WATER POLLUTION CONTROL RESEARCH SERIES •16130 ONE 03/71
Advanced Nonthermally Polluting
Gas Turbines in
Utility Applications
ENVIRONMENTAL PROTECTION AGENCY • WATER QUALITY OFFICE
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WATER POLLUTION CONTROL RESEARCH SERIES
The Water Pollution Control Research Series describes
the results and progress in the control and abatement
of 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.
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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
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EPA Review Notice
This report has been reviewed by the Water Quality
Office, EPA, and approved for publication. Approval
does not 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.
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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
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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
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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
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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
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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
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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
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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
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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
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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
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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
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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
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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
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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-
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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
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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.
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?. 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.
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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.
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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
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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
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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.
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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.
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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
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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
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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
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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-
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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
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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
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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
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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, $75 million program to solve the gas shortage. Potential output from
stimulated fields is placed at a trillion cu ft within 10 years from now and
ultimately 317 trillion cu ft. The quantity and distribution of proven and poten-
tial reserves of natural gas in various parts of the US, shown in Fig. k, highlight
the vast reserves that would be available for use in the South Central Region
from those areas denoted as D, E, F, G, and J in the Potential Gas Committee Survey.
Furthermore, there are proven reserves of 52 trillion cubic feet in Canada, and some
west coast utilities are even exploring the possibility of bringing in gas from
19
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South America (Ref. 19)- Liquified natural gas will be available for use on
the east coast, and studies are under vay to determine the costs of bringing
Alaskan gas via pipeline to the midwest or as LNG to the west coast.
In the early days of natural gas utilization after World War II, gas was sold
to pipeline customers at prices as low as 3 to 5# per Mcf. (Gas prices are often
quoted in cents per thousand cubic feet (Mcf) and gas from the well will generally
average 1075 Btu/Mcf. However, after processing, the heat content of the gas
available for sale is usually reduced to about 1000 Btu/Mcf, and this value has
been used throughout this report to avoid confusion. Consequently, prices in <£/Mcf
are equal to prices in ^/million Btu, another pricing quantity often encountered.)
Recently, there have been examples where gas has been bought in the field in the
South Central Region for more than 20$ 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 _
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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
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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
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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
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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
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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
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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
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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
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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
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(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.
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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.
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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
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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
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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
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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
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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
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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
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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
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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
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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
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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
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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.
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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
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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
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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.
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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.
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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
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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
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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
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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
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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.
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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.
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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
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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).
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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
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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
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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
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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
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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
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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
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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).
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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
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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
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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.
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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
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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
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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
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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
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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.
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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
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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
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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
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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.
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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.
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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
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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
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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.
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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
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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
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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
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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.
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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
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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.
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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.
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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.
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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.
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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.
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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).
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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
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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
$12/kw between the 1970-decade and early 1980's designs, the installed costs of
the gas turbine were reduced by approximately the same amount. In addition, the
net plant efficiency of the gas turbine station increased by more than 5 percentage
Points as described in Section VII, while the efficiency of the steam plant in-
creased by only 2 percentage points. Thus, even though gas costs are projected to
increase to the 30<£/million Btu level, the low installed costs of the gas turbine
Cation would offset the relatively high fuel and maintenance costs and still
provide a minimum-power-cost system.
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Sensitivity to Economic Factors
Although the values selected herein for capital charges, natural gas costs,
interest charges, and escalation rate are based on careful projections, a number
of outside influences such as the general economic condition could change these
factors considerably. Therefore, the sensitivity of the results presented in
Table XXVIII to variations in several important values vere considered. The
variation in busbar pover costs with changes in the capital charges and cost of gas
are indicated in Figs. 59a and 59b, respectively. The Fig. 59a results indicate
that even with a reduction in capital charges to the 10$ 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).
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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
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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.
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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*).
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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).
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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
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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.
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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.
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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.
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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
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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.
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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
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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.
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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,
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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
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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
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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
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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
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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
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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
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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
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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
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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
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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
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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
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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
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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
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. 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
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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.
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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 ' -
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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
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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.
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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
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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
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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
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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 $150,000 each. All turbine blade tooling costs were asstmed to be amortized in
five years.
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Discussions vith vendors revealed that if it vere necessary to cast tip
shrouds integrally with turbine blades , the price estimates present for solid
blades and for hollow blades would be increased by 10/S to allow for additional
casting complexity and machinery. Machining the root sections of the blades for
engine designs with varying hub diameters would add an additional 15$ 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
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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
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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.
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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.
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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*
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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
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