOL  C0£ Acceptor
      Process.  US Department of the Interior, Office of Coal Research Contract
      No.  lU-01-0001-1*15, Report No. 2, PB 176 910,  November  1967.

135.   Linden, H. R.:  Sources of Gas Supply for the USA  to the Year 2000.   Paper
      presented at the International Gas Union Conference, June  1970.

136.   Airbus is Ready, But the Airlines are Hot.   Business Week,  July 18,  1970,
      pp.  80-81.

!37.   tfatts, F.  A.:   Aircraft Turbine Engines.   Development and  Procurement  Cost.
      A Rand Corporation Report RM-H670-PR, November 1965.

3-38.   Allen, R.  P.,  and R. C. Petitt:  New Gas Turbine Design  for Large  Power
      Systems.  Presented at the American Power Conference, April 22, 1970.

139.   Cuffe, S.  T. ,  and R. W. Gerstle:   Emissions from Coal-Fired Power  Plants;
      A Comprehensive Summary.   Presented at the American Industrial  Hygiene Asso-
      ciation Meeting, May 1965-

Ik).   Bagwell, F.  A., et al.:  Oxides of Nitrogen Emission Reduction  Program
      for  Oil and Gas Fired Utility Boilers.   Presented  at the American  Power
      Conference,  April 21-23,  1970.
                                       125

-------
                             REFERENCES (Continued)
lUl.   Bell, A.  W., N. B.  deVolo, and B.  P.  Breen:   Nitric  Oxide Reduction "by
      Controlled Combustion Process.  Presented at the Spring Meeting of Western
      States Section, The Combustion Institute, April 20-21,  1970.

lU2.   Peters, G. T.; Editor:  Reference  Handbook of Prime  Mover Characteristics.
      United Aircraft Research Laboratories Report D-110287-1, 1966.
                                     126

-------
                                                    TABLE I
                    UNITED STATES  CONSUMPTION  OF ENERGY RESOURCES BY ELECTRIC UTILITIES
                                               Trillions of Btu's


Type of Fuel
Coal




Oil




Gas




Hydro




Nuclear




Actual
Consumption
1965
*-
-
6391(2)
6UOO
5880
_
_
890 (2>
TOO
7^3
_
—
2691 (2>
2UHO
2399
_
-
2039 <2;
2090
2050
-
-
52(2)
-
38

Projected
1970
8050
8U93
8035
-
_
1570
837
856
_
_
32*40
3336
2589
-
-
_
80 U
2193
-
-
1000
737
87U
-
-
1975
^
9,050
11, 13^
—
8,520
_
720
863
_
9^0
_
3,963
2,789
-
3,770
_
893
2,J*22
-
2,580
-
5,96U
1,803
-
U,260
1980
7,^50
9,6Uo
12,516
11,000
—
3,230
659
861
625
_
U,6oo
5,156
2,976
U,ltOQ
-
_
1,098
3,027
h,060
—
12,000
13,300
M78
11,300
-
1985
_
9,780
-
-
11,300
_
655
—
—
1,070
__
6,619
-
-
M50
_
1,286
-
-
3,200
—
25,913
-
-
15,500
1990
5910
—
-
-
_
37^0
_
-
—
-
6100
—
-
-
-
_
-
-
-
—
Ul,500
-
-
-
-
2000
_
_
18,720
-
_
_
_
861
_
-
_
—
U.128
—
-
_
-
5,056
-
-
_
-
H3,526
-
-


Reference
(7)
(8)
(11)
(9)
(10)
(7)
(8)
(11)
(9)
(10)
(7)
(8)
(11)
(9)
(10)
(7)
(8)
(11)
(9)
(10)
(7)
(8)
(11)
(9)
(10)
(l)  Converted from data in
     103 Btu/cu-ft for gas.
(2)  Data for 1966 inserted
 Ref.  7 using 26.2 x 106  Btu/ton  for coal,  5.8  x 106  Btu/barrel  for  oil,

,  natural gas liquids  included in oil figures

-------
                                     TABLE II
         REGIONAL ELECTRIC  GENERATION BY FUEL TYPE AND HYDROELECTRIC POWER

                                  Billion kwhr
                                 Data  from Ref, 1

Region

South Central




Southeast




West




East Central




West Central




Northeast





Fuel

Coal
Oil
Gas
Nuclear
Hydro' l)
Coal
Oil
Gas
Nuclear
Hydro
Coal
Oil
Gas
Nuclear
Hydro
Coal
Oil
Gas
Nuclear
Hydro
Coal
Oil
•Gas
Nuclear
Hydro
Coal
Oil
Gas
Nuclear
Hydro
Actual
1966

5-3
.03
110.5
-
-
161.1
18.0
1U.8
-
25.0
12.5<2>
11.6
61*. 1
0-89
120.0
199.7
0-2
0-2
0-9
2.2
99-9
.9
23.9
l.U
12.2
130.5
1*3. U
9-3
2.2
10.1
Projected
19TO

7.35
.12
16U.O
_
-
218.1
1U.7
23.2
7.2
28.7
29-8
17-9
91*. 0
7.U5
159-0
2UU.2
0-3
o.u
U.3
2.6
121*. 8
1.1
30.1
17.5
12.2
157.5
1*3.8
15-3
30.7
13. fl
1975 ' 1980

2U.30
-85
225-50
13.90
-
2U5.8
12. U
27-8
llU.5
35. U
_
-

U3.2
.82
289.0
73.0
-
259. U
12.1*
21*. 6
286.7
36.5
125.0
35.8
1 76.0
; 210.0
;; 18U.O
279-9 3W.9
o.i 1 o.i
0-2
57.5
5-9
- -
1U6.3
1.3
35-0
69.5
12.9
156.1
36.5
16.9
iue.5
15.5
«
0-3
110.0
6.6
ll*2.1*
1.5
3U.5
178.U
13.7
lUl.o
30.1
15.1
309-3
18. U
1985

70.30
.67
339.0
198.0
-
282.3
13.0
36.5
U87.2
38.5
_
-
„
_
-
371.3
»
o-i
23.6
7.0
15U.2
1.7
37-8
300.1
1U.U
122.0
26.0
13.3
1*92.7
23.5
1 ' • • -
1990

9U.O
.65
U31.0
UoU.o
-
316. U
12.3
U5-3
761.3
Ul.3
211.5
30.2
80.6
685.0
198.0 j
U28.1
0-2
0-5
1*20.1
6.6 J
155-3
2.0
39-5
U89-1
15-6
.— -
101.5
23.^
12.1
7U7.1
27-3
ClJ   Data not provided.
C2)   Estimates included for 1965.
                                        128

-------
                                   TABLE III
           REGIONAL FOSSIL FUEL COSTS FOR ELECTRIC ENERGY GENERATION
                                     1965
                                From Ref. 10
Per Cent Total
Fossil Btu
,_Coal
61
77
97
51
80
92
0
U9
0
	 1
Oil
36
16
0
1
12
0
0
U
19
Gas
1
3
Mid
7
East
3
West
1*8
Sov
8
East
8
West
100
U7
81
Average Fuel Cost
^/million Btu
Coal
ev England
32. U
Idle Atlantic
25. ^
North Central
23.7
North Central
25-6
th Atlantic
2U.8
South Central
18. U
South Central
Mountain
19.0
Pacific
Oil
3^. I*
32.3
66.2
50.8
33-7
62.8
-
26.2
32.0
Gas
3^.2
33-8
25-9
2U.2
32.3
23.8
19.8
?7.1
31. U
Weighted Average
Fossil Fuel Cost
^/million Btu
33.8
27.7
23.3
25.5
26.7
19.3
19-8
23.3
31.5
Rote:   Regions are not identical with those designated "by the FPC.
                                       129

-------
                                                             TABLE IV

                                        SULFUR CONTENT AND DISTRIBUTION OF COAL RESERVES

                                                           From Ref. 25
                                           (a)  Sulfur Content of all US Coal Reserves
bituminous
Sub-bituminous
Lignite
Anthracite
Total
Low Sulfur
(1% or less)
Billion Tons
215- T
387.2
1+06.0
H+.7
1023.6
% of Total
13.7
2k. 6
25.8
0.9
65.0
Medium Sulfur
(1.1 to 3.0%)
Billion Tons
19^-9
1.5
1*1.6
0.1+
238. 1*
% of Total
12. U
0.1
2.6
15.1
High Sulfur
(over 3.0$)
Billion Tons
311+.2


311+.2
% of Total
19.9
19-9
Total
Billion Tons
721+.7
388.7
1+1*7.6
15.2
1576.2
% of Total
1+6.0
21+.7
28. U
0.9. _
100.0
U)
o
                                (b)   Distribution  of US  Coal  Reserves According to Heat  Content*


Bituminous
Sub-bituminous
Lignite
Anthracite
Total


Low Sulfur
17.3
22.1+
16.6
1.2
57.5

Medium Sulfur
15.6
0.1
1.7
	
17.1+
t

High Sulfur
25.1
	
	
	
25.1

Total
% of Total
58.0
22.5
18.3
1.2
100.0

  *Based  on heating  values  as  follows
       Bituminous      -  26,200,000  Btu/ton
       Sub-bituminous  -  20,000,000  Btu/ton
       Anthracite      -  25,1+00,000  Btu/ton
       Lignite         -  13,1+00,000  Btu/ton

-------
                                   TABLE V

                       DISTRIBUTION OF COAL WITH SULFUR
                        CONTENT OF ONE PERCENT OR LESS
                                From Ref. 25

est of the Mississippi:
Bituminous
Sub-bituminous
Lignite
Anthracite

ast of the Mississippi:
Bituminous
Sub -b i fund nou s
Lignite
Anthracite
Total

Billion Tons
13U
387
U06
2
929

82
—
—
13
95
' Low-Sulfur Coal
% of Total
13.1
37.8
39.6
0.2
90.7

8.0
	
	
1.3
9.3
Distribution by States
West Virginia
Kentucky (eastern portic
Virginia
Alabama
Pennsylvania
Other
Total
Geological
Reserves in
Place*
Billion Tons
U7.5
an) 22.1
8.1
2.1
1.2
1.0
82.0
Low-Sulfur
Bituminous Output
Million Tons
89.0
UO.U
26.0
8.9
1.7
166.0
% of Total
State Output
63
90
82
62
i
% of Total
National
Output
18. U
8.3
5.U
1.8
O.U
3U.3
Only 50% of the geological reserves are recoverable,
                                      131

-------
            TABLE VI

SUMMARY OF PROJECTED FUEL COSTS
 IN SELECTED REGIONS OF THE US

      From Various Sources
1
Fuel Price Estimates (tf/ndllion Btu)
excluding transportation

Coal (in Rocky Mountain Region)
Natural Gas
Oil, No. 6
Oil, Lov Sulfur
Uranium
Thorium

Coal
Natural Gas
Oil, Lov Sulfur

Coal, Lov Sulfur
Coal, High Sulfur
Oil, Lov Sulfur

Coal
Natural Gas (Plus Gasified Fuel)

Coal
i
!0il, High Sulfur
Oil, Lov Sulfur

Coal, High Sulfur
Coal, Lov Sulfur
Residual-Oil, Lov Sulfur
Residual-Oil, High Sulfur

Regions
1968 1970 I960
West
16 15 16
30 31 3U
32 32 28
Ul UU
26 20 15
20 15
South Central
17 to 29
23
37 to 1*0
East Central
35 to HO UO
27-5 25
32
Southeast
25 30
25 30
West Central
32

35
-
Northeast
25-33 30
U5-55 55
50 U5
25 30


1990

17
36
32
kk
13
13

25 to 31.5
21 to 38
-

Ho
25.0
30.0

32
UO +

30

35
-

30
U5
^
35
             132

-------
                                              TABLE  VII

                 SUMMARY OF EXISTING AND EMERGING  REGIONAL WATER MANAGEMENT PROBLEMS

                                             From  Ref. 3>*
Vater Resource Regions
:?crth Atlantic
Scuth Atlantic-Gulf
3reat Lakes
Ohio
Tennessee
'Jyper Mississippi
lever Mississippi
Scuris -Red-Rainy
Missouri
Ar> an s a s -Wh i t e- R e d
Texas Gulf
?.ID Grande
,'pper Colorado
lever Colorado
"rc-at Basin
-dumbia-North Pacific
California
Alaska
~£";aii
?«rto Rico

Adequacy of* '
Annual Natural
Runoff
3
k
3
3
k
3
U
2
2
2
2
1
1
1
1
3
2
h
k
k

Ground Water \2'
Storage Depletion
3
3
U
1*
U
l+
14
1*
2
1
1
1
U
1
3
3
2
U
3


Water Quality
Wastes *3)
1
2
1
2
3
2
3
3
3
3
2
2
3
2
2
3
2
3
3


Heat^
1
3
1
2
3
2
Salinity* 5)
1+
U
u
1+
1+
14
U ! U
u ; 3
3 i 3
U 1
3 i 2
U 1
U
u
u
2
3
u
u


3
1
3
U
2
14
U


Sedinent*6^
3
2
3
3
3
3
1
U
2
2
2
1
2
1
3
k
3
h
I


Comparison of projected consumptive use with natural  runoff
which includes perennial yields of ground water aquifiers.

An indication of the extent that use of ground water  would
exceed recharge.

An indication of pollution loading and of investment
required for alleviation.

An indication of waste heat discharges from industrial  and
steam-electric cooling requirements and of investment
required for alleviation.

An indication of the relative severity of the salinity
problem from both natural sources and man-caused sources

An indication of the relative severity of sediment  from
land and stream bank erosion both natural and man-caused
Order of Severity:

Severe problem in some areas or
major problem in many areas

Major problem in some areas or
moderate problem in many areas

Moderate problem in some areas
or minor problem in many areas

Minor -problem in some areas
                                              133

-------
                        TABLE VIII

            LIMITING TEMPERATURE CRITERIA IN
WATER QUALITY STANDARDS FOR SOUTH CENTRAL POWER REGION

                       From Ref. 33

                 (Temperatures in Deg F)
States
and
Other
Juris-
dic-
tions
Ark.
Kan.
La.

Miss .
Mo.



Okla.
Tex.



Limiting Uses

Cold
Water
Fish
Temp.
68








70




Rise
5








5





Small-
mouth
Bass
Temp.
86








75




Rise
5








5





Warm
Water
Fish
Temp. Rise
95








93




5








5




All Waters




Temp.
95
90
97

93
90



93
96
93


Rise
5
5
5

10
5



5
5
5
k
1.5





Exceptions and Remarks

Not approved.
Except some rivers with 95° max. and
U° rise
10° rise not approved.
Except Des Moines vhere max. temp, is 93°
and except North Fork White, Current,
and Eleven Point Rivers where max. rise
is 2°

Except Canadian River and tidal waters.
Canadian River
Fall, winter, spring - tidal waters
Summer tidal waters

-------
                                              TABLE IX

              SCHEDULED ADDITIONS OF STEAM POWER ELECTRIC GENERATING CAPACITY BY YEARS

                                            From Ref.  2



Year
1969 (last 3 months)
1970
1971
1972
1973
197^
1975 and later
Total

Conventional
dumber of
Units
21
53
51
1+1
31+
lU
5
219
Total
Capacity, Mw
6,13)4
16,689
18,876
19,888
17,371
6,115
3,721
88,793
Average Unit
Capacity, Mw
292
311+
370
U85
511
1+36
731+


Nuclear
Number of
Units
1+
7
11
16
16
12
9
75
Tbtal
Capacity, Mw
1,817
U.865
8,6U3
13,531
15,396
11,009
8,U92
63,752
Average Unit
Capacity , Mw
1*55

785
81+5
960
920
935

(l)   Based on scheduled dates of commercial operation as  of October 1,  1969,
     in terms of manufacturers'  ratings  of the units.

-------
                                              TABLE X

         SCHEDULED ADDITIONS OF STEAM POWER ELECTRIC GENERATING CAPACITY BY REGIONS

                                            From Ref. 2

Region
Northeast
East Central
Southeast
West Central
South Central
West(2)
Total
Conventional
Number of
Units
2U
k2
37
33
61
22
219
Total
Capacity , Mw
10 ,771
19.6U5
17,999
9,025
20,303
9»050
88,793
Average Unit
Capacity, Mw
U50
U67
1*86
27^
330
1+11

Nuclear
Number of
Units
25
8
20
Ik
1
2
75
Total
Capacity, Mw
22,137
7,31*3
18,102
10,99^
903
^,373
63,752
Average Unit
Capacity, Mw
885
915
909
4*5
903
1,093

(l)   Based on scheduled dates of commercial operation  as  of October 1, 1969 in terms of manufacturers'
       ratings of units.   Regions defined by map shown in Fig.  3.

(2)   For purposes of simplification,  estimates  of the  power additions  from original 8 FPC power
       regions have been incorporated into present 6 regions.

-------
                     TABLE XI
ESTIMATED PERFORMANCE OF TYPICAL STEAM POWER STATIONS
1970 Decade and Early 1980's
Coal
Residual Oil
Natural Gas
! Late 1980 's Decade
i
Coal
Residual Oil
Natural Gas
Net Station Efficiency, %
@ 70/5 Load Factor
36.6
37.0
36.0


38.6
39.0
38.1
@ Design Point
38.1*
38.9
37.8


1*0.5
1*1.0
1*0.0
                        137

-------
                                    TABLE XII
              CAPITAL COST SUMMARY FOR COAL-FIRED STEAM STATIONS
Land and Land Rights
Structures and Improvements
Boiler Plant Equipment
Turbogenerator Units
Accessory Electrical Equipment
Misc. Power Plant Equipment
Transmission Plant
General Expense and Overhead
Station Equipment
Direct Design Cost and
Final Drawings
Other Expenses
Subtotal
Interest During Construction
Engineering, Design, Construction
Supervision, and Contingency
Escalation
Total
TVA Bull Run l^1'
$/kw 1.88
21.36
59.^7
22.07
6.31*
2.25
U.82
21. 8U
(3)
7.6U
(3)
$/kwiU7.67
9.97
(3)
(3)
$/kw!57.6U
t
Design Studv^2'
$/kw 0
13.07
61.00
36.08
11.05
0.52
(3)
(3)
1.72
(3)
1.25
$/kw 12^.72
19.80
12. 2U
20.20
$/kw 176.96
(l)  Represents costs for an actual plant which  has been  completed  (Ref.
(2)  Represents estimated costs  for a  plant which  could be built  (Ref. U5).
(3)  Not applicable due to the differences in  the  method  of reporting costs
                                        138

-------
                                                   TABLE XIIX

                INVESTMENT COSTS FOR ALTERNATE METHODS OF COOLING CONDENSER WATER DISCHARGES

                                                      $/kw
Cooling Water System " ""'— —— ^_
Once-Through River
Once-Through Ocean
Cooling Pond/Reservoir
Spray Pond
Spray Cooling Canals
Wet Cooling Tower - Mech. Draft
Wet Cooling Tower - Nat. Draft
Dry Cooling Tower - Mech. Draft
i
Dry Cooling Tower - Nat. Draft
Fossil-Fueled Plants
Data Source
^Eef, 63
6.25
6.11
8.50
—
—
8.00
11.25
•»
19.25
39.00
Ref. 56*
5.30-5-00
6.00-6.30
6.50
7.60
—
7.20
7.50-8.50
13.0
20.0
Ref. 60
Base
—
1.65
—
3.U1
3.75
6.92
19.07
20.82
Ref. 62
—
—
—
—
—
—
—
17
20
Nuclear-Fueled Plants
Data Source
Ref. 63
9.25
9.00
12.00
—
—
11.75
17.5
30.5
62.5
Ref. 56*
5.2U-5.88
6.2U-6.88
7.50
8.10
—
9.^0
11.50-12.50
15.00
22.00
Ref. 35***
8.00
9.68
9.65
—
—
11.9^
lU.17
29.90
—

Ref. 62**
—
—
—
—
—
—
—
23
27
*    Costs vary with temperature rise in condenser from 10 to 20 F.

**   Costs vary with condenser design pressure from 5-5 to 16.0 in. Hg abs.

***  Values presented in Ref. 35  are relative to a base for once-through river cooling.
     This base value was selected as $8.00/kw for comparison purposes only.

