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Home My WebLink AboutCoal Gasification Combined-Cycle System Analysis 1980Coal Gasification Combined-Cycle System Analysis AP-1390 Research Project 986-2 Final Report, April 1980 Prepared by UNITED TECHNOLOGIES CORPORATION Power Systems Division Governor's Highway South Windsor, Connecticut 06074 Principal Investigators S. Hamilton J. Garow S. J. Lehman Prepared for Electric Power Research Institute 3412 Hillview Avenue Palo Alto, California 94304 EPRI Project Manager B. Louks Engineering and Economic Evaluation Program Advanced Power Systems Division > a. ° © > cc < c a =i Alaska Power Authority 334 W. 5th Ave. Anchorage, Alaska 99501 DO NOT REMOVE FROM OFFICE ORDERING INFORMATION Requests for copies of this report should be directed to Research Reports Center (RRC), Box 50490, Palo Alto, CA 94303, (415) 965-4081. There is no charge for reports requested by EPRI member utilities and affiliates, contributing nonmembers, U.S. utility associations, U.S. government agencies (federal, state, and local), media, and foreign organizations with which EPRI has an information exchange agreement. On request, RRC will send a catalog of EPRI reports. Copyright © 1980 Electric Power Research Institute, Inc. EPRI authorizes the reproduction and distribution of all or any portion of this report and the preparation of any derivative work based on this report, in each case on the condition that any such reproduction, distribution, and preparation shall acknowledge this report and EPRI as the source. NOTICE This report was prepared by the organization(s) named below as an account of work sponsored by the Electric Power Research Institute, Inc. (EPRI). Neither EPRI, members of EPRI, the organization(s) named below, nor any person acting on their behalf: (a) makes any warranty or representation, express or implied, with respect to the accuracy, completeness, or usefulness of the information contained in this report, or that the use of any information, apparatus, method, or process disclosed in this report may not infringe privately owned rights; or (b) assumes any liabilities with respect to the use of, or for damages resulting from the use of, any information, apparatus, method, or process disclosed in this report. Prepared by United Technologies Corporation South Windsor, Connecticut ABSTRACT This report summarizes the results of the second phase of a study involving combustion turbine power plants using coal gasification. The study concentrated on systems integration and the optimization of power plant conceptual designs. In order of importance, the objectives of the study were to: 1. Determine potential levels of thermal efficiency for well integrated gasified coal combined cycle (GCC) systems employing current technology combustion turbines and near-commercial, oxygen-blown Texaco gasifiers 2. Quantify the effects of varying key design parameters of various components on overall plant performance 3. Project potential levels of performance made possible by using advanced, combustion turbines and advanced gasifiers in well- integrated gasification power plants. To meet these objectives emphasis was placed on effective waste heat management and practicality in:synthesizing overall power plant arrangements. Evaluations were confined to thermodynamic analysis and did not include equipment cost estimation. Current technology systems were defined which yielded thermal efficiencies in the range of 35 to 37 percent. It was found that approximately one percentage point in thermal efficiency could be realized by using either the British Gas Corpora- tion (BGC) slagging, fixed-bed gasifier or the air-blown Texaco gasifier in place of the oxygen-blown Texaco gasifier. Two percentage points were gained by increasing the gas turbine combustor exit temperature about 300°F. These results suggest the possibility of approaching 40 percent in practical future power plants. Cost-effective configurations with such performance levels are expected to offer a more than competitive alternative to conventional coal-fired steam power plants equipped to meet emission standards. In addition to the performance studies, a design study was made to determine the feasibility of installing a medium BTU coal gas combustor in an FT4 engine. The combustor configuration was evaluated in a component test as part of the EPRI RP985-3 program. It was concluded that a combustor of this type could be successfully installed with only minor modifications to the engine. 1ii EPRI PERSPECTIVE PROJECT DESCRIPTION Thermal efficiencies of a number of integrated coal gasification combined-cycle configurations were studied by the United Technology Corporation (UTC) under EPRI Research Project (RP) 986-2. PROJECT OBJECTIVES In order of importance, the objectives were to: 1. Determine potential efficiency levels of configurations employing a current technology combustion turbine (the UTC FT4) and the near- commercial Texaco oxygen-blown coal gasifier 2. Quantify effects of varying key design parameters on plant performance 3. Project potential levels of performance by using advanced combustion turbines and advanced gasifiers, specifically a Texaco air-blown gasifier and a British Gas Corporation (BGC) slagging fixed-bed gasifier. PROJECT RESULTS The study clearly shows that efficiencies as high as 37.3 percent can be attained using current combustion turbines and a Texaco oxygen-blown gasifier. If either a Texaco oxygen-blown gasifier or a BGC gasifier were employed, efficiency would increase by about 1 percentage point. If it were possible to increase the combustion turbine temperature to 2300°F (compared to 1984°F for the current FT4 engine), efficiency could increase by about 2 percentage points. The results also show that wide variances in design parameters (e.g., gasifier pressure, steam conditions, etc.) have minor effects on thermal efficiencies. These results suggest that an efficiency of 40 percent is attainable with moderate advances in gasification and combustion turbine technologies. B. M. Louks, Project Manager Engineering and Economic Evaluation Program Advanced Power Systems Division ACKNOWLEDGMENTS The helpful comments and assistance of B. Louks and Dr. M. J. Gluckman of EPRI, and Dr. R. I. Strough, J. H. Lewis and F. H. Boenig of UTC are gratefully acknowledged. vii CONTENTS ix SECTION PAGE 1. INTRODUCTION id Background i=1 Phase 2 Objectives L=3' Items of Work Performed 1-3) Assumptions and Constraints <5) THERMAL EFFICIENCY OF CURRENT TECHNOLOGY SYSTEMS 2-1 System Description 27 Results of System Variation Studies 2-2: EFFECT OF CONFIGURATION AND PARAMETRIC CHANGES ON PERFORMANCE 3-1 Heat Management Methodology - Analytic Techniques 3-1 Effect of Steam Conditions on Thermal Efficiency 3-7 Effect of Recycle on Thermal Efficiency 3-12 Effect of Gasifier Pressure on Thermal Efficiency 3-14 Effect of Second Fuel Gas Reheat 3-15 Effect of Oxygen Preheat Temperatures on Thermal Efficiency 3-16 Effect of Ambient Temperature on Performance 3-16 ADVANCED TECHNOLOGY COMBUSTION TURBINE AND GASIFIERS 4-1 Effect of Higher Combustor Exit Temperature 4-1 Comparison of Air-Blown and Oxygen-Blown Gasifiers 4-2 Comparison of Texaco Entrained-Flow and British Gas Corporation (BGC) Slagging Fixed-Bed Gasifiers 4-4 DESIGN COMPATABILITY OF MEDIUM BTU COMBUSTOR WITH FT4 5-1 CONCLUSIONS 6-1 RECOMMENDATIONS 1 References 7-2 SECTION APPENDIX APPENDIX APPENDIX APPENDIX APPENDIX APPENDIX APPENDIX APPENDIX APPENDIX APPENDIX CONTENTS (continued) PAGE INITIAL STUDIES OF FT4 WITH OXYGEN-BLOWN GASIFIER A-1 EFFECT OF EFFICIENCY OF MAJOR SUBSYSTEMS ON OVERALL POWER PLANT EFFICIENCY B-1 CONFIGURATION XC47 C-1 CONFIGURATION XC48 D=1 CONFIGURATION XC46 E-1 CONFIGURATION XC45 F-1 CONFIGURATION XC45A G-1 CONFIGURATION XC53 H-1 CONFIGURATION ACO3 Ta, CONFIGURATION XC50 J-1 FIGURE LT aa oe 3-7a 3-7b 3-8 359 3-10 4-1 4-2 LIST OF ILLUSTRATIONS Various Performance Levels for GCC Power Plants with Texaco Gasifiers Computed by Various Contractors Simplified Schematic of Gasified Coal Combined Cycle Using Oxygen-Blown Texaco Gasifier Texaco Oxygen Blown Gasifier/FT4 Gas Turbine Combined Cycle Configuration No. XC45A, n Thermal = 37.3% Energy Flow in Simplified Model of Air-Blown GCC Effect of Heat Passing Directly to Bottoming Cycle, Ner = 32% Influence Coefficients of Gas Turbine and Bottoming Cycle Efficiencies on Overall Thermal Efficiency Available Work Loss Due to A T in Heat Exchanger Maximum Temperature is Not Mean Effective Temperature Loss in Efficiency Due to A T Across Fuel Gas Regenerator Temperature-Energy Diagram for Steam Bottoming, Single Pressure System Temperature-Energy Diagram for Steam Bottoming, Multipressure System Recycle Configurations Effect of Gasifier Pressure on Overall Thermal Efficiency Effect of Ambient Temperature on Power Output Texaco Oxygen Blown Gasifier/Advanced Combustion Turbine (2300°F CET) Combined Cycle Configuration No. XC53, nN Thermal = 39% BGC Slagger Gasifier/FT4 Combustion Turbine Combined Cycle Configuration No. XC50, n Thermal = 37.3% Medium BTU Combustor in FT4 Engine Energy Flow in Gasified Coal Combined Cycle xi PAGE 2a 2-5 3-3 3-4 3-5 3-6 S=7; 3-8 3-8 3-13) 3-14 3-17 4-3 4-6 S21 B-3 FIGURE C-1 C-2 D-1 E-1 Tat J-1 LIST OF ILLUSTRATIONS (continued) Texaco Oxygen Blown Station Numbers for Texaco Oxygen Blown Cycle Configuration Texaco Oxygen Blown Gasifier/FT4 Combustion Turbine Configurations XC45, 45A, Gasifier/FT4 Gas Turbine No. XC47, n Thermal = 35. Gasifier/FT4 Gas Turbine Configuration No. XC48, n Thermal = 36.1% Texaco Oxygen Blown Cycle Configuration Texaco Oxygen Blown Cycle Configuration Texaco Oxygen Blown Cycle Configuration Texaco Oxygen Blown Gasifier/FT4 Gas Turbine No. XC46, n Thermal = 36. Gasifier/FT4 Gas Turbine No. XC45, n Thermal = 36. Gasifier/FT4 Gas Turbine 46, 47, 48 Combined 3% Combined Cycle Combined 3% Combined 8% Combined No. XC45A, n Thermal = 37.3% Gasifier/Advanced Turbine (2300°F CET) Combined Cycle Configuration No. XC53, n Thermal = 39% Texaco Air Blown Gasifier/FT4 Combustion Turbine Combined Cycle Configuration No. AC03, n Thermal = 37.1% BGC Slagger Gasifier/FT4 Combustion Turbine Combined Cycle Configuration No. XC50, n Thermal = 37.3% xi PAGE Ca2 Ba2 F-2 Ho? 1-2) J-2 LIST OF TABLES TABLE PAGE 2-1 Comparison of Power Plants Using FT4 Combustion Turbine and Oxygen-Blown Texaco Coal Gasifier 2-3 3-1 Effect of Steam Pressure on Overall Efficiency with Current Combustion Turbines and Cold Fuel Gas Recycle 3-9 3=2 Effect of Steam Reheat on Efficiency With No Fuel Gas Recycle 3-10 3-3 Effect of Fuel Gas Recycle on Efficiency 3-14 3-4 Effect of Second Fuel Gas Reheat (Following Expander Turbine) at 1200 psi Gasifier Pressure 5-15 3-5 Effect of Condenser Pressure on Performance at 110°F Ambient Temperature 3-17 4-1 Effect of Advanced Combustor Exit Temperature on Power Plant Thermal Efficiency 4-1 4-2 Effect of Air-Blown and Oxygen-Blown Gasifiers 4-4 4-3 Comparison of Power Plant Performance with Texaco Oxygen-Blown Gasifier and BGC Slagging Gasifier 4-5 4-4 Comparison of Power Plant Performance with Oxygen-Blown Texaco and BGC Slagging Gasifier with Steam Reheat 4-5 A-1 Identification of Performance Differences Between Phase 1 and Initial FT4 Studies A-3 C-1 Performance Summary for Configuration XC47 c-4 c-2 Energy Balance for Configuration XC47 Datum = 59°F C-5 C-3 Temperature for Configuration XC47 C-6 Cc-4 Mass Flow for Configuration XC47 C-7 c-5 Pressure for Configuration XC47 C-8 C-6 Enthalpy for Configuration XC47 c-9 xiii TABLE C-7 c-8 D-1 E-1 E-2 E-3 E-4 E-5 E-6 E-7 E-8 F-1 LIST OF TABLES (continued) Turbomachinery Summary for Configuration XC47 Heat Exchanger Heat Loads (MMBTU/HR) for Configuration XC47 Performance Summary for Configuration XC48 Energy Balance for Configuration XC48 Datum = 59°F Temperature for Configuration XC48 Mass Flow for Configuration XC48 Pressure for Configuration XC48 Enthalpy for Configuration XC48 Turbomachinery Summary for Configuration XC48 Heat Exchanger Heat Loads (MMBTU/HR) for Configuration XC48 Performance Summary for Configuration XC46 Energy Balance for Configuration XC46 Datum = 59°F Temperature for Configuration XC46 Mass Flow for Configuration XC46 Pressure for Configuration XC46 Enthalpy for Configuration XC46 Turbomachinery Summary for Configuration XC46 Heat Exchanger Heat Loads (MMBTU/HR for Configuration XC46 Performance Summary for Configuration XC45 Energy Balance for Configuration XC45 Datum = 59°F Temperature for Configuration XC45 Mass Flow for Configuration XC45 Pressure for Configuration XC45 Enthalpy for Configuration XC45 xiv PAGE c-10 C-11 D-9 D-10 E-3 E-4 E-5 E-6 E-7 E-8 E-9 E-10 F-3 F-4 F-5 F-6 F-7 F-8 TABLE F-8 G-1 G-2 G-3 G-4 G-5 6-6 G-7 G-8 H-2 He3 H-4 H-5 H-6 HT H-8 T-T 2 r3 I-4 E-5 I-6 C7, LIST OF TABLES (continued) Turbomachinery Summary for Configuration XC45 Heat Exchanger Heat Loads (MMBTU/HR) for Configuration XC45 Performance Summary for Configuration XC45A Energy Balance for Configuration XC45A Datum = 59°F Temperature for Configuration XC45A Mass Flow for Configuration XC45A Pressure for Configuration XC45A Enthalpy for Configuration XC45A Turbomachinery Summary for Configuration XC45A Heat Exchanger Heat Loads (MMBTU/HR) for Configuration XC45A Performance Summary for Configuration XC53 Energy Balance for Configuration XC53 Datum = 59°F Temperature for Configuration XC53 Mass Flow for Configuration XC53 Pressure for Configuration XC53 Enthalpy for Configuration XC53 Turbomachinery Summary for Configuration XC53 Heat Exchanger Heat Loads (MMBTU/HR) for Configuration XC53 Performance Summary for Configuration AC03 Energy Balance for Configuration ACO3 Datum = 59°F Temperature for Configuration AC03 Mass Flow for Configuration AC03 Pressure for Configuration AC03 Enthalpy for Configuration AC03 Turbomachinery Summary for Configuration AC03 Page F-9 F-10 G-3 G-4 Gan) G-6 G-7 G-9 G-10 H-10 LIST OF TABLES (continued) TABLE PAGE I-8 Heat Exchanger Heat Loads (MMBTU/HR) for Configuration AC03 I-10 J-1 Performance Summary for Configuration XC50 J-3 J-2 Energy Balance for Configuration XC50 Datum = 59°F J-4 J-3 Temperature for Configuration XC50 J-5 J-4 Mass Flow for Configuration XC50 J-6 J-5 Pressure for Configuration XC50 J-7 J-6 Enthalpy for Configuration XC50 J-8 J-7 Turbomachinery Summary for Configuration XC50 J-9 J-8 Heat Exchanger Heat Loads (MMBTU/HR) for Configuration XC50 J-10 xvi SUMMARY This report summarizes the results of the second phase of an EPRI sponsored study of combustion turbine power plants using coal gasification. In the Phase 1 study, (Ref. 1), it was concluded that a well-integrated gas- ified-coal, combined cycle power plant could attain an overall efficiency of 39% with an air-blown Texaco gasifier and a gas turbine combustor exit temperature of 2000°F. The 2000-degree temperature level represents current state-of-the-art. A 39% efficiency level would be more than competitive in efficiency with current, coal-fired steam power plants equipped to meet emission standards. However, to attain that efficiency would require the development of several major components including the air-blown Texaco gasifier and a heat exchanger to preheat gasifier feed air to 1000°F. Consequently, Phase 2 was undertaken to ascertain whether a similar high efficiency could be attained in an optimized gasified-coal power plant system using the FT4 current production combustion turbine and the near-current oxygen-blown Texaco gasifier. A simplified schematic of a gasified coal combined cycle (GCC) using an oxygen-blown Texaco gasifier is shown in Figure S-1. Sec- ondary objectives of the Phase 2 study were to: (1) quantify the effects of varying design parameters on plant performance when using an oxygen-blown Texaco gasifier and current technology combustion turbines, and (2) investigate the potential gain in efficiency achievable with advanced technology combustion turbines or advanced technology gasifiers such as the British Gas Corporation (BGC) fixed-bed slagger or the air-blown Texaco gasifier. EFFICIENCY OF CURRENT OR NEAR-CURRENT TECHNOLOGY At the beginning of Phase 2, a detailed performance study of a power plant config- ured with an oxygen-blown Texaco gasifier and the FT4 combustion turbine was made in collaboration with Fluor Engineers and Constructors at the request of EPRI. Thermal efficiency of that system was estimated to be only 33 to 34%. Because this result was far below the 39% efficiency estimated in Phase 1, it became important to isolate and quantify those effects causing the difference in performance, and to S-1 find the system modifications necessary to raise the efficiency level. Subsequent analysis identified differences which taken together add up to the total difference between the Phase 1 results and the early Phase 2 results. After considering which energy losses were inherent and which could be circumvented, it was concluded that a thermal efficiency of 36 to 37% appears to be a reasonable goal for a near- current technology power plant (i.e. oxygen-blown Texaco gasifier and FT4 combus- tion turbine). WATER CLEAN GAS A, A, COAL RESATURATOR SLURRY CLEAN ee oc! RAW GAS COOLERS COMPRESSOR SULFUR >—)+——_ GENERATOR TO ECONOMIZERS HOT CLEAN FUEL GAS HEAT RECOVERY —— EXHAUST STEAM GENERATOR (HRSG) Figure S-1 Simplified Schematic of Gasified Coal Combined Cycle Using Oxygen-Blown Texaco Gasifier EFFECTS OF VARYING DESIGN PARAMETERS In the course of Phase 2, several configurations were studied which yielded thermal efficiencies in the range of 35 to 37%. The best of these required a gasifier pressure of 1200 psi with 2400°F untempered fuel gas entering a raw gas cooling boiler. Such conditions may be too severe from the standpoint of heat exchanger durability. But a more feasible plant could be designed with only a small perfor- mance penalty by either reducing the gasifier pressure to 600 psi or tempering the hot, raw fuel gas to 1800°F by mixing it with cooled recycle gas. Even with this performance penalty, the resulting configurations would be more than competitive in efficiency with a conventional coal-fired steam power plant equipped to meet emis- sion standards. Performance values for these two cases are shown in Table S-1. Diagrams of these configurations are shown in the Appendices. Although the oxygen- blown Texaco gasifier is still undergoing development, it is referred to here as "current technology" to distinguish it from the air-blown Texaco gasifier upon which less development effort is being placed. Table S-1 PERFORMANCE OF NEAR-CURRENT TECHNOLOGY CONFIGURATIONS CONFIGURATION PERFORMANCE DATA XC45A xC46 Overall Thermal efficiency, % 37.3 36.3 Fuel Gas Temperature Entering Boiler, °F 2400 1800 Temperature of Recycled Gas, °F No Recycle 350 Steam Pressure, psi 1450 1450 Steam Temperature, °F 1000 1000 (No Reheat) Fuel Gas Reheat Temperature, °F 1000 1000 Gasifier Pressure, psi 1200 588 Oxidant Oxygen Oxygen EFFECT OF VARIATION IN STEAM CONDITIONS Consistent with a variation in heat exchanger technology, steam conditions were varied as shown on Table S-2. Component arrangement is similar to that shown in Figure S-1. Table S-2 EFFECT OF STEAM CONDITIONS ON GCC PERFORMANCE CONFIGURATION TTT XC47 XC46 XC54 XC54A Overall Thermal Efficiency, % 35.3. 