-------
                                                           TABLE XIV
                                   ADDITIONAL COST  FACTORS  FOR  ALTERNATIVE COOLING SYSTEMS
Cooling Water System
Once through River
Once through Ocean
Cooling Pond /Reservoir
Spray Pond or Canal
Wet Cooling Tower-Mech.
Draft
Wet Cooling Tower- Nat.
Draft
Dry Cooling Tower-Mech.
Draft
Dry Cooling Tower-Nat .
Draft
Auxiliary Power
Requirements * '
% Generator)
V Output /
O.U25
0.375
0.1*25
0.875
1.075
0.875
3. OU
0.91
Added Fuel Cost^2^
for Auxiliaries
/mills/kwhr)
0.0116
0.0102
0.0116
0.021*0
0.029^
0.021*0
0.0930
0.021*9
Added Cost^3^
for Auxiliary Power
/mills/kwhrj
0.0085
0.0075
0.0085
0.0175
0.0215
0.0175
0.0605
0.0182
Maintenance ,
tfater Treatment
mills/kwhr
o .0058
0.0050
0.0058
0.0120
0.01^7 •
0.0120
0.01U7
0.0120
Loss irT '
Capability
/mills/kwhr)
0
0
0.012
0.012
0.012
0.012
0.18
0.18
(5)
Added Fuel
Coats
mills/kwhr)
0
0
0
0 .0108
0.0108
0.0108
o .2705
o .2705
(l)  Based on data from Ref. 63.
(2)  Based on heat rate of 10,200 Btu/kwhr for dry tower-mechanical draft, 10,000 Btu/kwhr for dry
       tower-natural draft, 9,110 Btu/kwhr for all others, and fuel cost at 30<£/106 Btu.
(3)  Based on $100/kw, &0% load factor, lU/f capital charges.
(1*)  Based on $100/kw for incremental capacity.
(5)  Cost factors for loss in capability and added fuel cost will depend on climatic  conditions  (see Ref. 60).

-------
                                     TABLE XV
                         GAS TURBINE COMBUSTOR MATERIALS
Metal Surface
Temperature - F
1200*
1600
1800
2000
2200
2300
Material
AISI Type 310
Hastelloy X
Haynes 188
TD-Ni
TD-NiCr
Dispersed thoria
in nickel
Condition
Uncoated
Uncoated
Coated
Coated
Uncoated
—
Status
Industrial
Production
New Engine
New Engine
Experi-
mental
Experi-
mental
Primary Properties
Strength
Strength; oxidation resistant
Strength; creep; fatigue; and
oxidation resistant
Strength; creep; fatigue; and
oxidation resistant
Strength; creep; fatigue; and
oxidation resistant
Strength; creep; fatigue
*  Turbine inlet temperatures  are normally  some  600  F  above these  values in present
   engines.

-------
                                                          TABLE  XVI
                                   PROJECTED TECHNOLOGY FOR BASE-LOAD GAS TURBINE ENGINES
                                     Simple-Cycle and Regenerative-Cycle  Power  Systems
Parameter
Turbine Inlet Gas
Temperature - F
Compressor Pressure Ratio
Compressor Polytropic
Efficiency - %
Turbine Nominal Adiabatic
Efficiency - %
Regenerator Airside
Effectiveness - %
Regenerator Total
Pressure Drop - %
Turbine Cooling Technique

19 TO Decade
1600 to 220o(1)
2200 to 21400(2)
U to 28
89

90
TO to 90
h and 8
Advanced
Impingement-Convection
Time Period
Early 1980's
2000 to 2Hoo(1)
2UOO to 2800(2)
* k to 26
92

92
TO to 90
k and 8
Advanced
Impingement-Convection
Late 1980 's
2UOO to 2800J1)
2800 to 3100(2)
U to 36
93

93
TO to 90
h and 8
Advanced
Impingement -Convection
H
rv>
        (l)  Compressor bleed air uncooled

        (2)  Compressor bleed air precooled to levels of 125 to 250 F.

-------
                                    TABLE XVII
             INFLUENCE OF COMPRESSOR PRESSURE RATIO ON POWER COST
                                200-Mw Unit Size
Turbine Inlet
Temperature

2000 F
2^00 F
2600 F
2900 F
Technology
Time Period

1970 Decade
1970 Decade
Early-1980's
Decade
Late-1980's
Decade
Compressor
Pressure Ratio

13:1
20:1
28:1
13:1
20:1
28:1
12:1
20:1
28=1
20;1
28:1
36:1
Power Cost Differential*
mills/kwhr
Fuel Costs
30tf/106 Btu 50tf/106 Btu
+ O.llt + 0.25
base base
+0.22 +0.31
+ 0.2U +0.38
base base
+ 0.03 + 0.007
+0.380 +0.67
+0.091 +0.160
base base
+0.23 +0.381
+ O.OU + 0.09
base base
*  Based on 0.8 load factor, 15$ capital charges.

-------
                               TABLE XVIII
                  APPROXIMATE  CHARACTERISTICS  OF
                     COMPOUND-CYCLE GAS TURBINE DESIGN
                          Early-1980's Technology
 Turbine  Inlet  Gas  Temperature -F

 Reheat Gas  Temperature - F

 Cycle Pressure Ratio

 Low-Pressure Compressor Ratio

 Engine Airflow  Ib/sec

 Net Thermal Efficiency - %

 Compressor  Inlet Diametej. - ft

 Low-Compressor Stages

 High-Compressor Stages

 Gas Generator  Turbine Stages

 Power Turbine

     Stages

     Rotational Speed - rpm

     Last-Stage Tip Diameter - ft

     Last-Stage Blade Height-in.

     Last-Stage  Blade Root Stress - psi

Engine Selling  Price - $/kw
2200

2200

75:1

U:l

2005
11.7

7

13

3



3

1800

lU.2

26.9

UT,000

20
                                    Ikk

-------
                                              TABLE XIX





                      CHARACTERISTICS OF ADVANCED GAS TURBINE POWER STATIONS


Time Period


Turbine Inlet
Turbine
F

Engine Pressure
Ratio


Engine
Size
Mw
Engine
Thermal
Efficiency
%
Station
Thermal
Efficiency
*

Net
Heat Rate
Btu/kwhr
Engine
Selling
Price
$/kw

Station*
Prices
$/kw
                                       SIMPLE-CYCLE ENGINES
1970-decade
Early 1980's
Late 19bO's
2200 to 2HOO
2^00 to 2600
2600 to 2900
20:1
28:1
30:1 to 36:1
200
250
250
31.9
37. 3
ItO.I*
30.6
35.8
38.7
11,160
9,530
8,820
32.0
2l*.5
28.0
80.0
66.5
71.0
                                    REGENERATIVE-CYCLE ENGINES
1970-decade
Early 1980 's
Late 1980 's
2000 to 2200
2200 to 2l*00
2600 to 2800
10:1
12:1
1U:1
200
200
200
35.14
39.2
1*3.1
33.9
37.6
1*0.6
10,070
9,080
8,1*10
1*8.0
1*3.7
HU. 0
100
92.7
93.0
*  Indoor construction assumed, outdoor construction will be only about $1 to 2/kw lower than figures shown

-------
                                     TABLE XX
                           POWER  PLANT  CHARACTERISTICS
                          EARLY-1980 DESIGN TECHNOLOGY
Cycle
Nominal Output -Mw
Turbine Inlet Temperature - F
Compressor Pressure Ratio
Engine Airflow - rb/sec
Compressor Stages Low /High*
Compressor Inlet Tip Diameter _ ft
Compressor First-Stage Blade Height -in.
Compressor Exit Tip Diameter
Compressor Last^ Stage Blade Height -in.
Compressor Turbine Stages High/Lov*
Compressor Turbine First- Stage Blade Height -in.
Compressor Turbine Las'U-Stage Blade Height in.
Power Turbine Stages
Power Turbine Rotational Speed -rpm
Power Turbine Last-Stage Tip Diameter -ft
Power Turbine Last-Stage Blade Height -in.
Power Turbine Last-Stage Blade Centrifugal Stress -psi
Compressor Turbine First-Stage Vane Temperature -- F
Compressor Turbine First-Stage Blade Temperature - F
Compressor Bleed Flow for Turbine Cooling ** - %
Power Turbine Exhaust Temperature - F
Simple
250
2600
28:1
1295
13/5
9-5
20.
5-5
2.3
1/3
7.7
12.2
3
1800
12.5
22.8
36,200
1925
15^0 .
90
99i
Regenerative
200
2UOO
12:1
1175
Ik
8.9
20.
7.5
3.6
2
9.0
13.7
2
1800
13.0
2k
1*0,000
1925
]686
5.5
866
*   Twin-spool design used for simple-cycle power plants.

**  Compressor bleed air precooled to 200 F.

-------
                                                      TABLE XXI

                              1000-MW STEAM-ELECTRIC STATION COSTS - 1980-DECADE DESIGNS

                                            COSTS IN 1970 DOLLARS PER KILOWATT
Steam Conditions
Number of Units
Fuel
Type Construction
Location
Construction Time, Years
FPC
Account
No. Description
310 Land and Land Rights
311 Structures and Improvements
312 Boiler Plant Equipment*
3lU Turbine Generator Units**
315 Accessory Electrical Equipment
3l6 Miscellaneous Power Plant Equipment
; 353 Station Equipment
Total
Other Expenses
Subtotal
i Engineering Design, Construction
Supervision, and Contingency
Subtotal
Escalation
Subtotal
Interest During Construction @ Q%
Total
3500 psig/1000 F/1000 F
One
Coal
Indoor
East Central
It



$/kw 0.03
9-00
5U.83
3U.20
10.02
O.U6
1-55
110.09
1.2U
111.33
13.08

12U.U1
18.35
1U2.T6
22.83
£/kw 165.59
3500 psig/1000 F/1000 F
One
Oil
Indoor
Northeast
U



$/kw 0.22
T.6U
U8.98
33.69
9-52
0.50
1-59
102. lU
1.22
103.36
12.28

115. 6U
17.06
132.70
21.23
$/kw 153-93
3500 psig/1000 F/1000 F
One
Gas
Indoor
Northeast
U



$/kw 0.22
7-03
38.36
33.69
9-52
0.50
1.59
90.91
1.22
92.13
11.5U

103.67
15.26
118.93
18.98
$/kw 137-91
 * Includes cost of stacks, and dust collectors
** Includes cost of cooling tower

-------
                                                        TABLE XXII
                           1000-MW  STEAM-ELECTRIC  STATION CAPITAL COSTS - 1970-DECADE DESIGNS


                                           (COSTS IN 1970 DOLLARS PER KILOWATT)


                                                     Two 500-Mw Units

                                       Steam Conditions:  2UOO psig/1000 F/1000  F


                                                 Construction Time U Years

                                          Interest  Rate During Construction - B%
Fuel
Coal
Oil
Gas
Type
Construction
Indoor
Outdoor
Indoor
Outdoor
Indoor
Outdoor
Region
Northeast
$/kv 188
172
172
158
152
fc/kw 136
East
Central
$/kv 180
165
166
152
1U6
$/kv 131
South-
east
$/kv200
18U
isi*
168
161
$/kwlH5
West
Central
l/kv 173
16U
16U
150
lUU
$/kw 129
South
Central
$/kw 183
168
168
15U
1U8
$/kw 132
West
t/kv 183
168
1^4'
15U
1U8
t/kv 132
H
-p-
00

-------
                            TABLE XXIII

1000-MW STEAM-ELECTRIC STATION CAPITAL COSTS - 1980-DECADE DESIGNS

               (COSTS IN 1970 DOLLARS PER KILOWATT)

                         One 1000-Mw Unit
           Steam Conditions:  3500 psig/1000 F/1000 F

                     Construction Time H Years
Fuel
Coal
Oil
I
Gas
Type
Construction
Indoor
Outdoor
Indoor
Outdoor
Indoor
Outdoor
Interest
Rate During
Construction
6%
&%
10%
6%
aof
O/a
10 %
6%
Q%
10$
6%
Q%
10%
6%
Q%
10%
6%
Q%
10%
Region
Northeast
$/kw 166
172
178
151
157
161
1H9
15U
159
13U
139
lUU
133
138
1U3
119
123
128
East
Central
$/kwl60
166
171
1U6
151
155
1U3
1U8
15^
129
13U
138
128
132
137
11U
119
123
South-
east
^/kw!77
18U
190
162
168
172
159
16U
170
Ikk
1U9
15^
1U2
lU?
153
127
132
136
West
Central
£/kw!58
16U
170
lUU
1^9
153
1^1
1U6
152
128
132
137
127
131
136
113
117
121
South
Central
$/kwl62
168
17U
ihQ
153
158
1^5
150
156
131
136
lUi
130
13U
139
116
120
12U
West
$/kv 162
168
171*
ikQ
153
157
ife
150
155
131
136
lUl
130
13U
139
11^
120
12»i

-------
        TABLE XXIV
BREAKDOWN OF FIXED CHARGES
' Interest on Borroved Capital:




      Internal Capital




      Debt Capital




      Equity Capital










 Insurance




 Taxes




 Annual Maintenance




 Depreciation
         Subtotal
          Total
Fixed Charges,





     l.UO




     U.80





     3.00




     9.20$




     0.60





     1.U5




     0.50
                                                      15.00$
          150

-------
                                     TABLE XXV
                 DETAILED COST BREAKDOWN FOR 1000-Mw SIMPLE-CYCLE
                    AND REGENERATIVE-CYCLE GAS  TURBINE STATIONS

                              (COSTS IN 1970 DOLLARS)

                             Early 1980-Decade  Designs


FPC Account No. 3^1 -
Structures and Improvements
Site Improvements
Site Grading
Building Excavation
Borings
Landscaping
Fresh Water Supply
Fire Protection
Sewage Disposal and Drainage
Flagpole
Guard House
Railroad
Roads and Parking Lot
Fencing
Switchyard
Structures
Administration Building
Turbine-Generator Building*
Gas Meter Area
Subtotal
Four
250-Mw Units \
Simple
Cycle :

$ 25,000
10,000
6,000 :
23,000 i
12,000
100,000 ;
19,000 '
5,000 ;
7,600
50,600
20,900
15,000
10,700

1+07,500
3,230,000
2,700
$3,9^5,000
Five '
200-Mw Units
Regenerative
Cycle

$ 25,000
12,500
6,500
25,000
12,000
100,000
19,000
5,000
7,600
50,600
20,900
15,000
10,700

U07,500
3,916,000
2,700
$U, 636, ooo
*  Indoor construction.  Outdoor construction cost would equal J0% of this value,
                                       151

-------
TABLE XXV (Cont'd.)

FPC Account No. 343 -
Prime Movers
Gas Turbines
Start-Up Motors
Torque Converters
Lute Oil Purification and Storage
Lute Oil Fire Protection
Turbine Air Precooler System
Air Compressor Serv. & Inst .
Breeching Incl. Liners, Silencers, and
Insulation
Expansion Joints
Inlet Filter Screen
Turbine Enclosure Air Coolers
Emergency Cooling Water, Tank, Pumping
and Piping
Misc. Pump and Tanks
Control Boards, Inst. and Controls
Computer
Piping
Insulation
Regenerators
Subtotal
FPC Account No. 344 -
Generators
Generators
Hydrogen Seal Oil Coolers
Subtotal
FPC Account No. 345 -
Accessory Electrical Equipment
Auxiliary Transformers
Start-Up Transformers
8000A Insul. Phase Bus Duct
1200 A Insul. Phase Bus Duct
Potential Transformers
Surge Protection
Four 250 -Mw Units
Simple
Cycle
$ 24, 550 ,000
30,000
300,000
74,ooo
80,000
1,000,000
50,000

2,700,000
120,000
135,000
80,000

10,000
20,000
200,000
200,000
980,000
152,000
	
$ 30,681,000


$ 9,875,000
40,000
$ 9,915,000


29,100
86,200
	
62,100
39,000
19,855
Five 200-Mw Units
Regenerative
Cycle
$ 27,600,000
37,500
375,000
88,000
100,000
^37,500
50,000

3,160,000
150,000
150,000
100,000

10,000
25,000
250,000
200,000
1,000,000
150,000
16,100,000
$ 49,983,000


$ 9,875,000
50,000
$ 9,925,000


48,500
35,000
1,122,800
103,500
65,000
32,000
        152

-------
TABLE XXV (Cont'd.)

FPC Account No. 3^5 -
Accessory Electrical Equipment
U80 Volt Power Svitchgear
U80 Volt Motor Control Centers
[Remote Motor Controls
; Duplex Relay Switchboard
; Annunciator Panel
Control Console
Turbine Control Panel
Temperature Detection Panel
Equipment Connect
Testing
250 Volt DC Switchboard
250 Volt DC Panelboard
Station Battery and Rack
Battery Chargers
Four 250-Mw Units
Simple
Cycle
$ 77,325
33,235
2,625
68,000
16,500
3^,500
6,000
15,000
1,800,000
378,300
27,500
3,600
53,000
56,500
Cable Tray I 82,000
600 Volt Instrument Cable 60,UOO
600 Volt Control Cable 122,000
Grounding Systems 370,500
U80 Volt Valve Control Center j 25,600
Conduit - Fittings
600 Volt Power Cable
1000 Volt Power Cable
12,OOOA Insol. Phase Bus Duct
Subtotal

FPC Account No. 3^6 -
Misc. Power Plant Equipment
Laboratory and Sampling Equipment
Tools, Shop, Stores, and Work Equip.
Lockers
Emergency Equipment
Misc. Cranes and Hoists
Portable Fire Extinguishers
Communication Equipment
Lunch Room Equipment
Office Furniture and Machines
Subtotal
FPC Account No. 353 -
Station Equipment
Station Transformer
L_
1*8,200
23,935
35,125
1,073,900
$ u, 650, ooo



10,000
75,000
3,000
10,000
15,000
20,000
50,000
20,000
15,000
$ 218,000

_
$ 1,719,000

Five 200-Mw Units
Regenerative
Cycle
$ 11U.2T5
1*6,615
3,500
68,000
16,500
3^,500
6,000
15,000
1,800,000
378,300
27,500
3,600
53,000
56,500
82,000
80,000
160,000
370,500
25,600
61,205
31,330
^3,5^5
___
$ U,88U,270



10,000
75,000
3,000
10,000
15,000
20,000
50,000
20,000
15,000
$ 218,000


$ 1,719,000

         153

-------
           TABLE XXVI
1000-MW C-AS TURBINE STATION COSTS

    (COSTS IN 1970 DOLLARS)
     Early 1980 Technology
Cycle .
Turbine Inlet / Compressor /Regenerator
Temperature / Pressure Ratio/ Effectiveness
Number of Units
Fuel
Type Construction
Net Station Efficiency, % \
Construction Time , yr
Net/Gross Output, Mw

FPC Account

Number Description
31*0 Land & Land Rights
3l*l Structures & Improvements
3U3 Prime Movers
3l*l* Generators
3l*5 Acces. Elect. Equipment
| 3l*6 Misc. Power Plant Equip.
353 Station Equip.
\
Subtotal
Other Expenses
Subtotal
Engineering, Design,
Construction, Supervision
and Contingency
Subtotal
Escalation
Subtotal
Interest During
Construction
^ TOTAL
Simple-Cycle
2600 F/28:l

1*
Gas
Indoor
35.8
2
1000/lOUO




$ 100,000
3,91*5,000
30,681,000
9,915,000
^, 650, ooo
218,000
1.719,000

$51,228,000
1,250,000
$52,1*78,000


5,61*0,000
$58,118,000
3,500,000
$6l,6l8,000

U, 925, ooo
$66,5^3,000
Simple-Cycle
2600 F/28:l

1*
Gas
Outdoor
35.8
2
1000/101*0 *




$ 30,000
2,975,000
30,681,000
9,915,000
U, 650, ooo
218,000
1,719,000

$50,188,000
1,250,000
$51,1*38,000


5,529,585
$56,967,585
3,^20,000
$60,387,585

U, 820, 000
$65,207,585
Regenerative
21*00 F/12:l/80*

5
Gas
Indoor
37.6
2
1000/101*0




$ 100,000
1*, 636, 000
1*9,983,000
9,925,000
1*. 881*, 270
218,000
1,719,000

$71,1*65,270
1,250,000
$72,715,270


8,000,000
$80,715,270
5,110,000
$85,825,270

6,860,000
$93,685 ,270
Regenerative
21*00 F/12:1/80J

5
Gas
Outdoor
37.6
2
1000/101*0



L
$ 30,000
3,1*60,000
1*9,983,000
9,925,000
1* ,881*. 270
218,000 1
1,719,000

$70,219,270
1,250,000
$71,1*69,270


7,850.000
$79,319,270
5,000,000
$8U, 319, 270

6,71*0,000
$91,059,270

-------
                                                  TABLE XXVII

               POWER GENERATION COSTS FOR ADVANCED POWER SYSTEMS LOCATED IN SOUTH CENTRAL REGION

                                                1970-Decade Designs
                                                  70/5 Load Factor
                                               Outdoor Construction
                                            Gas Cost 23<£/million Btu(3;
                                                 1000-Mw Stations
Operating Conditions

Number of Units

Station Thermal Efficiency, %

Net Station Heat Rate, Btu/kwhr

Station Installed Cost(1), $/kw

Capital Charges'2', mills/kwhr

Operation and Maintenance, mills/kwhr

Fuel, mills/kwhr

Busbar Power Cost, mills/kwhr
                                                  Steam Turbine
Simple-Cycle
Gas Turbine
Regenerative-Cycle
   Gas Turbine
2UOO psig/1000 F/1000 F
Two 500-Mw
36
9^90
132.3
3.250
0.365
2.180
5-792
Pr = 20:1, 2>*00 F
Five 200~Mw
30.6
11,160
80
1.980
O.TOO
2.561
5.2U1
Pr = 10:1, 2200 F
Five 200-Mw
33.9
10,070
100
2.U50
0.800
2.318
5.568
(l)  8% interest rate during construction
(2)  15$ fixed charges
(3)  Gas cost estimates from Table VI

-------
                                                  TABLE XXVIII

               POWER GENERATION COSTS FOR ADVANCED POWER SYSTEMS LOCATED IN SOUTH CENTRAL REGION

                                                Early 1980's Designs
                                                  1Q% Load Factor
                                           Gas Costs 30$/million Btu(3}
                                               Outdoor Construction
Operating Conditions

Number of Units

Station Thermal Efficiency, %

Net Station Heat Rate, Btu/kwhr

Station Installed Cost(i;, $/kw

Capital Charges (2\ mills/kwhr

Operation and Maintenance, mills/kwhr

Fuel, ,mills/kwhr

Busbar Power Cost, mills/kwhr
(l)  Q% interest rate during construction
(2)  15$ fixed charges
(3)  Gas cost estimates from Table VI
                                                  Steam Turbine
Simple-Cycle
Gas Turbine
Regenerative-Cycle
    Gas Turbine
3500 psig/1000 F/1000 F
One 1000-Mw
38
8955
120. k
3.009
0.312
2.690
6.011
Pr = 28:1, 2600 F
Four 250-Mw
35.8
9530
66.5
1.630
0.700
2.860
5.190
Pr = 12:1, 2HOO F
Five 200-Mw
37.6
9080
92.7
2.270
0.800
2.722
5-792

-------
                                                      TABLE XXIX

              POWER GENERATION COSTS FOR ADVANCED POWER SYSTEMS LOCATED  IN SOUTH CENTRAL REGION

                                                 Late  1980's Designs
                                                  10% Load Factor
                                            Gas  Costs Uotf/million Btu(3)
                                               Outdoor Construction
Operating  Conditions

Number  of  Units

Station Thermal  Efficiency,  %

Net  Station  Heat Rate,  Btu/kwhr

Station Installed Cost(l;,  $/kw

Capital Charges ^2)

iOperation  and Maintenance,  mills/kwhr

Fuel, mi11s/kwhr

Busbar  Power Cost, mills/kwhr
Simple-Cycle Regenerative-Cycle
Steam Turbine Gas Turbine Gas Turbine
3500 psig/1000 F/1000 F
One 1000-Mw
39.1
87^0
120. U
3.009
0.312
3.^90
6.811
Pr = 30:1, 2800 F
Four 250-Mw
38.7
8820
71.0
1.7!*2
0.700
3.525
5.967
Pr = 1U:1, 2600 F
Five 200-Mw
U0.6
8U10
93.0
2.228
0.800
3.365
6.393
 [l)   Q%  interest  rate  during  construction
 [2]   15% fixed  charges
 [3)   Gas cost estimates  from  Table VI

-------
                                                                                             PROPOSED ASW SPECIFICATIONS FOR GAS ruRBKE


>e*J gnat las
/ \
|a^




)o, 2-GT



io, J-GT(e)



*o. U-OT






Grade of G«

'or ga* turbine* re-
ulring a more clean
urn Ing fuel than
o. 2-QT.
dl*tUl«t* fuel of
ow a*h and medium
ga* turbine* not re-
quiring Ho. 1-CT.
A 1« volatility, lov
ash fuel that **y
contain recidual com-
>on*nts ,
H low volatility fuel
containing residual
L'c«tpan*nt* and h*vitig
ilgher vanadium conten
than Ho. 3-CT.


Fla*h
Point.
.ld*< CJ
Kin
ioo( 38)
or legal



100 (3fl)
or legal


130(510
or legal


150(66)
or legal





Pour
Point
deg f
de* C)
Max
0
(-16)
Ml


M
(-7)
(d)


..



-.






Water
and
by Vol \me
MM
0.05



0.10



1.0



1.0






Carbon
Heal due
on 10*
Perrent
Max
0-15



0.35



_»



	






Ash.
erctnt
'ejgftt^
0.01



0.01



0.03



—






DUtlllHior
9O< Point.
Kin Ma*
550
(200)



51.0 675
252) (357)


—



	








(b)
Eajbolt VUcoslty. aecv
Universal Futol
at «t
•Ha
—



>2 6)



..



_




JO C
Max

3U. It)



(•5)



_
























l^<
Mln
—







«5



k5




Hu
—







300



300





Kinematic Vlaeoaity,
C-ntLtok.."1'
at 100 F '^fl r^ "* 1^>;) ' (sod
Min
l.li



2.0



(5.8)



(5.8)




Max
2.5



5.8












-*- - • * • r *
Max
"







(638)



638)






ravity
API
Mill
35



30



	



	






Vanadium
(V).
Height
Mai




2



2
(•)


500




Sodimt
plus
Potaasiua
(». * K).
ppin by
*•'«"* 	
Kax




5



5
(*)


10
(f)





C aid Hi
(C.).
ppa by
Max
5




10



10
(•)


10
(r)





Uad
(Pb).
pjm by
u.laftt
Max
5




5



5



5




Ma|».ll«-
to
Vanadlua
U.i»ht
Ratio
Min




—



-



3.0
'*'



p«ia




-



-



3.5
l*V



(») Ho. 1-CT:  Correapondi in g«n*r»l to ASTM D396, Gradt Io, 1  fu*l  and AStK D975 Grade  »o.  1-D dt«»el
    fuel In phyilcal properties
    So. 2-GT:  Correaponda in general to ASm D396, Grade Ho. 2  fuel  and AStK D975 Grade  Ho.  2-D die»»l
    fyel In ph/»lc*l propffrtl**.
    Ho. 3-GT and
    Ho. 1.-GT:  Viicoilty range bracketi ASW D396, Grades Ho. 1*. Ho.  5  (light).  Ho.  5  (twavy) and Ho.  6 and ASTW D975,
               Grade No. I*-D, dltiel fuel, vblch »*y bt aupplled provided utali coeipoiltlon  r«quire»*rit»
               are »et.

(bj Vi»co»ity value* in parenthe«e» are for Jj)/orni»tion only and are  not limiting.

(c) Recognlclng the ntceaiity of additional requirfBenta for certain  typei of gai turbinea, the  folloving  nay  be
    apeclfled for Ho. 1-CT fuel:
         lAwlnoMter nunber, nin • 1*0
         Sulfur* percent by weight, ttax * 1.0
         TbenteJ at«bility te»t for 5 bour» at 250 f (121 C) preheeter  t«p«r»tur«,  350 7 (177 C) filter t*«p*rature
           and at a flov rate of 6 Ib (2.7 kg) per hour:
              Filter preiaure drop, rnej » 12 In.of Hg  (30 cm of  Hg)
              Prcheater deposit code, aiax • 2.
U)  Lover or higher pour polntf may be apecified vhenever required by condition* for itor»«* of u«*.   l-heA a pour
    point le»t than 10 F (-18 c) la specified for Grade Ho.  2-GT, the minim* vlicoelty a&all be 1,5  c« (32.0 »«c.
    Saybolt Universal) and the mininum 90 percent a point shall be waived,

(«)  For ga* turbine* operating belov 1500 F (61*0 C) Maximum gai temperature, the limitation* on Tan*dluai, «o*U«i
    plu* potaail.ua, and calcium may be waived, provided that a ailtcon-ba»e additive, or equivalent,  ii ewplQjreil.
    The special i-equir«B«r.ti covering the addition of and the typ« of additive ahtll be apecified only by mutual
    agr«en«nt between purchase and seller.

(f)  Vtien vater vashing facilities are available at the point of use, th«*« requirement* ma/ b* waived by »«tu*l
    agreement between the purchaser and leller.

(g)  Special requirement* covering the addition of and the type of Mgjiftaiun-baie additive,or equivalent, to be
    used shall be specified only by mutual agreement between purchuer and aeller.

-------
                      TABLK XXXI
REPRESEMTATIYE COAL GASIFICATIOI PROCESSES SURVEYED
Process Home
Vellman - Galusha

Power Gas

Bamag - Winkler
Ruhrgas Vortex

Lurgi Dry Ash

Koppers Totzek

BiW - DuPont

Hummel Single Shaft

Texaco Partial Oxidation

USBM Pressure Gaslfier

BCR 2-Stage Gaslfier
1
IGT Pressure Gasifier

Kellogg Molten Salt
Consol C02 Acceptor
Description
Autothemal, fixed bed,
counter-current flow
Autothermal, fixed bed,
counter-current flow
Autothermal , fluidized bed
Autothermal, cocurrent.
up flow, slagging
Autothermal, fixed bed.
counter-current flow
Autothermal, cocurrent,
tangential flow, slagging
Autothermal, cocurrent,
up flow, slagging
Autothermal, cocurrent.
up flow, slagging
Autothermal, cocurrent,
down flov, slagging
Autothemal , cocurrent.
down flov, slagging
Autothermal , cocurrent ,
up flow, slagging
External Heating, cocurrent,
dovn flow, slagging
External Heating, salt bath
Fluidized bed, cocurrent
flow
Status
Commercial

Commercial

Commercial
Cosnercial

Comercial

Comercial

Comercial

Commercial

Pilot

Pilot

Pilot
Pilot

Pilot
Pilot
Press.
Atm.
1

1

1
1

20

1

1

1

lU

ItO

68
68

28
1
Gasifier Outlet
Temp., F
1250

1250

16OO
1900

950

2000

2200

1800

2200

17UO

1700
1500

1800
1600
Btu of Gas (HHV)
Btu of Coal (HHV)
0.82

0.85

0.66
0.66

0.82

0.67

0.68

0-79

0.72

0.61.

0.7li
0.83

0.82
0.60
Oniilble H*«t. of Gas
Dtu of Coal (HHV)
0.13

0.13

0.10
0.27

0.09

0.19

0.20

0.15

0.19

0.11

0.07
—

0.08
0.03
Gasification Bate
lb/hr-ft2 Ib/hr-ft3
U2.« —

63 -

170 IV
— 1.3

300 —

16

2*

21

— 57

- Ii30

- 5«
— ko

— to
— - —

-------
                               FROM REPS. 2,4


             ESTIMATED NEW ADDITIONS BY TYPE OVER 1969-78 PERIOD

                    FOSSIL STEAM                 58.0%

                    NUCLEAR                    28.2

                    GAS TURBINE                  5.3

                    CONVENTIONAL HYDRO           11

                    PUMPED STORAGE HYDRO         5.4


                    TOTAL                       100%



             (a) NEW GENERATING ADDITIONS AND YEAR-END CAPACITY IN SERVICE
   40

*
*:
*/>
z
2  30
_i
_j

2
 I
   20
a
a
0  10
~* 600
 Wl

 O

— I
    400
  I

 IU
       Of.
       Ul
         200
NEW GENERATING ADDITION
                       1
                                         ^GENERATING CAPACITY
                                           IN SERVICE
                                                              I
             56        60        64        68        72         76


             (b) KAtlO OF YEAR-END CAPACITY TO SUMMER PEAK LOAD
         1.36
     °   U2
         U8
     i
         1-24
         1.20
                                                      \
                                                                 \
             56
                 60
            64
  68

YEAR
72
76
                                                                  80
80
   FIG. 1. YEARLY ADDITIONS TO GENERATING CAEACITY

           AND YEAR-END  MARGINS IN ELECTRIC  UTILITY  INDUSTRY
                                   161

-------
                       D FPC DATA (1970)
                       L FPC DATA (1970) (NUCLEAR PLANTS)
                       A NEMA DATA (1970)
                       0 AEC REPORT TO PRESIDENT (1967)
                         (NUCLEAR PLANTS)
                       O EEI 46th SURVEY (1969)
     1960
1965
1995
2000
FIG.  2.  PREDICTED  GROWTH OF ELECTRIC  UTILITY GENERATION  CAPACITY
                                   162

-------
                                  SOUTH CENTRAL
                                       (8.8%)
          FIGURES IN ( ) INDICATE YEARLY GROWTH RATE AVERAGED OVER 20-YR PERIOD
   1200
   1000
    800
u.
o
«/>  600

O
CO
    400
    200
                 SOUTH CENTRAL-


               NORTH EAST-
                                                                  /SOUTHEAST
                                                                   WEST
                                                                    EAST CENTRAL
                                                                   WEST CENTRAL
                                           I
      1965
                  1970
1975
 1980

YEAR
1985
1990
1995
FIG.  3.  REGIONAL FORECAST OF ELECTRICAL GENERATION  IN THERMAL PLANTS
                                       163

-------
                                            POTENTIAL GAS COMMITTEE
                                               PUBLICATION AREAS
             ESTIMATED PROVEN MID POTENTIAL SUPPLY OF NATURAL GAS III
                     UNITED STATES AS OF DECEMBER 31, 1968

                    Trillion Cubic Feet @ 1^.73 psia and 60°F
                                           Potential
    Area
    A+B
     C
    D+F
E+G

 H
 I
 J
       \0nshore
      " (Offshore
Total
Proven
    8
    1
   18
  15U
    9
   15
   TO
    5
  _1
  287
Probable
1*1*
2
18
80
33
lU
12
29
22
6
Possible
20
1
32
55
90
26
3
75
18
15
Speculative
57
2
70
9
19
21
13
21
392
28
                                                                     Total
                            260
335
632
                  FIG. 4. LOCATION OF NATURAL  GAS RESEflVES
                                       161*

-------
                FROM REF. 1
                                  LEGEND
                              Lignite
                              Subbituminous coal
                              Medium-and high-volatile
                                  bituminous coal
                              Low-volatile bituminous coal

                              Anthracite and semianthracite coal
FIG. 5. COAL  FIELDS OF THE  UNITED  STATES

-------
                                                               FROMREF.  34
WHEN CONDENSER REQUIREMENT EXCEEDS WITHDRAWAL, FRESH-WATER SUPPLIES FOR ONCE-THROUGH COOLING ARE INADEQUATE





                        FIG. 6. PROJECTIONS OF REGIONAL FRESH WATER SUPPLIES FOR ONCE-THROUGH  CONDENSER COOLING

-------
       OONORA
                FROM REF. 37

NEW EAGLE    EL RAM A    CLAIRTON
                                                            McKEESPORT
                         10
             15          20          25
             DISTANCE (MILES DOWN RIVER)
30
                                                                               PITTSBURGH
35
FIG. 7. TYPICAL TEMPERATURE VARIATIONS ALONG  MONONGAHELA RIVER  DUE TO HEAT REJECTION
       FROM VARIOUS SOURCES

-------
                                              FROM REF. 42
   O
   UJ
-<
»-
Z
>-
H
U
o.
u
_J
    Z
    UJ
    u
    a:
    UJ
    a.
120 i	

110

100

 90

 80

 70

 60

 50

 40

 30

 20

 10

  0
            1968 FOSSIL
            AVG. SIZE - 271 MW
            TOTAL - 72 UNITS
               1968 NUCLEAR AVG
               SIZE - 444 MW
               TOTAL - 5 UNITS
                                J-
                                                     1971 NUCLEAR AVG.
                                                     SIZE-762 MW
                                                     TOTAL - 72 UNITS
                                                  1971 FOSSIL AVG.
                                                  SIZE - 494 MW
                                                  TOTAL - 29 UNITS
                                                                               I
I
I
             100  200  300  400  500  600  700 800   900  1000 1100 12001300 1400 1500 1600 1700 1800  1900
                                               UNIT SIZE, MW
FIG. 8. DISTRIBUTION OF UNIT  SIZE  FOR  1968-1971 NUCLEAR  AND FOSSIL STEAM  INSTALLATIONS

-------
                                                NOMINAL 500-MW CAPACITY
                                                        FROM REF. 45
ON
MD
200 FT
                                                         •340 FT-
                             FIG. 9. ELEVATION VIEW OF TYPICAL STEAM POWER STATION

-------
       INCLUDES COST OF LAND. ESCALATION, INTEREST DURING CONSTRUCTION, ETC.

                   ALL PLANTS USE ONCE-THROUGH CONDENSERS


                      	          REF. 27

                      	          REF. 52

                      	          REF. 51
   300
   250
   200
o
u

o
UJ
H-
tx)



U

ul

U
IU
a.
   150
   100
    50
                   \
                    \
                        %
                                                   NUCLEAR PLANTS
   OIL-FIRED BASE-LOAD PLANTS
^OIL-FIRED CYCLER PLANTS
                                                    I
                    500
          1000            1500


         UNIT RATING - MW
2000
      FIG.  10.  TYPICAL INSTALLED  COSTS  OF  STEAM POWER  PLANTS
                                    170

-------
 (a) ONCE- THROUGH COOLING SYSTEM
       STEAM
                               GENERATOR

        TURBINE
STEAM SURFACE
 CONDENSER
 COMPENSATE
  RETURN TO
 POWER CYCLE
                HOT-WELL
                   PUMP
                            HEATED WATER.
                                 85 F
 CIRCULATING PUMP

SUPPLY, 70F
 (b) CLOSED-CIRCUIT WET COOLING TOWER SYSTEM
       STEAM
        TURBINE
STEAM SURFACE
  CONDENSER
CONDENSATE
 RETURN TO
POWER CYCLE
                                 COOLING AIR
                                 COOLED WATER
 (c) OPEN-CIRCUIT WET TOWER SYSTEM
       STEAM
       TURBINE
STEAM SURFACE
 CONDENSER
 CON DEN SATE
  RETURN TO
 POWER CYCLE
                                                                WET-COOLING TOWER
                                                                    COOLING AIR
                                                               MAKE UP PUMP
                                                                    SUPPLY. 70 F
                                                                  WET-COOLING TOWER
                                                                     COOLING AIR
                                                       RETURN TO
                                                       RIVER AT 75 F I SUPPLY 70 F
  FIG. 11. SCHEMATIC  DIAGRAMS  OF ALTERNATIVE CONDENSER COOLING METHODS
                                        171

-------
 (a) CROSS-FLOW MECHANICAL-DRAFT TOWER
                     AIR
                   OUTLET
   WATER INLET
        A
FAN
 WATERJNLET
AIR INLETJWffFILL
        IR INLET
       WATER OUTLET
 (b) HYPERBOLIC NATURAL-DRAFT TOWER
  HOT WATER"!
    INLET

AIR INLET
       WATER OUTLET
 FIG.  12. TYPES OF WET COOLING TOWERS
                   172

-------
   LARGE EVAPORATIVE COOLING TOWER INSTALLATIONS




   COOLING WATER SUFFICIENT THROUGH 1980




   COOLING WATER SUFFICIENT THROUGH 1972




   EXTENSIVE COOLING TOWER INSTALLATIONS ALREADY EXIST
FIG. 13.  GEOGRAPHICAL AREAS OF COOLING WATER  SUFFICIENCY

-------
                                    INDIRECT SYSTEM
 STEAM
                            GENERATOR
     TURilNE
CONDENSATE
 RETURN TO
POWER CYCLE
            DIRECT CONTACT
              CONDENSER
                                                         t  t  t
                                                        HEATED AIR
                                         VARIABLE -
                                         PITCH  FAN-
                  COOLING AIR
          COOLED WATER
HEATED WATER
        COMPENSATE  CIRCULATING
           PUMP         PUMP
                                                                           — HEAT
                                                                            EXCHANGER
                                                                             ELEMENTS
COOLING AIR
                                                                               GAS
                                                                            PRESURIZER
                                         CONDENSATE
                                         STORAGE TANK
                                                        REFILL PUMP
         FIG.  14.  TYPICAL  MECHANICAL-DRAFT DRY COOLING TOWER  SYSTEM
                                         17U