36.3 36.6 36.7 Fuel Gas Temperature Entering Boiler, °F 1800 1800 1800 1800 Temperature of Recycled Gas, °F 350 355 355 355 Steam Pressure, psi 800 1450 1450 2400 Superheat Temperature, °F 800 1000 1000 1000 Steam Reheat Temperature, °F No Reheat No Reheat 1000 1000 Fuel Gas Reheat Temperature, °F 800 1000 1000 1000 Gasifier Pressure, psi 588 588 588 588 8-3 The first column (configuration XC47) represents a conservative design based on placing the superheater downstream of the FT4. The second column (Configuration XC46) represents a more advanced design based on placement of a superheater in the raw gas cooling circuit, downstream of a raw gas cooling boiler. The third column (XC54) shows a 0.3 percentage point benefit for steam reheat. In this configura- tion, 1788°F raw fuel gas enters the superheater. The fourth column (XC54A) is the same arrangement with a 2400/1000/1000 steam system for slightly increased effici- ency. EFFICIENCY WITH ADVANCED GASIFIERS AND INCREASED COMBUSTOR EXIT TEMPERATURE Performance was estimated for various gasifiers and combustion turbine combustor exit temperatures. (Table S-3). With current technology combustion turbines, 0.7 to 1.2 percentage points in thermal efficiency can be gained by using the British Gas Corporation (BGC) slagging, fixed-bed gasifier in place of the Texaco oxygen- blown gasifier. A gain of up to 2.2 percentage points results from increasing the combustor exit temperature about 300°F. Table S-3 PERFORMANCE COMPARISON NEAR CURRENT AND ADVANCED CONFIGURATIONS PERFORMANCE DATA XC54 XC55 XC56 XC57 Combustor Exit 1984 1984 1984 2300 Temperature, °F Gasifier Type Texaco Texaco BGC Texaco Oxidant Oxygen Oxygen Oxygen Oxygen Overall Thermal 36.6 3721 37.8 38.8 Efficiency, % Steam Pressure 1450 1450 1450 1450 Steam Temperature, °F SH/RH 1000/1000 1000/1000 1000/1000 1000/1000 Raw Gas Temp. Entering 1800 2400 282 1800 Heat Exchanger, °F Temperature of Recycled 355 No Recycle No Recycle 361 Gas, Fuel Gas Reheat 1000 1000 §55 1000 Temperature, °F S-4 Calculations were also made for an air-blown Texaco gasifier compared to an oxy- gen-blown Texaco gasifier. This comparison is shown on Table S-4. TABLE S-4 PERFORMANCE COMPARISON OF OXYGEN AND AIR-BLOWN GASIFIERS CONFIGURATION XC54B ACO3 Combustion Exit Temperature, °F 1984 1984 Gasifier Texaco Texaco Oxidant Oxygen Air Overall Thermal Efficiency 36.0 Sian Steam Pressure - psia 1450 1450 Superheat Temperature - °F 800 800 Reheat Temperature - °F 800 800 Temperature of Recycled Fuel Gas - °F 309) 362 Fuel Gas Reheat Temperature - °F 800 800 A design study was made to determine the feasibility of installing a medium BTU coal gas combustor in a current FT4 combustion turbine. The combustor configura- tion used in the study was tested as part of EPRI RP985-3 program. It was con- cluded that this combustor could be successfully installed with making only minor modifications to the engine. However, a new fuel control would be necessary. SUMMARY OF RESULTS Combustion Turbine Overall thermal efficiency increases 2.2 percentage points when combustor inlet temperature is increased from 1984°F to 2300°F. Gasifier Type Overall thermal efficiency with air-blown Texaco is 1.1 percentage points higher than oxygen-blown Texaco. Overall efficiency with oxygen-blown BGC gasifier is 0.7 to 1.2 percentage points higher than oxygen-blown Texaco. Steam Pressure (without Reheat) Increasing steam pressure from 800 to 1450 psi increases thermal efficiency by 0.8 to 1.0 percentage points. Further increases require reheat. Reheat At 1450 psi and 1000°F superheat temperature, reheat increases thermal efficiency by 0.3 percentage points. Steam Pressure (with Reheat) With reheat, raising steam pressure from 1450 psi to 1800 psi increases effiency 0.2 points. Further increases are ineffective at the gas temperatures assumed. Steam Temperature With a 1450 psia reheat system increasing both the superheat and reheat tempera- tures by 200°F increases thermal efficiency by 0.6 percentage points. Roughly half is due to each component. Gas Temperature Entering Raw Gas Cooler Recycling cold fuel gas to temper the raw gas temperature from 2400°F to 1800°F causes a 0.5 percentage point drop in thermal efficiency. Expansion Turbine and Gasifier Pressure A system with 600 psia gasifier pressure and a fuel gas expander turbine has 0.7 percentage points higher efficiency than a 300 psia gasifier with no expander turbine. Low Pressure Boiler Pinch Point Pinch point in the low pressure boiler influences stack temperature. A pinch differential of 70° gives a 300°F stack temperature. Reducing the pinch to 40° reduces the stack temperature correspondingly, and increases thermal efficiency by 0.4 percentage points. Ambient Temperature Increasing ambient temperature decreases power about 1% for each 2°F temperature rise. The effect on thermal efficiency is rather small. S-6 Oxygen Preheat Temperatures No significant effect. Second Fuel Gas Reheat Reheating fuel gas from 600°F to 1000°F following the expander turbine has no significant effect on performance. CONCLUSIONS As a result of the Phase 2 analysis, it was concluded that a gasified-coal, com- bined cycle power plant based on current technology that is either available or near final development could be competitive in thermal efficiency with modern steam plants. Advanced combustion turbine or gasifier technology is not necessary to make the gasified coal combined cycle competitive. However, advanced technology will improve efficiency, save fuel, and make future systems even more attractive. S-7 Section 1 INTRODUCTION BACKGROUND Over the past four years, EPRI has conducted a substantial effort to determine the performance characteristics of integrated gasification - combined cycle (GCC) power systems. Initial emphasis in the studies conducted was on the Texaco gasification system integrated with advanced (2400°F combustor exit temperature) combustion turbine combined cycle power plants. Subsequent studies conducted for EPRI by the three major U.S. combustion turbine companies indicate that integrated GCC power plants employing the Texaco gasifier and current technology combustion turbines operating in the vicinity of 2000°F combustor exit temperature have the potential for overall system efficiencies of 39% (Ref. 1). These studies led to the con- clusion that gasified coal combined cycles (GCC) using current technology com- bustion turbines could compete with conventional coal-fired steam plants equipped to meet emission standards. This information contradicted earlier results which indicated performance levels for Texaco based GCC systems employing current technology combustion turbines to be in the range of 30% to 34%. Such results are summarized graphically in Figure 1-1, which shows overall Texaco based GCC plant efficiency as a function of turbine firing temperature. The wide range of estimated power plant efficiency at 2000°F turbine firing temperature is confusing and indicates that a better understanding of the effects of GCC plant configuration on system efficiency is essential to ensure optimum design of such power plants. The data presented in Figure 1-1 point out another confusing result. The top line on the plot representing the most efficient designs and the bottom line repre- senting the least efficient designs indicate a markedly different influence of turbine inlet temperature on overall system efficiency. It is important to estab- lish the correct trend for the following reasons: i? LIBRARY COPY Alaska Power Authority 334 W. 5th Ave. Anchorage, Alaska 99501 DO NOT REMOVE FROM OFFICE ° To determine whether or not the development of high temperature com- bustion turbines is necessary for the construction of high efficiency GCC power systems. e To evaluate the incentives for the development of high temperature com- bustion turbines for integration with gasification plants. ae 7A 42 a or ie v 40+ Laer i v v OVERALL | -« aed THERMAL 3, ed EFFICIENCY 7 PERCENT —36+- 7o ye 34+ / Qo 7 oO 7 32h le 7 Os SOU a eg 1900 2000 2100 2200 2300 2400 2500 2600 FIRING TEMPERATURE - °F Figure 1-1 Various Performance Levels for GCC Power Plants with Texaco Gasifiers Computed by Various Contractors The issue was clouded further by a study conducted in the first weeks of Phase 2. The study was to determine the performance of a configuration which could be built at low risk using current or near-current equipment. The powerplant studied was based on a production FT4 combustion turbine engine manufactured by United Tech- nologies Corporation and an entrained-flow oxygen-blown coal gasifier now being developed by Texaco. Performance of several variants ranged between 32% and 34% overall thermal efficiency which is only marginally acceptable. Subsequent studies were made to identify those factors causing the difference between the approximate 33% efficiency and the 39% for an air-blown system shown in the earlier study. These performance differences are summarized in Appendix A together with a summary description of the system. The difference was found to be due to differences in gas turbine and steam cycle designs and in the losses assumed. Some of the losses are inherent in a real system; other losses can be circumvented. Considering which losses are inherent and which can be circumvented, a goal of 36 to 37% thermal efficiency appreared to be reasonable for the subsequent Phase 2 studies. PHASE 2 OBJECTIVES The major objectives of the Phase 2 effort, in order of importance, were: e To determine whether or not it is possible to configure oxygen-blown Texaco-based GCC power systems employing current-technology combustion turbines such that the overall system thermal efficiency (based on coal HHV) would fall in the range 35% to 40%, and to determine the maximum thermal efficiency of Texaco based GCC systems employing current tech- nology combustion turbines. e To determine and quantify the influence of GCC system configuration on overall plant performance. It is important to understand why certain design considerations affect system performance to a greater degree than others so that future designs of GCC based power systems can be opti- mized. For example, it is important for the system design engineer to understand how to apportion economizing, boiling, superheating and reheating duties between the gas coolers and the heat recovery steam generator (HRSG) downstream of the combustion turbine to obtain optimum plant performance. The influence of steam system integration and con- figuration is a key element affecting the ultimate performance of GCC power plants. e To determine the effect of advanced high-temperature combustion turbine technology and advanced gasifiers on the performance of integrated GCC power plants. When this objective has been met, it will be possible to evaluate the incentives for developing high-temperature combustion tur- bines and advanced gasifiers for GCC based power systems. The primary basis for these studies was the oxygen-blown Texaco coal gasification process. A configuration employing the air-blown Texaco gasifier and a configu- ration employing the BGC gasifier also were investigated for the purpose of meeting the third objective. ITEMS OF WORK PERFORMED To meet the first objective, a series of power plant configurations utilizing current technology combustion turbines and other components and the Texaco oxygen-blown gasifier were analyzed to establish the efficiency levels attainable. This work is reported in Section 2. pis) The second objective of determining the influence of GCC system configuration and parameters on overall power plant performance was accomplished by examining the effects of a number of design variables on the performance of the GCC power plant. These variables include: e The split of economizing, boiling, reheating and superheating duties between the gas coolers and the HRSG's. e Steam cycle conditions such as throttle pressure, superheat and reheat temperatures, pinch point differentials, intermediate pressure and non- reheat systems. e The effect of raw fuel gas temperature and quantity entering the gas cooler. The temperature of the raw gas was moderated by employing a recycle of cooled gas as was done in the Phase 1 studies. e The effect of having an expansion turbine between the clean, reheated high pressure fuel gas and the gas turbine combustor. e The effects of gasification system operating pressure on overall system performance. e The effects of ambient temperature. e Supplemental firing of clean fuel gas for superheating or production of additional high pressure steam. e Effect of clean fuel gas temperature to the gas turbine combustor. e The effect of oxygen temperature to the Texaco gasifier. This work is reported in Section 3. The final objective of determining the effect of higher temperature combustion turbine technology on the performance of GCC powerplants was attained by con- figuring and calculating the performance of a power plant based on a gas turbine with a 2300°F combustor exit temperature. In addition, the performance improvement offered by advanced gasifiers was evaluated. This work is reported in Section 4. The design problems of adapting existing combustion turbine equipment for use in gasification-combined cycle power systems were also explored. This work is report- ed in Section 5. 1-4 ASSUMPTIONS AND CONSTRAINTS General design conditions and constraints applied to all base case designs were as follows: e Performance estimates were determined using ambient conditions of 59°F, 14.7 psia and 0.0% relative humidity (ISO conditions). e Stack temperatures lower than 250°F were not considered for any case. e Direct transfer of heat between reducing and oxidizing gases was not permitted. e Temperature of clean fuel gas to either the combustion turbine or the expander turbine did not exceed 1000°F. e Temperature of oxygen between the air separation plant and the Texaco gasifier did not exceed 300°F. e Maximum steam temperature did not exceed 1000°F. e Maximum steam throttle pressure did not exceed 2400 psia. e Overall system energy balances must close to within 1/2%. e A condenser pressure of 2 1/2 inches Hg absolute was used for all base cases. e Except for the difference in ambient conditions, cooling tower operation (i.e. water losses, temperature differences, etc.) were consistent with those reported in EPRI Report AF-642. (Ref. 2) e Minimum temperature approach conditions in major heat exchangers was not less than 40°F. Section 2 THERMAL EFFICIENCY OF SYSTEMS USING CURRENT AND NEAR CURRENT TECHNOLOGY A series of configurations were investigated with overall thermal efficiency rang- ing from 35.3% to 37.3%. ied, they all fit the same general description. configuration with a 35.3% efficiency and the improvements required to bring the efficiency level of this system to 37.3%. SYSTEM DESCRIPTION A simplified schematic of a gasified coal combined cycle (GCC) using an oxygen- blown Texaco gasifier is shown in Figure 2-1. AIR IN WATER CLEAN GAS A, OXYGEN PLANT COMBUSTION TURBINE [4 RAW GAS COOLERS TO ECONOMIZERS HOT CLEAN FUEL GAS HEAT RECOVERY —— EXHAUST STEAM GENERATOR (HRSG) GENERATOR Figure 2-1 Simplified Schematic of Gasified Coal Combined Cycle Using Oxygen-Blown Texaco Gasifier 2-1 Although there were differences between the systems stud- This section presents the base SULFUR Fuel Gas Stream Coal and water at 140°F is partially combusted with 300°F oxygen to yield a 2400°F gas mixture. The hot gas is successively cooled by generating, and in some cases, superheating high pressure steam, reheating the fuel gas, generating low pressure steam, scrubbing with water, cooling in the fuel gas resaturator and finally, heating feed water. In some cases, the raw 2400°F fuel gas is mixed with cool fuel gas to obtain an 1800°F mixture before it enters the first heat exchanger. (Ap- proximately 10 to 13% of the hot fuel gas is carried along with the slag and quenched with water. This represents an inevitable loss in system efficiency.) Hydrogen sulfide is removed from the cooled fuel gas through Allied Chemical's Selexol® process. The HS is converted to elemental sulfur in the Claus plant. Fuel gas is reheated first in the resaturator which adds water vapor as well as sensible heat to 300°F. The gas is then reheated to 1000°F in the raw gas cooler, expanded through the fuel gas expander turbine and delivered to the FT4 combustor. Combustion Turbine Stream Air enters the compressor and is compressed to 13 to 14 atmospheres. It then enters the combustor to be heated by combustion of the fuel gas to a temperature of 1984°F. The combustion products are expanded through a gas turbine. The turbine exhaust stream at 908°F is used to generate and superheat steam, heat water in the high pressure economizer, and generate low pressure steam. Following the low Pressure boiler is a low pressure economizer and finally a deaerating boiler. For these studies, a combustor exit temperature of 1984°F was selected which corres- ponds to the peak load rating for the FT4. While this rating is not used for extended service at the present time, it does represent current technology. Oper- ation at the Electrical Base Load rating (60 degree lower combustor exit tempera- ture) would cause about one-half percentage point decrease in overall thermal efficiency. This difference is well within the spread obtainable by variations in the steam system. RESULTS OF SYSTEM VARIATION STUDIES Variations were made to the general system described above. Six of the most significant variations are shown in Table 2-1, together with the resultant overall thermal efficiencies. Table 2-1 COMPARISON OF POWER PLANTS USING FT4 COMBUSTION TURBINE AND OXYGEN-BLOWN TEXACO COAL GASIFIER CONFIGURATION FRRFORMANCE DATA | | lxce7 XC48 XC46 XC54 XC45 XC45A Thermal Efficiency, % 35.3 36.1 36.3 36.6 36.8 37.3 Steam Pressure, psi 800 1450 1450 1450 1450 1450 Steam Temp. ,°F* 800 1000 1000 1000/1000RH 1000 1000 Superheater HRSG HRSG Location Only supplem. HRSG and RAW GAS COOLER combust. Fuel Gas Temp. ,°F 800 800 1000 1000 1000 1000 Temp. into Raw Gas Cooler, °F 1800 1800 1800 1800 2400 2400 Gas Recycle Temp., °F 360 360 350 355 No Recycle No Recycle Gasifier Pressure, psia 588 588 588 588 588 1200 Stack Temperature, °F 250 250 263 275 263 257 * Steam is not reheated except as indicated The first column, Configuration XC47, represents the lowest level of required tech- nology and the lowest performance. For this case, it was assumed that the boiler in the gas stream is not capable of withstanding the 2400°F raw fuel gas effluent from the gasifier. The raw fuel gas is accordingly tempered to 1800°F by mixing it with gas which has been cooled and scrubbed. It is further assumed that (1) the raw fuel gas is too corrosive because of its H2S content for superheating steam and reheating fuel gas to 1000°F and (2) that 800°F is the maximum allowable fuel gas reheat temperature. Accordingly, fuel gas is reheated to 800°F in the raw gas cooler. The steam superheater is, therefore, placed in the gas turbine exhaust. Steam temperature is limited to 800°F by the combustion turbine exhaust temperature and steam pressure is limited to 800 psi for acceptable steam quality at the low pressure end of the steam turbine (steam turbine exhaust moisture is discussed in Section 3. In Configuration XC48, (second column) a supplemental combustor is used in part of the combustion turbine exhaust so that the steam conditions can be raised to 1450 psi, 1000°F, and as shown on Table 2-1, efficiency increases from 35,3% to 36.1%. In configuration XC46 it was assumed that by selecting proper superheater materials (e.g. 300 series stainless steel), the steam and fuel gas could be heated to 1000°F 2-3 by the raw fuel gas stream instead of by supplemental firing in the combustion turbine exhaust. However, efficiency increases only slightly indicating that a more expensive raw gas cooler may not be cost effective. In configuration XC45, a further improvement in efficiency is made by eliminating gas recycle on the assumption that the boiler can be designed to withstand 2400°F gas inlet temperature. Such a boiler taking raw fuel gas at 2400°F is used at Oberhausen, Germany (Ref. 3) In Configuration XC45A, the gasifier pressure is raised to 1200 psia to yield an increase in the power of the expander turbine, resulting in an efficiency of 37.3%. A schematic of Configuration XC45A with temperature and flow at significant sta- tions is shown in Figure 2-2. Schematics for Configurations XC45, XC45A, XC46, XC47 and XC48 showing temperature and flow at significant stations are given in Appendices C through J. With overall coal-to-busbar thermal efficiency in the range of 35 to 37%, it is concluded that the gasified coal combined cycle can be competitive in efficiency with current coal-fired power plants equipped to meet emission standards. 0# (NO RECYCLE) RECYCLE COMP. WATER eee: =| QUENCH | aay 1596F | 1081F Lan 450F ce 600F H 1 + WATER OXYGEN GS-2 AIR 59OF 61108 —— 105F Sf FLASH GAS <I) ee) meee CONFIGS xcas o CONDENSATE 46 PUMP is SULFUR RECOVERY CONDENSER SULFUR LEGEND # FLOWIN LBS/SEC F TEMPERATURE — °F P PRESSURE IN PSIA Figure 2-2 Texaco Oxygen Blown Gasifier/FT4 Gas Turbine Combined Cycle Configuration No. XC45A, 7 Thermal = 37.3% Section 3 EFFECT OF CONFIGURATION AND PARAMETRIC CHANGES ON PERFORMANCE This section includes discussions of heat management methodologies, the effect of steam conditions, gas recycle and other significant variables which influence the performance of current technology systems. The effects of advanced gasifiers and advanced combustor firing temperatures are discussed in Section 4. To evolve configurations meeting program goals, a number of configurations with various parametric conditions were analyzed using SOAPP, United's computer program for system performance. HEAT MANAGEMENT METHODOLOGY - ANALYTIC TECHNIQUES As an adjunct to the computer programs, insight into the factors leading to ef- ficiency changes can be gained by examining simplified system models and by con- sidering the thermodynamic effects of system variables. The following section deals first with the effect of major subsystems on overall thermal efficiency. This analysis points out the importance of steam system efficiency. A second analysis treats the effect on efficiency of temperature differentials across heat exchangers. This effect is particularly important in the steam system where pres- sures impact the temperature differentials across the heat exchangers which in turn impact overall thermal efficiency. Effect of Efficiency of Major Subsystems on Overall Power Plant Efficiency. Figure 3-1 shows the energy flows for a simplified model of an air-blown system. For this model, overall efficiency based on Lower Heating Value of the fuel is given by: Egy = (1 - b) Egy (1 - Egop) + aor @ where & it} overall efficiency based on lower heating value of coal. LHV b = fraction of energy bypassing combustion turbine and flowing as gasifier thermal output directly to the bottoming cycle. Eor = combustion turbine efficiency Egor = bottoming cycle efficiency COAL LHV = 1.0 FUEL GAS | COMBUSTION GASIFIER TURBINE -—> POWER Ect HEAT BYPASSING GT,b COMBUSTION TURBINE EXHAUST HEAT BOTTOMING CYCLE LATENT HEAT OF WATER Esor r— POWER IN COMBUSTION PRODUCTS | Qres ELuv =(1- b) Et (1 - Egor) + Egor Figure 3-1 Energy Flow in Simplified Model of Air-Blown GCC Efficiency as calculated from equation (1) is plotted in Figure 3-2 against b, the proportion of coal energy bypassing the combustion turbine, for a combustion tur- bine efficiency of 32%, and steam bottoming cycle efficiency of 20%. A typical bottoming cycle is about 20% efficient. A typical conventional steam plant is about 33% efficient. Because of the high furnace temperature in such a plant, a regenerative high pressure reheat steam cycle can be used. At constant bottoming cycle efficiency, Figure 3-2 illustrates the deleterious effect on efficiency of releasing heat in the gasifier which can be recovered only at bottoming cycle efficiency rather than at overall combined cycle efficiency. However, if this heat can be used at high temperature to improve the steam conditions the bottoming cycle and overall efficiency may be improved. The dotted line on Figure 3-2 represents such a condition in which the steam bottoming cycle efficiency is improved by high 1) on this " temperature heat bypassing the combustion turbine. The end point (b line represents a typical 33% efficient free-standing steam plant hot combined with a gas turbine and not limited by the low temperature exhaust of the combustion turbine. 