-------
                                  DATA FROM REF. 64

     BASED ON ESTIMATE OF MANUFACTURER FOR CONVENTIONAL TURBINE MODIFIED FOR OPERATION
                               AT HIGH EXHAUST PRESSURE
                                                            LOAD CAPABILITY
FUEL CONSUMPTION
            HIGH BACK PRESSURE
            DESIGN TURBINE
        NUCLEAR-FUELED
             PLANTS
             - -t—4
                                                    :NUCLEAR-FUELED
                                                         PLANTS
                                               ONCE-THROUGH
ONCE-THROUGH

     100  120
140   150
               160
                                   170
100  120
140
                150
160
                    170
         CONDENSING TEMPERATURE -° F
                                          CONDENSING TEMPERATURE -° F
                 _L
           _L
0246      8     10    12
    TURBINE EXHAUST PRESSURE - IN. HgABS
                                                    J_
                              J_
                                        246      8     10    12
                                      TURBINE EXHAUST PRESSURE -  IN. HgABS
                                    14
           FIG  15  ESTIMATED EFFECT OF CONDENSER BACK  PRESSURE ON
                   STEAM PLANT PERFORMANCE
                                       175

-------
o:
 I
IU
oc
K
U4
                     a:
                      I
                     u
                     a.
                     UJ
                     i-
                                        HEAT ADDITION
      ENTROPY,*

  (b) REGENERATIVE CYCLE
ADIABATIC
COMPRESSION
                             ENTROPY,*

                           (a) SIMPLE CYCLE
                                               ISENTROPIC EXPANSION
                                                 COOLING
                              ENTROPY, s

                       (c) INTERCOOLED CYCLE
       ENTROPY, s

  (d) REHEAT CYCLE
                            ENTROPY • s


                      (e) COMPOUND CYCLE
     FIG.  16. DIAGRAMS FOR SELECTED  GAS TURBINE CYCLES
                              176

-------
                                   COMBUSTOR
                        COMPRESSOR
                           POWER
                          TURBINE
                                     COMPRESSOR
                                      TURBINE
                            INTAKE           EXHAUST

                             (o) SIMPLE CYCLE- TWO SHAFT
                                                      POWER
                                                     COUPLING
   EXHAUST
            REGENERATOR
  COMPRESSOR
      INTAKE
                              TURBIN
                                POWER
                               COUPLING
   (b) REGENERATIVE CYCLE - SINGLE SHAFT
                                                     INTERCOOLER
                                                                         POWER
                                                                       COUPLING
                                                        COMBUSTOR
                        (c) INTERCOOLED CYCLE-SINGLE SHAFT
COMPRESSOR
         COMBUSTOR
         o
    INTAKE
     EXHAUST

 TURBINE
                                                                  REHEAT
                                                                 COMBUSTOR
 
-------
                         AMBIENT TEMPERATURE - 60 F

                      TURBINE INLET TEMPERATURE - 1500 F

                         CYCLE PRESSURE LOSSES - ZERO

                         COMPRESSOR EFFICIENCY-85%

                          TURBINE EFFICIENCY-85%
                     6?
                      I
                        80
                      z 60
                      UJ
                      y
                      u. 40
                      UJ


                      2
                      UJ
                       20
                      P  "0    5   10   15   20
                       COMPRESSOR PRESSURE RATIO

                           (a) SIMPLE CYCLE
  REGENERATOR EFFECTIVENESS = 70%
  fe? 40
   I         REGENERATIVE
      05    10   15    20
    COMPRESSOR PRESSURE RATIO

    (b) REGENERATIVE CYCLE
                                       6?
                                       I

                                       U
                                       Z
                                       UJ
                                       u
                                        01
                                        en
                                                       INTERCOOLED

                                                          SIMPLE^
        40


        30


        20


        10


        0
          0    5    10   15   20
        COMPRESSOR PRESSURE RATIO

          (c)  INTERCOOLED CYCLE
6?
 I  40
 U
 01
 U
 UJ
   30
    20
    10
     0    5    10   15  20
  COMPRESSOR PRESSURE RATIO

       (d)  REHEAT CYCLE
6?
 I
>-
U
Z
UJ
u
                                  u.
                                  UJ
                                   cc
                                   UJ
                                      40
30
                                    20
                                     10
   COMPOUND WITH
   REGENERATION.
NTERCOOLING& REHEAT
                                                               SIMPLE
                                        0    5    10    15   20   25   30
                                          COMPRESSOR PRESSURE RATIO

                                             («) COMPOUND CYCLE
FIG.  18.  THEORETICAL PERFORMANCE  FOR MODIFIED GAS TURBINE  CYCLES
                                  178

-------
   40
   35
   30
   25
H
<
U
    20
    15
    10
                                           TWIN-SPOOL

                                           COMPRESSOR
                                                     SINGLE-SPOOL COMPRESSOR
                                                      WITH VARIABLE STATORS
        SINGLE-SPOOL
        COMPRESSOR
                     N.G.T.E. 109
                    COMPRESSOR
                                               O ACTUAL OR PROPOSED
                                                   POWERPLANTS
         ~*/7,
     1940
1950
                                   1960
                              1970
                                                                 1980
1990
                                          YEAR
          FIG.  19.  PROGRESSION  OF AIRCRAFT  COMPRESSOR  TECHNOLOGY
                                       179

-------
              ALL SYMBOLS REFLECT ACTUAL OR PROPOSED ENGINE DESIGN
UJ
u
u.
u.
UJ
Ul
O
^-ESTIMATED LEVEL ACHIEVED
  WITH 3 - YEAR DEVELOPMENT
  REF.
               1600
                            STAGE PRESSURE RATIO - P2 /P,
            UJ
                80
                 1940      1950      1960       1970

                                       YEAR
   1980
1990
        FIG. 20. ADVANCES IN COMPRESSOR  PERFORMANCE PARAMETERS
                                      180

-------
o. AIRCRAFT DESIGN
 INLET GUIDE-
   VANES
                  -BLADE
STATOR
                                                        SEALS
                                                       __ff
                          -DISCS
b. DISK DRUM
             INLET GUIDE
               VANES
                                    STATOR
  AIRFLOW
                          -DISCS
:. DRUM ROTOR
   AIRFLOW
                                                    DRUM
                        '—INLET GUIDE VANES
      FIG.  21.  COMPRESSOR CONSTRUCTION TECHNIQUES
                             181

-------
             3600
oo
                                                 MILITARY AIRCRAFT





                                                            A
                                                                                             COMMERCIAL TRANSPORTS
                                                                                             INDUSTRIAL APPLICATIONS
              1200
                1950
1958
1966
1974
1982
1990
                                                                YEAR
                                FIG. 22,  ESTIMATED TURBINE INLET TEMPERATURE  PROGRESSION

-------
      *   BLADE MATERIALS

      A   VANE MATERIALS
    UJ
    DC
    3


       "
    UJ Si
    U Q.
    O O
    Of o
    a. —
    IU
    2s
    Ul
    Q.
    -s.
    ui
             3000
             2600
2200
1800
             1400
                     CHROMIUM-

                  -AND COLUMBIUM
                    3ASE ALLOYS
.CAST
                  B- 1900 -

                  INC0713-
                   WROUGHT
                                    SM 200

                                    MARM 509
                 CM,
          DIM ET 700
        -UDIMET 500
            I
        VACUUM-MELTED WASPALOY
                                      CARBON GRAPHITE
                                         COMPOSITES
                                          WLWWL
                                    SILICON-NITRIDE
                                     COMPOSITES
                1950
              1960
           1970
1980
1990
2000
                                        YEAR
(b)   IN-100 STRESS PROPERTIES
                                   TEST DATA

                                   EXTRAPOLATED
 UJ
 oe.
 Of

 O
 H
 oo
 to
 Ul
 Oi
     80,000
     60,000
     40,000
     20,000
U
n
U
ft
U




\
\

>v
\



\
\
\
V

\
•^

\
N,

«»
\
• 100,000


s.



\_ 1000-HR LIFE
XT
10,000 ^


V
^^H






^-^_
130o 1400 1500 1600 1700 1800 19
                                 METAL TEMP., F- -

          FIG.  23. ADVANCES IN TURBINE BLADE MATERIALS
                                    183

-------
   tttl
              K0Tli BASED ON LARSON-MILLER PARAMETER EXTRAPOLATIONS


                                    PRESENT MATERIALS
                                          1970 DECADE
                                          1980 DECADE
   50
                                                   CHROMIUM-AND COLUMBIUM-BASE

                                                   ALLOYS
 I
te.
z

8
o


I  201
a.
1U
IU
&.
U

6?

IU
U
=}


i  101
0.
IU
a:
        COBALT-BASE ALLOYS
                 NICKEL-BASE ALLOYS
                1
                         1
    1200
1400
2200
                   1600         1800        2000
                      METAL TEMPERATURE - F

FIG.  24. SUMMARY  OF PROJECTED CREEP STRENGTH PROPERTIES

         FOR ADVANCED TURBINE BLADE  MATERIALS
2400
                                    18 U

-------
  (<0 DISC COOLING CONFIGURATION
COOLING AIR
                                                 COMBUSTION
                                                    GASES
                                             DISCS
  (b)  BLADE AND VANE COOLING
                    -BLEED FROM COMPRESSOR
                          FOR COOLING
                               BLADE
-SEALS
                                                  COMBUSTION
                                                [>    GASES
                                                STATOR
               FIG. 25. TURBINE COOLING SCHEMES
                              185

-------

       FILM-COOLED
IMPINGEMENT-CONVECTION    TRANSPIRATION COOLED
FIG. 26.    ADVANCED  BLADE COOLING CONFIGURATIONS FOR AIRCRAFT  POWERPLANTS
                                     186

-------
                                CONVECTION
                                       IMPINGEMENT
                                                                                 FILM
EARLY DESIGN'
    PRESENT
CONFIGURATIONS
 TEMPERATURE
 CAPABILITIES
2200 ° F                                    2400 ° F

          FIG. 27.  TURBINE BLADE COOLING BLADE IMPROVEMENTS
2600 °F

-------
                                                  FUEL-METHANE (HHV - 1000 BTU/FJ3)

                                                      AMBIENT - 80 F AND 1000 FT

                                    TURBINE COOLING CONFIGURATION:ADVANCED IMPINGEMENT-CONVECTION

                             ELECTRIC GENERATOR EFFICIENCY AND AUXILIARY POWER REQUIREMENTS NOT INCLUDED
CD
CD
                35
            z
            UJ
            U
            DC
            UJ
            Q.
            u
            z
            UJ
            u
            UJ
            UJ
            X
            UJ

            1  25
            o:
             */>
             <
             O
                30
                20
                  40
                                     'THEORETICAL VALUE IF MATERIAL COULD WITHSTAND TEMPERATURES
                                                               NO COMPRESSOR BLEED-
                                                              FOR TURBINE COOLING*
                    COMPRESSOR
                   PRESSURE RATIO
PRESENT-DAY GAS

   TURBINES
                              1800
                                                16
                                                                                       2200
                        COMPRESSOR
                         BLEED AIR
                       PRECOOLED TO
                           125 F
                           TURBINE INLET GAS TEMPERATURE - F
                                               COMPRESSOR BLEED AIR UNCOOLED
                                  I
                              I
I
I
    60
80        100       120        140       160        180      200

      SHAFT HORSEPOWER PER UNIT AIRFLOW - SHP/LB/SEC
                             220
                  240
                    FIG. 28. ESTIMATED  1970-DECADE SIMPLE-CYCLE  BASE-LOAD GAS  TURBINE PERFORMANCE

-------
                   COMPRESSOR BLEED AIR PRECOOLED TO 200 F
                   FUEL-METHANE (HHV = 1000 BTU/FT 3)
                   AMBIENT-80 F AND 1000 FT
                   TURBINE COOLING CONFIGURATION! ADVANCED
                         IMPINGEMENT-CONVECTION
                   ELECTRIC GENERATOR EFFICIENCY AND
                    AUXILIARY POWER REQUIREMENTS NOT INCLUDED
                                                            TURBINE INLET
                                                           GAS TEMPERATURE - F
                              19        21        23
                           COMPRESSOR PRESSURE RATIO
                                                        27
                                                       29
34
33
32
31
30
                                                            TURBINE INLET
                                                           GAS TEMPERATURE - F
  13
15
17
                                                                   27
                              19        21        23
                           COMPRESSOR PRESSURE RATIO
FIG  29   ESTIMATED  1970-DECADE SIMPLE-CYCLE BASE-LOAD  GAS TURBINE
             PERFORMANCE  WITH  SUPPLEMENTARY  COOLING
                                                                 29
                                   189

-------
VO
o
            z
            LU
            u
            tx.
            LU
            O.
             I
            >-
            U
            z
            UJ

            u
u.
UJ
            DC
            UJ
            Ul
            CO
            Of
            =9
            O
                                                  FUEL-METHANE (HHV = 1000 BTU/FT3)

                                                      AMBIENT 80 F AND 1000 FT

                                                   TURBINE COOLING CONFIGURATION:
                                                 ADVANCED IMPINGEMENT-CONVECTION


                                            ELECTRIC GENERATOR EFFICIENCY AND AUXILIARY

                                                  POWER REQUIREMENTS NOT INCLUDED

                                           INLET AND EXHAUST GAS PRESSURE LOSSES INCLUDED


                                              COMPRESSOR BLEED AIR PRECOOLED TO 200 F
                                            COMPRESSOR PRESSURE RATIO = 36
                                                                                            LATE 1980'S TECHNOLOGY
                       TURBINE INLET GAS

                        TEMPERATURE-F
                                         2400    2500
                                                        2600     2700
                                              EARLY 1980'S TECHNOLOGY
                                                280        300        320       340       360

                                              SHAFT HORSEPOWER PER UNIT AIRFLOW-SHP/LB/SEC
                                                                                              400
420
                  FIG. 30.  ESTIMATED 1980  - DECADE  SIMPLE-CYCLE  BASE-LOAD GAS  TURBINE PERFORMANCE

-------
(°) SIMPLE AND REGENERATIVE CYCLE
               COWPRESSOn
             COMPRESSOR BLEED
              AIR FOR TURBINE
                COOLING


1 	 — -\
1 PRECOOLER 1
y) (OPTIONAL) 1
1 |
i 	 (VWAi 	 '
TO AMBIENT HEAT
REJECTION SYSTEM

— cc
T
"•^


IU

                                                       rUHSINE
                          .      (OPT IUNAL)    '
                    TO EXHAUST STACK
 (b) COMPOUND CYCLE
                               ELECTRIC
                               GENERATOR
                   AIR
          AUPIFfNT AIR
          FORCOOL NG
                           LOW-PRESSURE
                            COMPRESSOR
                               IMTERCOOLER
POWER TURBINE
                                      MICH-PRESSURE   „ L°W "
                                        ™«"«    "SnV
                     HIGH-PRESSURE
                     COMPRESSOR
                                            REHEATER
                         FIG. 31.  GAS  TURBINE FLOW DIAGRAMS
                                             191

-------
              FUEL-METHANE (HHV = 1000BTU/FT3)

              AMBIENT-80 F AND 1000 FT

              TURBINE COOLING CONFIGURATION: ADVANCED IMPINGEMENT-CONVECTION

                                          COMPRESSOR BLEED AIR UNCOOLED

              ELECTRIC GENERATOR EFFICIENCY AND AUXILIARY POWER REQUIREMENTS

                                    NOT INCLUDED

              REGENERATOR TOTAL PRESSURE DROP           4.0 %

                                              	8.0

              TURBINE INLET GAS TEMPERATURE= 2000 F
                                                                       REGENERATOR

                                                                     EFFECTIVENESS, %

                                                                           70	
                                                                           90 \	
                                    8        10        12

                                 COMPRESSOR PRESSURE RATIO
tu
u
ot
IU
0.
 I
>•
u
z
UJ
u
u.
u.
UJ
Of
Ul
UJ
z
CD
ee.
o
     31
      29                             8        10         12

                                COMPRESSOR PRESSURE RATIO

      FIG. 32.  ESTIMATED 1970-DECADE  REGENERATIVE-CYCLE  BASE-LOAD  GAS

                               TURBINE PERFORMANCE
                                        192

-------
                   EXHAUST GAS
  AIR FROM

  COMPRESSOR-
1400
1200
1000
u.
I
IU
«
3
>-
<
K
UI
CL
                                    NO FLOW
                                                 TEMPERATURE
              HOT GAS IN
                                            AIR TO
                                          COMBUSTION
                                                            HEAT EXCHANGER LENGTH


          FUEL-METHANE (HHV = 1000 BTU/FT3)


          AMBIENT-80 F AND 1000 FT


          TURBINE COOLING CONFIGURATION! ADVANCED IMPINGEMENT-CONVECTION


                                       COMPRESSOR BLEED AIR UNCOOLED


          REGENERATOR PRESSURE LOSS  =4.0%


          REGENERATOR AIRSIDE EFFECTIVENESS =80%

          TURBINE INLET GAS TEMPERATURE  =2000F
      HOT-SIDE INLET GAS
        TEMPERATURE
      COLD-SIDE EXIT GAS
        TEMPERATURE
      HOT-SIDE
        TEMPERATURE
      COLD-SIDE INLET GAS
         TEMPERATURE
                                 6          8         10

                             COMPRESSOR PRESSURE RATIO
     FIG. 33. REGENERATOR GAS TEMPERATURES FOR 1970-DECADE DISIGNS
                                      193

-------
                                   REGENERATOR INLET
                                   GAS TEMPERATURE
                                                                       HEATED AIR TO COMBUSTOR
                                                REGENERATOR EXIT
                                                GAS TEMPERATURE
                                                        INLET AIR FROM COMPRESSOR
MD
                          UJ
                          DC
                          Q£
                          UJ
                          Q.