3-2 EFFICIENT STEAM CYCLE OVERALL CYCLE EFFICIENCY PERCENT 20 TYPICAL BOTTOMING CYCLE 10 i 1 all J 0 0.25 0.5 0.75 1.0 b, PROPORTION CF ENERGY BYPASSING COMBUSTION TURBINE Figure 3-2 Effect of Heat Passing Directly to Bottoming Cycle, ngt = 32% At constant bottoming cycle efficiency, it is desirable to minimize the amount of gasifier thermal energy bypassing the combustion turbine and flowing directly to the bottoming cycle. But reducing the amount of energy which flows directly to the bottoming cycle may conflict with improving the bottoming cycle efficiency by superheating and reheating steam. Improvement in the bottoming cycle efficiency is generally more important. Whether the overall efficiency increases or decreases depends on the details of how the cycle is configured. Partial differentiation of equation (1) yields influence coefficients: AE. (2) LHV aaa = (l = bd) 1G - FE) BE or BOT AE (3) LHV = (1 -b) (1 - E,) BE por GT AE. (4) LHV _ a = Fer © ~ Egor? Coefficients (2) and (3) are plotted against b, the fraction of sensible heat in the gasifier output stream bypassing the combustion turbine in Figure 3-3 for typical values of Ror & Eor: The value of b is about 0.11 for an oxygen-blown Texaco gasifier, essentially zero for the BGC gasifier, and 0.24 for an air-blown Texaco gasifier. Depending on the value of b, it can be seen that it may be more important to improve the bottoming cycle efficiency than the gas turbine effi- ciency. Figure 3-3 shows in magnitude how the influence coefficients respond to increasing the sensible heat in the gasifier output stream as more heat flows directly to the bottoming cycle. The effect of bottoming cycle efficiency accordingly increases in direct proportion to the power output of the bottoming cycle. 1.0 npor = 02 AE LaV AEBoT 08 INFLUENCE COEFFICIENTS 0.6 1 i = 1 04 0.1 0.2 0.3 0.4 0.5 PROPORTION OF FUEL ENERGY BYPASSING COMBUSTION TURBINE AND GOING DIRECTLY TO THE BOTTOMING CYCLE, b Figure 3-3 Influence Coefficients of Gas Turbine and Bottoming Cycle Efficiencies on Overall Thermal Efficiency A more complete model (Appendix B) includes the oxygen plant compressor and the effect of the small Brayton cycle consisting of the oxidant boost compressor and the fuel gas expander. This model indicates a benefit from higher gasifier pres- sure, higher fuel gas reheat temperature, and reduced oxygen requirements. Identification of Component Heat Exchangers Causing Loss in Overall Efficiency In systems as complicated as gasified coal combined cycles, interaction between various parts of the system can sometimes render performance plots against in- dividual variables almost meaningless. It is not, in general, possible to hold all 3-4 other variables constant while one variable is changed. While there tends to be a grouping of several good configurations having nearly equal performance, plots may show scatter in performance. The reasons for this scatter do not become clear until the configurations are examined in detail. When this is done, it may be found for example, that a poor configuration has heat exchangers which have a low pinch differential at one end, but a high temperature difference between the hot and cold stream at the other ena. In that event, the configuration should be modified by changing the flow in one side or the other of the heat exchanger, i.e. providing alternate parallel flow paths where necessary, so that the temperature differentials can be minimized. This effect may be more powerful than simply a change in steam pressure or pinch differential. Efficiency of actual combined cycle power plants is significantly below the Carnot efficiency associated with the maximum and minimum temperatures involved. Part of this reduction in efficiency is caused by the AT required to transfer heat. This AT reduces the mean temperature at which heat is transferred into the working fluid (Tc on Figure 3-4). Efficiency based on Tc is significantly below the Carnot efficiency based on Tyax: Figure 3-4 shows the available work with actual heat exchangers compared with the available work when ideal heat exchangers with no temperature differential between hot and cold sides is assumed. (The derivation is shown on Figure 3-5.) The loss in available work is directly proportional to the mean temperature differential between hot and cold sides of the heat exchanger. Twax ps I I I i HEAT EXCHANGER SW le ee a OTRANSFERRED Ty Te Figure 3-4 Available Work Loss Due to aT in Heat Exchanger 3-5 WORK ea TRANSFERRED ' T sink I; ( {i Tam at 1 T sink T sink = = Je =| T, - T2) T,-T, Ty gn T, Figure 3-5 Maximum Temperature Is Not Mean Effective Temperature Taking into account the proportion of the fuel input energy which is transferred in a given heat exchanger, the loss in overall efficiency may be expressed as: tf. 7 Q (5) AE = -e X ee : = X Tsink X aaa H ic Fuel Input where AE = Change in efficiency. (.01 is one percentage point) e = Engine efficiency. (0.8 is a representative value for steam systems and depends primarily on turbine efficiency) Ty = Log mean temperature in the heat exchanger hot stream Ty, = Log mean temperature in the heat exchanger cold stream Tonk = Minimum cycle temperature e.g. 560 R Qa RANSFERRED = Heat transferred per unit time QeuEL INPUT = Fuel input energy per unit time All temperatures are absolute temperatures. (Ty-T) incidentally is not equal to the log mean AT used in heat exchanger calculations. Figure 3-6 presents an ex- ample using Equation (5). The fuel gas regenerator causes a 0.5 percentage-point loss in efficiency. (LMAT = 300) SINCE 5% OF COAL’S ENERGY TRANSFERRED i 0.8 x AT X Tsinx x 0.05 Ty X Te AE = 0.005 Figure 3-6 Loss in Efficiency Due to aT Across Fuel Gas Regenerator EFFECT OF STEAM CONDITIONS ON THERMAL EFFICIENCY The steam cycle furnishes a good example of the use of equation (5). A steam cycle does not fit perfectly under the temperature/enthalphy diagram of the exhaust gas. (Figure 3-7a). Even though the pinch differentials may be very low, a very high temperature differential exists between the water and the hot gas at the hot end of the boiler. This condition gives rise to a loss in thermal efficiency as shown by equation (5). This loss may be reduced through the use of a multi-pressure steam system. Figure 3-7b illustrates the temperature-energy diagram of a multipressure system. By dividing the system into three boilers, it is possible to maintain a lower mean AT between the hot gas and the steam and also to recover more heat from the exhaust gas. Such systems are used by Turbo Power and Marine Systems Division of United Technologies and are described in Reference 4. ond WORKING FLUID TEMPERATURES =F EXHAUST GAS BOTTOMING FLUID 0 700 HEAT ABSORBED — PERCENT Figure 3-7a Temperature-Energy Diagram for Steam Bottoming, Single Pressure System WORKING FLUID TEMPERATURE °F EXHAUST GAS SUPERHEATER HIGH PRESSURE BOILER HIGH PRESSURE ECONOMIZER LOW PRESSURE BOILER LOW LEVEL ECONOMIZER DEAERATING BOILER 0 700 HEAT ABSORBED — PERCENT Figure 3-7b Temperature-Energy Diagram for Steam Bottoming, Multipressure System 3-8 A roughly equal division of temperature drop between the condenser, the boiler and the superheater outlet will maximize efficiency. In some instances, it may be desirable to use parallel flow paths to obtain a more uniform AT from the hot to cold side of a heat exchanger and thereby, reduce the (Ty-T,)- In any event, after computer runs are made, the temperature across each heat exchanger should be in- spected to determine whether or not the system should be reconfigured to obtain a lower (Ty-T,)- When complicated systems are configured, it is very possible that some of the heat exchangers will have excessive values of (Ty-T,) leading to loss in overall system performance. This effect is probably the major reason for the high scatter shown in Figure 1-1. Effect of Steam Pressure, Steam Temperature, and Steam Reheat on Overall Thermal Efficiency Optimum steam conditions depend on the temperature of the exhaust gas. With a higher gas temperature, there is more incentive to increase steam pressure. The conditions listed in Table 3-1 are for gas temperatures of current gas turbine engines. In addition to the high pressure boiler with its varying pressure, each of the various systems uses a 115 psia low pressure boiler for steam injection into the turbine and a 30 psia boiler for deaerating. Temperature differentials across heat exchangers are held to consistent levels in all instances. TABLE 3-1 EFFECT OF STEAM CONDITIONS ON OVERALL EFFICIENCY WITH CURRENT COMBUSTION TURBINE AND COLD FUEL GAS RECYCLE CONFIGURATION XC47 XC54B XC48 XC46 XC54 XC54A Steam Pressure, psia 800 1450 1450 1450 1450 2400 Superheat Temperature, °F 800 800 1000 = 1000 1000 1000 Reheat Temperature, °F None 800 None None 1000 1000 Gas Temp into SH or RH, °F O12 O12/ L568 lt OOM TALS 1788 1790 Stack Temperature, °F 250 267 250 263 276 276 Turbine Exit Moisture, % 14.4 10.5 TSa6 ales eels Ha 10.0 % of HP Steam Generated in 39 17s 48 25 * * Gas Cooler Overall Thermal Efficiency, % BD.S) 36.0 36.1 86.5 36.6 36.7 *Essentially zero. Heat is used in reheater and superheater. 3-9) Steam cycle efficiency increases as high pressure boiler pressure is increased from 800 psito the range of 1200 to 1450 psia. With a givensteaminlet temperature, turbine exhaust moisture increases with increasing inlet pressure. Conversely, for a given allowable turbine exhaust moisture and pressure, the steam inlet temperature must be increased as inlet pressure is increased. With a 1000°F steam inlet temperature, inlet pressure can be increased to 1200-1450 psia before an acceptable level of exhaust moisture is exceeded. Maximum moisture in the turbine should not be allowed to exceed a value in the range of 10 to 14% unless special moisture separators are used. Some steam turbine manufacturer's limit moisture to 10% as a rule of thumb. This condition would allow a pressure of up to about 1200 psia with 1000°F superheater outlet temperature and 2 1/2 inches Hg condenser pressure. One large domestic manufacturer, on the other hand, has designed a steam turbine for Louisiana Power and Light combined cycle plant at Sterlington, La., with steam conditions of 1320 psia and 995°F with a turbine exhaust pressure of 2 1/2" Hg (Ref. 6). These conditions would result in a turbine exit moisture slightly greater than that obtained with the inlet condtions of 1450 psia and 1000°F inlet used in the configurations on Table 3-1. However, the difference in performance between 1200 psia and 1450 psia is not very large. Accordingly, in most instances, (and in all cases when the superheat temperature was 1000°F), an absolute upper limit for steam turbine exhaust moisture was held to 14%. To observe this limit when pressure is increased from 800 to 1450 psi, either the turbine inlet temperature must simultaneously be increased, or the steam must be reheated. Specifically, it is practical to used steam conditions of 800 psi and 1000°F, or 1450 psi and 1000°F, or 1450 psi and 800°F/800°F reheat. But the 1450 psi and 800°F non-reheat combination results in excessive exhaust moisture which would be impactical. For this reason, a table of comparative performance for feasible steam turbines can show inlet pressures and temperatures which change simultaneously ona _ consistent basis. Consistency is achieved by holding a practical upper limit to moisture content. Such,is the case with Table 3-1. Configuration XC47, shown in Table 3-1, represents a current combined cycle with 800 psi and 800°F. When pressure is increased to 1450 psi, changes must be made to steam temperature in one way or another to avoid exceeding 14% moisture at the turbine inlet. Configurations XC54B, XC48, XC46, XC54, and XC54A (summarized in Table 3-1) represent different ways in which such changes can be made. The configuration changes represent successitely increasing levels of technology, and show correspondingly increased levels of performance. 3-10 XC54B represents the lowest level in which the 800°F steam temperature is retained. Excessive moisture is avoided through the use of reheat. Overall thermal efficiency increases 0.7 percentage points. The efficiency gain is slightly higher if a non-reheat cycle with 1000°F superheat temperature is used. This cycle can be configured in two ways. The first, XC48, uses a supplemental combustor in the gas turbine exhaust. Its overall efficiency is 36.1%. In the second, XC46, 1000°F steam temperature is obtained by placing the superheater in the raw gas cooler where the temperature is 1419°F. This system achieves an overall efficiency of 36.3%. Placement of a superheater in the raw gas cooler is not state-of-the-art and may possibly present corrosion problems. However, if by the use of suitable materials, corrosion life is adequate, then it may be possible not only to superheat but also to reheat in the fuel gas stream. Configuration XC54 represents the addition of a reheater to the system. Increasing steam temperature and/or reheating the steam not only permits higher steam pressure, but also produces an additional efficiency benefit. As shown in Table 3-1, at 1450 psi and 1000°F, the use of reheat increases thermal efficiency by 0.3 percentage points (XC54 vs. XC46) to 36.6%. For a 1450 psia reheat steam system, a simultaneous increase in both superheat and reheat temperatures from 800°F/800°F to 1000°F/1000°F increased overall efficiency by 0.6 percentage points (XC54 vs. XC54B). However, the quantity of heat required to superheat and reheat is such that no heat is available for a high pressure boiler in the raw gas cooler. Without the boiler to cool the gas, the superheater and/or reheater is directly subjected to the 1800°F corrosive and erosive raw gas stream. These conditions are somewhat similar to but more severe than conditions experienced in hydrocarbon reformers used in the petrochemical industry. The final colum in Table 3-1 (Configuration XC54A) shows the effect of increasing the steam pressure to 2400 psi. For these conditions the additional improvement in performance is only 0.1 percentage points better than XC54 with its lower pressure. This is the result of having steam conditions out of balance with the available gas turbine exhaust temperature, as discussed in the previous section. Table 3-2 indicates 0.3 percentage point improvement for reheat (Configuration XC45 and XC55 without fuel gas recycle). Only 0.1 percentage point is gained by increasing steam pressure from 1450 psia to 2400 psia (XC55 vs. XC55A). Spl TABLE 3-2 EFFECT OF STEAM REHEAT ON EFFICIENCY WITH NO FUEL GAS RECYCLE CONFIGURATION XC45 XC55 XC55A Steam Pressure HPT 1450 1450 2400 IPT <= 350 460 LPT 115 15 115 Superheat Temp - °F 1000 1000 1000 Reheat Temp - °F None 1000 1000 Gas Temp Entering 1608 2179 2316 Reheater - °F Overall Thermal 36.8 371 37 2 Efficiency - % EFFECT OF RECYCLE ON THERMAL EFFICIENCY Hot gas leaves the gasifier at 2400°F. In Figure 3-8a, this gas is passed directly to the high pressure (H.P.) boiler. To protect the H.P. boiler from this high temperature gas, it has been proposed that cold gas from the scrubber be recircu- lated and mixed with the raw fuel gas to produce a mixed temperature of 1800°F entering the H. P. boiler as shown in Figure 3.8b. Recirculating at temperatures below the high pressure steam boiler temperature makes less heat available for raising high pressure steam and more heat available for low pressure steam which is converted into work at lower efficiency. More heat is also rejected in the scrub- ber. As a result, in an oxygen-blown system, the overall efficiency with cold gas recycle is one-half percentage point lower than that of a system without recycle. The loss may be reduced by recirculating warm gas at about 600°F to 700°F after it has been used for superheating as shown in Figure 3.8c rather than passing it to the low pressure boiler. This reduces the flow and the temperature of the gas entering the low pressure boiler. Less low pressure steam is generated and more heat is made available for the high pressure boiler. Such a system gives almost the same performance as a system with no recycle. But it suffers from high parti- culate loading in the fuel gas which may erode the recirculating blower. The problem could be alleviated through the use of a cyclone separator, although this needs to be proven. A similar effect can be achieved by recirculating scrubbed gas which is reheated to approximately 600°F in a heat exchanger between the super- heater and the low pressure boiler as shown in Figure 3.8d. Of these four systems, 3=12 cold gas recycle is the easiest. The no recycle scheme realizes the highest poten- tial performance, about 0.5 percentage point higher overall thermal efficiency. Table 3-3 shows the performance difference between cold gas recycle (XC46) and no recycle (XC45). The desirability or need for gas recycle systems therefore depends on the hot side superheater/reheater temperature limits. (a) NO RECYCLE aso (b) COLD GAS RECYCLE RECYCLE BLOWER ~(ons (c) HOT GAS RECYCLE — (d) REHEATED RECYCLE — (com po Figure 3-8 Recycle Configurations 8=13 TABLE 3-3 EFFECT OF FUEL GAS RECYCLE ON EFFICIENCY CONFIGURATION XC46 XC45 Recycle Gas Temp - °F 355 None Steam Pressure psia HPT 1450 1450 IPT -- -- LPT 115) 115 Superheat Temp - °F 1000 1000 Reheat Temp - °F None None Gas Temp Entering Superheater °F 1419 1608 Overall Thermal Efficiency - % 36.3 36.8 EFFECT OF GASIFIER PRESSURE ON THERMAL EFFICIENCY The gasifier pressure was varied between 300 psi and 1200 psi by‘increasing the pressure ratio in the oxygen compressors. Increasing the system pressure has an effect on the thermodynamics of the fuel gas scrubber. As pressure was varied, the scrubber conditions were set for a constant ratio of water added to gas entering the scrubber. Overall thermal efficiency increased from 36.8 to 37.3% when the pressure was increased from 600 to 1200 psia as shown in Figure 3-9. This is due to an increase in the fuel gas expander power output which exceeds the increase in the oxygen compressor power consumption and a slight decrease in both the gas turbine and steam turbine power output. The circled end point at 300 psia on Figure 3-9 repre- sents a configuration with no expander turbine since 300 psia is the required fuel pressure for the combustion turbine. 