                          UJ
                          O
                          H
                          UJ
                           at
                           o
                           o
                           UJ
                           ex
                             1800
1600
1400
                                                                     REGENERATOR EFFECTIVENESS = 80%
                                          FURBINE INLET GAS
                                          TEMPERATURE - F
                                UII1UU
                              INCOLOY 800
                                 STAINLESS STEEL
                             1200
                              1000
                                                    8        10        12         14
                                                      COMPRESSOR PRESSURE RATIO
                    FIG. 34.ESTIMATES OF  MATERIAL TYPES  REQUIRED IN REGENERATIVE-CYCLE ENGINE

-------
     25
                                            FUEI	METHANE (HHV=]OOOBTU/FT3 )
                                                AMBIENT 80 F AND  1000FT
                             TURBINE COOLING CONFIGURATION: ADVANCED IMPINGEMENT-CONVECTION
                      ELECTRIC GENERATOR EFFICIENCY AND AUXILIARY POWER REQUIREMENTS NOT  INCLUDED
                                       COMPRESSOR BLEED AIR PRE COOLED T0200F
                                                         TURBINE
                                                           2400 F
                                                         TEMPERATURE
                                                                                                EFFECTIVENESS
                                                                          COMPRESSOR PRESSURE RATIO
                                                                 EARLY 1980'S TECHNOLOGY
                  1970-DECADE TECHNOLOGY
       120
140
160
180       200        220       240       260       280
     SHAFT HORSEPOWER PER UNIT AIRFLOW-SHP/LB/SEC
300
FIG. 35.  ESTIMATED 1970 - AND EARLY 1980 - DECADE  REGENERATIVE-CYCLE  BASE-LOAD  GAS TURBINE  PERFORMANCE

-------
            INLET AND EXHAUST GAS PRESSURE LOSSES INCLUDED


                      FUEL. METHANE (HHV = 1000 BTU/FT3)

                          AMBIENT: 80 F AND 1000 FT

        TURBINE COOLING CONFIGURATION.- ADVANCED IMPINGEMENT-CONVECTION

  ELECTRIC GENERATOR EFFICIENCY AND AUXILIARY POWER REQUIREMENTS NOT INCLUDED

                   COMPRESSOR BLEED AIR PRECOOLED TO 200 F

                    REGENERATOR TOTAL PRESSURE DROP = 4%
   50
Z
iu
u
tt
Ul
0.
 I
>-
u

Ul

u
u.
u.
Ul
ee
LU
Ul
Z

5
oc
=3
45
   40
35
   30
     300
                        -AIRSIDE EFFECTIVENESS-%



                 -COMPRESSOR PRESSURE RATIO
           310        320       330       340       350       360


             SHAFT HORSEPOWER PER UNIT AIRFLOW-SHP/LB/SEC
370
  FIG. 36. ESTIMATED  LATE 1980'S - DECADE  REGENERATIVE-CYCLE

          GAS  TURBINE PERFORMANCE
                                    196

-------
                    FUEL-METHANE (HHV = 1000 BTU/FT )
                         AMBIENT SOF AND 1000 FT
      TURBINE COOLING CONFIGURATION: ADVANCED IMPINGEMENT-CONVECTION
ELECTRIC GENERATOR EFFICIENCY AND AUXILIARY POWER REQUIREMENTS NOT INCLUDED
 U           INLET AND EXHAUST GAS PRESSURE LOSSES INCLUDED
                 COMPRESSOR BLEED AIR PRECOOLED TO 200F
 CO

 0.
 I
o
z
LU
O.
Q£
UJ

O
0.
UJ
               EARLY-1980'S TECHNOLOGY
                           .TOTAL CYCLE PRESSURE RATIO, (CPR) = 100:1
                            LOW- PRESSURE COMPRESSOR
                            PRESSURE RATIO. (LPR) = 5:1
               EARLY-1980'S TECHNOLOGY
     1800
                  2000
2200
                                         2400
                                                      2600
2800
                      TURBINE INLET TEMPERATURE - F
  FIG.  37.  ESTIMATED 1980 - DECADE  COMPOUND-CYCLE BASE-LOAD
           GAS TURBINE  PERFORMANCE
                                 197

-------
                    FUEL-METHANE (HHV = 1000 BTU/FTJ)

                        AMBIENT 80F AND 1000 FT

       TURBINE COOLING CONFIGURATION: ADVANCED IMPINGEMENT-CONVECTION

  ELECTRIC GENERATOR EFFICIENCY AND AUXILIARY POWER REQUIREMENTS NOT INCLUDED

               INLET AND EXHAUST GAS PRESSURE LOSSES INCLUDED
(o)
                  COMPRESSOR BLEED AIR PRECOOLED TO 200F
      500
                       TURBINE INLET TEMPERATURE. F
  U
  IU
IU
0-
      400
£g   300
      200
                                        TOTAL CYCLE PRESSURE RATIO
                                    	TOTAL CYCLE PRESSURE RATIO = 50
                                              I            I
       44
 IU
 U
 DC
 IU
 0.

  I

 >•
 U

 IU

 U
 u.
 U.
 Ul
 at
 ui
                                   TURBINE INLET TEMPERATURE. F
40
      38
      36
                      	1	1	
                      •TOTAL CYCLE PRESSURE RATIO = 100
                  	L TOTAL CYCLE PRESSURE RATIO = 50
                                  2600
                                  2200
                                  1800
                                                      1800
                     4567

                     LOW-PRESSURE COMPRESSOR PRESSURE RATIO
 FIG. 38.  EFFECT OF  LOW-PRESSURE  COMPRESSOR. PRESSURE RATIO
         ON COMPOUND-CYCLE  GAS  TURBINE  PERFORMANCE
                                  198

-------
 (o)
Of.
Ul

Q.
   u
   UJ
O  m
a.  -i

o:
o
   I
iu
IU
o:
       1.3
       1.1
       0.7
                   TURBINE INLET TEMPERATURE = 2200/2200F



                    TOTAL-CYCLE  PRESSURE RATIO = 75


                    LOW COMPRESSOR PRESSURE RATIO = 5


                  TURBINE COOLING PENALTIES NOT INCLUDED
                           GRANGE OF WORK SPLIT
                             USED IN STUDY
(b)
     1.05
                                           RANGE OF WORK SPLIT

                                           USED IN STUDY
  RATIO OF
          0.2          0.4          0.6          0.8         1.0


WORK EXTRACTED TO TOTAL WORK AVAILABLE IN GAS GENERATOR TURBINE
 FIG.  39. EFFECT  OF  WORK  SPLIT ON  COMPOUND-CYCLE PERFORMANCE
                                    199

-------
                                    SELLING PRICE IN 1970 DOLLARS BASED ON CONSTANT TOTAL MARKET;

                                          INCLUDES DEVELOPMENT. ASSEMBLY. AND TEST COSTS.

                                             GENERAL AND ADMINISTRATIVE EXPENSES. AND

                                                PROFIT IN SELLING PRICE  OF EACH UNIT

                                             COMPRESSOR BLEED AIR PRECOOLED TO 200F

                                                BLADE STRESS LEVELS  AS INDICATED
fV)
8
      IU
      O
      o
      z
Ul
«/»
Ul
z
m
      O
      Ul
      to
      UJ
                          36,000 PS I

                               POWER TURBINE OUTPUT SPEED

                                 3600 RPM
                                                                                                          COMPRESSOR
                                                                                                           ?RESSURE
                                                                                                            RATIO

                                                                                              TURBINE INLET
              PRESENT
               DAY
              ENGINES
                                                                                             -..	TEMPERATURE, F
                                                                                              >^ ^"" •» _    i
                                                      1970-DECADE TECHNOLOGY
                                                                     EARLY-1980'S
                                                                     TECHNOLOGY
               TWO EXHAUST ENDS
                                                          150              200


                                                         UNIT CAPACITY - MW
                            FIG.  40. EFFECT OF GAS TURBINE  UNIT CAPACITY ON SELLING PRICE

-------
                 40
                                                 TWIN-SPOOL  COMPRESSOR  DESIGNS

                                                  APPROXIMATELY 200-MW OUTPUT

                                         SINGLE EXHAUST END ON POWER TURBINE UNLESS NOTED

                                           TURBINE BLADE STRESS  < 35,000 psi UNLESS NOTED

                                                          1970 DOLLARS
(V)
o
H
               I
              LU
              y
              Q£
              a.

              o
              LU
              GO

              LU
CO
Of
=1
              O

              Q
              01
              LU
                 35
      TURBINE INLET TEMPERATURE-F
                         2000
   30
1970-DECADE TECHNOLOGY

 ©  I   '
 w  ' 2000F ENGINE DESIGN

      WITH BLADE STRESSES
l\
                               ( ABOVE 35,000
                  25
                 20
                                                                             'TWO EXHAUST ENDS REQUIRED
                                                                                       3100
 LATE-

 1980'S	

TECHNOLOGY



2900
                                                                                 EARLY- 1980'S TECHNOLOGY
                   10
                    15
                                                     COMPRESSOR PRESSURE RATIO
                            FIG.  41.  EFFECT OF COMPRESSOR PRESSURE RATIO AND TURBINE  INLET

                                        TEMPERATURE  ON GAS   TURBINE SELLING  PRICE

-------
                             EARLY-1980'S TECHNOLOGY
                           APPROXIMATELY 200-MW OUTPUT
                               SINGLE EXHAUST END
                                  1970 DOLLARS
                  TURBINE INLET TEMPERATURE.
                                 4-STAGE POWER TURBINE
                                                      THREE STAGES
                            r-LOW-PRESSURE TURBINE
                            \ -«—TWO STAGES	•-
                                   LOW+HIGH COMPRESSOR
                                      HIGH-PRESSURE TURBINE

                                                        25
10
15              20
   COMPRESSOR PRESSURE RATIO
FIG. 42.  COMPONENT COST  DISTRIBUTION  OF ADVANCED GAS TURBINE ENGINES
                                     202

-------
(o)
                       1970-DECADE TECHNOLOGY

                       OUTPUT CAPACITY = 200 MW

                        OUTPUT SPEED = 1800 RPM
                       TURBINE INLET GAS TEMPERATURE = 2200 F
1
UJ
u
0.
£ 30
i
UJ
UJ
i 20
Z i
UJ
SINGLE SPOOL DESIGNS





..— •
— — — .




«
1 	 ^- 	



TIP

	 • 	 , 	



SPEED = 100
	 0.4

0 FT/SEC
ODp
	 1150
0.4

5 Dp

! 68 10 12 14 1
(b)
                  COMPRESSOR PRESSURE RATIO


              TURBINE INLET GAS TEMPERATURE = 2000 f-
               COMPRESSOR PRESSURE RATIO = 20 TO 1
       20
   I/O
   UJ
   O
   q 10
   UJ
   Q£
   Q.

   O
   U
(c)
        1000       1100      1200      1300       1400

                         COMPRESSOR TIP SPEED - FT/SEC
                                                1500
      30
   UJ
   5 25
   o
   z
   in
       20
          r INDUSTRIAL GAS TURBINE TECHNOLOGY
                             I
                          ADVANCED AIRCRAFT TECHNOLOGY
1000       1100      1200      1300       1400

                COMPRESSOR TIP SPEED - FT/SEC
                                                         1500
1600
1600
       FIG. 43.  EFFECT OF  COMPRESSOR DESIGN  PARAMETERS

                     ON ENGINE SELLING PRICE
                                  203

-------
                             OUTPUT: 200 MW
                              SPEED: 1800
                     TURBINE INLET TEMPERATURE : 2400 F
                      COMPRESSOR PRESSURE RATIO 20:1
     («) HUB-TO-TIP RATIO EFFECT
            30
         I
         IU
         g  »
         0.
         o
         IU
            20
       0.4      0.5       0.6       0.7       0.8
                POWER TURBINE EXIT HUB-TIP RATIO
(b) EXIT VELOCITY EFFECT
                                                            0.9
           400       500       600      700       800       900
                  POWER TURBINE EXIT VELOCITY - FT/SEC

FIG.  44. EFFECT OF POWER  TURBINE  DESIGN PARAMETERS ON  ENGINE
                         PERFORMANCE AND COST
                                 20k

-------
ro
o
VJ1
              I
             LLJ
             y
             a;
             a.

             o
             */>

             ai
             z

             CD
             Di
             O

             O
             01
             LU
                   35
30
                   25
                   20
                                                    APPROXIMATELY 200 MW OUTPUT

                                                       1800 RPM OUTPUT SPEED

                                                           1970 DOLLARS

                                                 COMPRESSOR BLEED PRECOOLED T0200F

                                                  POWER TURBINE INLET HUB/TIP RATIO;


                                                          <^ 0.85   	


                                                          <_ 0.875	


                                            BLADE STRESS LEVELS NOT EXCEEDING 35,000 PSI
                        COOLED NICKEL ALLOY BLADES AND VANES
       NICKEL ALLOY

     BLADES AND VANES
                                         COLUMBIUM ALLOY

                                         BLADES AND VANES
                                                                                       LATE 1980'S

                                                                                      TECHNOLOGY
                                                   ITURBINE INLET
                                                   (TEMPERATURE
                                                        3100 F
                           EARLY- 1980'S TECHNOLOGY

                           TURBINE INLET TEMPERATURE =2600 F
                      10
                  15
20              25'


 COMPRESSOR PRESSURE RATIO
30
35
40
                FIG. 45. EFFECT OF  POWER  TURBINE MATERIALS AND DESIGN PARAMETERS ON  SELLING PRICE

-------
                2600 F TURBINE INLET TEMPERATURE
                 28: 1 COMPRESSOR PRESSURE RATIO
                   BLEED AIR PRECOOLED TO 200 F
                     EARLY 1980'S TECHNOLOGY
                                                                80.000
             SPECIFIC POWER OUTPUT
                                    - VALVE USED IN STUDY TO
                                     CALCULATE PERFORMANCE
       THERMAL
       EFFICIENCY
                                  COATING LIFE
               1800                    1900

               VANE METAL TEMPERATURE - F
FIG.  46.  EFFECT OF VANE  COOLING  REQUIREMENTS ON ENGINE
             PERFORMANCE AND  COATING  LIFE
                             206

-------
                   TURBINE INLET TEMPERATURE, 2000 F
                   PRESSURE RATIO, 8: 1

      0.1               }	 11.1 STRIP FINS PER INCH ON BOTH SIDES
        UARL ESTIMATES (
            USING      (	11=1 PLAIN FINS PER INCH ON HOT SIDE
                       '           19.8 PLAIN FINS PER INCH ON COLD SIDE
  10'
U
 I
UJ
o

UJ
O£
O
U
  103
                     MANUFACTURER'S ESTIMATES
          TOTAL PRESSURE
               DROP*
                4%
                               *CORE PRESSURE DROP ACCOUNTS FOR 75%
                                     OF TOTAL PRESSURE DROP
                                                                      2.0
                                                                      1.0
                                                                         O
                                                                         U
                                                                      0.5 *
                                                                     I
                                                                          O
                                                                          U
                                                                          UJ
                                                                         <
                                                                         _l
                                                                         UJ
                                                                     0.1
                 70                       80
                     AIR SIDE EFFECTIVENESS - %
                                                                   90
   FIG. 47. VARIATION OF  REGENERATOR  SIZE WITH EFFECTIVENESS
                                    207

-------
o
03
                 FUEL COST - SOf/lO6 BTU
                  MANUFACTURERS'COST
        CAPITAL CHARGES    f




                  FUEL COST - 50
-------

                       ouTT
                       /  I
       GAS _

       FLOW




     AIR FLOW
             '
                          8^
CROSS-COUNTERFLOW ARRANGEMENTS

TURBINE INLET TEMPERATURE = 2000 F

ENGINE PRESSURE RATIO = 8 '• 1
                     t)NE PASS
o

 I
a:
O
U
                — LENGTH

                — VOLUME



                    7200
                                                      6800
                                                    t 640°

                                                 U-   I

                                                 I   UJ

                                                 X  Z


                                                 CD  ~"^

                                             300  5  > 6000

                                                 -I  UJ
                                                 *  ex
                                                    u

                                                    <
                                                      5600
                                                      5200
                                                      4800
           TOTAL PRESSURE LOSS,— - %
                                                                   20
                                                                  PERCENT OF TOTAL-p-  ON GAS SIDE
                                                                                                         320
                                                                           I

                                                                           z
                                                                           t-
                                                                           o

                                                                           UJ
                                                                                                             o
                                                                                                       100
           FIG. 49. EFFECT  OF PRESSURE LOSS PARAMETERS  ON  REGENERATOR SIZE CHARACTERISTICS

-------
                           APPROXIMATELY 200- MW OUTPUT

                           80% REGENERATOR EFFECTIVENESS

                        REGENERATOR TOTAL  PRESSURE LOSS = 8%
    8000
   6000
                                                   1970- DECADE TECHNOLOGY
                                                   TURBINE INLET TEMPERATURE
                                                           = 2000 F
  I

 IU
   4000
 o

 IU
 tt
 O
 O
   2000
EARLY- 1980'S TECHNOLOGY
2400 F TURBINE INLET
TEMPERATURE
                                                .TWO-PASS CROSS-COUNTER-
                                                 FLOW DESIGNS
                                                • COUNTERFLOW DESIGNS
                                           10
                                              12
U
FIG.  50.  EFFECT OF COMPRESSOR PRESSURE RATIO  ON  RECUPERATOR  SIZE
                                    210

-------
(o)
Of
o
l/l
o
U
5
4
3
2
1
0
8(












EXCH.





kNGER C



	 	
OST = ($



^^

/SQFT)E
(SEE



^

USE MA
TEXT)


/


TERIAL


X


X COST

x



FACTOR

X















JO 900 1000 1100 1200 1300 14
 (b)
                           MAXIMUM METAL TEMPERATURE - F
                         APPROXIMATELY 200-MW OUTPUT

                        80% RECUPERATOR EFFECTIVENESS

                       RECUPERATOR TOTAL PRESSURE LOSS = 8%
                                      TEMPERATURE

                                      COST
      1400
   ui
   OC.
   Ul
   0.

   Ul
   »-

   _1
   <
   »-
   Ul
   X
   <
      1200
                                              _EARLY 1980'S TECHNOLOGY

                                                TURBINE INLET

                                                TEMPERATURE = 2400 F
           1970-DECADE

           TECHNOLOGY TURBINE

           INLET TEMPERATURE = 2000 F
       1000
                                  8          10

                            COMPRESSOR PRESSURE RATIO
14
     FIG  51 EFFECT  OF TEMPERATURE  AND  COMPRESSOR PRESSURE  RATIO
                            ON  RECUPERATOR COST    -
                                      211

-------
                           NOMINAL 200- MW OUTPUT

                      SINGLE EXHAUST END ON POWER TURBINE

                     RECUPERATOR TOTAL PRESSURE DROP - t*

                        RECUPERATOR EPPECTIVENESS - 80%
*u
5
^
1
Ul
U TA
j; JO
a.
3
j
Ul 1?
jg •"
z
0
HI
O 21
|y *•
>-
S
«*
Ul
91


COMPRESSOR TIP
SPEED -PT/SEC =
1000











— 	




1150



1100














"\_
-^
1200














r*
^^ 	



\

V 1970-DEC





:ADE TECHNI





3LOGY
^/TURBINE INLET
[ TEMPER-

/


^S
^^^
- — •—
kTURE =2200




] EARLY 198
f




B'S
'/TECHNOLOGY TURBINE
I INLET TEM
' = 2400 F
PERATURE
                          8         10        12        U

                           COMPRESSOR PRESSURE RATIO
                                                               16
(b)
    100
     80
Ul
o
0.

u
o
Z
UJ
    60
    40
                               COUNTERFLOW DESIGN
                      1970- DECADE TECHNOLOGY

                      2000 F TURBINF INLET TEMPERATURE'
                                      EARLY-1980'S TECHNOLOGY:
                                        2400 F TURBINE INLET

                                          TEMPERATURE
                                                                   42
                                                                   40
 38
                                                                   34
                                                                   32
                                                                   30
                              8           10

                        COMPRESSOR PRESSURE RATIO
                                                     12
14
      18
                                                                       ui
                                                                       u
                                                                       ec
                                                                       tu
                                                                       a.
                                                                       u

                                                                   36  n
     u.
     ui

     _1


     u
     Ul
     X
    FIG.  52. EFFECT  OF  DESIGN PARAMETERS ON  REGENERATIVE-CYCLE

                   GAS TURBINE ENGINE SELLING   PRICE
                                    212

-------
                                                    EARLY 1980'S TECHNOLOGY

                                               TURBINE INLET GAS TEMPERATURE   2400 F
                                                  COMPRESSOR PRESSURE RATIO  32:1

                                                       AIRFLOW    1176 LB/SEC

                                              SEE TABLE XXFOR REFERENCE CHARACTERISTICS
AIRFLOW-
                                                                         POWER TURBINE
                                                                                               EXHAUST DUCT
                            LOW SPOOL
                            COMPRESSOR
HIGH SPOOL COMPRESSOR
             COMBUSTION
              CHAMBER
                                                                            BEARINGS
                                                                    LOW SPOOL COMPRESSOR
                                                              HIGH SPOOL TURBINE
                            FIG.  53.  CONCEPTUAL DESIGN OF  200 - MW  BASE-LOAD  GAS TURBINE  ENGINE

-------
IV)
           AIR
        AMB =0.9644 ATM*
       'AMB
           = 80 F
      Wa = 1295 LB/SEC
                                                           NOMINAL. 250 MW
                                                  TURBINE INLET TEMPERATURE = 2600 F
                                                  COMPRESSOR PRESSURE RATIO = 28 TO I
                                                      EARLY 1980'S ENGINE DESIGN
                                               COMPRESSOR BLEED AIR PRECOOLED TO 200 F

                                                    SEE TABLE VI FOR REFERENCE
                                                                   P = 5.65
                                                                   T = 1665.5
                                                                   Wg = 1319.2
                                              WB= 121.4
                                                PRECOOLER
                                               260       T = 160
                                                 WH20=239
                                                                                            -•— TO EXHAUST STACK
P = 0.9790
T = 991.3
Wg = 1319.2
                                                                                                               ELECTRIC
                                                                                                              GENERATOR
                                                                   BURNER
                                                                               W f = 24.2
                                        P = 26.5
                                        T = 1017.5
                                        Wa = 1173.6
       P =PRESSURE IN ATMOSPHERES
       T = TOTAL TEMPERATURE IN  F
       W = FLOW RATE IN LB/SEC
       •1 ATM= 14.7 PSIA
                                  FIG.  54, HEAT BALANCE  FOR SIMPLE-CYCLE GAS TURBINE