3-14 38 POWER PLANT 27 EFFICIENCY PERCENT 36 ONO EXPANDER TURBINE 35 1 | Pe 1 4 200 400 600 800 1000 1200 GASIFIER PRESSURE — Psia Figure 3-9 Effect of Gasifier Pressure on Overall Thermal Efficiency EFFECT OF SECOND FUEL GAS REHEAT Since the expansion through the fuel gas expander with a 1200 psi gasifier pressure is greater than with a 600 psi gasifier pressure, it was expected that the accom- panying lower expander exit temperature would favor a second reheating of the fuel gas before admission to the gas turbine combustor. This comparison is shown in Table 3-4. Table 3-4 EFFECT OF SECOND FUEL GAS REHEAT (FOLLOWING EXPANDER TURBINE) AT 1200 PSI GASIFIER PRESSURE 6 CONFIGURATION ‘THERMAL FUEL GAS FUEL GAS E EFFICIENCY TEMP @ TEMP @ wa == % EXPANDER COMBUSTOR 2 68 EXIT, °F INLET, °F —~S oe E36 2 O2 +4 bh XC 31 35.9 609 609 oO ae <2 onu< XC 32 35.8 609 1000 =. -@ < 9° = ® a>oe em or SG ME dF ae oe aie el <Care i ‘ 3-15 | Reheating the fuel gas stream regeneratively to 1000°F produced no significant change in efficiency. Combustion turbine power increased; but in the the steam cycle, the amount of high pressure steam (efficient cycle) decreased and the amount of low pressure steam (inefficient cycle) increased. EFFECT OF OXYGEN PREHEAT TEMPERATURE ON THERMAL EFFICIENCY It was anticipated that increasing the temperature of the oxygen entering the gasifier might improve the overall system efficiency. Accordingly, this temp- erature was increased from 300 to 800°F by using gas turbine exhaust heat, but there was no effect on the system performance. In reality sensible heat in the oxygen between 300 and 800°F is a small fraction of the chemical energy. (This is not true for air-blown systems in which the oxidant flow is five times higher.) EFFECT OF AMBIENT TEMPERATURE ON PERFORMANCE Evaluation of the impact of ambient temperature on power plant performance requires the knowledge of the characteristics of the individual components at the off-design conditions. The controls required to maintain the proper operating conditions must also be defined. Since both of these requirements are beyond the scope of this study, the results of the ambient temperature evaluation must be considered pre- liminary. It was assumed that the combustion turbine, the off-design characteristics of which are known, would operate with a constant turbine inlet temperature. Since com- pressor air flow decreases as ambient temperature increases, coal feed rate must change with changing ambient temperature. The heat exchangers were assumed to have a constant UA at the off-design conditions except for the steam condenser whose off-design characteristics were not modeled. The gasifier and fuel gas expander were also not modeled for off-design performance. For these assumptions, there is very little change in overall thermal efficiency with ambient temperature. There is, however, a significant dropoff of power with increasing ambient temperature. (Figure 3-10). This follows a normal combined cycle trend and is due primarily to the combustion turbine. This trend can and should be ameliorated by increasing the gas turbine's turbine inlet temperature rating on a hot day. A temperature exceeding 88°F occurs in Chicago only 2% of the time, and 10% more power is obtained by increasing the turbine inlet temperature to the peak rating as compared to electrical base load ratings. Since most power is required on the hottest days, 3-16 and turbine degradation due to operating at peak rating occurs only 2% of the time, the increased rating on hot days is justifiable. With increasing ambient temper- ature, the condenser coolant temperature would obviously increase. This increases the condenser pressure and reduces the output and efficiency. The extent of the increase in condenser pressure would depend on details of the particular installation which are beyond the scope of this study. The effect of condenser pressure was, therefore, treated parametrically. Pressures of 2.5, 4.0 and 10.0 inches of mercury which correspond to condenser temperature of 109°F, 125°F, and 161°F respectively were investigated at an ambient temperature of 110°F. The results shown in Table 3-5 indicate a substantial loss in powerplant efficiency (2.8 points) when condenser pressure and temperature increase to 10 in Hg and 161°F. (XC43A vs XC43C) 110 POWER ain OUTPUT =—s-*10 PERCENT 5” He (109° OF =m CONDENSER a5 he OG DESIGN 80 PRESSURE 4” Hg (125°F) _ 10” Hg (161 °F) -o i ! J 0 20 40 60 80 100 120 AMBIENT TEMPERATURE — °F Figure 3-10 Effect of Ambient Temperature on Power Output Table 3-5 EFFECT OF CONDENSER PRESSURE ON PERFORMANCE AT 110°F AMBIENT TEMPERATURE RMAN' CONFIGURATION FEREO ee XC43A XC43B XC43C Condenser Pressure, Inches Hg 255 4.0 10.0 Condenser Temperature, °F 109 125 161 Overall Thermal Efficiency, % Spiel, 35.0 32.9 3=17 Section 4 ADVANCED TECHNOLOGY COMBUSTION TURBINES AND GASIFIERS EFFECT OF HIGHER COMBUSTOR EXIT TEMPERATURE Combustor exit temperature has traditionally increased with time as technology advances. Since concurrent advances are made in materials, cooling techniques, and component efficiency, computed performance with higher combustor exit temperatures at either constant level of cooling airflow or at a constant metal temperature would not reflect the true performance effects which accompany the increase in temperature. For this reason, the true effect of increasing combustor exit tem- perature on efficiency is not directly calculable in a parametric sense. Different results will be obtained depending on the accompanying assumptions which are made or implied. With these reservations, the effect of a 300° increase in combustor exit temperature is shown in Table 4-1. A temperature of 2300°F was selected as the approximate limit for advanced technology with an acceptable level of re- liability. Table 4-1 EFFECT OF ADVANCED COMBUSTOR EXIT TEMPERATURE ON POWER PLANT THERMAL EFFICIENCY CONFIGURATION CE EDA XC46 XC54 XC57 Xxc53 Combustor Exit Temperature, °F 1984 1984 2300 2300 Combustion Turbine Exhaust Temp., °F 910 910 1096 1096 Steam Pressure, psi 1450 1450 1450 1800 Steam Temperature, °F 1000 1000 1000 1000 Reheat Temperature, °F No Reheat 1000 1000 1000 Stack Temperature, °F 263 276 296 301 Overall Thermal Efficiency, % 36.5 36.6 38.8 39.0 Recycle Temperature, °F 355 355 361 591 4-1 The comparison between 1984°F and 2300°F combustor exit temperature is made on several bases. Since 1984°F combustor exit temperature represents a near term power plant, a simple 1450 psi, 1000°F non-reheat steam cycle is appropriate (Configuration XC46). Since 2300°F represents an advanced engine, a 1800 psi 1000/1000°F reheat cycle is appropriate (Configuration XC53). With a higher gas temperature entering the boiler, the optimum boiler pressure increases. For this reason, an 1800 psi steam pressure was selected for the higher temperature engine. With 1800 psi inlet pressure, reheat is required to maintain adequate steam quality at the low pressure end of the steam turbine. It is also desirable to show the effect of changing turbine inlet temperature without changing the steam cycle. For this direct comparison, a 1450P 1000/1000°F reheat cycle was analyzed (Con- figuration XC54 and XC57). All four configurations are compared on Table 4-1. The gain in thermal efficiency with increased gas turbine firing temperature is 2.2 to 2.4 percentage points depending on the basis for comparison. A schematic of the 2300°F combustor exit temperature power plant (Configuration XC53) with temperature and mass flows at significant stations is shown in Figure 4-1. Schematics and complete data printouts for Configurations XC46 and XC53 are given in Appendix H. COMPARISON OF AIR-BLOWN AND OXYGEN-BLOWN GASIFIERS For the condition investigated, the air-blown gasifier produces a higher (1.1 points) overall power plant efficiency than the oxygen-blown gasifier. (Table 4-2.) (Other investigators, Ref. 5, have shown that oxygen-blown gasifiers produce a slightly more efficient power plant at a different set of conditions.) Non- performance factors however, favor oxygen-blown gasifiers. More test data is available on the oxygen-blown Texaco gasifiers than on the air-blown. Medium BTU gas (from an oxygen-blown gasifier) integrates better with energy parks where MBTU gas may be used as a chemical feedstock. MBTU gas may also be piped from central gasifier plants to satellite combined cycle power plants located close to sites where power plant exhaust heat can be utilized cogeneratively by local industries or communities. Remote gasifiers can be used efficiently if the gasifier heat is utilized to generate power. This heat was not efficiently utilized in some of the early studies. Schematics of the air-blown system and complete data printouts are given in Appendix I. 4-2 €-¥ WATER 1 2400F 400F a eee 144 WATER WATER WATER a. RECYCLE COAL & n 7 HEATER WATER 98# | 23 | 1850F 1087F 1294 105F |CLEAN BOIL REGEN| coot Hy 632F 559F 137 68 69 81 1000F 300F 83 MN 6 105F 155 # " CONDENSATE t HS To OXYGEN PROCESS SULFUR PLANT 62 a ‘29 RECOVERY PROCESS| 528P 1108 ry | 1000F 1.2P CONDENSER Y AIR tee SULFUR EXPAND. 781F 1708 TOPROCESS 106# ier 109F 19123 o¥ 137 WATER 88| 204P r THROTTLE Hated CONDENSATE 1654 ' 16P 176F : 98 AIR 1096F 153¢ 9264 Y ~ DEAER, nL 52 a cs 39 128 1 1000F |1ss¢ J 51 250F * 621F 621F 165# COOLER FLASH SH HP. HP. bog 107 tio 9274 | BOL] gga ECON. asp FEED PUMP 102 yw} vw = STACK 608% 126 | 127 130 MIF 1102F 9274 ate, tA lhe 486F FROM LEGEND i tA — 108 an | 125 3 PROCESS # FLOWIN LBS/SEC 115 1654 F TEMPERATURE — °F TER YW m P PRESSURE IN PSIA TO COND. 114 707F TO LPT 1000F Figure 4-1 Texaco Oxygen Blown Gasifier/Advanced Combustion Turbine (2300°F CET) Combined Cycle Configuration No. XC53, 7 Thermal = 39% Table 4-2 EFFECT OF AIR-BLOWN AND OXYGEN-BLOWN GASIFIERS PERFORMANCE DATA TEXACO TEXACO OXYGEN-BLOWN AIR-BLOWN GASIFIER (XC54B) GASIFIER (AC03) Combustor Exit Temperature, °F 1984 1984 Steam Pressure, psi 1450 1450 Steam Temperature, °F SH/RH 800/800 800/800 Fuel Gas Reheat Temperature, °F 800 800 Temp. of Recycled Fuel Gas, °F 355 362 Stack Temperature, °F 267 245 Overall Thermal Efficiency, % 36.0 37.1 COMPARISON OF TEXACO ENTRAINED-FLOW AND BRITISH GAS CORPORATION (BGC) SLAGGING FIXED-BED GASIFIERS The BGC gasifier is more efficient than the Texaco gasifier for combined cycle power generation principally because the gas produced contains a lower fraction of sensible heat and it uses less oxygen. On Tables 4-3 and 4-4 the BGC gasifier is compared with the Texaco gasifier on different bases with and without steam reheat. In Table 4-3 the BGC configuration XC50 is compared with Texaco configuration XC47 because it has nearly the same steam conditions. On the other hand, the higher temperature leaving the Texaco gasifier may allow high superheat temperature and, therefore, higher boiler pressure which yields the higher efficiency shown for Texaco configuration XC46. The use of steam reheat as shown in Table 4-4 increases the advantage of the BGC system relative to Texaco slightly over the non-reheat system advantage. Comparisons were also made of the two gasifiers with a 1450 PSI 1000/1000 reheat steam bottoming cycle (XC54, XC55, and XC56, Table 4-4). On this basis, the BGC is 0.7 to 1.2 percentage points higher in efficiency than the Texaco gasifier. Other studies (Ref. 2) have shown a 2 percentage point advantage to the BGC slagger. Reference 2 also showed a capital cost benefit for the BGC gasifier. The choice between the two could be affected by the relative state of development. Available information indicates that the Texaco gasifier operates satisfactorily. Little information has been published on the BGC. A schematic of the power plant with BGC gasifier (Configuration XC50) with temperatures and mass flows at significant stations is shown in Figure 4-2. Schematics and data printouts are given in Appendix J. 4-4 Table 4-3 COMPARISON OF POWER PLANT PERFORMANCE WITH TEXACO OXYGEN-BLOWN GASIFIER AND BGC SLAGGING GASIFIER PERFORMANCE DATA TYPE OF GASIFIER TEXACO TEXACO BGC (XC46) (XC47) (XC50) Combustor Exit Temperature, °F 1984 1984 1984 Oxidant Oxygen Oxygen Oxygen Steam Pressure, psi 800 1000 Steam Temperature, °F (No Reheat) 800 800 Fuel Gas Reheat Temperature, °F 800 508 Fuel Gas Temp. Entering Boiler, °F 1800 1800 282 Temperature of Recycled Gas, °F 350 None Stack Temperature, °F 250 270 Overall Thermal Efficiency, % 35-3) 37.3 Table 4-4 COMPARISON OF POWERPLANT PERFORMANCE WITH OXYGEN-BLOWN TEXACO AND BGC SLAGGING GASIFIER WITH STEAM REHEAT TYPE OF GASIFIER TEXACO TEXACO BGC XC54 XC55 XC56 Oxidant OXYGEN OXYGEN OXYGEN Combustor, Exit Temperature, °F 1984 1984 1984 Steam Pressure - psia 1450 1450 1450 Superheat Temperature, °F 1000 1000 1000 Reheat Temperature, °F 1000 1000 1000 Fuel Gas Reheat Temperature, °F 1000 1000 1000 Fuel Gas Temperature Entering Boiler, °F 1800 2400 282 Temperature of Recycled Gas -°F 305 None None Stack Temperature 276 278 271 Overall Thermal Efficiency 36.6 Sid 37.8 9-97 COAL OXYGEN T AIR 7491 AIR 59F 325P 95 620F 69F 414Pp 17) 508F 388F 214P 22 & 1984F 27 (& 905F 7855# 756F 65754 7 Gasir. |_820F JACKET 314 WATER "1 12 282F 199F 139F A CLEAN. 3734 371# uP 8784 70F 164 355# CONDEN- SATE TO PROCESS To PROCESS 50 PSIG: PA 100 PSIG 1604 115P 323F 58 9834 434 iip 338F 336F SH aan ECON THROTTLE 99 100 101 102 378F 103 104 98 92 829F| —459F 69 157 FROM PROCESS LEGEND # FLOWIN LBS/SEC 484F F TEMPERATURE — °F 94 P PRESSURE IN PSIA Figure 4-2 BGC Slagger Gasifier/FT4 Combustion Turbine Combined Cycle Configuration No. XC50, 7 . sermal = 37.3% H2S SULFUR RECOVERY SULFUR Section 5 DESIGN COMPATABILITY OF MEDIUM BTU COMBUSTOR WITH FT4 A design study was made to determine the physical compatability of a medium BTU coal gas combustor in an FT4 engine (Figure 5-1). The combustor was tested in June, 1979 as part of EPRI RP 985-3 program. Part load is accomplished by means of staging the combustion. The primary combustion zone is the small combustor section located on the horizontal centerline. Lightoff, idle and low power are accomplish- ed in this zone only. For higher power output, fuel is admitted to the outer por- tion of the combustor (secondary zone) and this flow is increased until full power is reached. A new fuel control and change in turbine nozzle vane area would be required to adapt the turbine section to 300 BIU gas. The nozzle vane area change can be accomplished by cutting the vanes at a slightly different angle from the present bill-of-material vanes. FIRST NOZZLE VANE PRIMARY COMBUSTION ZONE ‘SECONDARY COMBUSTION ZONE Figure 5-1 Medium BTU Combustor in FT4 Engine 5-1 Section 6 CONCLUSIONS The studies conducted under this program confirmed that gasified coal combined cycle powerplants utilizing near-current technology coal gasifiers such as the oxygen-blown Texaco unit and current technology combustion turbines can attain thermal efficiencies in the range of 36 to 37%. These current technology power- plants could successfully compete in performance with conventional coal-fired plants equipped to meet emission standards. The studies also have shown that further improvement in overall plant efficiency can be attained by the use of advanced technology gasifiers and combustion turbines. An improvement of about one percentage point in thermal efficiency over the 36 to 37% range can be achieved by the use of the British Gas Corporation fixed-bed slagging gasifier in place of the Texaco unit. The replacement of current combustion turbine with an advanced technology combustion turbine having a combustor exit temperature of 2300°F, i.e. about 300°F higher than current engines, would improve the thermal efficiency of the power plant by about 2.2-2.4 percentage points. These improvements are additive and suggest the possibility of advanced technology plants with thermal efficiencies approximating 40%. The studies conducted in Phase 2 emphasize that the overall thermal efficiency of the gasified coal combined cycle power plant is very sensitive to changes in the management of the heat entering the steam system. In particular, the efficiency of the plant is very sensitive to the mean temperature difference between the gas and steam sides of the boilers and other heat exchangers. This mean temperature dif- ference must be minimized to attain the maximum overall power plant efficiency. Section 7 RECOMMENDATIONS In view of the high efficiencies which have been shown to be attainable in current or near current technology gasified coal gas turbine combined cycle power plants, a conceptual powerplant design program should be undertaken to establish a baseline plant design and to obtain a first estimate of the powerplant capital and operating costs for comparison with conventional steam plants. Such a conceptual design study would constitute the first step toward the definition of a gasified coal, combustion turbine combined-cycle demonstration plant which is necessary to eval- uate the economics of the approach and gain operational experience necessary for the introduction of this type of powerplant into utility systems. Studies should also be made to evaluate optimum steam conditions from the cost-effectiveness standpoint. The gains in efficiency of gasified coal gas turbine combined-cycle power plants which have been shown to be attainable through the introduction of advanced gasifier and gas turbine technology are attractive from the standpoint of the fuel savings they represent. On this basis, continued research and development related to advanced gasifiers and gas turbines is recommended. rat REFERENCES Coal Gasification System Analysis S. Hamilton, United Technologies Corporation. EPRI Report AF-992. February 1979. Economics Studies of Coal Gasification Combined Cycle Systems for Electric Power Generation. K. Chandra, B. McElmurry, E. W. Neben and G. E. Pack, Fluor Engineers and Constructors, Inc. EPRI Report AF-642. Jaunary 1978. The Near Term Potential for Gasification-Combined Cycle Electric Power Generation M. J. Gluckman, N. A. Holt, S. B. Alpert, and D. F. Spencer, Electric Power Research Institute, Sixth Energy Technology Conference, April, 1979. A Unique Combined Cycle System To Meet Utility Intermediate Cycling Loads C. R. Boland and R. D. Patterson, Turbo Power and Marine Systems, United Technologies Corporation, American Power Conference, 1972. Economics of Texaco Gasification Combined Cycle Systems B. McElmurry and S. Smelser, Fluor Engineers and Constructors, Inc. EPRI Report AF-753. April 1978. Fossil-Fuel Central Station Power, November 1973, Page S-4. APPENDICES APPENDIX A INITIAL STUDIES OF FT4 WITH OXYGEN-BLOWN GASIFIER APPENDICES INTRODUCTION Appendix A summarizes differences between the initial studies of the FT4 with an oxygen-blown gasifier and the Phase 1 air-blown study. Appendix B contains a description of the effect of major subsystems on overall power plant efficiency. Appendices C through J contain detailed information for the eight typical cycles of interest: Appendix Configuration c XC47 FT4 with 0,-Blown Texaco Gasifier - Lowest Required Technology D XC48 FT4 with 0,-Blown Texaco Gasifier - Supplemental Combustor E XC46 FT4 with 0,-Blown Texaco Gasifier - Improved Superheater Material F XC45 FT4 with 0,-Blown Texaco Gasifier - No recycle. G XC45A FT4 with 0,-Blown Texaco Gasifier - High Pressure Gasifier H XC53 2300°F Combustor Exit Temperature Gas Turbine with O2 Blown Texaco Gasifier I ACO3 FT4 with Air Blown Texaco Gasifier J AC50 FT4 with BGC Gasifier For each of these appendices, the following is presented: e Description ° Schematics with station numbers, showing temperatures, flows and pres- sures at some of the more important stations. For computation purposes, low pressure steam is shown as expanding through a separate turbine. In reality, the low pressure steam would generally be injected into the same low pressure turbine as the main steam flow. e Performance Summary e Energy Balance e Printouts of: Temperature in degrees Rankine at each station Flows in pounds per second at each station Pressure in pounds per square inch absolute at each station Enthalpy in Btu/lb at each station In some instances, pressure drops across a group of heat exchangers in series have been lumped together. Some station numbers listed on the table but not shown on the diagrams are fictitious of stations used as a basis for calculating enthalpies. e A summary of temperatures, pressures, flows, enthalpies and power for all the turbomachinery. e Heat loads for the heat exchangers. A-ii The FT4 gas turbine was studied in conjunction with the oxygen-blown gasification system specified by Fluor Engineers and Constructors. A design rating was selected for the FT4 consistent with this gasifier application, which is representative of a current market product. Based on FT4 exhaust conditions and the interface steam requirements specified by Fluor for the combined cycle power block, a waste heat recovery bottoming system was defined and its performance calculated. Overall efficiency for the powerplant was calculated to be 33-34% depending on configuration. System performance was calculated for two pressure levels in the high pressure boiler: 2,400 psia, and 1,450 psia. Thermal efficiency with 2,400 psia was slightly higher (34%) than that for 1450 psia (33.5%). Since this efficiency level was far below the 39% estimated for a system using an air-blown gasifier in Phase 1, it was considered important to isolate and quantify those effects causing the difference in performance, and to make necessary system modifications to raise the efficiency level. Subsequent studies identified differences which taken together add up to the total difference between the Phase 1 results and the early Phase 2 results. These differences are shown on Table A-1. DESCRIPTION OF SYSTEM Fuel Gas Stream The 140°F coal and water is partially combusted with 300°F oxygen to yield a 2400°F gas mixture. The hot gas is successively cooled by generating high pressure steam, scrubbing the water, cooling in the hot side of the fuel gas resaturator and finally heating feed water. Hydrogen sulfide is removed from the fuel gas through