-------
                                                           ELEVATION
                                                    EARLY 1980'S TECHNOLOGY
                                                                  EXHAUST STACK
                                                                   SILENCER
                                                                      ENCLOSURE
                                                                      (SOUND ATENU ATH3N)
                                         L_
AUX. TRANSF.
  SEAL
OIL COOLER
                                           TURBINE AIR
                                           PRECOOLER
                                                                  TURBINE
                                                                  ENCLOSURE
                                                                  AIR COOLER
  OIL
COOLERS OIL
        RESERVOIR
  SEAL-
OIL UNIT
  PUMP
                                                                              REVOLVING
                                                                              TYPE FILTER
                                                                                                                          LEVEL
                                                                                                                    GROUND LEVEL
                                                  GROUND
                                                  TRANSFORMER
                                          FIG.  55. 1000-MW GAS  TURBINE  POWER PLANT

-------
                 PLAN VIEW
          EARLY  1980'S TECHNOLOGY
                                                           200 FT.
     nnn
        AIR COOLERS
FIG.  56. 100Q-MW GAS TURBINE POWER PLANT
                   216

-------
     AIR
PAMB = 0.9644 ATM*
  AMB ~
      SO P
W0 ^ 1190 LB/SEC
 P- PRESSURE IN ATMOSPHERES
 T =TOTAL TEMPERATURE IN F
 W = FLOW RATE IN LB/SEC
 *1 ATM - 14.7 PSIA
                                                     NOMINAL 200 MW
                                            TURBINE INLET TEMPERATURE -= 2400 F
                                            COMPRESSOR PRESSURE RATIO - 10 TO I
                                                 EARLY I980'S TECHNOLOGY
                                         COMPRESSOR BLEED AIR PRECOOLED TO 200 F
                                          REGENERATOR TOTAL PRESSURE DROP = 8%
                                             REGENERATOR EFFECTIVENESS = 80%
                                                            P = 3.82
                                                            T - 875.5
P = 3.76
                                         PRECOOLER
                                         (OPTIONAL)
                                  P-9.46
                                  T = 632.9
                                  W_  = 1132.5
                                                                P = 9.17
                                                                T = 1162.8
                                                                Wa = 1132.5
                                                                                                         ELECTRIC
                                                                                                        GENERATOR
        P = 1.03
        T = 1303.8
        Wg = 1211.2
                          TO EXHAUST STACK
                         FIG.  57.  HEAT BALANCE  FOR REGENERATIVE-CYCLE  GAS  TURBINE

-------
                                                         EARLY 1980'S TECHNOLOGY



                                               CAPACITY 200 MW     TURilNE INLET TEMPERATURE Z400P

                                         COMPRESSOR PRESSURE RATIO 12i1   REGENERATOR EFFECTIVENESS 10%
CO
H
CO
                      VIEW AA


        SHOWING REGENERATOR NOZZLE ARRANGEMENT
         MODULE LINES SPACED AROUND HALF CIRCLE

*v -i
m I
REGEN
'
MODULES
L-,0
ILJJJJJ

i
_ __^ —
T— EXHAUST STACK
r, -h ^ ^ TTL >,
ii ill ill m m
Uo. I !
1 i
i i
i ! . . i
|_t i -JJ-JJ-il Jil J/I J/LJJL



-s\
111

J'l J
— V- ~z

                                                                                     CAS PLOW
           COMPRESSOR SPOOL
                               COMBUSTION CHAMBER
            POWER TURBINE

GAS GENERATOR TURBINE
                                           FIG.  58. REGENERATIVE - CYCLE ENGINE  FLOW  PATH

-------
                   OUTDOOR CONSTRUCTION IN SOUTH CENTRAL REGION
                               EARLY  1980'S DESIGNS
                                 1000-MW STATIONS
(o) EFFECT OF CAPITAL CHARGES
   6.5
   6.0
   5.5
 */>
 O
 o
 Si 5.0
 o
 Qu
 _,
 CO
    x 5
    4.3
   4.0
                     STEAM TURBINE STATION
                REGENERATIVE-CYCLE
                     GAS TURBINE

          SIMPLE-CYCLE
           GAS TURBINE
                                               VALUE USED IN STUDY
      10
11
12
   13         14
CAPITAL CHARGES %
15
16
17
(b) EFFECT OF NATURAL GAS COSTS
     8
                                     STEAM TURBINE STATION
                                  REGENERATIVE-CYCLE
                                      GAS TURBINE
            SIMPLE-CYCLE GAS TURBINE
                          COST PROJECTED FOR EARLY 1980'S
                               30                       40
                             NATURAL GAS COSTS-< PER MILLION BTU
 FIG,  59. EFFECT OF  CAPITAL  CHARGES AND GAS COSTS ON  STATION POWER  COSTS
                                       219

-------
                     OUTDOOR CONSTRUCTION IN SOUTH CENTRAL REGION
                                 EARLY 19M'S DESIGNS
                                 CAPITAL CHARGES 15*
                             GAS COSTS 30< PER MILLION BTU
  (o) SENSITIVITY ANALYSIS TO SHOW EFFECT OF INCREASED GAS TURBINE STATION COSTS ON
     POWER COSTS
  6.5
oe
x
*
                          £.
                              STEAM PLANT, 1% INTEREST DURING CONSTRUCTION
                             6% INTEREST
  6.0
                                           1
o
u
ee
ui
ae
co
*/>
                                                SIMPLE-CYCLE GAS TURBINE STATION, 8%
                                                   INTEREST DURING CONSTRUCTION
  4.0
         BASE
     0                  10                 20                 30
                     GAS TURBINE STATION COSTS ABOVE BASE - PERCENT
   (b) SENSITIVITY ANALYSIS TO SHOW EFFECT OF IMPROVED STEAM PLANT PERFORMANCE ON
      POWER COSTS
     6.5
   ee
   z
   *
   5 6.0
   I
   O
   u
  §5'5
  Q.
  oe
  ca
    5.0
               •VALUE USED IN TABLE
                      SIMPLE CYCLE GAS TURBINE WITH
                     PERFORMANCE SHOWN IN TABLE VIII
              WITH PERFORMANCE 5% POORER-\
         BASE
                  39.0
                 	L_
                                                   42.5
45.0
      9000                   8500                      8000                   7500
                          STEAM PLANT HEAT RATE - BTU/KW HR
    FIG. 60.  EFFECT  OF GAS TURBINE PERFORMANCE AND COST CHARACTERISTICS
            ON  BUSBAR  POWER COSTS
                                      220

-------
ro
                                                       EARLY 1980'S DESIGNS

                                                       FIXED CHARGES= |5%

                                                        1000-MW STATIONS

                                                     	 STEAM TURBINE

                                                     	SIMPLE-CYCLE GAS TURBINE
     i(a) NORTHEAST REGION - INDOOR CONSTRUCTION

         36
      «   ->n —
     Q
     U
     tt
     LU
     *
     o
     a.
     *   12 -
     eo
     to
     =>
     CO
                       COAL 4 304 PER MILLION BTU
OIL « 30* PER MILLION BTU
                      GAS 8 504 PER MILLION BTU
                      0.2          0.4           0.6

                         LOAD FACTOR-PERCENT
                                      (b) WEST REGION - OUTDOOR CONSTRUCTION

                                         361	
                                                                      32
                                                                      28
                                                                      24
                                     8
                                     O
                                     Q_

                                     DC

                                     m
                                     m
                                         20
                                         16
                                                                         _      \
                             0.8
                                                                                 \
               -COAL @ 16« PER MILLION BTU


                            -OIL 0 28<



   •GAS » 34* PER MILLION BTU
                                                                                    I
0.2          0.4         0.6

   LOAD FACTOR-PERCENT
0.8
                      FIG. 61. COMPARISON  OF  BUSBAR POWER  GENERATION  COSTS  IN  SELECTED  REGIONS

-------
                                                   EARLY WO'S DESIGNS



                                                   FIXED CHARGES  15%


                                                    1000-MW STATIONS



                                                  -  STEAM TURBINE
(o) NORTHEAST REGION - INDOOR CONSTRUCTION

   40
tx.
   36  -
   32  —
   28  —
*  24

 I
8  2°
ae
§
a.

a:
-<
«o
   16
   ,„
   12
                                                 	SIMPLE-CYCLE GAS TURBINE


                                                               (b) WEST REGION - OUTDOOR CONSTRUCTION

                                                                 40
                  COAL 0 25f PER MILLION BTU
              GAS (t 40< PER MILLION BTU
                               I
                                            I
                  0.2          0.4           0.6

                   LOAD FACTOR - PERCENT
                                                                 36 —
                                                                 32
                                                                 28
                                                                 24
                                                              8
                                                              tt
                                                              iu
K

-<

CO



m
                                                                  12
                                                       0.8
                   COAL (S 30* PER MILLION BTU
                                                                             GAS @ 40* PER MILLION BTU
                                            I
                  0.2          0.4          0.6


                     LOAD FACTOR - PERCENT
OJ
                            FIG.  62. COMPARISON OF BUSBAR POWER GENERATION COSTS IN SELECTED REGIONS

-------
                           FORCED OUTAGE RATE =0.02
                                (NUMBER OF UNITS - UNIT CAPACITY AS A PERCENT
                                                 OF INSTALLED CAPACITY
          TYPICAL INDUSTRY VALUE
         r- i i TIV.AU inuuj i n i
10-5
               80          82          84          86          88
              ANNUAL PEAK LOAD/SYSTEM INSTALLED CAPACITY - PERCENT
            FIG.  63.  EFFECT OF UNIT SIZE  ON LOSS-OF-LOAD
                                    223

-------
  (•) IPFICT OF SYSTEM UNIT MIX AND FORCED OUTAGE  RATE OH LOSS-OF-LOAD

          5
    et
    ui
    >•
    0
        10°
    o
    o
    u.
    o
2X10-
       10
         -l
       10
         -2
            FORCED OUTAGE
            RATE
            OF 2% UNITS
                     0.04
                     0.03

                     0.02
                                           FORCED OUTAGE RATE
                                           FOR 10% UNITS = 0.02
             (FIGURES INDICATE NUMBER OF UNITS + UNlf. CAPACITY AS A PERCENT OF
             INSTALLED CAPACITY OF UTILITY SYSTEM)
                                   i
                                   i
                         	(EIGHT -W%) AND(TEN-2%)
                        	(TEN-10%)
           78           80          82          84          86          88

               ANNUAL PEAK LOAD/SYSTEM INSTALLED CAPACITY - PERCENT

  (b) ALLOWABLE FUEL COST INCREMENT BETWEEN MIXED UNIT SYSTEM AND HOMOGENOUS UNIT
     SYSTEM
    BE
    Ui _
    8*
3
•"
         -1
    u
      U
                 I            I
          FORCED OUTAGE RATE |
         OF UNITS IN MIXED SYSTEM
              -0.02
              0.03
              0.04
T
 FIXED CHARGES  = 15%
 THERMAL EFFICIENCY = 36%
                      20           40           60           80

                     LOAD FACTOR OF ADDITIONAL UNIT - PERCENT
                                                               100
FIG.. 64. EFFECT OF  SYSTEM  UNIT CAPACITY COMPOSITION ON RELIABILITY  AND
        FUEL  COST
                                    22U

-------
  (o) EFFECT OF EXPANSION UNIT COMPOSITION ON LOSS-OF-LOAD

        1
 Ul
 §
 >-
 a
 O
 o
  I
 u.
 o
 J.
   -1
2X10
      10-
      10
       -2



"7
/




s
/ /
<£-.
(ELEVE
(TEN-1
.,' (TEN -10%)

(TEN -2%)
A
N-10%) // I
0%) AND//
(FIVE-2%) /
(EIGHT-10%)AND'
(FIFTEEN -2%)


) AND
/




(NUMBER OF UNITS - UNIT CAPACITY AS A PERCENT

OF IN
STALLED CAPAC
:ITY)
**
'^ —
ss^ —
s
FORCED OUTA
0.02 FOR 1C
0.03 FOR 2





•^
"

— 	
GE RATE =
)% UNITS
% UNITS





         78          82          86         90          94          96         100
              ANNUAL PEAK LOAD/SYSTEM ORIGINAL INSTALLED CAPACITY - PERCENT
  b) ALLOWABLE FUEL COST INCREMENT BETWEEN MIXED UNIT SYSTEM AND HOMOGENOUS
     UNIT SYSTEM
SYSTEM ALLOWABLE FUEL COST
INCREMENT - ^/MILLION BTl1
K> — • O — « K>

( EIGHT -10%) »
(TEN -10%)
_AND(FIVE-2?

iND (FIFTEEN-
^^-^
) 	

2%)
	 	 	 _


1
FIXED CHARGES = 15%
THERMAL EFFICIENCY = 36%
	





) 20 40 60 80 10
                        LOAD FACTOR OF ADDITIONAL UNIT - PERCENT

FIG. 65= EFFECT OF EXPANSION  AND SYSTEM COMPOSITION ON  RELIABILITY AND
        FUEL COST
                                     225

-------
   e) AVERAGE CREDIT DUE TO SYSTEM EXPANSION WITH SMALL SIZED GENERATING UNITS
     WHICH MORE CLOSELY MATCH  THE DEMAND CURVE
    0.30
     0.23
0. _l
X -I
u 5 0.26
> I 0.24
2n
    0.22
     0.20
FIXED CHAF
LOAD FAC
- 11

-

UNIT 5
IGES= 15%
'OR = 70%
	

— — — • 	

IZE AS APERCE
SMALL UNC
CAPITAL COST
SMALL UNIT =
11 -•••••
==•••
OF CAPI
170/KW u

— ~™^"™™ss=:
TAL COST OF
iRGE UNIT

= S165/KW
NT OF INITIAL INSTALLED CAPACITY
PS - 2% LARGE UNITS - 10*
1 1
                   20          40          60          80         1Q<0
                 TOTAL INCREASE IN PEAK SYSTEM LOAD - PERCENT
   b) ALLOWABLE FUEL COST INCREMENT RESULTING FROM EXPANDING UTILITY SYSTEM
     IN SMALL UNITS
SYSTEM ALLOWABLE FUEL COST
INCREMENT - ^/MILLION BTU
=» M *- o. «• «





V
\
\




x
^

1
FIXED CHARGES = 15%
CAPITAL COSTS
SMALL UNIT = S70/KW
LARGE UNIT = J165/KW
THERMAL EFFICIENCY = 36%

' 	 -— _




                  20          40         60          60
                     LOAD FACTOR OF SYSTEM - PERCENT
100
       FIG. 66. EFFECT OF EXPANSION  AND SYSTEM UNIT SIZE ON
                     RELIABILITY  AND FUEL COST
                               226

-------
                                  SINGLE UNIT
CASE I

SINGLE-UNIT CENTRAL
GENERATING PLANT LOCATED
NEAR SYSTEM GRID; POWER
TRANSPORTED TO LOAD
CENTER
            CENTRAL
           GENERATING
             PLANT
  STEP-UP
TRANSFORMER
                            •X" MILES
                                             STEP DOWN
                                            TRANSFORMER
                                                               SYSTEM
                                                                GRID
                           SUBSTATION DISTRIBUTION BUS
                                  (LOAD CENTER)
                                           •X" MILES-
        MULTIPLE UNIT
  CENTRAL GENERATING PLANT
                                   STEP-DOWN
                                  TRANSFORMER
                                         SYSTEM
                                          GRID
          BUS  DISTRIBUTION
            (LOAD CENTER)
CASEH
MULTIPLE-UNIT CENTRAL GENERATING PLANT LOCATED NEAR LOAD CENTER;
RESERVE POWER TRANSPORTED TO LOAD CENTER
                FIG. 67.  POWER  TRANSMISSION  SYSTEMS
                                   227

-------
                                          TWO CIRCUITS

                                          SINGLE CIRCUIT

                                          (SEE FIG. 47)
oe
•<
UJ
   2X10
Q

O
O
a
                                                         FORCED OUTAGE RATE

                                                          FOR GAS TURBINE
                                                     TRANSMISSION CIRCUIT FORCED

                                                        OUTAGE RATE =0.002
                                                     TRANSMISSION CIRCUIT CAPACITY

                                                    •JIVALENT TO A SINGLE UNIT OUTPUT
                   80          82           84           86           88

                 ANNUAL PEAK LOAD/SYSTEM INSTALLED CAPACITY - PERCENT
          FIG.  68. EFFECT OF FORCED  OUTAGE RATE ON LOSS-OF-LOAD

                                        228

-------
                   ISOD nand anavMomv waisxs
                                        o
                                        cc
nia Nomiw/* -
ISOD land aiavMcmv
229

-------
                               r
                                                 DENOTES OPTIONAL EQUIPMENT FOR HIGH BTU GAS
       (a) AUTOTHERMAL HEATING FOR HIGH OR LOW BTU GAS
                                 COAL  STEAM
                                  LJ
            AIR   I   OXYGEN
                 I SEPARATION
                 I	I
GASIFICATION


SCRUBBING
                         SHIFT   |

                      | CONVERSION I
                      I	I
    GAS

PURIFICATION
                                                                                                   I PRODUCT GAS
                                                                                        METHANATION .	«•-
ro
U)
o
                                       DUST AND TAR
                                   SULFUR COMPOUNDS
        (b) EXTERNAL HEATING FOR HIGH BTU GAS
               COAL
               STEAM
                       GASIFICATION
    GAS

PURIFICATION
                          HEAT      SULFUR COMPOUNDS
                                                                                 METHANATION
                                                                                              PRODUCT GAS
                      FIG.  70.  SIMPLIFIED SCHEMATIC  DIAGRAMS  FOR COAL GASIFICATION PROCESSES

-------
(a) ESTIMATED DEVELOPMENT COSTS
     200
  ae

  _i
  -i
  o
  a
  z
  o
o
u

K
Z
UJ
  O
  _J
  LU
  >
  1U
  a
   150
     100
      50
       0

       100
                              -ESTIMATED TOTAL DEVELOPMENT COSTS

                              FOR ESTABLISHED MANUFACTURER
                                    PRODUCT IMPROVEMENT FOR 15 YEARS
                  INITIAL DEVELOPMENT

                  FOR FIRST 5 YEARS
150                      200

    ENGINE OUTPUT - MW
                                                                              250
 (b) MARKET PENETRATION ON ENGINE SELLING PRICE
      20
      15
  I
y   10
ce
a.

z    5
S    0

UJ
z
13   "•*)

z
UJ

z  _]0

LU
o
  x
  u
     -20
                                      250-MW ENGINE-TURBINE INLET TEMPERATURE 260 ° F

                                                EARLY 1980 TECHNOLOGY



                                  DEVELOPMENT COSTS OF 250 MILLION DOLLARS



                                         125 MILLION DOLLARS
                                                     VALUE USED IN STUDY
                1000
                        2000
    3000       4000       5000


 MARKET PENETRATION - MW/YR
6000
7000
        FIG.   71.  ESTIMATED  DEVELOPMENT  COSTS  OF  ADVANCED GAS  TURBINES
                                         23i

-------
                             PRESENT TECHNOLOGY
UJ
0
UJ
O
U
*
O!
O
  POLYTROPIC     S
EFFICIENCY, % = 81 '
                            0.5                     0.7

                                 FLOW COEFFICIENT
          FIG.   72.  TYPICAL MULTISTAGE COMPRESSOR EFFICIENCY
                                     232

-------
                      PRESENT TECHNOLOGY
   12
   10
Z
UJ
"-  8
UJ  °
o
or
O
          STAGE EFFICIENCY, 'i  88
     0.4
0.8                1.2

  FLOW COEFFICIENT
1.6
  FIG.  73. TYPICAL  HIGH-PRESSURE TURBINE PERFORMANCE
                             233

-------
                         TMETAL- TCOOLANT

                           TGAS ~ TCOOLANT
     IU



     HI
     u
     IU
     Ul

     o
     o
     o
     u
         PRODUCT OF COOLING FLOW FRACTION AND GAUGING,^*- A - IN.
FIG.  74.  COOLING EFFECTIVENESS CORRELATION FOR  ADVANCED

            IMPINGEMENT-CONVECTION COOLED BLADES
                             23U

-------
                                          TWIN - SPOOL CONFIGURATION
           BEARINGS       TURBINECASING
                                      /•"
                                      V.
                          \
(     LOW  A
LOW
COMPRESSOR
HIGH -*
COMPRESSOR
BURNER(S)
        A
HIGH ->
TURBINE
                                                                            EXHAUST ELBOW
                                                                         m
                                                                                    Y
POWER
TURBINE
                 FIG.  75.  SCHEMATIC DIAGRAM OF MODEL  USED IN GAS TURBINE COST ANALYSIS