the Selexo1® process. The HS is converted to elemental sulfur in the Claus. plant. Fuel gas is reheated first in the resaturator which adds water vapor as well as sensible heat to 300°F. The gas is then reheated to 1000°F in the fuel gas re- generator, fed through the fuel gas expander turbine, and delivered to the FT4 combustor. A-1 Combustion Turbine Stream The air enters the compressor, is compressed to 13-1/2 atmospheres, then enters the combustor where it is combusted with fuel gas. The combustion products are ex- panded through the gas turbine. The combustion turbine exhaust stream at approxi- mately 950°F (hot day condition) superheats and reheats high pressure steam, heats water in the high pressure economizer, and generates steam in a low pressure boiler. (In subsequent configurations, a high pressure boiler was placed in the gas turbine exhaust for higher thermal efficiency.) Following the low pressure boiler is a low pressure economizer and finally a deaerating boiler. Steam System High pressure steam is generated in the fuel gas stream. This high pressure steam is superheated in the gas turbine exhaust stream. Low pressure steam is evaporated in the gas turbine stream. This steam is also superheated in the gas turbine stream before expanding through the low pressure turbine. The biggest shortcoming of the configuration (TOF4-10) studied early in Phase 2 was its failure to generate more high pressure steam. (This defect was corrected in later runs). Two effects causing performance loss as compared to the Phase 1 runs are inherent in the real system: the 0.7% loss due to the requirement for carrying some fuel gas with the slag which is quenched with cold water, and the 0.1% loss due to the pressure drop across the fuel control. Considering which losses were inherent and which could be circumvented, 36 to 37% thermal efficiency was set as a goal for the subsequent Phase 2 studies. Table A-1 IDENTIFICATION OF PERFORMANCE DIFFERENCES BETWEEN PHASE 1 AND INITIAL FT4 STUDIES PENALTY PHASE 1 INITIAL PHASE 2 (C7G)- CYCLE PARAMETERS (C7G) (TOF4-10) (TOF4-10) AIR-BLOWN OXYGEN-BLOWN A% POINTS Gas Turbine Parameters 1 Gas Turbine Pressure Ratio (1) 10 13..5) Combustor Exit Temperature, °F 2,000 1983 Steam Conditions and Flow 2.3 Steam Pressure (Excl. DA)-PSI 1850/480 1450/480/115 Steam Flow Generated at each Pressure - Lbs/Sec 1168/56 323/572/24 Effect of Power For O2-Plant Air Compressors 3.3 Effect of Heat Bypassing Combustion Turbine and Passing Directly to Bottoming Cycle (Equ. 1 and 4, Section 3. b=24% for air-blown; 11% for O2-Blown) -3.1 Pinch Differential in LP Boiler, °F 40 70 0.4 Quenching of Slag - 10% of Fuel Gas (2) 0 10% 0.7 Fuel Control AP (2) --- 50 psi 0.1 Gasifier Pressure (2) 1200 600 0.6 Oxidant Preheat (2) 1000° 300° 0.3 Overall Thermal Efficiency, % 38.9 33).2 Sa? (1) Higher pressure ratio increases combustion turbine efficiency because heat is rejected at a lower temperature. Lower gas temperature entering the inlet however, reduces the steam temperature and pressure and reduces the bottoming cycle efficiency unless supplemental firing is used. (2) The number tabulated is the effect which each variable would have individually on an oxygen-blown system. The effect of these variables on an air-blown system would be significantly larger. A-3 APPENDIX B EFFECT OF EFFICIENCY OF MAJOR SUBSYSTEMS ON OVERALL POWER PLANT EFFICIENCY 20idsO IWOUS SACWAY LUN UU Loge6 Byseiy ‘eieioyouy "OAY UIS "M PEE Aysouiny 1eMod eyse|y AdOOD AYVHEIT The energy flow in a model of a gasified coal combined cycle is shown on Figure B-1,, This is a more complete model than that discussed in Section 3, and it takes into account two additional subsystems: where: BOT — ica} i) o Ml isa " o iT the oxygen plant the small Brayton cycle which consist of the boost compressor, the net sensible heat added to the gas in the gasifier/cleanup/heat exchanger, and the fuel gas expander. In the simplified model this Brayton cycle is neglected because its efficiency is nearly zero (It may be negative.) The equation for efficiency of the complete model is: 1 _ [Egr (1+b) (1-E, op) + Enor - 0.2 (1-C) Fo,Blown, HHV ~ - Eye, (1 - EE, - E (B-1) ct ~ Bpor * EgrEpor? ~ 4! Main Compustion Turbine Efficiency Bottoming Cycle Efficiency Efficiency of small Brayton cycle consisting of oxidant boost compressor, fuel gas expands and net sensible heat added between them. Proportion of coal lower heating value supplied as sensible heat to the small Brayton cycle. Proportion of fuel energy unavailable as latent heat. (1+L= HHV+LHV) Proportion of coal lower heating value required for auxiliary power. Proportion of coal lower heating value which bypasses the gas turbine and goes directly as gasifier sensible heat to the bottoming cycle. Proportion of coal lower heating value appearing as chemical energy in the fuel gas. The 0.2 (l1-c) term represents the power consumed by the oxygen plant. This term should be omitted when using an air-blown gasifier. Oxygen requirements increase as the proportion of sensible heat increases and as the proportion of chemical energy c, decreases. The constant 0.2 is derived from Fluor data (AF642) relating to the Texaco gasifier. B-1 The term E,e (1 - or - Rot + For +E Bor) represents the net power increment with the small Brayton cycle. The third term contains more than simply (e, E)) because the heat rejected by this cycle is utilized by the other cycles. Equation (B-1) and Figure B-1 includes provision for the auxiliary power require- ments "a". The magnitude of "a" has not been rationalized in the model as a function of other variables such as "c" and E,or to which it may be related. The following influence coefficients are derived from equation (B-1): AE(1+L) _ i i =e, (1 - Egy - Egor * EgrE por? AE, AE(14L SEGUE) = (1 - Egop) (1 - b= exB1) AE or AE CIE) ~ 4 = = b) — e,F, (1 - EL) iz GT GT BOT AIR BLOWN ONLY AE(1+L) _ ie i : = Ey (1 - Egy - Egor * Eggs) C1 AE (1+L) AECL) 2 Rais Be) a cT BOT B-2 ORE) HHV OF COAL = 1.0 + L LHV OF COAL 10 = ey +e9+bdec ey SENSIBLE HEAT FUEL GAS EXPANDER -BOOST COMPRESSOR By e,(1-Ey) CHEMICAL ENERGY, C BOTTOMING CYCLE Egor ORE e1Ey ey(1- Eq) + epee NET POWER POWER TO 02 PLANT = 0.2(1-c) AUXILIARY POWER REQUIREMENTS ORE y Figure B-1_ Energy Flow in Gasified Coal Combined Cycle B-3 APPENDIX C CONFIGURATION XC47 DESCRIPTION - (See Figures C-1 and C-2. Figure C-1 contains station numbers and Figure C-2 contains temperatures and flows) Fuel Gas Stream Coal and water at 140°F is partially combusted with the 300°F compressed oxygen to yield a 2400°F gas mixture. Ten percent of the hot fuel gas is carried along with the slag and quenched with water. The remaining hot gas is mixed with cool fuel gas that is recycled from downstream of the scrubber. The resultant 1800°F gas mixture is then successively cooled by generating high pressure steam, reheating the fuel gas, generating low pressure steam, precooling prior to scrubbing, and scrubbing with water. It is further cooled in the hot side of the fuel gas re- saturator, the feedwater heater and in the last cooler. Hydrogen sulfide is re- moved from the cooled fuel gas through the Selexo1® process. The fuel gas is then reheated first in the resaturator which adds water vapor as well as sensible heat and then to 800°F in the fuel gas regenerator. It is then expanded through the fuel gas expander turbine and delivered to the FT4 combustor. The oxygen plant was not modeled in detail. Power requirements for air com- pression in the oxygen plant were obtained from Ref. 2 and adjusted for a 90% compressor efficiency. Oxygen is compressed from 17 psia to 735 psia with 5 stages of intercooling. (Intercoolers are located between stations 9 and 10, 11 and 12, 13 and 14, 15 and 16, and 17 and 18. Combustion Turbine Stream Air enters the compressor and is compressed to 14.5 atmospheres. It then enters the combustor where it is heated by combustion of the fuel gas to a temperature of 1984°F. The combustion products are expanded through the turbine. The gas turbine exhaust stream at about 910°F is used to generate and superheat high pressure steam, heat water in the high pressure economizer, and generate low pressure steam. Following the low pressure boiler is a low pressure economizer and finally a deaerating boiler. C-1 o-9 COAL& WATER AIR: RECYCLE COMP. WATER WATER WATER WATER CONFIGS XC45 450 46 NOT USED FOR XC47 &XC48 WATER CONDENSATE PUMP SULFUR FLASH GAS 140 142 Figure C-1 Texaco Oxygen Blown Gasifier/FT4 Combustion Turbine Station Numbers for Configurations XC45, 45A, 46, 47, 48 €-9 RECYCLE COMP. 346F 206# WATER 1 WATER 2400F 50# / quence 400F Ne WATER Lal WATER COAL& y | WATER 455¢|1800F 863F 532F 378F 362F scrue P5833 505¢ = 661F | yp. iF 366F 7398 BOIL on BOIL com. 525F 294F |< wr at rw) | 387#| 1 conics 4424 wn 105F 45A — 525F - 101F | CONDENSATE cor | 1944 OXYGEN SULFUR PLANT RECOVERY 800P fate | 800F 155P AIR per 393F 400F CONDENSER - SULFUR _ 109F 458F 800F ay 11418 ‘4424 GAS THROTTLE 700 EXPAND. aa7r [445] 459F| 109F PSIG. | 6I6F CONFIG XC48 ONLY ii q X | | 214P I IDEAER) | 1 i 518F 525F 253F AiR soa j sos YY ea fw 250F! 59F ‘aa S.H WP. HP. THROTTLE FEED ( a BOIL ECON. ieas cu L_ gize soor | sae PUMP 320F 250F Cc 1 | -— 66528! 1 380F 6552# : = 338F =e J Lm i ] Tat] “We | LEGEND 105F SH, # FLOWIN LBS/SEC 5# F TEMPERATURE — °F FLASH GAS 2967 438F P PRESSURE IN PSIA Figure C-2 Texaco Oxygen Blown Ga: 1/FT4 Gas Turbine Combined Cycle Configuration No. XC47, 7 Thermal = 35.3% Table C-1 PERFORMANCE SUMMARY FOR CONFIGURATION XC47 Combustor Exit Temperature 1984°F Fuel Gas Reheater Temperature - Raw Gas Inlet 863°F Clean Gas Outlet 800°F Steam Temperature - Superheat/Reheat 800°F/NRH Coal Rate 10,000 tons/day Thermal Efficiency 35.3% Heat Rate 9669 BIU/KW-HR Combustion Turbine Power 742.4 MW Steam Turbine 481.9 MW Fuel Gas Expander Power 33.1 MW Air Compressor Power (-)106.0 MW Oxygen Compressor Power (-)43.7 MW Recycle Compressor Power (-)0.7 MW Utility Power (Including Steam Auxiliaries) (-)52.4 MW Net Power 1054.7 MW Coal Heating Value 12,235 BTU/LB Clean Gas (Higher) Heating Value 258.3 BTU/SCF Combustion Turbine Overall Pressure Ratio 14.68 Steam Turbine Pressures - High Pressure 800.0 psia Low Pressure 115.0 psia c-4 Table C-2 ENERGY BALANCE FOR CONFIGURATION XC47 DATUM = 59°F Sensible Latent Chemical Heat Total % Heat Heat Energy Losses Heat Total Feed Feeds MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR % Coal +1955 +10196.0 +10215-5 Slurry +321 +32.1 Feeds Subtotal +10247.6 +100.00 Net Electrical Output -3762.7 36.72 Condenser =i) -3606.9 -3638.2 -35.50 Stack Gas -1096.4 - 677.9 -32.8 -1807.1 -17.63 Gas Cleanup rll oe 120 -203.8 -136.2 O42 a1 3534 Ash Loss 52.5 =O) -0.51 Generator Losses -57.8 -57.8 -0.56 Net Process Heat Loss -82.6 -82.6 -0.81 Miscellaneous Losses (1) -6.0 -6.0 -0.06 Resaturator 31 —Sieo =0531 0, Plant -473.3 473.3 -4.62 TOTAL =652 -0.06 PERCENT IMBALANCE = Qi8-Q0ut = _9 64 Qin (1) Primary bleed flow from the low pressure cooling and ducted overboard. c-5 compressor exit used for turbine TEMPERATURES °R STAT RD Ba me ee UNWANCOwWS 3 ft GEARAAUdHANEN HHH ASS DCD ne AYASSRUGAINE NE GPRS Tee Dane iz 1215. es as a GI: Oru spec S2R e838 B8aA8SRRo: i 4 8 0 Cae “sy BAF “DO Anus voo 1075.9 2078.5 1363.6 565_ 60 1373.3 1975.9 1259.6 1018.2 998.07 779 69 520.00 1323.2 806.10 565.00 wa = @ wa Ww 149 TABLE C-3 TEMPERATURE FOR CONFIGURATION XC47 uo ~ NE Nooo CAND Italo tue RABAR=SSKT ow Nw won au ee -OOoo ONS TARO Roe ee SOUNWHW ee auw os 143 146 152 155 158 S520. C-6 8 7] en Sesuaeaee BSS8322 AA Onan 15a oo - © 06 g 3 % @ ™ So 918. a wa ‘D aS =O Sn ow sv oo uw = ao a a 1199 ies 1363 1367 1373 1373. 1018. 5 a 22838 ages Bsn Son auwon—w § 150 SRISVRASA ee et tt te et Rs egs NSH OUN Ou 984 67 =—Ow sa nae ahaRRgaay3 BES Bin SRRalin AAD— Ras aown eo ua aor-o 568.7 568.27 2444.3 1775.1 565.88 1373.3 1373.3 1373.3 1018.2 888.67 840.26 789.99 1323.2 1323.2 915.18 520.08 N o N nwa Ne _ 2 N a o TABLE C-4 MASS FLOW FOR CONFIGURATION XC47 MASS FLOW LBS/SEC STal 1 4 ” ‘ 10 13 16 19 22 25 28 31 34 3? 40 43 46 49 S2 3S 58 6124.7 6 194.43 194.43 194 43 194.43 194.43 505.44 205 63 1149.8 1140.9 108 34 Re BS8 + I SO oe +O Gl 41. 112.06 549.63 13.995 S67 83 $49 63 2s, 295 1140.8 9 6313.0 6546.9 5.1014 6551.9 8 6551.9 6306.2 245.77 6551.9 6551.9 .66142E-02 205 _63 29.518 149 ee et ee tA DO TOTS ee ee KONINUOHWOAY ee aeuw WO-is 146 152 155 158 604.98 n 2 y SEER aNe3 RE MABE PR SES TS SLDIRI SEE BRSBRBS SRAAANSRASLE 120 123 126 129 132 135 133 141 144 147 153 iss PRESSURE PSIA STAT 1 $ 19 13 16 19 nD Se S2BRRIAASITS mo o-3 ee 2s2c tad fe ne et te ee WWM Shee DADNA-—GuUT on oe Pasa @eunw 151 134 157 rs) 28a Beuse oe oe ee oo oe, Aaaw SEENAAHRS :33 S433 BBSRSSVVIA88 ee ee ROR ER BOUUNwN 149 TABLE C-5 PRESSURE FOR CONFIGURATION XC47 STAT PSWGUAN Vly Nee S-GUNIGANVONS WU 4 aA o-. REBARASAKS AA WAIN fo fue ee f—ODUNNWVAWS c-8 800 6a oe. aeseee saaeezeazese bh ee ome ot oo AAA SRLASSS wear nn ates 2QO7SnO0 — ee eee UT =: iNMODDOOVONEO Ah aOomty WEMVUGhWAUety WA Ntyto QO Haas ee ae <ind AAD in 14.755 44 755 14 755 14 755 14.755 158 105 DNV KK 851. Bee 783: s$seae eee eons aan ah gaaegesee + ed BRAGG nu NWO oe ENTHALPY STAT BATOGAWO-a— GEGSRI DY Toe ee S- SSDS 'ONWON BRIaASVL2AS aN D'D'0O ike OD af 10G 183 106. 169 112 115 113 121 124 127 130 133 136 139 142 145 148 151 154 157 BTU/LB 33 = ar Sues rr od aoe 83 oe oe te oe WI DOWWUN 86 Panu - oe - ee NIA—wo 23 Why Rg Ba BR noes 3 VOW TABLE C-6 ENTHALPY FOR CONFIGURATION XC47 149 wn 4 =—OMIONIONS Oct Ll BAA AT CUD Re ee RABY ee ee ee Who tute eee OOO WDM KDUINWAWOWS=— DANO ee AGW Out 143 146 152 155 158 4 124 us: mmm Gs NTO 399. aw S = 'D'O DOWD BeOS PTO me oe See SRue-=S mney! oe 88 BNWUWHOOGNN TUude = BE RUS SAB? BOB. 22 aw 2 24.9 is 158 & Nee Ko OnS-O 491. Oo 112.61 168.30 126.86 168.75 130.68 1160.8 1168.8 298.99 389.12 eS —aM ogres 88 Boon GIN @Giairor 8 = Nawe WBBSEsRGss otal ABNas gN3 OI-9 Table C-7 TURBOMACHINERY SUMMARY FOR CONFIGURATION XC47 Stat. Flow Rate Pressure Temperature Enthalpy Total Power No. LB/SEC PSIA °F BTU/LB HP Combustion Turbine 1,008,220 Comp. Inlet 56 6124.7 14.55 59.0 124.0 Comp. Disch. 58 5376.3 213.63 1859.8 295.3 Turb. Exhaust 124 6546.9 15.92 908.6 340.4 Steam Turbine 655,119 HPST Inlet 66 992.3 800.0 800.0 1399.3 MPST Inlet 104 13.1 460.0 458.5 1204.3 LPST Inlet 99 & 107 1104.3 115.0 399.5 1224.9 Expander Turbine 45,299 Inlet 94 441.7 527.6 800.0 467.1 Exit 95 441.7 263.6 615.9 394.6 Recycle Comp. 993 Inlet 154 205.6 567.0 346.1 292.6 Exit 155 205.6 588.0 355.1 296.0 Air Compressor AIR COMPRESSOR POWER CONSUMPTION WAS TAKEN FROM 142,083 for 0, Plant REFERENCE 2 AND ADJUSTED FOR COMPRESSOR EFFICIENCY OF 90% 0, Compressor 58,541 Inlet 8 194.4 17.0 90.0 119.2 Exit 19 194.4 734.7 300.0 165.8 Table C-8 HEAT EXCHANGER HEAT LOADS (MMBTU/HR) FOR CONFIGURATION XC47 Fuel Gas Stream High Pressure Boiler Fuel Gas Reheater Low Pressure Boiler Scrubber Precooler Resaturator Feedwater Heater Clean-up Precooler Combustion Turbine Exhaust Stream High Pressure Steam Superheater High Pressure Boiler High Pressure Economizer Low Pressure Steam Superheater Low Pressure Boiler Low Pressure Economizer Deaerator 949 140 715 27 -65 310. -23 13. 296. 208. -65 59 95 98 14 <5 1482. 746. -01 288. 356. 407. 76 65 70 26 01 c-11 APPENDIX D CONFIGURATION XC48 DESCRIPTION - (See Figures D-1, C-1) Configuration XC48 differs from Configuration XC47 (described in Appendix C) in that a supplemental combustor is used in part of the gas turbine exhaust so that steam conditions can be raised to 1450 psi, 1000°F. e-a 346F RECYCLE COMP. 2064 WATER QUENCH mt 400F 532F WATER WATER WATER REGEN up BOIL Hy CONFIGS XC45 45A 46 1450P 1000F AIR THROTTLE 600F HP. HP. ae BOIL 632F ECON. Bon THROTTLE 385F LEGEND # FLOWIN LBS/SEC F TEMPERATURE — °F P PRESSURE IN PSIA Figure D-1 Texaco Oxygen Blown Gasifier/FT4 Gas Turbine Combined Cycle Configuration No. XC48, 7 Thermal = 36.1% Table D-1 PERFORMANCE SUMMARY FOR CONFIGURATION XC48 Combustor Exit Temperature Fuel Gas Reheater Temperature - Raw Gas Inlet Clean Gas Outlet Steam Temperature - Superheat/Reheat Coal Rate Thermal Efficiency Heat Rate Combustion Turbine Power Steam Turbine Fuel Gas Expander Power Air Compressor Power Oxygen Compressor Power Recycle Compressor Power Utility Power (Including Steam Auxiliaries) Net Power Coal Heating Value Clean Gas (Higher) Heating Value Combustion Turbine Overall Pressure Ratio Steam Turbine Pressures - High Pressure Low Pressure 1984°F 863°F 800°F 1000°F/NRH 10,000 tons/day 36.1% 9451 BTU/KW-HR 687.8 MW 564.3 MW 33.1 MW (-)106.0 MW (-)43.7 MW (-)0.7 MW (-)55.8 MW 1079.0 MW 12,235 BTU/LB 258.3 BTU/SCF 14.68 1450.0 psia 115.0 psia D-3 Table D-2 ENERGY BALANCE FOR CONFIGURATION XC48 DATUM = 59°F Sensible Latent Chemical Heat Total % Heat Heat Energy Losses Heat Total Feed Feeds MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR % Coal +19'35 +10196.0 +10215.5 Slurry +32:.1 +32.1 Feeds Subtotal +10247.6 +100.00 Net Electrical Output -3872.6 -37.79 Condenser -30.9 -3595.6 -3626.5 =35 .39 Stack Gas -1022.4 - 677.9 =32...8 -1733\.1 -16.91 Gas Cleanup 1,1 -1.0 -203.8 -136.4 -342..3 -3.34 Ash Loss =52..5 =52).5 -0.51 Generator Losses -61.0 -61.0 -0.60 Net Process Heat Loss -82.6 -82.6 -0.81 Miscellaneous Losses (1) -5.6 -5.6 -0.05 Resaturator -31.5 =31.5 -0.31 O02 Plant -473.3 -473.3 -4.62 TOTAL =33). 1 -0.33 PERCENT IMBALANCE = Qi8-Qout _ _9 339 Qin (1) Primary bleed flow from the low pressure compressor exit used for turbine cooling and ducted overboard. TABLE D-3 TEMPERATURE FOR CONFIGURATION XC48 TEMPERATURES °R STAT STAT STAT 1 519 00 2 513.00 3 4 513 09 5 S19 00 6 ° 528.00 3 350.00 9 19 5e5 ao ML 738 33 12 13 737 38 14 335 .@0 15 16 S85 69 17 737.71 18 1 768.01 20 768 O1 21 <2 cveg 0 23 2see 9 24 25 815.19 26 315.10 7 ze 519 90 29 Se1 00 2a 31 612 63 2 612.63 33 34 612.11 35 710.34 3 37 709.93 38 712.86 39 40 795.06 41 795 06 42 43 795 06 44 795. a6 45 46 1059.5 ? 1959.5 48 49 1051.7 50 1051.7 51 32 1059.5 33 1051.7 34 SS 1051.7 S36 319.@8 3? so 1215.5 $9 1323.2 68 61 1323.2 62 1323.2 63 64 1468.0 65 1468.8 66 67 877.54 68 877.54 69 78 79? .62 71 797 .62 v2 73 798.07 74 798.87 73 76 798 .67 7? 798.07 768 7? 992.15 88 837.98 81 82 868.68 83 826.63 o¢ 83 606.69 86 606.69 97 88 $31.09 33 563.00 99 91 365.00 92 520.69 93 94 1268 9 35 1073.9 96 9” SIL Fe 938 8gr.59 99 196 8s7 59 io 5é3 71 192 103 797 .O3 104 913.49 105 106 568 0 107 397 S59 1938 19s 568.70 110 S63. 7 111 112 97 62 113 58 78 114 115 $14.53 116 518.53 117 113 1075.9 119 1199 9 120 12 2073 5 122 W751 123 124 1462 5 125 1763 5 126 127 565.68 123 1367 8 129 13a 1373.6 131 1273.6 132 133 1075.9 134 1 e 135 136 1152.1 137 1 1 133 139 1091.7 140 1 Z 141 142 898 .0S 143 144 145 783 17 146 147 148 520.00 149 9 158 151 1323.2 152 153 154 806 18 155 156 157 563.00 158 159 nh Ja RRR GSES BBxRinStes 1978.9 aun Bnne me So888rrs DBS gE Resryzse g nN BR TABLE D-4 MASS FLOW FOR CONFIGURATION XC48 STaT STAT 2 Sér2.8 3 8 s Sé6r2.8 6 194.43 3 134.43 © 194.43 11 124 43 12 194.43 14 194.43 15 194.43 Le 194.43 18 194.43 20 194.43 21 585.44 23 454 22 24 50.544 26 205.63 27 668.53 29 1125.1 26 1125.1 32 1219.0 33 106.17 a 30 106.17 36 1125.1 1 38 4125.1 39 1125.1 1 41 1002.9 42 122.24 9 44 13.096 45 989.79 9 <a $13 15 48 476.64 5 50 513.15 Si 476.64 33 76.64 54 @ Sé $672.8 $7 SS75.3 so 668.53 68 668.52 9E-B2 62 668 .S2 683 989.79 65 989.79 66 989.79 68 989.79 69 989.79 71 78.648 v2 44.198 74 34.7353 7 23.295 7? 98.943 7 668.52 98 668.52 91 668.52 83 731.79 84 738.735 3 86 285.63 87 47.63 0 39 425.54 98 396.82 92 3396.02 93 441.74 35 441.74 96 441.74 98 2) 548.65 191 102 548.65 104 105 13.096 107 183 548.68 110 111 23.295 113 114 1125.1 116 117 409.15 119 128 5388.8 122 12 5984.4 125) 1265 5.1014 128 129 6068 9 131 122 3206.2 134 135 2895.3 137 138 6101.5 149 141 $392.1 143 144 6101.5 146 147 6101.5 , 149 6101 5 150 .66142E-82 5142E-82 152 66142E-02 153 .66142E-82 t 155 205.63 156 285.63 158 29.518 159 29.518 D-6 PRESSURE PSIA STAT 115 Giger fe Pere WSN Pee eee AGMDAALW NS ODUnw 1713.9 age 28 mo oo oo oe Nae 8333233 — Kan — PAN A-INh wuoOd ago 9a 115.00 149 TABLE D-5 PRESSURE FOR CONFIGURATION XC48 np 4 = OMNWONS— OUND REBAR SAV SRAS NSA SaNERUSTEe Wty tute eee OOGWIDDO =—DAMOANWSUND=—OUN'o ee a uw o-1S 143 146 152 155 158 D-7 14 696 14.696 17 080 63.974 111.94 395.87 S88 00 332.00 St a ao 2 o >>, ecre é] o psemeneses eo ee ed de SgbSRsaaagse ages sie: SS BBSSSSZao8SE 00 n000S3 uae Cowen hy tics - HN wast hy ee ee RR s AUNAG’UIDONON wee ty: HLELLLOwly- AS IU ON MARUNMODODOOWON 180 588. eae Bd AGae muag ge ehna: B3Zo08 Qannnw ppanae eee 8332 - meme Gl SasGanss Ne eee UE a eae GNA NN AN Nw TABLE D-6 ENTHALPY FOR CONFIGURATION XC48 ENTHALPY BTU/LB ‘STAT STAT STAT 1 2 124 63 3 8 4 5 124 03 6 112.61 7 3 113.17 9g 168.38 1a y) 161 a0 12 126.86 13 14 126.86 15 168.75 16 i? 160 @1 18 138.88 iy 20 165 84 al 1168.8 22 22 1160.9 2 1168.0 23 26 236 O1 27 898.99 2 29 63 961 30 69.065 31 32 120.45 33 128.45 34 35 1163.3 36 218.86 ° 38 225.32 39 399.12 49 4 309.12 42 309.12 43 44 309.12 45 309.12 46 47 616 93 43 616.93 49 50 1178.4 Si 616.93 S2 53 1170.4 54 1178.4 SS 36 124.83 57 209.32 58 39 491.63 68 491.63 61 62 491.63 63 1178.4 64 65 1491.8 66 1491.8 6? 