-------
                                   MATERIAL-AMS 4928
                               STRAIGHT PEDESTAL DESIGN
    10'
    103
1U
o
 I
01
2
a.
ui
in
5
QC
    101
    10'
                            PRICE
A-VENDOR NO. \
B-VENDOR NO. 2


BLADE PRICE ALSO WILL VARY WITH:

   a)  MATERIAL
   b)  PEDESTAL DESIGN

   c)  MIDSPAN ROD STIFFENERS
                                    I
                          10        15        20        25        30
                                AIRFOIL LENGTH OF BLADE ,  X- IN.
                    35
40
  FIG.   76.  MANUFACTURERS' PRICES FOR SOLID  COMPRESSOR  AND TURBINE  BLADES
                                         236

-------
             NO CORRECTION FOR PRODUCTION VOLUME
                      AMS 5661 PRICE = 118.5D
                                                 AMS 6304 PRICE
                                                    = 21.4D3'25
                          DISK D'AMETER - FT
                                                10'
FIG.   77.  PRICE  ESTIMATES FOR FORGED  COMPRESSOR  DISKS
                            23T

-------
rv>
UJ
c»
PEDESTALS
                                                                                    (NOTE: SCALE CHANCE)
                 FIG. 78.  ILLUSTRATION OF  TYPICAL IMPINGEMENT-COOLED  TURBINE  BLADE DRAWINGS

                                        SENT TO BLADE MANUFACTURERS

-------
ro
U)
vo
                  103
                lit
                o
                CD
 I

UJ
U

a:
a.
                   102
                                                     MATERUL-B-1900



                                                         AMS 5661




                                                TOOLING COSTS NOT INCLUDED
                                                          PRICE = e(°'067X + 5.03)
                                             5                       10


                                                 BLADE AIRFOIL LENGTH , X - IN.
                                                                           15
                     FIG.  79.  MANUFACTURERS'  PRICES FOR IMPINGEMENT-COOLED  TURBINE  BLADES

-------
                  103
                                              TOOLING COSTS NOT INCLUDED
ro
.p-
o
               ai
 I

tu
O
5
a.
                  102
                                                      10


                                                    VANE LENGTH  X- IN.
                                                             15
20
                      FIG.  80.  MANUFACTURERS' PRICES  FOR IMPINGEMENT-COOLED  TURBINE VANES

-------
                 NO CORRECTIONS FOR PRODUCTION VOLUME
    105
y   104
oc
0.
       10°
                               AMS5661 PRICE
                                                      AMS6304 PRICE

                                                          = 19.9 D3'17
          5            10'


DISK DIAMETER - FT
     FIG.  81.  PRICE  ESTIMATES FOR FORGED TURBINE  DISKS
                                 21*1

-------
                    GAS TURBINES NORMAL-RATING AT 100 FT ALTITUDE, ZERO INCHES WATER INLET PRESSURE DROP

                                                    AND 80 F

                        (b) EFFECT OF AMBIENT TEMPERATURE ON GAS TURBINE OUTPUT
                                                  ALTITUDE = 1000 FT
                                                                     VARIATION FOfl HIGH
                                                                     PRESSURE RATIO
                                                                     TWO-SHAFT ENGINES
                              VARIATION FOR LOW TO MODERATE

                                  PRESSURE RATIO ENGINES
                             -40
        40       80       120

        AMBIENT TEMPERATURE-F
(a) PART-LOAD PERFORMANCE FOR TWO-SHAFT ENGINE
          (c) EFFECT OF AMBIENT PRESSURE AND INLET PRESSURE  ON GAS TURBINE OUTPUT

                                                                8000
    40          60          80

       PERCENT OF NORMAL-RATED POWER
100
                                                  o
                                                  a.
                                                   a:
                                                   O
                                                   •z.

                                                   H-
                                                   Z
                                                   Ul
                                                   U
                                                   a:
                                                   01
                                                   a.
                                                             I         I
                                                            INLET PRESSURE
                                                             DROP CURVES
               0.8
                                                      0.7
                                                                6000
                                                                                                           Ui

                                                                                                       4000 §
                                  2000
 14       13        12

BAROMETRIC PRESSURE-PSIA
                           FIG. 82. OAS TURBINE OFF-DESIGN CHARACTERISTICS

-------
                                  SECTION XIII
                                  PUBLICATIONS
     As a result of the work performed under Contract No.  1^-12-593, the
following publications have been produced.

          Utility Applications for Advanced Gas Turbines to Eliminate
          Thermal Pollution.  ASME Preprint TO-WA/GT-9.  Presented
          at the ASME Winter Annual Meeting, November 29-December  3, 1970,
          New York,. New York.  Authors:   F. R. Biancardi and  G.  T.  Peters.
                                       2U3

-------
                                   SECTION XIV
                                   APPENDICES
                                   APPENDIX A
         EMISSIONS OF NITROGEN OXIDES FROM GAS TURBINE-TYPE  POWER SYSTEMS
     Although nitrogen oxide (NO )  emissions from a power system depend upon many
                                JC
design and operating factors, gas turbine-type power systems  usually  emit  less
NO  (on a ppm stack gas concentration, or pounds  per unit heat  input  basis) than
  Jv
do reciprocating internal combustion engines or large steam power "boilers.  Gas
turbine-type power system NOX emissions usually range from 75 to 130  ppm (Ref.
138), whereas coal-fired steam boiler NO  emissions may be as high  as 1200 ppm
(Ref. 139).  The NO  emissions from oil- and gas-fired steam  boilers  are generally
                   Jv
lower than from coal-fired units, and values as low as 150 ppm  have been reported
in some gas-fired steam boilers (Ref. 1^0).   However, efforts to reduce NO
emissions from steam boilers have also resulted in reduced operating  efficiencies .

     The combustion of a fossil fuel with air in all types of power systems
results in the formation of nitrogen oxides.  Because 90$ or  more of  the NOX  in
the stack gases is present as the relatively unreactive nitric  oxide  (NO), methods
to control NOX emissions by stack gas removal are quite complex (Ref. 1^1) and
appear to be unattractive unless they might  possibly be combined with sulfur  oxide
stack gas removal.  Thus, the most practical method of limiting NOX emissions  is
to control their formation in the combustion process itself and in  the subsequent
processes by which the combustion products undergo cooling before being emitted
from the stack.

     Theory predicts that values of NO  concentration in the  combustion products
are dependent on design and operating conditions, flame temperature,  excess  air
(atomic concentrations), and residence time.  The formation of  NOX, which  begins
with the onset of combustion in the primary combustion zone where temperature is
a maximum, continues at a relatively slow pace (due to low chemical reaction  rates)
as colder bulk gas enters the recirculating flame zone.  The  NOX formation cannot
proceed until high temperatures are achieved through the combustion of the hydro-
carbon fuel with air, and therefore, the hydrocarbon chemistry  is virtually
completed before NOX formations begins.  The amount of NOX formed thus cepends  on
the conditions in the primary zone and the subsequent temperature and concen-
tration history of the combustion products with time.
                                        2U5

-------
      Since the NOX ia formed after the hydrocarbon chemistry has been  completed,
 the potential exists for controlling NOX while achieving  good combustion
 efficiency.  Although present theory is useful in pointing  out the  approaches
 vhich may be employed to limit NOX emissions,  it  is not comprehensive  enough
 to explain all the interactions involved in the chemical, thermodynamic,  and
 fluid dynamic processes.  Experience has shown that NOX emissions may  vary
 extensively between power systems having similar  operating  conditions, and may
 even be widely different in identical equipment.

      Design and operating factors found to affect the power plant NOX  emissions
 include:  fuel type and composition, heat release rate, burner configuration,
 excess air, and air inlet temperature.   All of these factors,  in turn, affect
 flame temperature, excess air, and residence time and thus  dictate  the amount of
 NOX formed.  Low-excess air combustion, multiburner combustors, steam  and water
 injection, reduction of air preheating, and recirculation of stack  gases
 have been used with some success  on some steam boilers to reduce NOX emissions.
 However, the use of these methods usually results in some degradation  in  operating
 efficiencies.   Stack gas recirculation was used to reduce the  temperature in the
 primary combustion zone and thereby achieve the aforementioned 150-ppm KOX con-
 centration in the exhaust of gas-fired steam boilers (Ref.  lUO).  Some, but
 obviously not  all, of the above methods could  be  applied  to reduce  the NOX emissions
 from gas turbine-type power systems without compromising  good  operating efficiencies,

      In present-day gas turbines, the primary  zone in the combustor is operated
 at  near-stoichiometric fuel-air ratios, and a  great deal  of recirculation in
 the flame zone is designed into the combustor  to  ensure proper combustion.  These
 characteristics  make for a hifcl- primary zone temperature  and long residence time,
 both of which  tend to promote the formation of NOX.  This type of operation is
 necessary in aircraft jet engines to vaporize  the fuel droplets and ensure stable
 fxame  propagation.   It is generally difficult  to  have a lean fuel-air  ratio in
 the primary zone, and to reduce recirculation  in  the aircraft  application, and
 still  meet the requirements  for altitude restart  capability  and flame  stability
 and propagation  at low pressures  (high  altitude and idle  power settings).  In a
 stationary gas turbine power system,  these  requirements do not apply and  stable
 flame  propagation with a lean primary combustion  zone and little recirculation
 could  be  achieved by premixing the  fuel  and air.   This approach would  not be
 considered for the  aircraft  application  because the widely varying  operating
 condition  would  enhance  the  possibility  of  an  explosion.   Another possible method
 of  reducing NOX  emission  in  a stationary gas turbine-type power system involves
water,  injection,  which  is  used to  cool  the  primary  combustion gases before
 substantial  amounts  of NOX have time  to  form.   It  is claimed (Ref.  lUl) that
water  injected in a mass  equal  to the fuel mass would reduce HOX emissions from
gas  turbines by  90JC.

-------
     Based on the previous discussion, and the Corporate- and Government-sponsored
work "being conducted by UA Research Laboratories and Pratt & Whitney Aircraft., it
appears that the principles and techniques required to reduce KOX emissions from
gas turbines are understood.  Furthermore, it appears that the NOX emissions
achievable in advanced gas turbine-type power systems can be lower than in present-
day engines and that these lower NOX emissions can be obtained without compromising
gas turbine performance.

-------
                                    APPENDIX B
                    DESCRIPTION OF GAS TURBINE DESIGN PROGRAM
     Power plant design parameters and constraints, reflecting the design
technology and materials improvements projected in the main body of this report
to be available during the 1970 decade, early 1980's, and late 1980's were used
to determine the performance and power plant dimensions of simple-cycle power
plant designs incorporating either single-spool or twin-spool turbomachinery
configurations, regenerative-cycle power plant designs with single-spool con-
figurations, and compound-cycle designs with three-spool configurations.

     The primary independent variables (namely, turbine inlet gas temperature,
compressor pressure ratio, and regenerator effectiveness) and the various turbine
cooling techniques considered are given in Table XVI for each type of power plant
design.  Turbine blade cooling configurations commensurate with anticipated
advances in the state of the art for the projected time period were investigated,
and the associated penalty to the gas turbine power system performance was
appropriately assessed.

     A high-speed digital computer program previously developed under Corporate
sponsorship was used to facilitate these parametric studies and also to provide a
realistic assessment of turbine cooling flow penalty effects on power plant thermal
efficiency and specific output (hp/lb/sec).  A brief description of this gas
turbine computer program is outlined in the following paragraphs along with a
definition of the basic assumptions.
                       Gas Turbine Design Computer Program

     Once the primary independent variables, i.e., the time period for the
power plant design, thermodynamic cycle, number of compressor spools, turbine
cooling technique, turbine inlet gas temperature, compressor pressure ratio, and
power plant airflow rate have been specified in the program, all combinations
of the design parameters (incorporated into the program) are investigated, and the
appropriate combination of parameters that satisfy specified design constraints
are determined.  The power plant flow path, pertinent dimensions, number of
compressor and turbine stages, turbine cooling flow requirements, and allowable
metal temperatures are then computed.  After these parameters have been determined,
the computer program assesses the effect of the calculated turbine cooling flow
rate, including the collective contributions of blade and vane cooling flows and
disk cooling flow, on specific horsepower and thermal efficiency.

-------
                                  Basic  Assumptions

      Certain restraints vere necessary  in the use of  the computer program.  For
 Instance, constant-ciean-dianeter flow passages vere assumed for all turbomachinery
 performance coasputations to provide  some  control over the number of possible
 design concepts.   The number of compressor stages in  each spool was computed on
 the premise of a constant flow coefficient per stage.  This assumption, together
 vlth the preceding one concerning constant mean diameter, resulted in a constant
 axial velocity and hence constant stage work for each of the respective compressor
 •pools.   Compressor stage performance representative  of current advanced-design
 aircraft technology (high stage loadings)  is presented in Fig. 72.  For the 1970-
 decade base-load compressor designs  the polytropic efficiencies and flow
 coefficients depicted in Fig.  72 were used, but the work coefficients were derated
 to  80$ of the Fig. 72 values.   The early  1980- and late 1980-technology compressor
 designs  vere assumed to exhibit the  work  coefficient  and flow coefficient per-
 formance depicted in Fig.  72,  but at the  correspondingly higher polytropic
 efficiencies presented in Table XVI.  The  aspect ratio  in the first stage of all
 compressor designs was not allowed to exceed 3.0, thus avoiding the need for costly
 stiffening rods  that would also contribute to increased pressure losses and
 degradation in compressor efficiency.   The last-stage hub/tip ratio for all
 cmpressor designs was not  allowed to exceed 0.93 in order to minimize blade end-
 losses.

      In  all twin-spool gas turbine designs, the high-pressure turbine was assumed
 to  comprise a single stage capable of delivering up to a maximum specified value
 of  stage work.  The high-pressure turbine  performance was described by Fig. 73 for
 the 1970-decade power plant designs.  The  early- and late-1980's designs were
 assused  to exhibit the same work coefficient and flow coefficient performance
 depicted in Fig.  73,  but  at the correspondingly higher efficiency values shown in
 Table XVI.

     The low-pressure turbine  consists  of  high-efficiency, stages each capable of
providing the  same work output  and whose number was determined from the appropriate
 design parameters  and constraints.   Designing for high stage work and hence
 a relatively high  temperature drop in the  high turbine reduces the cooling flow
requirements and possibly  the high-temperature material requirements in the low
turbine.   The number  of turbine  blades  and vanes in both the high-turbine and low-
turbine  stages were estimated on  the basis of velocity triangles and lift coeffi-
cients comparable with  current  aircraft engineering design procedures and con-
sistent with the assumed stage efficiency levels.   The ratio of blade height to
axial width for the unshrouded high-pressure turbine was limited to a range of
values between a maximum of  2.5  and  a minimum of 1.0.   High-turbine minimum mean
axial widths of 1.0 and 1.5  in. were assumed for the blade and vane, respectively.
                                        250

-------
     The lov-turbine stages and the power turbine stages vere assumed to be
shrouded, and the maximum blade height-to-axial-width ratio was not allowed to
exceed a value of 5.5-  Low-turbine minimum mean axial width of 1.5 in.  was assumed
for both the blades and vanes.  The aforementioned basic assumptions are consistent
with related aircraft propulsion system design technology.

     Turbine blade and vane cooling flow requirements were obtained from correlations
of cooling effectiveness (n) with cooling flow as shown in Fig. fh for advanced
impingement-convection cooling techniques.  Cooling effectiveness is defined as
the difference between average blade or vane metal temperature and cooling air
temperature divided by the difference between gas temperature and cooling air
temperature.  The correlations for advanced impingement-convection cooling shown
in Fig. 7^ are comparable to the best current cooling designs for aircraft gas
turbine propulsion systems.  Cooling flow fraction is not used directly in Fig. 7U
but is multiplied by the gauging (defined as the airfoil passage throat width).
Hence the cooling flow required depends on geometry in addition to temperature.

     Disk cooling flow requirements were estimated on the basis of related
engineering experience, since accurate determinations of this parameter would have
entailed detailed heat transfer and seal leakage analyses.  Disk cooling flow
requirements of 0.75$ per face for the high turbine and 0.315% per face for each
stage of the low turbine were assumed for all design conditions.  It was assumed
that disk metal temperatures could be maintained at from 1200 F to lUoo F or lower
with these cooling flows.  The contribution of cooling flow to the degradation in
turbine adiabatic efficiency was estimated in the following manner.  A 1% penalty
in high-turbine stage efficiency for each percent disk cooling flow in excess of 1.0%
was assumed.  Further, a. 1% penalty in low-turbine stage adiabatic efficiency for
each percent disk cooling flow in excess of 0.25$ was also assumed.  The blade
and vane cooling flows also contribute to a loss in adiabatic efficiency.  A 0.5$
decrease in stage adiabatic efficiency was assumed for each percent of cooling
flow for the blades plus vanes in each stage.

     The long-time (l% creep in 100,000 hr), steady-state, creep strength-
temperatures properties for the 1970-decade turbine blade alloys (Fig. 2U) were
estimated from Larson-Miller-type extrapolations of short-time (1000-hr) l$-creep
data corresponding to nickel-base alloys.  The long-time, steady-state,  creep
strength-temperature properties assumed for the early 1980- and late-1980-decade
technology turbine blade alloys were estimated by assuming a 20 F/yr improvement
in allowable metal temperature with the 1970-decade alloy serving as the
reference material.  Relatively short-life (l% creep in 1000 hr) aircraft materials
have improved at a rate of 30 F/yr.  The projected early 1980-decade creep strength
properties (see Fig. 2U) agree reasonably well with Larson-Miller extrapolations
of an advanced nickel-base alloy and a unidirectionally solidified eutectic alloy
currently under development for advanced aircraft propulsion systems.  Similarly,
the projected late 1980-decade material data also compare with preliminary data
for columbium alloys as well as with projections for future high-temperature
chromium alloys.


                                       251              '  -

-------
      The average "blade metal temperatures used to  determine  cooling effectiveness,
 defined in the preceding discussions,  were computed vith  the aid  of the creep
 strength properties in Fig.  2U, used in conjunction vith  a simplified blade root
 stress relationship, defined as follows:
                  S, g     1 |                     AT
                        = -  1 + H/T - 2  (H/T)2  + -i  (2 - H/T  -  (H/T)2
                          6

 where      S^ * blade root stress,  lb/ft2

             g = gravitational constant,  32.2  ft/sec2

            p,  = "blade material specific  weight, lb/ft3
             o

            Vrp = blade tip speed, ft/sec

           H/T = rotor hub/tip ratio

        A /A  = blade taper ratio

 The blade allowable  stress  was  assumed equal  to the calculated value of blade root
 stress, and the allowable metal temperature was determined from Fig. 2U as a
 function  of the blade allowable stress.  The  allowable vane metal temperatures
 vere  obtained from Fig.  2U,  assuming  an  allowable stress of 5000 psi in the high
 turbine and 10,000 psi  in the low turbine.

      Adaptation of Fig.  2k  to the computer program placed the emphasis on the use
 of the best material  available  during  the specified time period to minimize cooling
 flow  requirements as  an  alternative to placing  emphasis on less-expensive alloys
with  comznensurately higher  cooling  flow requirements.  This design philosophy
was also  used to determine  the  power turbine  configurations.  As a result, only
a few of  the power turbines designed in this  study required cooling.  The
discussion in SECTION VII indicated that avoiding cooling the power turbine
appears to be the most economical approach.  A maximum allowable blade root stress
up to 60,000 psi was imposed  on  all power turbine designs.