68 1234.7 69 1234.7 78 71 389.12 72 389.12 ?3 74 1189.4 7 1389 4 76 7? 1189.4 78 491.63 73 88 382.12 81 296.17 82 93 299.53 84 292.59 ss 86 292.59 87 26.35 88 39 198.81 98 196.31 a1 92 133.83 93 271.75 94 95 394 56 96 394.56 oF 38 1240.1 99 1240.1 1n0 101 969.14 102 981.97 1633 104 1204.3 103 985 .37 1WE 107 1240.1 103 976.83 1ua 110 978.14 111 309.12 We 113 264 29 114 76.584 115 lig 26.922 117 394.56 118 119 302.82 12e 648.55 121 122 452 46 123 452.46 1e4 125 340 36 126 128.98 127 128 Ha.19 129 341.73 128 131 B41 72 132 341.73 133 134 342.32 135 393.57 136 137 214.18 133 267.63 139 140 267 .63 141 215.93 142 143 216.00 144 204.74 145 146 171.14 147 171.14 146 149 124.68 150 491.63 151 152 491.63 153 491.63 134 155 296.81 156 296.01 157 158 186.88 159 186.88 D-8 6-d Table D-7 TURBOMACHINERY SUMMARY FOR CONFIGURATION XC48 Stat. Flow Rate Pressure Temperature Enthalpy Total Power No. LB/SEC PSIA oF BTU/LB HP Combustion Turbine 934,124 Comp. Inlet 56 5672.8 14-55 59.0 Comp. Disch. 58 4979.7 213.63 155.0 Turb. Exhaust 124 6063.8 15592 908.5 Steam Turbine 767,806 HPST Inlet 66 989.8 1450.0 1000.0 1491.8 MPST Inlet 104 Sieg 460.0 458.5 1204.3 LPST Inlet 99 & 107 1088.8 115.0 427.6 1240.1 Expander Turbine 45,300 Inlet 94 441.7 527.6 800.0 467.1 Exit 95 441.7 263.6 615.9 394.6 Recycle Comp. 993 Inlet 154 205.6 567.0 346.1 292.6 Exit 155 205.6 588.0 3551 296.0 Air Compressor AIR COMPRESSOR POWER CONSUMPTION WAS TAKEN FROM 142,083 for 0, Plant REFERENCE 2 AND ADJUSTED FOR COMPRESSOR EFFICIENCY OF 90% 0, Compressor 58,541 Inlet 8 194.4 17.0 90.0 119152 Exit 19 194.4 734.7 300.0 165.8 Table D-8 HEAT EXCHANGER HEAT LOADS (MMBTU/HR) FOR CONFIGURATION XC48 Fuel Gas Stream High Pressure Boiler Fuel Gas Reheater Low Pressure Boiler Scrubber Precooler Resaturator Feedwater Heater Clean-up Precooler Combustion Turbine Exhaust Stream High Pressure Steam Superheater High Pressure Boiler High Pressure Economizer Low Pressure Steam Superheater Low Pressure Boiler Low Pressure Economizer Deaerator 949 310. -03 14. 296. 208. -65 140 1145 1022. 1096. «35 247. 339. 398. 3 -64 60 16 98 14 -50 49 86 33. 37 67 D-10 APPENDIX E CONFIGURATION XC46 30ldsaO WOU SAOWSY LON Od LoS66 eyse\y ‘eHeioyouy ‘OAV UIS "M PEE Ayoulny seMo0g eysely AdOO AuVuEN DESCRIPTION - (See Figures E-1, C-1) In Configuration XC46 it is assumed that the steam and fuel gas can be heated in the raw gas cooler to 1000°F by selecting the proper heat exchanger materials. Configuration XC46, therefore, differs from Configuration XC47 (Appendix C) in that a high pressure steam superheater has been added in the fuel gas stream to permit final superheat to 1450 psi, 1000°F. The fuel gas is also reheated to 1000°F. E-1 e-a OXYGEN PLANT t AIR AIR 59OF 62284 206# 346F RECYCLE COMP. WATER 355F } WATER WATER WATER 50f 2a00F [auencn ng 400F WATER { — 800 1419F Aaa __1088F 630F 378F 362F__ [366F] 5 yp lt | 6614) HP. | SH REGEN Lp 7324 7394 533 coor} BOlL ! | Bolt [ae ret mm iat pa CONFIGS XC45 294F mt 450 + ~ 600F 46 101F | CONDENSATE SULFUR RECOVERY 1450P 1000F ‘7964 418F 443F CONDENSER —_ SULFUR — 5284 1000F / Gas \ THROTTLE 0 EXPAND. sur 3374 [445 ]485F PSIG 791F|-— — — — — — — 4 t PROC. ' CONFIG XC4B ONLY 796# | q X | | 24P | 0 | 735F ' 255# 335°, 253F I 600F J 335F AA ase { HP. np, | 796# Le. THROTTLE = | LP. a o BOIL | 632F ECON. 474g] BOIL ECON. reo Fl 263 Vt} - LW _ T T 378F 324F 6655F 338F 2318 | rer | LEGEND SH. # FLOWIN LBS/SEC 1A F TEMPERATURE — °F aor 438F P PRESSURE IN PSIA Figure E-1 Texaco Oxygen Blown Gasifier/FT4 Gas Turbine Combined Cycle Configuration No. XC46, 7 Thermal = 36.3% Table E-1 PERFORMANCE SUMMARY FOR CONFIGURATION XC46 Combustor Exit Temperature Fuel Gas Reheater Temperature - Raw Gas Inlet Clean Gas Outlet Steam Temperature - Superheat/Reheat Coal Rate Thermal Efficiency Heat Rate Combustion Turbine Power Steam Turbine Fuel Gas Expander Power Air Compressor Power Oxygen Compressor Power Recycle Compressor Power Utility Power (Including Steam Auxiliaries) Net Power Coal Heating Value Clean Gas (Higher) Heating Value Combustion Turbine Overall Pressure Ratio Steam Turbine Pressures - High Pressure Low Pressure 1984°F 1088°F 1000°F 1000°F/NRH 10,000 tons/day 36.3% 9403 BTU/KW-HR 752.4 MW 499.0 MW 38.5 MW (-)106.0 MW (-)43.7 MW (-)0.7 MW (-)55.1 MW 1084.4 MW 12,235 BTU/LB 258.3 BTIU/SCF 14.68 1450.0 psia 115.0 psia E-3 Table E-2 ENERGY BALANCE FOR CONFIGURATION XC46 DATUM = 59°F Sensible Latent Chemical Heat Total x Heat Heat Energy Losses Heat Total Feed Feeds MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR % Coal ti9s5) +10196.0 +10215.5 Slurry 32 ook +3200 Feeds Subtotal +10247.6 +100.00 Net Electrical Output 3762.7 -36.72 Condenser =29 52 -3413.1 -3863.8 -37.70 Stack Gas 1189.4 - 677.9 -32.8 -1900.1 -18.54 Gas Cleanup -1.1 -1.0 -203.8 “136.5 -342.4 -3.34 Ash Loss 52.15 525 =0-5i1 Generator Losses -59.8 -59.8 -0.58 Net Process Heat Loss -82.6 82.6 -0.81 Miscellaneous Losses (1) =6.1 -6.1 -0.06 Resaturator =31.5 =O La) =0531 O02 Plant -473.3 -473.3 -4.62 TOTAL -6.8 -0.06 PERCENT IMBALANCE = 218-Qout _ _9 6% Qin (1) Primary bleed flow from the pressure compressor exit used for turbine cooling and ducted overboard. E-4 TEMPERATURES °R ES BaRSAaS O's sar 199 S19 S13 520. 585. 34> e a eengeee | R o eS 260 519. AND ae aon nO etOneo Gaoa wakes NR ANANE Sagagaiss S8aSSRSv0. $3 BBeseenanegaegass. BS Ragga eSG5 oo i 565 _ 0 7, ao ao NNNWOTIIDOON Brad S65. 149 TABLE E-3 TEMPERATURE FOR CONFIGURATION XC46 STAT tytulu tee DAAONS DUN WSO TS Nome ee oe 520. E-5 ~3 wn oO “ z a aaa ‘ONDA nog 3433s BSnoa2a 2 Ane +0 aaa =o on NNGARH ao- 33° ED oe ee Ce ee I UD > SBBLI For a8 W335 span RRA Bxu8 FRSLFS SENOBS ay - th - 3 © : ANS sg BOW SASSADS 158 25 aRRERES gax8a-8ne Bets Sasa 797 .62 nN NN 2444.3 1773.9 565.68 1372.7 1372.7 1372.7 1091.7 936.48 838.06 722.84 1879.4 1879.4 815.12 520.68 TABLE E-4 MASS FLOW FOR CONFIGURATION XC46 MASS FLOW LBS/SEC ‘STAT STAT STAT 1 €223 1 2 6228.1 3 8 q 9 5 6228 1 6 194.43 CR 194.43 3 134.43 9 194.43 19 194 42 MW 194.43 12 194.43 13 194.43 14 194.43 15 194.43 16 194 4% 17 194 43 18 194.43 19 194. 43 20 194 43 21 505.43 ee Su5 43 23 2 24 58.544 23 205 64 26 2? 668.53 22 1063 1 29 3a 1863.1 31 1063 1 33 97 .549 34 9g? 543 35 36 1963.1 i. 1063 1 33 39 1063.1 42 1063 1 41 42 254.55 43 795 Se 44 45 795.50 46 795.50 4 438 199.53 49 595.97 50 Si 199.53 52 9 53 54 9 55 199 53 S56 ? 6121.0 58 5467.1 $9 60 668.52 61 .68339E-82 62 63 795.58 64 795.58 65 66 795.38 67 795.38 68 69 793.58 7e 234.55 71 72 72.764 v3 181.79 74 73 23.295 76 72.764 77 78 669.52 7 668 .S2 68 81 668.52 82 71.273 83 84 738.76 8s $33.13 86 87 457.63 83 425.97 89 98 396 . 62 91 23.518 92 93 4941.74 94 441.74 95 96 441.74 ar 231 25 93 99 492.29 16003 534.46 101 102 492.29 102 13.0996 104 105 13.096 16 505 39 10 188 534.46 18:9 334.46 11 111 23.295 Mle 23 235 1 114 1063.1 115 WWEZ 1 11 117 441.74 113 a M1 120 5908.9 121 6412.2 12 123 6562.9 124 6650 0 12 126 5.1014 a 5 1014 les 129 6655.8 12a 6655.0 13 132 8 133 8 13 135 6655.1 136 6655.1 13 138 6655.1 139 6164.5 140 141 6164.5 142 498.57 143 144 6655.8 145 6655.8 146 147 6655.0 148 6655.0 149 : 150 .66142E-82 151 .66142E-02 CLNEIS2 .66142E-82 153 .66142E-02 154 205.64 WRCY155 205.64 156 205 .64 157 29.518 oOTS9158 29.518 iS9 29.518 E-6 PRESSURE PSIA STAT 5 te LO OQ ms de ‘oS > | eee ee Ne rere tery BBSBSSSoSSoa0n%0 : RSS asgss VOU ts 3 UA2oe 333433 Pete ~ yD Dee RUT UR _ = au a ID a2 5 ‘ 247d 213 63 106.49 15.921 110.00 14.756 213.63 14 756 14 756 14.756 14.736 147 588 08 56? oe SS2.e0 149 TABLE E-5 PRESSURE FOR CONFIGURATION XC46 STAT 2 14 696 5 14 696 3 17 a00 11 63 974 14 111.94 17 335.87 2 S88 00 23 S38 00 26 S828. 00 22 1.2279 32 38.000 35 20 900 33 1713.9 41 1713 9 44 1713 9 rig 1542.6 50 1450 9 53 1458.6 Sé 14.549 so 588.80 62 388.80 65 1458.9 68 115.08 71 115.00 74 113.68 7? 115.68 88 567.68 83 367.68 96 367.68 89 552.80 92 341.00 95 263.63 33 115.00 101 1.2279 104 460.00 107 115.900 110 1 2279 113 116 119 122 iZo 123 131 134 137 7 140 14756 143 14 756 146 14.756 14.736 152 588 00 155 588.00 158 552.08 E-7 150 BRASVLLSANSR ooo ood NheKHKOOo AON OU 126 129 132 135 133 141 144 147 153 156 159 588. Nee eee ASAGAN oeasan eonnnua Soe DLOAAHAO aeegsccgg Banna B388888338a00 NEN ae a S °S Table E-8 HEAT EXCHANGER HEAT LOADS (MMBTU/HR) FOR CONFIGURATION XC46 Fuel Gas Stream High Pressure Boiler High Pressure Steam Superheater Fuel Gas Reheater Low Pressure Boiler Scrubber Precooler Resaturator Feedwater Heater Clean-up Precooler Combustion Turbine Exhaust Stream High Pressure Steam Superheater High Pressure Boiler High Pressure Economizer Low Pressure Steam Superheater Low Pressure Boiler Low Pressure Economizer Deaerator 397 230 14 296 587 1187 881 87 254) 893° 438. -58 eo -98 208. -65 24 TM) 14 -40 oo -47 +28 576. 320. 365. 19 80 LZ E-10 TABLE E-6 ENTHALPY FOR CONFIGURATION XC46 ENTHALPY BTY/LB STAT STAT STAT 1 124 93 ie 124 03 3 8 4 <6. 935 S 124.93 6 112.61 < 112.61 3 119017 9 168.30 iG 126 3€ M1 161 60 2 126.86 13 160.95 14 126.96 15 168.75 16 126.26 iy. 160.81 18 130.08 19 165.34 20 165 84 21 1160.8 Par 1160 8 23 1160 9 24 1168.8 2 296 04 26 236.01 ti 898.99 <8 223 . 36 <9 63 961 29 69.065 31 123.45 32 123.45 33 123.45 34 123 45 35 11623 3 36 218.86 37 218.86 33 225.32 39 389.12 49 389.12 41 309.12 42 399.12 43 309 12 44 389 12 45 389.12 46 616.93 re 616.93 48 616.93 49 616.93 58 1170.4 Si 616.93 S2 616 93 S3 1170.4 54 1178.4 SS 1170.4 56 124.83 Ss? 209.32 38 295.28 $9 723.82 69 723.82 61 723.82 62 723.82 63 1178.4 64 1375.5 65 1491.8 66 1491.8 6? 123.9 68 1234.9 69 1234.9 78 389.12 71 389.12 v2 389.12 73 1169.4 74 1189.4 7 1189.4 76 9958 7? 1189.4 78 563.68 7 16 98 382.18 81 296.17 TE an aE an 68 196.45 89 198.31 98 fse-31 91 117 63 92 133.83 93 271.75 94 S47 66 25 462.35 96 463.35 a 1294 2 93 1248.2 99 1248.2 190 1248.2 10 973 90 162 996.84 103 309 12 194 1204.3 195 983.37 106: 934 7 107 1248.2 163 981.68 10:3 931.68 119 933.12 111 389.12 112 399 12 113 968 35 114 76.584 115 26_ 922 116 26 922 1d 463.35 113 463 25 119 307 85 120 648.26 121 539.13 122 451 92 12 451.98 124 240 15 125 340 16 126 128.98 rd 123 9A 128 340 60 129 341.48 12¢ 341 40 131 341 40 132 341.48 133 463.35 134 341 40 135 341.48 136 316.69 137 216 $2 133 267 .32 139 267 32 140 267 32 141 227.68 142 217.98 143 226.89 144 202.84 145 189 46 146 174 21 147 174.21 148 124.57 149 124 $7 1s@. . 723.82 151 723.82 152 ©2382 153 723.82 154 292 68 155 296.81 1S6 296.01 157 117.68 158 106.88 159 186.88 E-8 Table E-7 TURBOMACHINERY SUMMARY FOR CONFIGURATION XC46 6-4 Stat. Flow Rate Pressure Temperature Enthalpy Total Power No. . LB/SEC PSIA °F BTU/LB HP Combustion Turbine 1,021,820 Comp. Inlet 56 6228.2 14.55 59.0 124.0 Comp. Disch. 58 5467.1 213.63 755.5 2953 Turb. Exhaust 124 6650.0 151-92 908.0 340.2 Steam Turbine 678,452 HPST Inlet 66 795.5 1450.0 1000.0 1491.8 MPST Inlet 104 13).0 460.0 458.5 1204.3 LPST Inlet 99 & 107 1026.8 115.0 442.8 1248.2 Expander Turbine 52,673 Inlet 94 441.7 527.6 1000.0 547.7 Exit 95 441.7 263.6 790.7 463.4 Recycle Comp. 993 Inlet 154 20576 567.0 346.1 292.6 Exit 155 205.6 588.0 355%. 1 296.0 Air Compressor AIR COMPRESSOR POWER CONSUMPTION WAS TAKEN FROM 142,083 for O02 Plant REFERENCE 2 AND ADJUSTED FOR COMPRESSOR EFFICIENCY OF 90% O, Compressor 58,539 Inlet 8 194.4 17.0 90.0 119).2 Exit 19 194.4 734.67 300.0 165.8 APPENDIX F CONFIGURATION XC45 DESCRIPTION - (See Figures F-1, C-1) Configuration XC45 differs from Configuration XC46 in that it is assumed that the boiler could be designed to withstand 2400°F raw fuel gas. Accordingly gas recycle was removed. c-i 0#(NO RECYCLE) RECYCLE COMP. WATER 1 WATER WATER ER Ee 2000F __ouencn ae WATER bis Mt | { 1608F a2F 378 359 _ [365°] say AS | zs 1216 [| rose | 505F 2400F | ‘Hp. cael P 526 5a1F 457# 4264 CLEAN uP BOIL | gop BOIL RESAT- AWA — JURATOR! 70F CONFIGS 30 XCa5 293F ees oe 7 ou WATER 101F | CONDENSATE 294 # y PUMP yes SULFUR OXYGEN PLANT RECOVERY 1450P 101F | 1000F AIR ~ convensen - SULFUR 109F _ 10524 Lome, 151F Gas THROTTLE 700 EXPAND, osu 337F [445 aser( tpt | oor ‘| Psic | PROC. 812F 338F 234 335F 253 IVa ot TT seus TSF pe 250F| SH. HP. HP. LP. THROTTLE SO FEED i BOIL ECON. BOIL . ~— 912F B12F 632F 466F PUMP = 325F 263F LWt—~ = WWW tv = ~— 6655F Tt 378F 6655# | 151 338F| I = 7 | N= LEGEND SH. # FLOWIN LBS/SEC 1A F TEMPERATURE — °F sau or P PRESSURE IN PSIA 105F Figure F-1 Texaco Oxygen Blown Gasifier/FT4 Gas Turbine Combined Cycle Configuration No. XC45, 7 Thermal = 36.8% Table F-1 PERFORMANCE SUMMARY FOR CONFIGURATION XC45 Combustor Exit Temperature Fuel Gas Reheater Temperature - Raw Gas Inlet Clean Gas Outlet Steam Temperature - Superheat/Reheat Coal Rate Thermal Efficiency Heat Rate Combustion Turbine Power Steam Turbine Fuel Gas Expander Power Air Compressor Power Oxygen Compressor Power Recycle Compressor Power Utility Power (Including Steam Auxiliaries) Net Power Coal Heating Value Clean Gas (Higher) Heating Value Combustion Turbine Overall Pressure Ratio Steam Turbine Pressures - High Pressure Low Pressure 1984°F 1088°F 1000°F 1000°F /NRH 10,000 tons/day 36.8% 9283 BIU/KW-HR 752.1 MW 512.6 MW 38.4 MW (-)106.0 MW (-)43.7 MW (-)0.0 MW (-)55.0 MW 1098.4 MW 12,235 BTU/LB 258.8 BIU/SCF 14.68 1450.0 psia 115.0 psia F-3 Table F-2 ENERGY BALANCE FOR CONFIGURATION XC-45 DATUM = 59°F Sensible Latent Chemical Heat Total % Heat Heat Energy Losses Heat Total Feed Feeds MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR % Coal +119 /-:5) +10196.0 410215 <5 Slurry +32.1 +32.1 Feeds Subtotal +10247.6 +100.00 Net Electrical Output =s911-5 =38,17 Condenser -28.9 -3365.8 -3394.7 =33.13 Stack Gas -1193.0 - 674.7 =32.8 -1900.5 -18.55 Gas Cleanup -1.1 =120) -203.8 -133.6 339-5 3-31 Ash Loss =52..5 521.5) -0.51 Generator Losses -60.8 -60.8 =0-59 Net Process Heat Loss -82.6 -82.6 -0.81 Miscellaneous Losses (1) -6.1 -6.1 -0.06 Resaturator -33.1 33/1 -0.32 0, Plant 473.3 =473.53 -4.62 TOTAL S760) -0.07 PERCENT IMBALANCE = @i8-Qout _ _9 974 Qin (1) Primary bleed flow from the pressure compressor exit used for turbine cooling ‘and ducted overboard. TABLE F-3 TEMPERATURE FOR CONFIGURATION XC45 TEMPERATURES °R STAT STAT 1 519.00 2 519.00 4 519 90 5 5190 ? 520 80 3 550 09 10 585.00 11 738 53 13 P37 3S 14 585.00 16 585.00 17 73? 70 19 76800 20 760 00 22 2860.8 23 2860 8 25 2102.2 26 2102 2 2 519 a0 29 561 00 31 615.31 32 615.31 34 614-7 35 710 34 37 709.93 38 712.86 40 795 86 41 795.06 43 795 26 44 795.06 46 1059.5 4? 1059.5 49 1051.7 50 1051.7 52 1059.5 53 1051.7 55 1051.7 56 519.00 58 1215.5 59 2068.8 61 2068.8 62 2068.8 64 1272.7 65 1463.0 6? 877.72 68 877.72 78 797.62 71 797.62 73 798.e7 74 798.07 76 738.07 7? 7368.07 7 971.93 a6 637.66 3 068.08 83 823.63 005.61 86 685.61 83 581.90 89 565.08 31 565 90 92 530.00 94 1469. 35 1250.6 9” 991 72 38 294,23 109 23 101 68 7 103 104 313.49 106 107 894 23 109 110 563.7 112 113 568 71 115 116 518.53 118 119 1219.4 121 122 1973.7 led 125 1268 0 127 128 1367 3 130 131 1372 7 133 134 13°2.? 136 137 1272.9 139 140 1031.7 142 143 926.07 145 146 723.44 - 148 149 520.08 159 151 152 2068. 154 155 2102.2 157 158 528.08 F-5 793.06 “3998 88 NON® ed “833 NAGS REE on Ou o Nis TABLE F-4 MASS FLOW FOR CONFIGURATION XC45 MASS FLOW - LBS/SEC STAT STAT STAT 1 6229 3 é 6229.9 3 8 4 9 5 6229 9 6 194.43 . 194.43 3 194.43 9 194.43 1g 194.43 M1 194.43 12 194.43 13 194.43 14 194.43 15 194.43 16 194.43 17 194.43 18 194.43 1g 194.43 29 194 43 al 595.43 ae SHS 42 2s 454 932 24 58.544 Zc B6T40E-07 26 36740E-@7 e 4954.89 26 1052 23 1952 9 24 1952.8 31 1052.9 32 255..15 33 96.824 34 96.324 35 36 $24 36 1852.8 37 1052.0 38 1952.9 39 1852.8 40 1052 8 41 977.26 42 174.72 43 864.16 44 13.096 45 864.16 46 864.16 ? S70 42 48 293.74 49 576 42 SO SvO 42 Si 293.74 S2 8 53 293.74 34 8 33 293.74 56 6229.9 57 6122.8 58 5468.7 39 454.89 68 454.89 61 .46387E-82 62 434.89 63 964.16 64 864.16 63 864.16 66 864.16 67 964.16 68 864.16 69 964.16 78 174.72 71 168.@5 ?2 6.6689 73 168.85 74 144.76 7 23.295 6 6.6688 7? 151.42 78 434.89 e 8 fears gs Sae.te eT 8s 331.17 86 8 87 87.13 838 425.97 39 425.53 98 396.82 a1 29.517 32 336.02 93 448.89 94 440 93 93 440.39 96 448.89 ” 151 42 98 1015.6 99 489.82 193 534 76 101 480.82 102 488.82 103 13.996 104 13.096 105 13.096 106 493 92 107 534.76 163 534.76 169 534.71 110 1023.7 M1 23.295 112 23.295 113 1952 9 114 1952.2 115 1052.0 116 1052.0 117 4428.89 118 8 119 5309 6 128 5909.6 121 6413.1 122 6563 7 123 6563.7 124 6650 9 125 6650 2 126 5.1014 127 5.1014 123 6656 0 129 6656.8 138 6656 0 131 6656 8 132 9 133 9 134 6656 0 133 6656.8 136 6656 @ 137 cése 0 133 6656.8 139 6334.7 14Q 321.29 141 6334.7 142 321.29 143 6656 8 144 6656 .8 145 6656.0 146 6656 8 147 6656.8 148 6656.0 149 6656 0 150 . 45551E-82 151 -43551E-02 152 45551E-82 153 -45551E-82 134 .453551E-82 155 .45551E-82 156 .45531E-82 157 29.517 158 29.517 159 29.517 PRESSURE PSIA STAT ~SRBIaAasI ee Soowow MWS ge ASNAKOUMEN et ee Wintt hy eee wae Pou AMOH 148 151 154 157 © STN O'S 'D'D BSNONN Rotetetne Winer ierte Saeegeugeees estononwow TP wou ve fe 4 3233455 S888833 fh “NI ‘oO ee Dee OU o 2 24730 213.62 106 47 15.920 110.60 14.755 213.62 14.755 14.753 14.755 14.755 14.755 588 08 SS2 72 SS2.ee =. Un 149 TABLE F-5 PRESSURE FOR CONFIGURATION XC45 n + UONS GANT fete ae ee et iho fy fy eee ee SO ISR DOUNWANONS - w F-7 RRSnSSe spose Soonvvws ry gs aaaagee 238888888088 a RLASHS wet n ft 115.00 150 SRASQRASANBRASAE pene aan {11 117 128 123 126 129 132 135 133 141 144 147 153 156 159 - - a 598. RAGANN otsssa a USSESBBSBSSZamnofLananvww PECL piers Rae ae: oo Spee ee me) TABLE F-6 ENTHALPY FOR CONFIGURATION XC45 ENTHALPY BTU/LB STAT STAT | sTaT 1 124.03 2 124.83 3 8 4 26 3985 5 124.83 6 112.61 ? 112.61 3 119.17 3 160.30 10 126 86 11 161 00 12 126.86 13 160 25 14 126.86 15 168.75 16 126.86 17 160 81 18 136.09 19 165.84 20 165.84 21 1168.9 22 1160 0 23 1168 0 24 1168.0 25 862.86 26 262 86 2? 1160.8 28 228 36 29 63 961 24 69.065 31 123.12 32 123.12 33 123.12 34 123.12 3 1163.3 36 218.86 37 218.86 38 225.32 39 309.12 40 309.12 41 309.12 42 389.12 43 389.12 44 WI 12 45 389.12 46 616.93 ? 616 93 48 616.93 49 616.93 sO 1178.4 S51 616.93 S2 616.93 53 1178.4 54 1178.4 55 1170.4 56 124.83 57 289.32 58 295.28 s9 882.59 68 802.59 61 802.59 62 802.59 63 1178.4 64 1375.4 65 1491.8 66 1491.8 67 1234.8 68 1234.8 69 1234.8 7e 389.12 71 389.12 v2 389.12 73 1189.4 74 1189.4 7 89.4 76 1169.4 7? 1189.4 78 1.46 73 313.78 CS] 308.87 81 293.92 g2 338.66 63 298.98 a4 291.94 6s 291.94 Bs 291.94 87 255.81 88 196.45 39 198.81 90 196.31 a1 1ir @3 32 183.83 93 271.23 94 547.41 35 463.12 96 463.12 7 1294 2 33 1243.6 99 1243.6 199 1243.6 101 971.26 102 984.13 192 309.12 104 1204.3 105 905.37 106 932 04 107 1243.6 108 978.47 163 978.47 110 930.18 M11 309.12 112 309 12 113 965 32 114 76.584 115 26 922 116 26.922 117 463.12 118 463.12 119 307 8a 128 648.17 121 539.04 122 451 37 123 451.87 124 340.11 125 340.11 1265 128.98 127 123 90 128 329.95 129 341.35 139 341 35 131 341 25 132 341.35 133 463 12 134 341.35 135 341.35 136 314.72 137 314 72 138 267.29 139 267 .29 140 267.29 141 225.30 142 217.88 143 224.95 144 202.72 145 189.48 146 174.34 147 174.34 148 124.56 149 124.56 150 802.59 151 882.59 152 802.59 153 802.59 154 882.59 155 817.55 156 817.55 157 117.68 158 106.88 159 186.80 F-8 6-5 Table F-7 TURBOMACHINERY SUMMARY FOR CONFIGURATION XC45 Stat. Flow Rate Pressure Temperature Enthalpy Total Power No. LB/SEC PSIA cE BTU/LB HP Combustion Turbine 1,021,460 Comp. Inlet 56 6229.9 14.55 59.0 124.0 Comp. Disch. 58 5468.7 213.63 155-5) 295.3 Turb. Exhaust 124 6650.9 15.92 908.0 340.1 Steam Turbine 697,005 HPST Inlet 66 864.1 1450.0 1000.0 1491.8 MPST Inlet 104 veya 460.0 458.5 1204.3 LPST Inlet 99 & 107 1015.6 115.0 434.2 1243.6 Expander Turbine 56,562 Inlet 94 440.9 527.6 1000.0 547.4 Exit 95 440.9 263.6 790.6 463.1 Recycle Comp. NOT APPLICABLE Air Compressor AIR COMPRESSOR POWER CONSUMPTION WAS TAKEN FROM 142,083 for O02 Plant REFERENCE 2 AND ADJUSTED FOR COMPRESSOR EFFICIENCY OF 90% 0, Compressor 58,541 Inlet 8 194.4 17.0 90.0 119.2 Exit 19 194.4 734.7 300.0 165.8 Table F-8 HEAT EXCHANGER HEAT LOADS (MMBTU/HR) FOR CONFIGURATION XC45 Fuel Gas Stream High Pressure Boiler High Pressure Steam Superheater Fuel Gas Reheater Low Pressure Boiler Scrubber Precooler Resaturator Feedwater Heater Clean-up Precooler Combustion Turbine Exhaust Stream High Pressure Steam Superheater High Pressure Boiler High Pressure Economizer Low Pressure Steam Superheater Low Pressure Boiler Low Pressure Economizer Deaerator 585 21 LL 637 1136. -58 57. 532. -36 957 317 362. «23 362. 438. -L3 -39 292. 204. -65 12 35 99 71 -98 50 15 56 56 F-10 APPENDIX G CONFIGURATION XC45A DESCRIPTION - (See Figures G-1, C-1) Configuration XC45A uses a 1200 psi gasifier pressure as contrasted to the 588 psi gasifier of XC45. This higher pressure increases the power output of the expander turbine and increases overall efficiency. LIBRARY COPY Alaska Power Authority 334 W. 5th Ave. Anchorage, Alaska 99501 Wi o | Ti ° = oO cc had > ° = Li ce - o a °o a 0# (NO RECYCLE) OXYGEN PLANT t AIR RECYCLE COMP. WATER Y WATER WATER WATER 50# 2400F esl Gouna ne 400F WATER 4 +, 421F| 1596F j 1081 4s0F | | 4iaF scrus (4 4454 4244 Ll a 1 aceka Lp 525# 6304 294F | ea | 120F BOIL | goog | BOIL RESAT- tA rrr URATOR} 79¢ my CONFIGS XC45 44st 450 — 352F 46 WATER 101F | CONDENSATE 298 # 4454 | 1000F SULFUR 109F RECOVERY 1450P 5344 101F 1000F 115P | 8587 Aer aar CONDENSER ~< SULFUR — 109F 532F 1042# 1478 THROTTLE 337F PSIG PROC. 338F 234 335F 253F 35 SOF a5F 17148 py VW esa HP. HP. LP. THROTTLE Enon BOIL ECON. BOIL . 632F 465F a WW TTY 378F — 338F \ Pol LEGEND i SH. # FLOWIN LBS/SEC yw F TEMPERATURE — °F FLASH GAS ae 43eF P PRESSURE IN PSIA Figure G-1 Texaco Oxygen Blown Gasifier/FT4 Gas Turbine Combined Cycle Configuration No. XC45A, 7 Thermal = 37.3% Table G-1 PERFORMANCE SUMMARY FOR CONFIGURATION XC45A Combustor Exit Temperature Fuel Gas Reheater Temperature - Raw Gas Inlet Clean Gas Outlet Steam Temperature - Superheat/Reheat Coal Rate Thermal Efficiency Heat Rate Combustion Turbine Power Steam Turbine Fuel Gas Expander Power Air Compressor Power Oxygen Compressor Power Recycle Compressor Power Utility Power (Including Steam Auxiliaries) Net Power Coal Heating Value Clean Gas (Higher) Heating Value Combustion Turbine Overall Pressure Ratio Steam Turbine Pressures - High Pressure Low Pressure 1984°F 1081°F 1000°F 1000°F/NRH 10,000 tons/day 37.3% 9147 BIU/KW-HR 743.3 MW 508.3 MW 72.1 MW (-)105.3 MW (-)48.7 MW (-)0.0 MW (-)55.0 MW 1114.7 MW 12,235 BTU/LB 256.9 BTU/SCF 14.68 1450.0 psia 115.0 psia Table G-2 ENERGY BALANCE FOR CONFIGURATION XC45A DATUM = 59°F Sensible Latent Chemical Heat Total - Heat Heat Energy Losses Heat Total Feed Feeds MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR ib Coal +195 +10196.0 +10215.5 Slurry +32...1 +32).1 Feeds Subtotal +10247.6 +100.00 Net Electrical Output -3967.3 -38.71 Condenser -28.6 -3330.6 -3359.2 -32.78 Stack Gas -1138.6 - 697.9 =32).8 -1869.3 -18.24 Gas Cleanup =I =0.5 -205.4 -122.6 -329.6 =3.22 Ash Loss -52.5 -52.5 -0.51 Generator Losses -62.1 -62.1 -0.61 Net Process Heat Loss -82.6 -82.6 -0.81 Miscellaneous Losses (1) -6.1 -6.1 -0.06 Resaturator -38.0 -38.0 0.37 O2 Plant -490.8 -490.8 -4.79 TOTAL -9.8 -0.10 PERCENT IMBALANCE = 2i8-Qout ~ _9 10% Qin (1) Primary bleed flow from the pressure compressor exit used for turbine cooling and ducted overboard. 6-4 TEMPERATURES °R uw ~ aG-.a—zD PS mo ee eA om Sek faaipe paso es nee ee WW Nee BS-IB=— GASH é 519 519 520 535 rie 7s? Ou nO no oG 35 og ag ZU 0 2083 § S19 -a> a g gs a 3 709. a R ~ a > n USSaN GRRE BS822 senses a @ agoea8 ANDI owes Zog3 was a= omy NN & a0 eo? 93 ANNAN “NM o-1 @ ONeo TABLE G-3 TEMPERATURE FOR CONFIGURATION XC45A wn = ONS Kw J Rue ee RAVARASAVSTS 'D'D'DD DUN'D ee 200 a= Alia ty by Dy ee ee DPRK DANWAWS* 4 513.00 519.00 550.00 738 54 535.00 327 > es ¢ 759 99 ecen O 2029 5 561 90 ' ¢ ° ‘4 +> 4710s —— OD tw 4 O« nen iN wD an R goad QSL AeN38N UR SURESMIIT > Gd Z u83B RSS ow 313 Si DAG ma TD DOWW (AMINO LIND on'0oo meh NOOO aN NUNN ON OWS “OR ee ee eee =—MONMAWNWW Te NOMA NO $20. & Mtoty B38 Sa G-5 158 pocaeee BH BERSRASVRAFaw on 85 123 126 129 132 135 138 141 144 147 153 156 159 1051 g- daha gas iga9- aBBRBXoIBdonvdni BESKS Noooa HIses “ N un 3 ad 3R8 a a 1373.7 1373.7 1373. 1091. 926.06 837 .62 717.42 2055.7 2055.7 2089.5 520.08 “IN TABLE G-4 MASS FLOW FOR CONFIGURATION XC45A MASS FLOW LBS/SEC w = Sees 'OBiad a ee tors fo ae ests nw ae ~ he fe ee ee ABwwict New CANS IA GR 467 82 $50.30 eopcos 1841.4 .o 6392.5 6535 3 S 114 6540.8 8 6540.3 6238.1 318.74 6548.8 6540.8 149 45436E-02 .45436E-02 29.512 uy = > GHATS Pohl fy ee RMON NS wre Dd WWl lite ee RG DK GBM GMWIS AS qd 6109.7 6109 7 133 133 193. 193 193 452 75 267 40E-07 141 4 938.19 103.23 1041.4 971.47 13.095 560 51 560 51 297 86 6109.7 4 Le ee et ee MAAR ofasand3 erage waa Dh ae _ ™“M RA aD wos 4£o0u a oI o e - tod 2 3 1913 1041 1941 $3803 6450 6575 6540 5540 6540 E540 31Q 74 6540 8 6540 3 6540 3 45436E-02 .43436E-82 29.512 a 5) CoM coco Mh Bee 150 Bernt B000008 Be SASS LOGE SUNHAUNAGUAU 45. TABLE G-5 PRESSURE FOR CONFIGURATION XC45A PRESSURE PSIA STAT STAT STAT 1 14.696 2 14.696 3 4 14.696 5 14 696 6 v 17 ave 3 17.900 9 19 33 993 11 62 974 12 12 114 34 14 111.94 15 16 203.90 17 335.87 18 19 1335 9 28 1289.09 21 <2 1200 9 23 1200 9 24 Zs 1280 @ 26 1290.9 it ee 1.2279 <9 1 2279 2a 31 30.900 32 39 9D 33 34 30.060 35 20 000 36 37 30 900 32 1713.9 39 40 1713 9 41 1713.9 42 43 1713.9 44 1713.9 45 46 1542.6 4? 1542.6 48 49 1458.8 50 1450 .@ S1 32 1542.6 33 1450.8 S4 35 1450.0 Sé 14.549 3? $8 213.63 $9 1208.8 ] 61 1200.8 62 1200.8 63 64 1438.9 63 1438.0 66 67 115.08 68 115.08 69 72 113.00 71 115.08 72 73 115.00 74 115.00 73 76 113.00 77 113.09 7 7 4157.3 se 1157.1 81 62 tise i 83 1157.1 a4 6 157.1 86 1157.1 87 30 1125.5 33 1126.5 98 1 11c6.5 32 1104.1 93 44 Ws & 95 263 63 96 sie 15 a9 33 115.00 99 yy 115 90 101 1.2279 102 List 4ed iny 194 460 60 163 dim, 1 3 17 115 00 108 Lene 1.2 11g 1 2279 11 112 1 113 1.2279 114 116 4°70 116 .24729 tiv. liz 213 6% 119 213.62 128 te) 10 63 122 5) 743 123 ted 1% 2133 15 14 620 126 Ar 119 68 123 14 754 129 138 14.744 131 14754 iz 133 213 63 134 14.754 135 136 14 754 137 14 754 133 139 14 754 140 14 754 141 142 14.754 143 14754 144 145 14 75 146 14.754 147 i4g 14 754 149 14.754 150 151 1203 8 152 1208 9 153 154 1128.6 155 1208 8 156 157 1126.5 158 1126.5 159 73.094 ome aN Ganges : auSngaay ISS as SES SSF ouH eH e888o000 SRGRaNsad ey wo 9.616 - en eee et ee Le) ee ee te ef) et De oa Ot oe ee Ok oe On SBMKOLsssspit TABLE G-6 ENTHALPY FOR CONFIGURATION XC45A ENTHALPY BTU/LB STAl STAT STAT 1 124 03 2 124 93 3 8 4 £6. 385 5 124 02 6 112.61 ic 112.41 3 LIS Zt 9 168.30 1a 126 ..86é 11 161 oa 12 126.86 12 160.34 14 126.36 15 160.75 16 126.26 17 160 31 18 183.96 19 165.984 20 165 34 21 1159.7 ee 1159 7 ed 1159) 7 24 1159.7 25 855.41 26 355.41 7 1159.7 28 223 36 <9 63 961 2a 69.065 31 114.94 32 114.94 33 114.94 34 114.94 35 1162 3 36 218.86 37 218.86 33 223.32 39 389.12 49 309.12 41 389 12 42 309.12 43 309.12 44 389 12 435 309.12 46 616.93 47 616.93 48 616.93 49 616.93 50 1178.4 Si 616.93 S2 616.93 53 1178 4 34 1170.4 SS 1170.4 56° 124.03 iT 209.32 58 295.28 $9 796.39 68 796.39 61 796 .39 62 796.39 63 1170.4 64 1376.1 65 1491.8 66 1491.8 67 1234.8 68 1234.8 69 1234.8 78 389.12 71 389.12 v2 309.12 73 1189.4 74 1189.4 75 1189.4 76 1169.4 7? 1169.4 78 577.46 7 32? .26 88 313.58 81 313.58 82 35? .66 83 319.5 ad 313.52 8S 313.52 86 313.52 87 268.18 88 195.64 39 120.05 99 199.55 91 116.493 92 133.11 93 293.35 94 548 11 35 39172 96 391.72 Si 1234 2 33 1243.5 99 1243.5 16909 1243.5 101 971519 182 984.07 193 209 12 104 1204 3 185 995.37 1b 331 86 107 1243.5 183 978.34 19:3 978.34 119 979.96 111 309.12 112 399 12 113 964 95 114 76.584 15 26 922 116 26 922 117 391.72 11s 391.72 119 302.63 126 649.13 tel $40.15 122 453 13 123 453.13 14 240.73 125 240 73 126 128.98 127 123 9A 123 340 87 129 342.00 13a 342 00 131 342 00 132 342.00 123 391 72 134 342.00 135 342.08 136 315 00 13V 215 a0 138 267 57 139 267.57 140 267 S7 141 225.16 142 218.09 143 224 83 144 202.98 145 189 56 146 73 02 147. 173.82 148 124.66 149 124.66 150 796.39 151 796 39 152 796 39 153 796.39 154 796 39 155 911.21 156 811.21 157 116.49 158 106.26 159 106.26 G-8 6-9 Table G-7 TURBOMACHINERY SUMMARY FOR CONFIGURATION XC45A Stat. Flow Rate Pressure Temperature Enthalpy Total Power No. LB/SEC PSIA °F BTU/LB HP Combustion Turbine 1,009,550 Comp. Inlet 56 6109.7 14.55 59.0 124.0 Comp. Disch. 58 5363.1 213.63 TODO 295.3 Turb. Exhaust 124 6535.8 15492 908.9 340.7 Steam Turbine 691,018 HPST Inlet 66 858.4 1450.0 1000.0 1491.8 MPST Inlet 104 1351 460.0 458.5 1204.3 LPST Inlet 99 & 107 1005.0 115.0 433.9 1243.5 Expander Turbine 98,574 Inlet 94 445.6 1076.6 1000.0 548.1 Exit 95 445.6 263.6 608.5 391.7 Recycle Comp. NOT APPLICABLE Air Compressor AIR COMPRESSOR POWER CONSUMPTION WAS TAKEN FROM 141,123 for O, Plant REFERENCE 2 AND ADJUSTED FOR COMPRESSOR EFFICIENCY OF 90% 0, Compressor 65,293 Inlet 8 193.1 17.0 90.0 1193 Exit 19 193) 1335\.7 300.0 165.8 Table G-8 HEAT EXCHANGER HEAT LOADS (MMBTU/HR) FOR CONFIGURATION XC45A Fuel Gas Stream High Pressure Boiler 593.45 High Pressure Steam Superheater 357.62 Fuel Gas Reheater 408.69 Low Pressure Boiler 22235) Resaturator 349.73 Feedwater Heater 171.99 Clean-up Precooler 8.54 Combustion Turbine Exhaust Stream High Pressure Steam Superheater 635.77 High Pressure Boiler 1116.74 High Pressure Economizer 951.16 Low Pressure Steam Superheater 55.35 Low Pressure Boiler 516.22 Low Pressure Economizer 314.19 Deaerator 389.59 G-10 APPENDIX H CONFIGURATION XC53 DESCRIPTION - (See Figure H-1) The higher turbine exhaust temperature associated with the higher turbine inlet temperature (2300°F) increases the optimum boiler pressure to 1800 psi. With this pressure it is necessary to reheat the steam in order to maintain adequate steam quality at the low pressure end of the steam turbine. Accordingly, Configuration XC53 differs from Configuration XC46 in that a steam reheater has been added in parallel with the superheater in the gas turbine exhaust. In the fuel gas stream the steam superheater was removed (final superheat is ac- complished in the gas turbine exhaust) and a recycle reheater was added. The heat removed in the scrubber precooler is transferred to the process feedwater. The heat removed during turbine cooling is rejected in the condenser. The effi- ciency of this configuration could be increased about one-half percentage point by generating low pressure steam for expansion through a turbine using the heat rejected from the turbine cooling air and the turbine cooling water. Additional 115 PSIA steam could also be generated in the HRSG. c-H WATER } | QUENCH 76 WATER WATER WATER RECYCLE HEATER 70 515F n up CLEAN) . cooL 632F 559F BOIL 137 up 69 81 300F 83 i oF 105F Has To ! raat EROCERS SULFUR PLANT RECOVERY PROCESS| = 5# t GAS sal FROM CONDENSER ¥ An EXPAND. s4op REHEAT SULFUR 781F att TOPROCESS 106 109F q x 4 496r 60 | 58 137 WATER 8 | 204P aT] “ THROTTLE ee 1654 ! 16 ar CONDENSATE eee : 98 AIR 1096F 153# | 34 9264 re eu) 486F DEAER — TT 155# ‘1 124 4 3 TCA 705F 4 = : 2M | sor % 621F 21 F 61 486F 1 COOLER (Wh 86 FLASH sH He. r He. LP. Le 654 GAS A ste oz | 28! aeie ECON. szor | BOIL ECON.| 4,,¢~—FEED PUMP 102 = wor [YY | Tis 127 128 maser YY Fa30 in| me 1102F 9274 in 3194 ae pe FROM LEGEND it nn | 125 37 PROCESS # FLOWIN LBS/SEC WATER 115 1654 in F TEMPERATURE — °F 114 707F P PRESSURE IN PSIA TO COND. { TO LPT 1000F Figure H-1 Texaco Oxygen Blown Gasifier/Advanced Combustion Turbine (2300°F CET) Combined Cycle Configuration No. XC53, 7 Thermal = 39% Table H-1 PERFORMANCE SUMMARY FOR CONFIGURATION XC53 Combustor Exit Temperature Fuel Gas Reheater Temperature - Raw Gas Inlet Clean Gas Outlet Steam Temperature - Superheat/Reheat Coal Rate Thermal Efficiency Heat Rate Combustion Turbine Power Steam Turbine Fuel Gas Expander Power Air Compressor Power Oxygen Compressor Power Recycle Compressor Power Utility Power (Including Steam Auxiliaries) Net Power Coal Heating Value Clean Gas (Higher) Heating Value Combustion Turbine Overall Pressure Ratio Steam Turbine Pressures - High Pressure Low Pressure 2300°F 1087.4°F 1000.0°F 1000°F/1000°F 10,000 tons/day 39.0% 8746 BTU/KW-HR 814.6 MW 905-3 MW: 40.9 MW (-)106.0 MW (-)43.7 MW (-)0.8 MW (-)52.4 MW 1167.9 MW 12,235 BIU/LB 255.2 BTU/SCF 14.0 1800.0 psia 540.0 psia H-3 Table H-2 ENERGY BALANCE FOR CONFIGURATION XC53 DATUM = 59°F Sensible Latent Chemical Heat Total % Heat Heat Energy Losses Heat Total Feed Feeds MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR % Coal 19.5 10196 10215.5 99.7 Slurry 32.1 32.1 0.3 Feeds Subtotal 10247.6 100.00 Net Electrical Output -4131.5 -40.32 Condenser -26.6 3305.6 3332.0 32.51, Stack Gas -1028.3 696.8 32.7 1757.9 -17.15 Gas Cleanup 1.1 1.0 204.2 13:9 -328.5 3021 Ash Loss 52.6 -52.6 -0.51 Generator Losses 63.7 -63.7 -0.62 Net Process Heat Loss 90.3 -90.3 -0.88 Miscellaneous Losses (1) 20.5 20.5 -0.20 Resaturator 27.9 -27.9 -0.20 O02 Plant 472.5 472.5 -4.62 TOTAL 29.8 -.29 PERCENT IMBALANCE = 2i8-Qout ~ _ 999 Qin (1) Primary bleed flow from the pressure compressor exit used for turbine cooling and ducted overboard. H-4 TABLE H-3 TEMPERATURE FOR CONFIGURATION XC53 TEMPERATURES °R OTOP POONA OA See e” - OS NR NO RNS ae i ee WOO eet | HNO OEM © CNN -NO@ -mM- so. . NOMMK HOV -HONNCMNOBDONN DO AON AND ETOR “NWWWOOWWOONO OND ANMOMNDOMOMM At FOOT FE HOO HOOT ATRNUONN THT TONDOOUAN MOK PK DANNNDVOR DAHA RAAN AH MOON HRN AON SAN MM MINIMAL NIOIID o © Pe) = IMONNNORTROMOANNO=@TROMUN QMWHNAUNOAKTROMOHDN Onw = SEAN ANMMMM SS SDD OODORRDOSDAAAAGOO — HAAG POSES: n RAMA tated 135 DOONOHNVO TOR TL AH AHQDOMM HOH HOAHROQOOh AAMTMMOWDOVOAN A UMS BDOONOOD - -QDWOMAN : : AHN -WOnSOG -- °° os RNR NO WMS Cr Om me Mt OMRON EHMO = Mss AKHDOMNN- DOM -ODNOVNMNM DOOM NOAATAUNNOSK “ONNNONK ODOM O—= OS AANMOMODOVOOM HF LOOC ST TT HOO AH TOON ~DONONNN A TFOONHODAMY DOOR ORL NANN OR DDN SRSA ANDIN BM OORINRIN AN tt tt I IIN MONON IOID 2 nN re) = IEANPOHKTFROMOHNUNOKTROMUOHNNNOAKATROMOANNOKTROMUNNNO. hon Be MMANANNMMM SST TONNNOGOORAADDDONANHQOVOAM MMMM Mes oo PMS tet tte ttt 134 DOOOWM OVO TOL TTO=—MOONMM An TONOBON D-H - MTOMODOODOHNSOOKRNS SOOQVO®D - -POWOOMNA = + = -AAN -OMeS@ :---- @:- RN: --OOne aera ane aL OM ete NR OOM UENO = Me ee AKHONN NWONHONOMNDOBON NTN TNNOST T—-NNNONKRODHNNNONMOIN IAM DOO OOM OM at MF OOOA ST FT FINN MOMMA NODDNOON NA TOON ANDO WNONOOKL DORN ANN ORAR DH HHA HDHD HKHODNN SHUN SANSA ANAM ONKRINID e THTROMUNNNOM TR OMVANN OK TROMYDNN OAK TREMUDNNDOMATNROMODN e NAM MM TEE TDN PODOORKAARRVDGANHOOBVO A HAAN MMS wo tt td et ed Ot H-5 TABLE H-4 DAAAHVUTONMANTONTEFANATOOTAVAUANMANNNABDNAAAHAAAA wo QONMM ONL NONOHNHDN AON DONA ANDANMONMODNUOWAUNANAN cD WOODOONA «vet Ret COCO + G0 A AHOWD re Det ss fe fa fa Rane tec PPROSNO WO OR PEO ON ONS» PRN OO : WANNODOO AN OO te 29 EMonuNeneram WNDNNO MSR OMONNIND< FP OMUNNNOMKTROMOHN Onw MMMANAUNMMMM ST TOWMOODOORRADOONANHVOOKMKMNUNNNM MrT MMMM ttt ete & te a Mm 3 “= ~ Zz a a ANANAAAHAHNNOOMAHMFOPM FAUT CMMAANUAUAHAMNNHDOOOWNWAN Ww- RRBBOMOONODOONAN OUD OLANOLOAANNMMONININDOGUOUN Box a SOD ODOOAD Oat et ON etOON HORINID MA -- ss DO OPS ONE er ONO er ww™ NOONIOIN Onn Oar o- MMRRANRNOMW -NOVSNKRNDG «MUN THADOOMMNAANOOQNDONMANNA - DOMINOS at OMAN MO MON ODOOAAT MOO et MAAS a9 n S nN o n ea = : EUN OK TROMOHVNO HK TROMOANN OK TFROMOHNNOMTROMUONNNO. hom MR ANNNNM MMT ST TONNMOOORRROO ADHOOOaAAMAMAUNAM Mee = o AAA tte etetet 5 3 fe “_ n z Ni AAAAHDVNON AMAT SCUST TON TT TONG HAANAUAAANNABDOHOHDAAHWOWO— He NNN NOWMKH ODHDDADDAMNDN OR NANG VDWAANNMUBNINDHOVNONNNDO— + WWOOONS = es st WWD: OPRININ Oa od Site Bl RAG ed WN cre eree AN Soest CTIHetnnm...-:- TTO OCOhOhRTN -OnRLOHK MORKRKAEAAKOONNONNNRINN +s NNT HNDOOUMMNADHUONWONANAN sw - OD MMMMM NT ARS BONER SHA TORODDONT KOAAANMAAAT AID ATROMUONNNDATROMUOHANN OM TROMOD WNDAHTROMODNNOATROMUAN WANN MM PTs TNINVOURARKR DOONANNOVOSOAAAMANNMMMM Ss Me SRR MASS FLOW LBS/SEC STAT 1 1 1 1 é H-6 TABLE H-5 PRESSURE FOR CONFIGURATION XC53 PRESSURE PSIA WONMOKRD DOOOKNAN®OVOVDIVOS WOMNMMBONNGOVOARNNN ROO AOHKADD SOo0o PS DOVSSSBGSONAROVSSNNSSOON MINI RMS WOAH SOOnTSS Sa ee ee A CONN ARAN IN “DOC ‘N=OOO RRAAGR T= -O8 - -OOOON - asu n eoorn ; SENCOK AHH HSHOO A AHHOS OSS OOOH HN VNOAAN OAT TOOTS were AEMONMNNNMMMMN AAA OD OONNNNONAA ANAM AOOID amar + rs ~ ~ EMOAAINOU TL OMOAUN HTN SODA DAT OOD MID NOM ONY Omg SAH MANUAUNMMMM T STTMMMNDODOOORARADVOAANANGVOOAMAMMUNNNM Mest & A tt tet etet tented no Mm “ WWOT TROVWNOAOHKAK POOP OVOVSHOVDONVONMTONNNGAARMOMMNM O29 ANOKV-HWOSOSNMOOOD :- - - - DWASVQVDOWVDAIOMANHAOMSWNNN Sen Nh ini ene WOON eles By fe N@OSOnr eos 9 3 ee es ets > -NNRRE ~- SENMANHDHO “S9SrTOoOSSSt SOSOHMOOANT OTT TOT s otPTN MND vr _ EAMMORWLNOMOAUIS aT amonuiye BATROMODNNOATROMUANNO= KOM MM ANNNNMMM ST TTNONNMOOORARADODDONANHGOOKRAAMNNNM MIT & Ce vv ™ CWVOM IF OK. DOHOVOHSHKL-L/HMVOOVOOVDOOS COSCO SOM HM VON KN VOAARN MN NM DAD NAHODAHHOSSK-OOD.-- SOS SOSA SSSA SOOO SSR Teese WWON ss es NOOBORKRTOO - e e e e e ‘AN NNARRE si b ED OI ID 1D A PD POD NN tt OOO OID IID IDI N tD tt OD tt tt tt tt LD EAT NOM WNAINY Eh BOON NID EP MON MIND eat BLOT UN OT BION CY MAMAN AMM TEE TOM NYOURAARADDODHDAHVOOO ae A QVUAMMMM st S A H-7 TABLE H-6 ENTHALPY FOR CONFIGURATION XC53 ENTHALPY BUT/LB =DOMONNOMWOWORAARETNRROTMAROMR YON TNOOVOTMMRYTOM WONW WMOND = -HWOINNOR Re. ORR UPAR NNO MMOR Met ee Onn NK ope er OO Oi HO Me te wt 8 tt + OMAN 8 EH NOOCTOVONM “MMOH AHNUTOKHHOVNNOOORNOVWNAHDOV ON seems NV 0 OKONOM KH HHT THRROAMRA NOR MAHDAHTOWNMNHANAAUMNNMO-ToONMe\o A ONO OEE OE FNMA ID SLUM meet tes POA a mes n o MONVUNOATROMUHN Oar nROMw DWAAHHOOVOKKAMUNNNM Meer A tet MK MAUNNMMMM ST TTNNMMODODOORKANK STAT 135 MMR OOMKTNUD@NMOOKKRA TORR OMMOR DN HAMUWODONNOTTMARTAM cw ees ee oe . Gateerer. Y-aie foe eee ti i oes 3 : ANVAOVNDO A DONTHRKA OH ERA AUNOAHNAHKRDDVOHNDAH A MNOWOat OOOO Rt tt rt rt rt rt DOD mt OSE NO tt et PN Nt Ut POR TEMP ot ttt) Ct COP Ot o oa vt N _ — INNOHK TFHROMOKNNNOK FKROMUOKDNNNOK TROMUOHNNOKLTROMUHnNNO= Ron - MMI HNUNANNMMM ET TNNNMDOOORKRADDDOANAHVVOK KKM AUNNM Mer o PRA ttt ttt 134 MO BMOTOTNUDOMUOMARRATNRRONTARORANNONOHOUT IMRAN TOMOeTTM DLODOHDHO -OMOWM Dries . RN -FONNDORARMMNAAAD - - - -OMND -O@ BOO el etl se Le Oe t9) se GLEN Lee LS Le ee le se MAKRNW. : : OI WO NL RHVNOND HON NTHNUROOAMRA VON RANDOOOOHNODVNA KH MN DOR DNA RN SSSR M ND SHAN TOO St eR MM Nee ERRMM Ht - THT OMOANNNOAMTHAMVWANN VK TRAMYANNOATROMOANN HAH TROMUAY = PAA NAM MMS TT PTOMNDOOORARARADDNAHHOOVO KHANNA MMM ST ory te tt H-8 6-H Table H-7 TURBOMACHINERY SUMMARY FOR CONFIGURATION XC53 Stat. Flow Rate Pressure Temperature Enthalpy Total Power No. LB/SEC PSIA °F BTU/LB HP Combustion Turbine 1,106,318 Comp. Inlet 65 4338.4 14.59 59.0 124.0 Comp. Disch. 66 3864.5 204.20 7051 282.4 Turb. Exhaust 97 4785.8 15.96 1096.4 394.5 Steam Turbine 700,791 HPST Inlet 53) 799.8 1800.0 1000.0 1480.8 MPST Inlet 116 & 120 852.8 540.0 1000.0 1519).3 LPST Inlet N/A Expander Turbine 55,970 Inlet 84 447.4 527.6 1000.0 549.2 Exit 85 447.4 254.2 780.9 460.7 Recycle Comp. 1,035 Inlet 76 212.2) 567.0 352.0 296.3 Exit 99 2122 588.0 361.0 299.7 Air Compressor AIR COMPRESSOR POWER CONSUMPTION WAS TAKEN FROM 142 ,083 for 02 Plant REFERENCE 2 AND ADJUSTED FOR COMPRESSOR EFFICIENCY OF 90% 0, Compressor 58,578 Inlet 8 194.6 17.0 90.0 119.2 Exit 19 194.6 734.7 300.0 165.8 Table H-8 HEAT EXCHANGER HEAT LOADS (MMBTU/HR) FOR CONFIGURATION XC53 Fuel Gas Stream High Pressure Boiler High Pressure Steam Superheater Fuel Gas Reheater Low Pressure Boiler Resaturator Feedwater Heater Clean-up Precooler Combustion Turbine Exhaust Stream High Pressure Steam Superheater Steam Reheater High Pressure Boiler High Pressure Economizer Low Pressure Boiler Low Pressure Economizer Deaerator Turbine Cooling Air Cooler Turbine Cooling Water Cooler 795 68 41 320. 228. 10. 954. 499. 608. 548. 56. -44 -06 751 231 13% -62 247 -07 442. 26 -09 04 56 61 19 28 96 23 31 66 95 H-10 APPENDIX I CONFIGURATION ACO3 DESCRIPTION - (See Figure I-1) Configuration ACO3 differs from Configuration XC46 in that it has an air blown gasifier. The gasifier air stream is split off the gas turbine compressor dis- charge and is cooled in an intermediate pressure boiler before going through the boost compressor. The intermediate pressure steam reheater is in the fuel gas stream downstream of the high pressure boiler. LIBRARY COPY Alaska Power Authority ci! 334 W. 5th Ave. Anchorage, Alaska 99501 Ww o u u o = oO oc rm tay > © om ua cc - Sc =z °o Qa ol 4574 457# 333F 20 WATER a2F 1 18 aut =| QUENCH Aen. WATER % q COAL & ae 58 { WATER ” 1195F 862F 1 _478F 372F 13 ul ae 2400F 11794 [Hp : 70 375F 1200# [CLEAN 13104 on REHEAT| REGEN|85 = 119 BOILER! soar = 532F aor 33aF} BOIL | 117 iad pa 59 666F 11257 ar 83 800F 105F > 334 121 100PSIG 1167 | ios 2ay PROCESS CONDENSATE STEAM PUMP SULFUR 1046# 234 RECOVERY 5 ap 115P 338F 794 15#_ FROM va 55 PROCESS CONDENSER < 1450P SULFUR 993# 458F{ 56 644 — IP. 458F 109F | 11724 11 756F BOIL 9934 27 800F oe IPT | 109F PROCESS STEAM 1 414P us 8434 CONDENSATE 154F 105# r read 54 To 154F 10464 380# me rag —_ SHOCEES 47 | 46 44 | a2 38 36 _458F 800F 592F 592F sa Bolt econ.| toeer i125 918F 778F 632F ‘ 5008 A 1125¢ Yt AA} + TACK 5412F 109 ~ | 110 11 rep) 22 ri 102 PUMP m zee 5a12# LEGEND # FLOW IN LBS/SEC F TEMPER ATURE — °F P PRESSURE IN PSIA Figure I-1 Texaco Air Blown Gasifier/FT4 Combustion Turbine Combined Cycle Configuration No. ACO3, 7 Thermal = 37.1% Table I-1 PERFORMANCE SUMMARY FOR CONFIGURATION ACO3 Combustor Exit Temperature 1984°F Fuel Gas Reheater Temperature - Raw Gas Inlet 862°F Clean Gas Outlet 800°F Steam Temperature - Superheat/Reheat 800°F/800°F Coal Rate 10,000 tons/day Thermal Efficiency 37.1% Heat Rate 9198 BTU/KW-HR Combustion Turbine Power 612.7 MW Steam Turbine 595.7 MW Fuel Gas Expander Power 76.2 MW Boost Compressor Power (-)118.2 MW Recycle Compressor Power (-)1.3 MW Utility Power (Including Steam Auxiliaries) (-)56.4 MW Net Power 1108.7 MW Coal Heating Value 12,235 BTU/LB Clean Gas (Higher) Heating Value 95.8 BTU/SCF Combustion Turbine Overall Pressure Ratio 14.68 Steam Turbine Pressures - High Pressure 1450.0 psia Low Pressure 414.0 psia I-3 Table I-2 ENERGY BALANCE FOR CONFIGURATION ACO3 DATUM = 59°F Sensible Latent Chemical Heat Total % Heat Heat Energy Losses Heat Total Feed Feeds MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR % Coal 19.5 10196 10215.5 99.69 Slurry Sze. o2e0 Sey Feeds Subtotal 10247.6 100.00 Net Electrical Output -3949.3 -38.54 Condenser Slee =3937 20 -3969.2 -38.73 Stack Gas -898.3 -851.4 -19.6 -1769.3 17. 