     The performance penalties  attributable to turbine cooling were computed in
the gas turbine design computer program as follows.   The mixed flows were
calculated at the appropriate stations in the turbine by a mass balance between
the main stream and the cooling flows.  The temperature of the mixed stream at
each station was calculated by a simple heat balance  between the main stream
                                        252

-------
and the cooling flows.  The effects of cooling flow pumping losses on the net
turbine work and the effect of blade and vane cooling flow and disk cooling flow
on the adiabatic stage efficiency were also incorporated into the program.
                                        253

-------
                                    APPENDIX C
                       DESCRIPTION OF GAS TURBINE COST MODEL

     The gas turbine engines "being considered as prime movers  for  advanced-cycle,
base-load electric power generating stations vhich could be commercially available
in the next two decades represent significant advances in the  state  of the  art,
both from the standpoint of operating conditions and that of unit  output power.
It is difficult to predict accurately the impact these engines vould have on
the future of electric power generation without an accompanying economic analysis
since these future gas turbine engines could be characterized  by entirely new
design criteria.  By contrast, many present-day industrial gas turbine engines
are merely adaptations of aircraft engines in which light weight and high specific
power (ib thrust/lb airflow) were emphasized.  As a result, few of the general
economic correlations developed for evaluating present-day industrial designs
could be expected to hold true for future designs, primarily because many of  the
required geometric and economic scaling factors are not well defined.  Further,
these future industrial-type gas turbine designs would make use of innovations
not applicable to current machinery in order to attain the performance necessary
to be competitive with alternative methods of generating base-load electric power.
The significance of these observations is that, except for the design of aerodynamic
flow passages (primarily the vanes and blades), and use of aircraft-type advanced
materials, future base-load large power output gas turbine engines could be
designed and manufactured with somewhat different philosophies than  are considered
common today.
                                Economic Analysis

     Descriptions of engine cost estimating procedures developed under United
Aircraft Corporation sponsorship and presented in this section are concerned
primarily with the economic analysis which was used to estimate costs and selling
prices of future gas turbine engines incorporating all standard accessories such
as the fuel system, inlet housing, and exhaust stack.   Cost estimates for
peripheral equipment such as pumps, generators, housings, mountings,  etc., are not
included here but were made and taken into account in the overall system costs
presented in SECTION VIII.

     A survey was made of present-day gas turbine engines manufactured by Pratt
& Whitney Aircraft, General Electric, and others to determine whether any
correlations oetween price and engine characteristics existed, and if so, whether
these could be used in a gas turbine model synthesized to reflect technological
developments anticipated for the forthcoming years.  Unfortunately, no general
correlations of this type were found to exist, primarily because most of the
                                       255

-------
 current engine models vere designed for entirely different  types  of service where,
 as noted, differing sets  of design criteria ultimately  dictate  final configurations.
 As a result, it was found necessary to establish a model  for  the  economic analysis
 which synthesized ah engine design and summed the costs of  the  various major
 individual components in  such a manner that overall  costs could be developed in
 "building-block fashion.   These costs for the various major  individual components
 were obtained from the literature where possible, but principally from vendors'
 estimates and from qualified United Aircraft personnel.

      It is interesting to note that practically  every source  of economic data
 indicated that the primary variables affecting gas turbine  engine component
 costs were:   (l) the individual component geometry (primarily its linear dimensions);
 (2)  the material selected; and (3) the production volume.   Other  variables were
 estimated to have little  effect on cost estimates in a generalized economic
 program suitable for analyzing many different gas turbine engine  designs.

      A schematic diagram  of the model developed  to estimate manufacturing costs
 of advanced gas  turbine engines is shown in Fig.  75-  Detailed  cost relationships
 were developed for the low and high compressors,  the burner,  the high, low, and
 power turbines,  the shafts,  the bearings,  and the casing and  exhaust elbow.
 The  costs of the inlet housing, fuel system,  and miscellaneous  small parts were
 lumped together  and considered to comprise a constant fraction  of the overall
 manufacturing cost.   Provisions in the economic  program were made to analyze
 engine designs incorporating compressors with either constant hub, mean, or tip
 diameters.   Different blade and disk materials can be specified for each compressor
 and,  in addition,  provision is made to accommodate a change in  blade and disk
 materials in the high-pressure compressor.

      The burner  was  sized using volume  flow,  reference velocity, and length-to-
 diamter (or  height)  relationships,  and either  an  annular (indicative of advanced
 engines)  or  a cannular (indicative  of present-day engines) design can be
 accommodated.  Each stage of the turbine section  was analyzed individually as
 were  the compressor stages.   Varying hub and tip  diameters as well as differing
 blade and disk materials  were accommodated.  Provisions also were made to estimate
 costs  for impingement cooling in the turbine blades  and the use of tip shrouds
 where  necessary.

     As many as  three shafts  (two:in the gas generator section  and one in the
 power  turbine) were  included in the  overall analysis.  The shaft diameters were
 calculated by  using material stress  relationships  and rotor torque.   Casings
 for each  section of  the engine were  assumed to be  cast in halves, and final
machining was  assumed to be  conducted in the engine manufacturer's facilities.
 Casing thickness was  based on either the minimum  thickness required to withstand
 internal  gas pressures, or the minimum thickness required for sound casting
techniques.  In the  former case, a metal working stress  of 30,000 psi  and a factor
                                       256

-------
of safety of 5 have "been assumed.  The exhaust elbow vas  considered  to  be made
from sheet steel and to be similar in design and fabrication to  those elbows
used on present-day engines.  A choice of bearings was  considered, including
Kingsbury thrust, roller, and ball types, the prices of which were found to vary
primarily as the bore diameter of the type selected. In  those engine designs
not requiring bearings at a particular location, provisions  were made to omit that
particular component cost estimate.  As mentioned, several  specific  engine
components (fuel handling equipment, inlet housing, nuts, bolts, etc.)  were
combined, and the associated cost of these parts was assumed to  be equal to l8j?
of the total manufacturing costs.
                           Component Cost Information

     Once the basic economic model was established,  vendors were  consulted  to
obtain costs for the different major engine parts.   Each vendor contacted was
supplied with illustrative drawings and a list of materials which  were  suitable
for the part to be manufactured, and then was asked  to supply  price  information
in 1969-70 dollars.  Where necessary, tables of nominal dimensions were  also
supplied to the vendors in order to provide them with  additional  information
which would cover a range of parts sizes to be considered in this study.

Compressor Section

     Drawings of solid compressor blades with aerodynamic blade lengths  from
6 in. to 36 in., were supplied to representative blade manufacturing organizations
to assist in obtaining price estimates. These prices (see Fig. 76) are primarily
for blades with straight pedestals, (i.e., a constant  hub diameter design)
manufactured from AMS ^928 material.  Mid-span rod stiffeners  would  be required
on blades with lengths greater than 2k in. and length-to-chord ratios  greater
than 3.0.  These stiffeners are quite expensive if cast integrally with  the blade,
but for the industrial designs considered, round rods  inserted through,  and welded
to each blade would be used and would increase the cost of each blade  by only
about $2.00.  Sloping pedestal blade designs would be  approximately  15$  more
expensive than the prices for constant hub diameter  designs, primarily because
of the increased machining required on the pedestal.  It was recommended that
changes in blade materials could be accommodated by  assuming that Uo$  of the
blade price would be material cost and 6Q% labor charge.  Different  blade
materials were handled in the analytical program by  taking the overall blade
dimensions of length, height, and width, calculating the volume of this  block
of material, estimating the price of this block if made from the  "new" material,
then replacing the original material fraction with the new estimate  in the  basic
blade price analysis.  Tooling costs for each blade  design were estimated at
$28,000, based on data obtained from the representatives, a value  which was  assumed
to be written off over a five-year production run.   It was assumed that  individual
                                       257

-------
 compressor vanes would be formed from strip stock at  a price  of $2.^0 per  foot.
 The vanes vould "be inserted into a vane ring holder,  and the  cost of material,
 insertion, and welding vas estimated to be $7.10 per  inch of  rim circumference.
 Finally, the cost of spacers between adjacent stages  was estimated to average
 $1*00 per stage.

      Drawings for typical compressor disks were  sent  to a manufacturer of  some
 of the largest forgings in the US.   Representatives from this vendor, in turn,
 supplied the price estimates used in this  study  (see  Fig.  77).   These estimates
 were based on production of at least 50 disks per year in one run.  The estimates
 would increase by 25# for a run of 20 disks, but would decrease  linearly for
 productions rates between 20 and 50 disks.   Tooling costs  for compressor disks
 with diameters less than 8 ft were  estimated for disks produced  from both AMS
 6301* and AM3 5661 materials.  Above a diameter of 8 ft, open  dies would be
 necessary and these would be maintained by the manufacturer at no cost to the
 purchaser.  It was recommended that the forged disk prices be increased by 30%
 to cover finished machining costs.

 Engine Burner

      The costs of the engine burners  were  estimated to be  proportional to the
 surface area of the burners.   Surface area,  in turn, was  computed for a given
 burner volume, design reference velocity,  and burner length-to-diameter (or
 length-to-annular height)  ratio.  Combustion chambers  were estimated to cost
 $400 per ft2 and $600 ft2  for cannular and  annular burner  designs, respectively.
 Accessories and manifolds  would add an estimated $35 per Ib of airflow to this
 cost.   Provisions  were included in  the analytical program to maintain the maximum
 burner diameter within certain present limits, and adjustments in burner length
 were made if this  overall  diameter  was  exceeded.

 Turbine  Section

     Many of the same  costing  techniques used for compressor blades were applicable
 for  turbine  blades.  However,  impingement-cooled turbine blades  (and vanes)
 require  entirely different manufacturing techniques  from those used to produce
 solid blades  and,  consequently, separate price estimates were needed.   Therefore,
 blade and vane  drawings were supplied to several manufacturers (Fig.  78).   Price
 estimates were received from two of these vendors for blades and vanes,  respectively,
produced  from B-1900A and Inconel 713C materials as shown in Figs.  79  and 80.
 It was soon  evident from these prices that  impingement-cooled turbine  blades would
be considerably more expensive than solid blades with the same dimensions and
materials.   Tooling costs for blade and vane designs  with airfoil lengths less than
 11 in. for blades and 7.5 in. for vanes were estimated to be $100,000  each, while
 tooling costs for blade and vane designs beyond these lengths  were  estimated to
be $150,000  each.  All turbine blade tooling costs were asstmed to  be  amortized in
 five years.
                                       258

-------
     Discussions vith vendors revealed that if it vere necessary to cast tip
shrouds integrally with turbine blades , the price estimates  present for solid
blades and for hollow blades would be increased by 10/S to allow for additional
casting complexity and machinery.  Machining the root sections  of the blades for
engine designs with varying hub diameters would add an additional 15$ to the
turbine blade prices.  No estimates were given for the variation of direct
manufacturing costs with production volume, but vendors indicated that the
volume of blades and vanes produced should be sufficiently high to preclude  any
upward adjustments in the prices per unit.

     Turbine disk drawings were supplied to vendors, and the price estimates
summarized in Fig. 8l were obtained for disks with diameters ranging from 3  ft  to
11 ft manufactured from tvo materials, MS 630^ and AMS 5661.   Since most of the
dimensions on gas turbine disks would be essentially proportional to disk diameter,
irrespective of the absolute value of disk diameter, the fact that the prices of
both the compressor and turbine disks would be roughly proportional to the third
power of the diameter (or essentially, to the disk volume) is not surprising.  The
tooling costs for turbine disks would be about the same as those for compressor
disks, varying only with disk material.  In addition, if disks  with diameters
larger than 11 ft were foreseen, entirely new forging facilities might be necessary
since no commercial facility exists, other than that presently  operated by the
US Air Force, which would be capable of producing such large  parts.  The cost of
such a new facility could be in the several-million-dollar range, all of which
would have to be charged to the production of engine disks should no further
commercial applications be developed.

Engine Bearings

     Additional vendors were contacted to obtain price estimates for large,  anti-
friction, roller and ball bearings as a function of bore diameter.  All cost data
obtained were for oil-damped  bearings manufactured from M-50 carburized steel  in
lots of at least 50 bearings per year.  If lots of less than 50 bearings were
made, the unit prices would increase by approximately 50$.  Estimates received
indicated that purchase prices for sleeve bearings and Kingsbury thrust bearings
would be approximately kQ% and 50$, respectively, of those for  roller bearings
with the same bore dimensions.  To each bearing price must be added a charge for
bearing supports, estimated to be approximately 50/S of the price of the bearing
alone.

Shafts and Casings

     The price of turbine engine shafts was estimated to be  $10/lb based on
available data.  Shaft weights were calculated from an estimate of shaft torque,
shaft length (a function of engine power rating), and the material density.
                                        259

-------
      Consultations with corporate personnel  revealed that, based on recent
 experience,  casings would cost approximately $1.50/lb.  A charge equal to 30% of
 the raw casting cost was added for setup  and  in-house machining.  As noted
 previously,  casing thickness would be  dictated either by the internal gas working
 pressure or  the minimum thickness necessary  to obtain sound castings.  The cost
 of the exhaust elbow was estimated to  be $10,000 for an engine with a throughflow
 of 250 Ib/sec, and proportional to the square root of the airflow for larger
 engines.

 Assembly Costs

      Assembly  and  test  costs are difficult to assess accurately.  Based on recent
 industry experience,  these costs should  vary  between $75,000 and $150,000 per
 engine, depending  on engine complexity and size.  In circumstances where complete
 engines would  be too large to  ship  to  the erection site in a completely assembled
 package, it  was assumed that subassemblies would be shipped to the site and
 assembled in position.   Whether the engine were shop-assembled or field-erected,
 the  cost of  assembly is  included in the manufacturing cost as if all units were
 shop-assembled.  X-ray  costs might  be  as high as $15,000 per engine inspected,
 and  when extremely  large units  are  built, it  would be necessary to X-ray every
 unit assembled.  For small, open-cycle engines, X-ray inspections would not be
 required as  frequently,  and consequently, the X-ray cost per engine produced would
 be less.

 Selling Price

     The summation  of individual  engine  component costs, assembly and test charges,
 and  X-ray  costs, comprise the manufacturing cost per engine.   In the analytical
 program,  it was  assumed that many of the component parts would be supplied in
 finished form by the  vendors and, where not,  appropriate additional machining
 charges  (including  overhead) were added.   Manufacturing overhead charges were
 assumed  to be similar to those  of industrial manufacturing plants,  and not to
 those which have been found ncessary in high-technology organizations.

     Although it would be essential that  a manufacturer know  his production costs,
 the ultimate electric utility purchaser would be concerned primarily with the
 final purchase price he must pay for a piece of machinery.  Therefore,  in order
 to develop a selling price which must be  charged by an  engine  manufacturer,
 additonal economic  data was specified.  These include:  the total development cost
of the engine (including additions to existing manufacturing  facilities  needed
 for engine production);  the average engineering expenses needed to  support the
production and operation of a particular  engine design;  the gross profit per
unit produced;  and  the general and administrative (G&A)  expenses allocable to the
particular program.  In addition, an estimate of market  penetration must be made
                                       260

-------
to schedule production volume and to estimate amortization of tooling  costs  during
production.  By taking the simplified approach of allocating development,  con-
tinuing engineering, profit, and G&A among all engines  produced,  the selling
price was developed directly from manufacturing costs.   In this analysis,  expenses
such as sales and field engineering support are treated as separate costs, and,
in all likelihood, the manufacturer vould increase the  selling price per unit  to
cover these expenses.
                                       261

-------
                                     APPENDIX D
                       GAS TURBINE OFF-DESIGN CHARACTERISTICS
                               Part-Load Operation

     The part-load performance characteristics of advanced-design  gas  turbines
considered in this study would "be similar to that shovn in nondimensional  form
in Fig. 82 for present-day, two-shaft (free power-turbine) configurations  operating
at constant rated output speed.  The correlation in Fig,  82,  is  depicted as  a
fairly wide "band, particularly at the reduced load settings,  because of the  scatter
in the original data points.  This scatter is due to differences in the operating
conditions (cycle pressure ratio, turbine inlet temperatures, etc.) of the different
engine designs represented.

     If high performance at part-load power were desired, the power turbine  could
be equipped with variable inlet guide vanes which would add to the complexity
and cost of the engine.  In most gas turbine engine designs,  the turbine inlet
guide vanes are fixed, and part-power is achieved by reducing turbine  inlet
temperature.  Although with this mode of operation engine airflow  rate decreases
slightly at part power setting, the principal factor contributing  to  a reduction
in output power is the decrease in heat content of the working fluid,  resulting
from the lower gas temperature.  If part-load power is achieved  by utilizing
variable geometry guide vanes  (maintaining choked flow) to control the airflow
rate through the engine constant turbine inlet temperature is maintained at  a
wide range of power settings.  Thus, although some degradation in  component
performance would result from the reduced airflow rate and from  changes in turbine
inlet guide vane angle, the loss in power plant performance would  not  be as
significant as that indicated in Fig. 82 for the fixed-geometry  operation.
However, there is a limit imposed by mechanical and aerodynamic  considerations  to
the maximum reduction in airflow rate which can be achieved with a variable-
geometry turbine.

     The part-load characteristics of the gas turbine are of  secondary importance
in this study because the power station design load factor selected requires the
engine to be essentially base-loaded, i.e., operating most of the  time at  rated
power.  To analytically determine the part-load performance of a gas  turbine
requires the use of a relatively complex engine matching procedure and involves
a detailed knowledge of the off-design performance of each component.
                                        263

-------
                           Effect of Ambient Conditions

     Gas turbine performance is directly affected by operating altitude and ambient
temperature.  As previously mentioned the performances in this study vere computed
in accordance with NEMA specifications; i.e., an altitude of 1000 ft above sea
level and an ambient temperature of 80 F.  Correlations of the effect of changes
in ambient temperatures on output pover for a number of present-day industrial-
and modified aircraft-gas turbine designs are also presented in Fig. 82.  The
data vere plotted for an altitude of 1000 ft and were normalized with respect to
a reference temperature of 80 F.  Output power decreases at higher operating
altitudes due to the lower density of gas and thus reduced mass throughput
capability of the compressor.  Generalized approximations showing the effects of
altitude as well as variations in inlet pressure losses on output power are
depicted in Fig. 82c.
                                       261*

-------
1
V
5

Accession Number
V
2

Subject Field & Group
024A
SELECTED WATER RESOURCES ABSTRACTS
INPUT TRANSACTION FORM
Organization
United Aircraft Research Laboratories of the United Aircraft Corporation,
    East Hartford,  Connecticut 06108
     Title
      ADVANCED NONTHERMALLY POLLUTING GAS TURBINES IN UTILITY APPLICATIONS
•JQ Authors)
Biancardi,
Peters, G.
Landerman,
F. R.
T.
A.M.
16
Project Designation
U.S. Environmental Protection Agency
21

Contract 14-12-593
Note
 22
     Citation
     Water Pollution Control Research Series,  16130DNE03/71, 264 p., March 1971,
     82 fig.,  31 tab.,  142 ref.
 23
   Descriptors (Starred First)

   Thermal power plant, Thermal pollution, economics
 25
   Identifiers (Starred First)

   Gas turbine* Base load
 27
   Abstract Detailed performance, size, and cost  estimates were  made  for advanced simple-,
   regenerative-, and compound-cycle gas turbine  engines for turbine  inlet temperatures
of 2000° F and above as anticipated to be commercially  available  in the next two decades,
Conceptual designs for 1000-Mw central power station utilizing gas  turbines and compar-
isons of complete gas turbine and steam turbine power station  installed costs  and total
busbar power costs were made for the various regions of the US.
   It is shown that the gas turbines in the 1970  decade could  produce electric power
at lower costs than steam turbines in the South Central region of the US where natural
gas is readily available.  Elsewhere in the US  the  gas  turbines would be economically
competitive if moderately priced clean fuels are  available.  Advanced gas  turbines will
become more competitive in the 1980 decade as anticipated  increases in turbine inlet
temperature, component efficiences and larger engine designs lead to  more  efficient and
lower-cost engines and power stations.
   Although the development costs for large, advanced gas  turbines  would approach from
100 to 200 million dollars, the total amount that utilities are expected to expend for
cooling devices to combat thermal pollution over  the next  two  decades will exceed more
than ten times this amount.  Thus advanced gas  turbines should be given serious
consideration for increased research and development support.
   This report was submitted in fulfillment of  Contract 14-12-593 under the sponsorship
of the Environmental Protection Agency, Hater Quality Office.  (Shirazi - EPA)	
Abstractor
  F.  R.  Biancardi
                              Institution
                                     United Aircraft  Research Lab
 WR:I02 (REV. JULY 1969)
 WRSI C
                           SEND WITH COPY OF DOCUMENT TO: WATER R ESOUR C ES SC I E NT I F I C INFORMATION CENTER
                                                    U.S. DEPARTMENT OF THE INTERIOR
                                                    WASHINGTON, D. C. 20240
                                                                                SPO: 1 970-389-930

-------