27, Gas Cleanup aed -6.4 -228.7 -220.8 $457.1 -4.46 Ash Loss 5265 5200) =0 51 Generator Losses -64.9 -64.9 -0.63 Net Process Heat Loss =109.-3 -109.3 =1-07 Miscellaneous Losses (1) 520 -5.0 -0.05 Resaturator +91.0 +91.0 +0.89 TOTAL -38.0 -0.37 PERCENT IMBALANCE = 2i8-Qout _ 379 Qin (1) Primary bleed flow from the pressure compressor exit used for turbine cooling and ducted overboard. I-4 Table I-3 TEMPERATURE FOR CONFIGURATION ACO3 TEMPERATURES °R SOO eee ER OR SSE eno acacia VSTMO0 bet RAGROO: RO: a bacco soe ee ee —=-@ -- -@Oo :-:.-.-.- - Th + -cCOCn# DN KH HOOHOTMORKLDPWODDDDDUOMNNNMO Th -HrMN MANO OM OD et MOON SMH UU MMAUMETHA-“MUM DDN MHD VOR. ND Ht Ht OID SSN DOKR NN HN HOW et 969.24 MODNNOATN DMUOHDNUNNOHKTFROMUODNNNOKTPLOMUHNNOAT GMWH AM AUNAMMMM TT TNONMWOWWW) KRRRODDAHHNHNOOOK— NUNS Bate Stet STAT 117 SQOMGOM Ol OV OY OR AAAANSSO=SOQEovrorvarersns hROOMe Oo - - - ROWTMAOW - - STN ‘-ROOrOO® ‘HN SrKND - MNOS. sss Na “Oe ee ae ONR@ rosoncoeiey Ce hae DAM AOO AAOMONN DN TODOUOHDONONRINORKROARRKR aT WMODDE PANN FOOD OS SA HHO DOHN RH HAVO UMUNOO NAR MAMMMOAOONWW AY DWN HH HN ON DORK NH HAAN DAHNHDNNOR DD Hee Heh Or NININW on} a yn - INNOATFRAMODUN OH TROMODUNNOKTROMWOANNIORKTREAM AANWO e BRB NUANAM MMT ET TNNNNOOORRRODDOANHOSORA BUNNY a Met tetetetet wo edi a BONN VOH-OHh OMLMARK DARN SOOT NOABON AT AT TAAHN OLN OM® BOO» gMOWTITMON | NTT ORE CARNOO 2 ‘NOSORIN®G iW] Ome mcincaTplicaales ; Sane) sisuioeraye Ca NEG! LOOCOEDOD) sat ae DAH LOPAASVMMENNDHGWOHODONM ANH SROR SRR MMINSODDO® BRR NF OO 0 4 MODAN MAO UOOMMNDD NODOMOMMNNAUAHO AY PNDH ANON NOOR AKA KAAADAHND HOR ODK ANA OR AA ANDORRININW ROMOHANNVThROo HOMO SHAM UNAM = AA tt ttt E=TROMoAUOHaTreN MON SSSA NNNMMMS Set 35 79 73 T=5 Table I-4 MASS FLOW FOR CONFIGURATION ACO3 MASS FLOW LBS/SEC = eo NONWOVENS TORE RR WT MMM ADO co co'tcoc0 TTR WD ON ON OO NR - 1m > Ow -M RR OD Wan -ININLIN -DINMON : inwe Mannan: aaa Soro MANOOMRAUTUN (POTTS = NUNUNMURDWOOMN eeistetee ON OA DAMMGAHOAANHODOSOTCAAOD VON HH VNOMWHOT TNT TAMMY WMD A DOR NN IOS WUD IDI wo vv MUNNNOANTROMUOANNOM THRO MMMMTTTNNN MUOANNDATROMUONNNOKT BMH MANNY nN DOOAAHHOVOVO st NAN Rae Hee STAT 117 ss eo enn a ee ee ee eee Rie Bea ne ce MMONAKKRR ‘SUS TO TS 7 "UMMAMS AS SOOM ates 7 NNR MVAPAKHN AMM AMOOVHOOK VON TH TTOVOMANNVNAMTOTWN ST TOMMY WIT OND et St tt tt tt DOO NN tet ete TININ ID = IID FUN tee S411 MOK ThROMODNNNVKTROMOHNNIOANTROM ANNO Att tet NWOeKTrROMUHnN MOND STAT 116 ie Oo ON ae ODM MMO MOCO AOS S nn -™m “co a Ad ol I on ae es ef ® IN) ON INI -ININID (ONUMONM - -DWONKMMN- non hlen@ NNR NNT TIONT | MAUMUMMTNWDD KAR tetris ON DANY AA HOA MOOR AAN OH THOVONHK ANAT TST TT PTOM SY MUODNNIOMTHhO DOH ANNUM ttt ttt Serre eee WANPOKTKhO AAA NUNAMMM TTT FOMMNMDODOOR & 85 838 91 34 190 1-6 i=1 ee ANNE OODQWOVVODOOVNNNDANUUU Ae AAWUNNME EE SNSPOANGAUSNS-DANVGATSVRASANGATS ISH BAVGAISVaHD eae ee in nn an ete et ND TRV Os BC Ba Ba Be ee eA A Gal a CI Ne AD. ee LA LAONOMAAAAAAD: © =HARAHRAUNOOOS: DOHA WHA Ranma - - 1 O- DAANUNNNNNNNANSOOMUAH. «| maDMN: - NN NININAINIAG © OM | DONWOSOON: =: . OND DUYVQWQUMUATNUNNVDODUNOOVOOVNVNSOVV®. ©» | DHVODYVOWVOHO0OY DOVOQIUAUAANGS NAVAIPIPPODOVOVAIWDVIOWVOHVOOSVHOVIISSwWAAH _ = n eee ee ANN] — = SODWOVWWODDDONVVADNDUNUNUUEseWUUWNNNNRE FKFDUND NTNONEEOUNVHWOVER-DUNOANOVA-OUNYDVONA wun ~ > a Senor eer ranadgaae pea baer ee oman AA LALALLO-GANATADD. © -ADAHDRAANNSOS: DOD AH Narva YL ONRRNNNY RN BOSSSUIGa R= MMM SNM NVANNaNe 5 so GONWWODON: AH OWNN® AUMAADAGADAD®OVONNS2OO: vos QOVVOOOANHNWOW DOVVOD GTUUNUUVHANNAVWVVOSVV AGP VVIOVAWHDH VO SVSGVAOVIWWUAHK _ _ NN me eee NVOVM —K—VOCWOWOWOWDDOYVNNHAAHNHANMUS S BWANA ee VAND HLKeDUNOANTOVNA|-$DUNUWHWOVNAKDOUNUDUAWNSVNA-DOUNOHG _ _ nn Te me ee St te ee me CF Be NC TUT OT Bo De De ee ee I II CUT UID RO ee u- FS PHHAHHAHLADOVOHRADAA- oe DN D BB BN NDSD SDD ODO ro “ nn: NNNANO: °. ‘ . tw: a) SOOUNsss9: <-> = CA evuvoo aU AAS DA —-VGIHHOSO SOO SSSOHSHOVSSSSSSSTAoSR 2 a = VISd aaNssaud €O0OV NOILVENDIANOD YOA FANSsSaAdd S-I 9T9eL Table I-6 ENTHALPY FOR CONFIGURATION ACO3 ENTHALPY BTU/LB MOMHOHNMMOTEMTHNOTMOOON Tarn trat woos TON BAS OCH BO OOD ‘ 1s 5 ss RMPTRHWODWO— YOWWWO “WINDOW te Dt te COT TODO ee eee nin POHAAA HAODHCROODSAAMOOOAISUATTI Rarwn SOON QNA TH ANDAR MMRKAMANMNT ST TOMICMNT -NNMOTOTMR NATO HANNAN VSP ANS TO St Stet Pete NSS MTN MMM Ahan EMODAING ATK SMODNIND AS OMONUN Oa OMONAIOD i amon MM ANVUNAMMMM TT TNNMOWOOORLRODODAHDHAMNVOOK CUNNY LA a 117 MMDOOMOT MMU TMMTOTMOONNNOHOMNKOMDO WOWOOtTT TOHNN BDONUNVNWWOHWWD =F CODVOV * "+ + + - @ TOTHKWETOWOW WOWWOCO wD NN wig ¢ aio. sw No -m OS PHOT tt TTNNMO-—-M - "AWN NOOR ONON TOT OTAORN AGAR RNIP-O- OOO -r NAVDDNN AMOHDUKR UM eM UMUNTOOROMN ROHN TOTTIRADATOWS PRHNAMDNNVO RANT OW Set tee SUNN Se MATE MMMM eel n N - e EVM OM TRODMOKDNUN OH TROMOHNNYOKMTROMUHNNNVKTROM DNNOA— a AP ANAUNNMMM TT TNMNONOOORRRODDDONAHOQOH— HUNNM oO a 116 12IO WOM OO IMM OT TMT OLMON DAD LT AN IO TIN OT OOO ONE TMA Se Eee > FONT TIN TAOMOUOMOOOWEN «cI Sie OS SOME Od Ie EE IS le Sy ee ie eel eee mete ce ats Ri : SF KDI DANO MI MOANHON ONAL FT OOHNOAUINAN HORINRNRADANNOAG ‘Nh ANLODN AMMAN ACA MA AAMAUARNNDMMON sat IM TTOODNVA HOS ANN DNA DAA TOSS ER TONNM NN RR NMINM NM MMM ato EHLROMCOANOS ATL OMAN GAT OMONA STL SMONAINOAERS PSMA NNANMMM TTT TNMNMOOORARAARDOHDHNHAHDOOVOAAMUNNM & RRR ttt I-8 6-1 Table I-7 TURBOMACHINERY SUMMARY FOR CONFIGURATION ACO3 Stat. Flow Rate Pressure Temperature Enthalpy Total Power No. LB/SEC PSIA °F BTU/LB HP Combustion Turbine 832,093 Comp. Inlet 6 5137.2 14.55 59.0 124.0 Comp. Disch. 8 4509.5 213.63 755.5 295.3 Turb. Exhaust 94 5411.8 16.00 918.4 347.7 Steam Turbine 811,495 HPST Inlet 48 1045.5 1450.0 800.0 1366.9 MPST Inlet 62 & 66 1125.3 414.0 800.0 1416.6 LPST Inlet 120 23 «A 115.0 338.1 1189.4 Expander Turbine 104,260 Inlet 86 1283.6 527.56 800.0 369.1 Exit 87 1283.6 263.63 614.5 311.6 Recycle Comp. 1,768 Inlet 78 457.3 567.0 332.7 232.4 Exit 96 457.3 588.0 341.7 235.1 Air Boost Comp. to Gasifier 158,506 Inlet 12 992.5 213.63 559.0 245.7 Exit i3 992.5 662.00 1000.0 358.6 0, Compressor 58,578 Inlet N/A Exit N/A Table I-8 HEAT EXCHANGER HEAT LOADS (MMBTU/HR) FOR CONFIGURATION ACO3 Fuel Gas Stream High Pressure Boiler Steam Preheater Fuel Gas Reheater Low Pressure Boiler Resaturator Feedwater Heater Clean-up Precooler Combustion Turbine Exhaust Stream High Pressure Steam Superheater High Pressure Boiler High Pressure Economizer Low Pressure Boiler Low Pressure Economizer Deaerator 1326; 635 701 186. 730. 213. 22 739. 756. 666 177. 869 393): 67 -43 -05 82 33 75 45 83 71 -65 08* -05 10 *Heat Exchanger is in air stream to gasifier. I-10 APPENDIX J CONFIGURATION XC50 DESCRIPTION - (See Figure J-1) Due to its low temperature, the fuel gas stream leaving the BGC gasifier in Con- figuration XC50 does not transfer any heat directly to the bottoming cycle. Inter- mediate pressure steam is generated in the gasifier jacket. This steam is further heated by the gas turbine exhaust and is fed to the gasifier. High pressure steam (1000 PSIA) is generated and superheated in the combustion turbine exhaust. Low pressure steam generation and deaeration are also accom- plished in this exhaust stream. The efficiency of this configuration could be increased about one-half percentage point by improving the steam conditions. Steam pressure could be raised from 1000 to 1450 PSIA if the steam leaving the high pressure turbine were reheated to 800°F before expansion through the intermediate and low pressure turbines. (Alternatively supplemental combustion could be used to superheat steam to 1000°F which would allow the use of 1450 PSIA steam pressure. The effect on efficiency of increasing steam pressure from 800 to 1450 PSIA with the Texaco gasifier is shown as 0.8 points on Table 3-1. For these cases however significant quantities of steam were generated in the fuel gas cooler. With the BGC, there is less steam generation so the effect of steam pressure will be pro- portionately diminished.) cL COAL OXYGEN PLANT f AIR AIR SOF GASIF. JACKET 4544 QUENCH WATER { \ "1 12 13 199F 139F T] tose —=} NAA —iCLEAN-| 3834 373# 3714 |uP FWH cooL 16 508F 388¢ 70F 355 108F 4994 CONDEN- SATE PUMP — TO PROCESS 1002# | 108F 129F 8784 COND SULFUR RECOVERY SULFUR PROCESS COND. (pa oga 338F TO FUEL PROCESS 50 PSIG R 43# 54 100 PSIG. 1604 115P 323F 5g FROM PROCESS # FLOWIN LBS/SEC F TEMPERATURE — °F P PRESSURE IN PSIA Figure J-1 BGC Slagger Gasifier/FT-4 Combustion Turbine Combined Cycle Configuration No. XC50, 7 Thermal = 37.3% 336F 983# Lp, | 338F 336F | 1p. ele poi, | THROTTLE ECON ne 102 378F = 103 104 Tos” STACK 270F 7855# LEGEND Table J-1 PERFORMANCE SUMMARY FOR CONFIGURATION XC50 Combustor Exit Temperature Fuel Gas Reheater Temperature - Raw Gas Inlet Clean Gas Outlet Steam Temperature - Superheat/Reheat Coal Rate Thermal Efficiency Heat Rate Combustion Turbine Power Steam Turbine Fuel Gas Expander Power Air Compressor Power Oxygen Compressor Power Recycle Compressor Power Utility Power (Including Steam Auxiliaries) Net Power Coal Heating Value Clean Gas (Higher) Heating Value Combustion Turbine Overall Pressure Ratio Steam Turbine Pressures - High Pressure Intermediate Pressure Low Pressure 1984°F 282°F 242°F 800°F/NRH 10,000 tons/day 37.3% 9160 BTU/KW-HR 859.5 MW 382.9 MW - MW (-)58.0 MW (-)22.3 MW (>> Hw (-)46.8 MW 1115.3 MW 12,235 BTU/LB 342.5 BTU/SCF 14.68 1000.0 psi 460.0 psia 115.0 psia Table J-2 ENERGY BALANCE FOR CONFIGURATION XC50 DATUM = 59°F Sensible Latent Chemical Heat Total % Heat Heat Energy Losses Heat Total Feed Feeds MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR MMBTU/HR % Coal 19.5 10196 10215.5 100.0 Slurry 0 0 0 Feeds Subtotal 10215.5 100.0 Net Electrical Output -3958.4 -38.75 Condenser 0.4 2933.5 -2933.9 =28). 72 Stack Gas 1489.3 731.8 24.3 -2245.4 -22.00 Gas Cleanup * * Ash Loss * * Generator Losses -59.0 -0.57 Net Process Heat Loss -768.6* -7.52 Miscellaneous Losses (1) -7.3 -0.07 Resaturator -19.7 -0.19 O, Plant -245.8 -2.40 TOTAL -22.6 = 22 PERCENT IMBALANCE = Si8-Qout _ Qin * BGC gasifier was not modeled in detail. Input for the gasifier and the cleanup system was taken from Ref. 2 and is available from that document. (1) Primary bleed flow from the pressure compressor exit used for turbine cooling and ducted overboard. Table J-3 TEMPERATURE FOR CONFIGURATION XC50 TEMPERATURES °R SIOSSQVNOVO 6 oS PIS LTAUDNNOMMGEOWNHONON Moo COCOSS |: - -AMMMMMMANHANH - - OO Sha . ie =I<Gy os a st ie © el PAT ee tt INO = -O-: o + AKDAUANAADODGHOSGOMMONN SSL SH-DODGO POMS moon A AOD tt) D0 wt tt Ht OD I MIO NOOQUNAN HH MONUMMOOWO WMOKRDR DK KH BMMNNDORRARALAR DD HOR HAAN HHA AANN n = 2 EMONUNOUTROMOAUNO WSR OMoONVNO@Tr OMonVN =TK S SHI ANUAUNMMMM ST PTNNMMNMOOWOORLADODHHHAHDD tet tt 108 DOOVMOMNM OOOH HK MRAULVTONUNVARMMMMRAWONHORNN O90- BOSOOWO - = - -MMAHOTMAAHD - - = - - OO Ht RROM tee WOO ee ROOMMnM - ow - +: AKRAYDSNAGGGDHONNNSANOHNSSSSODGOOOOOO Geowa BVA TNMONOMMDOOONANN HDAH KH OOVOVOAAM KH MOVNHHKOWOOW POKROUNKDHKVAH NNW SERENE ctet MANN MMH OND onw o n xn EUMMOATROMENUNS “TL OMONUNS-ereMonwwoat eonwun MAMAN AUNNMMM EST TNNMNOOORLAADODDDANHSD wt & St ttt xn eo = SIPS DHOM OOK AADVOTORVARKNMMMUROOOTRKE OOVTOwow |- : -NMMOSOMONAD .... .- 0@ ee en No ae tse) SONT ONIN CO aoe we DWAIN ININWI et SMO. SRM DDO HH NUN OODOQVOAAVNAH KH TANOMN DORK NDOVNAANDNNNDNDER.O) wt wt wt RR DDD tO wt J Mm nN e Ie thOMUNNIDth ve NOK ThROM BDNND e ARRON MMMM n DONHHOID Drees oo St ttt J-5 Table J-4 MASS FLOW FOR CONFIGURATION XC50 MASS FLOW LBS/SEC ATOMMONMONNVNOORKRHAAMNMOTTTNORARATANMMM nwnwst OMMe i . NONNT ST (MO -HNOTTTTRE -oMwoNMn -. -. - ne + SOM owt. 9 PNe OD F888 9 | DREN Sonen dl MTMORONNON “AN ONMONAUOSOOMN * > -wINWw nro OHH NOL AK OOATOOMN GG TONOOCO TT HUSUNDODOOND HS EMMNORN AO mt mm ADNAN TRI TOON OND a = i WAND wth DAAQQ cere et ttt WANN Om ThOMON RM MNNNMM EM STAT 3 81 84 87 90 93 188 MMNUNUMATMHAOOD MAUHDMAATTTRMNARATITNOHRMM onwrn MOTO pO UTT NA CM Mere TORMSNMVMDONM — - oro ION A UN OO ON -O CMME “NN - AAAMAIWOLOHOOHE HeANM WOO INH wl ANS: TIMONDINNDAORKONDOO=H-DOOUNO=TANNH-HDO-O0T RRTMMMORRMAHOO “ANH RANINR RR ANTM RH BOMRADR AA = KR = e CANO HTROMOANNOAKTROMUODNNAATROMONANOAT OMWN eK MM ANANAMMM PTE TONNNOOORRROODOANHOO aes o aa Ra nt © = M DAKMTOTRCBO MAUHAAMNBOTTMHOWOnTTNAMM onNND - ONES NT RA CM HOOTTERMM MMI - - ONNO -_ NTOOe - NOM © - ses ee eee OMe “mM so D_ MNAQGOHMOAO | “MOAVOOOUMRINRM WI HINT TONSK OMAK NDOR@QAGMOAONNOON SHR ONH-OOMAONT RST MMRORR RAD CNSR RIROARRAINNSRR TROMRA CRO re) in 2 ~ b TATROMUONANIND AR TROAMUDANNOATRAMUNNINOATROEM NUNC & AANA MMM EES PTODMOUOORRRROADANHOO Osa no A SMe J-6 Table J-5 PRESSURE FOR CONFIGURATION XC50 PRESSURE PSIA oe sO S SSO WOSSOOIID DNNH ™ Nowe ee ee SOOO ESSS em ‘OOor 8 8 On tans OS OO Oe SSIS S. 'o EMONUINBU TR SHON WIND <5 ROMUDNNNO=“TrOMONnNIN =r MMANANNMMMM ST TTINNINDOWORAA.DDODAAHANADD te & et tte 168 WWOQWNMNMAMN®DVOO® QQNOVOANSOSSS PLS Senn AAD ANDOOSWONSNMNBOOO2SG :.:.: - - - 2a isan RN: WW ORO900 nNnNe®...- - “RRR NNO ‘OOow KN ANS - “INININMMMD - - - ees mo: SR COM RaGaMIN - “oor PFINAHMHN HK HOOT OVONHNOVOH A ADOWWNO A ATOOKTTITO - SEM MANN Ra SM MMO OM at SE EE IE EE ID ttt nm ~ “ Eqnonsren 26 29 MIMO MNMMTTTN RRDODDOHNHDHOQD setts o St tte i oe _ WWSVDOOTF HR NHOVMDDOVDONHNG POAHDOVOOMVOOMNN BHM ANDOVONOHRMKR OOOO :-:- - - - BSooovooopnwy CRN ww 'O -ArK-NOS® SOSH ONS ae . sie ANN FSR e SBR ee ee eee eee ene B nn in = - ZH TROMUODNNOAMTROMUANINVARTROMUNNIORMTNROM NAAN in MSA ANA NM MMS SSE TNOINMOWOORARA DOONNHOD Orica AS Steet NYO FROM ANNNOKTROMURNNOMT BMW MNNWWY J-7 Table J-6 ENTHALPY FOR CONFIGURATION SC50 ENTHALPY BTU/LB SOMRESLVBANMMEYAQuMMAMoonrcoOOnM—esn Won BDWHNO=-MAMOO - NINN OOM... - - ces AON ad a tas te mie wwnMnm..--.- PRIS ISO Oe . rr) WTOTHRAND - (DONOMOAAR CO HODOHOO ONO =n BOOHKVOUATMVV OHH DONDHOMM TMMMNVNAHAMNNM AA MMND NNN AN EM NRE St DNON NMS PIN et tet te PT MN tt ON ° ° N - EMONANHT Th OMONMOS R= SH OM onUNS= sh OMonwwo ath MHI ANNUUM MMME TTMOMMOWOUOORKRARODOAAMRHAOD rere 5 aa ae @ =) ed NUT UVDOOUNANDONTOANAMAHOOOTTAUNMAOMM WRAD DOWODMONOHAODO -ACOMAAOO - - . s =O ALM MMO {he Se =e lew DON nn RG RR PO “Oo SPOON TODD Ran OMAnKnORRRMRDOANOWON@ NaN ANRMNOHMMM OR AOUVARDBIOM TF AMMM HAMMAM HK TATONIW MANN MNNMM PDI NTS es et tO I NIN @ = = = ENNOAMTROMONUNOATROMUANNOANTROMYAANNOAKT SMON e ABANUNAMMM TET TONMNMOOORRARDDDOAAHOD aes o aa Ree tn iJ = MOVOAOURNVTAD OOOHAAAMADOOUNTS Sor Dorn DYONQHMAMADLS MMOHMAaODo aes PON DOD DA DOOD STS Et WS SIMO ere dese yas Sk ry ; . TF MONTARKRG - - - piolaisini muha wets cp Knieel nwo. - NOR ODVS T TM VON ONAN AAMOMTTMMMORAHMMNANOMT TOO BANANAMAN TYRE D (ANNAN TI MAM ge MUNN OANRR © n~ “ = TATROMONNNOATREMOANNAATROMUNAHMOATROEM DANO - MASA NAUNM MMP ETT TOMDOWOORAARRDDDAHHOO Ga a Sat tts J-8 6-£ Table J-7 TURBOMACHINERY SUMMARY FOR CONFIGURATION XC50 Stat. Flow Rate Pressure Temperature Enthalpy Total Power No. LB/SEC PSIA or BTU/LB HP Combustion Turbine 1,167,340 Comp. Inlet 18 7490.6 14.55 59.0 124.0 Comp. Disch. 20 6575.3 213.63 139 %9 295.3 Turb. Exhaust 26 7854.7 16.00 904.6 339). 1 Steam Turbine 521,610 HPST Inlet 69 766.4 1000.0 800.0 1389.8 MPST Inlet 76 & 111 753.4 460.0 623.3 1316.6 LPST Inlet 108 718.9 115.0 355.9 1200.3 Expander Turbine Inlet N/A Exit N/A Recycle Comp. Inlet N/A Exit N/A Air Compressor AIR COMPRESSOR POWER CONSUMPTION WAS TAKEN FROM 77,750 for 0, Plant REFERENCE 2 AND ADJUSTED FOR COMPRESSOR EFFICIENCY OF 90% 29,850 O, Compressor Inlet OXYGEN COMPRESSOR POWER CONSUMPTION WAS TAKEN FROM Exit REFERENCE 2 Table J-8 HEAT EXCHANGER HEAT LOADS (MMBTU/HR) FOR CONFIGURATION XC50 Fuel Gas Stream Resaturator Feedwater Heater Clean-up Precooler Fuel Reheater Combustion Turbine Exhaust Stream High Pressure Steam Superheater Gasifier Steam Reheater High Pressure Boiler High Pressure Economizer Medium Pressure Economizer Low Pressure Boiler Low Pressure Economizer Deaerator 21S 65. 23% 145. 548. 23). -02 LA, 357 462. 682. 328. 427. 69 12 75 21 81 048 2a 56 31 25 16 J-10 0002-SS8-Si¥ COEKE VO ‘OLIV 18d ‘ZL ¥OL XO 09110 1804 BLNLILSNI HOUVASSY YAMOd 91N19373 0002-SS8-SI¥ EOEFE VO ‘OLY 18d ‘ZL¥OL xO 09110 1800 BLNLILSNI HOUVASIY HIMOd 91H19373 EPRI AP-1390 Below are five index cards that allow for filing according to the four cross-references in addition to the title of the report. A brief abstract describing the major subject area covered in the report is included on each card. For information regarding index card subscriptions to past and future EPRI publications contact the Research Reports Center, P.O. Box 50490, Palo Alto, California 94303. Telephone (415) 965-4081. B[OAD-PEUIQWIOD UO!}EDIJISEH [EOD oeel-dv dda og6! Indy vodey jeul4 Thermal efficiencies of gasified-coal combined-cycle systems using current- technology combustion turbines and near-commercial Texaco oxygen-blown gasifiers, and the effects of configuration and parametric changes on their performance were studied. The effect of a higher combustor exit temperature was studied. Comparisons were made between air- and oxygen-blown gasifiers and between Texaco entrained flow and British Gas Corporation slagging fixed-bed gasifiers. The design compatibility of a medium-BTU combustor with an FT4 engine was studied. | EPRI Project Manager: B. Louks Cross-references: 1. EPRI AP-1390 2. RPOB6-2 3. Engineering and Economic Evaluation Program 4. Coal Gasification 7-986du Z 232 3 A < pid 3 EP 38 3 g > EPRI AP-1390 Coal Gasification Combined-Cycle Rh 2 RP986-2 System Analysis Z D> Final Report 3 o April 1980 Contractor: United Technologies Corporation a a D S WVY5Oud NOILVNIVAR SINONODS GNV ONIYSSNISNS “pelpnys sem euj6ue p14 ue YIM JO}SNqWoo NLE-wWn|pew e yo Ay|1QedWoo UBysep ey) “s19!)18e6 peq-pex!; Buj66e\s Uo}}B10dJOD SBH Ysijg PUe MO}j PeUles}Ue COBXE| UBEeMjeq pUe si9}j/se6 uojyB0d105 sejBojouyse, peu 10}0B1}U0D ELECTRIC POWER RESEARCH INSTITUTE Post Office Box 10412, Palo Alto, CA 94303 415-855-2000 UMO}G-UEBAXO puke -1/2 USEMJeg epeW e1eM SUOSLeEdWOD “pe|pnys sem @iNJesEdW9} }1xe JOJSNQWOD JeYBiYy & JO JOeJje OY) "PE|PNYs eseM EOUBWWOJJEd 4104} UO SeBURYO OLJeWeRJed pUe UOI}BINBYyUOO Jo S}OejJe ey) puke ‘si9}j|Se6 UMO}|q-UEBAXO OOBXe] |B/JEWWOO-JeEU PUR SEUIGIN} UO!JsNqwoo ABojOUYIe} -JueND Bujsn swe}shs ej0Ad-peu|qwiod jeod-pe}sjse6 jo sejoue/d|jje ;EWIEY) EPRI AP-1390 EP} weiBoid Uo}En}eAz 9/WOU0D puke Bueeuj6u3 *¢ EPRI AP-1390 Coal Gasification Combined-Cycle RP986-2 System Analysis Final Report April 1980 Contractor: United Technologies Corporation Thermal efficiencies of gasified-coal combined-cycle systems using current- technology combustion turbines and near-commercial Texaco oxygen-blown gasifiers, and the effects of configuration and parametric changes on their performance were studied. The effect of a higher combustor exit temperature was studied. Comparisons were made between air- and oxygen-blown gasifiers and between Texaco entrained flow and British Gas Corporation slagging fixed-bed gasifiers. The design compatibility of a medium-BTU 2-986du @JOAD-PEUIQWIOD UO!}PEDIJISED [LOD ob l-dv 1Wd3 og6t Indy yodey jeul4 RP986-2 EF = © no combustor with an FT4 engine was studied. 23 g m 3S O° EPRI Project Manager: B. Louks O>3 3 ro o a 2 — @ > Cross-references: So 3 = 3 a 1. EPRI AP-1390 2. RP986-2 3. Engineering and Economic Evaluation Prograr 38 3 $ > o 4. Coal Gasification 2 - 5 2 = = > ELECTRIC POWER RESEARCH INSTITUTE 2 —_— a Post Office Box 10412, Palo Alto, CA 94303 415-855-2000 s 3 S a > 6 = ® a oO : a : 3 > = = 2 fe} a z EPRI AP-1390 Coal Gasification Combined-Cycle “pepnis sem eu/Bue p14 Ue YyJM JO}SNqWOCD NLE-wn|pew e jo Ayiquedwoo uBjsep ey, “sie}}!se6 peg-pexi; bu66e\s uo}Je10di05 seD YsI[Jg pue MOY PeujesjUe CORXe| UEEMJeq PUe S191)|Se6 RP986-2 System Analysis Final Report April 1980 Contractor: United Technologies Corporation uo}}210dJ09 sejBojouyoe| peyUuN :10}981}U0D Thermal efficiencies of gasified-coal combined-cycle systems using current technology combustion turbines and near-commercial Texaco oxygen-blowr gasifiers, and the effects of configuration and parametric changes on their performance were studied. The effect of a higher combustor exit temperatu was studied. Comparisons were made between air- and oxygen-blown gasifiers and between Texaco entrained flow and British Gas Corporation slagging fixed-bed gasifiers. The design compatibility of a medium-BTU combustor with an FT4 engine was studied. EPRI Project Manager: B. Louks UMO}IG-UEBAXO puke -4ye UBEMJEg ePEW eJeM SUOS|JeEdWOD ‘peipnjs sem @inyesedwe} 11x JOYSNQWOD JeYBIY e 4O 199)j0 EYL “PeIPNys elem eOURWOJIEd Cross-references: 1. EPRI AP-1390 2. RPS86-2 3. Engineering and Economic Evaluation Progr: 4. Coal Gasification 4}@y} UO SeBueYS OJeWeJed PUB UO}}EINB}jUOD 40 S}9ejje 4) puke ‘sidj!seB UMO}G-UEBAXO OORXE| /2/DJEWWOD-JBEU PUB SEUIGIN} UO!}sNqwoo ABojOUYoe} -JueUNS Buysn sweyshs @/9X9-peujqWiod |BOd-paljiseB JO Se;OUGJO}Jjo JEUIEY) wes6oig UO}EN|eAgZ 9}WOU0D, puke Buyeeu/6u3 *¢ ELECTRIC POWER RESEARCH INSTITUTE Post Office Box 10412, Palo Alto, CA 94303 415-855-2000