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THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS 345 E. 47th St., New York, N.Y. 10017 97-GT-490 The Society shall not be responsible for statements or opinions advanced in papers or dlip.ission at meetings of the Society or of its Divisions or Sections, or printed In its publications. Discussion is printed only if the paper is published in an ASME Journal. Authorization to photocopy material for Internal or personal use under circumstance not fairing within the fair use:provisions of the Copyright Act is granted by ASME to libraries and other users registered with the Copyright Clearance Center (CCC) Transactional Reporting Service provided that the base fee of $0.30 per page is paid directly to the CCC, 27 Congress Street Salem MA 01970. Requests for special permission or bulk reproduction should be addressed to the ASME Techniosl Publishing Department Copyright 0 1997 by ASME All Rights Reserved Printed in U.S.A Process Optimization of an Integrated Combined Cycle The Impact & Benefit of Sequential Combustion Dr. Walter Jury ABB Power Generation Ltd. Baden, Switzerland 1111111111 E1111111111 David E. Searles ABB Power Generation Inc. Richmond, Virginia USA Abstract Introduction Advanced gas turbine designs require revisiting the optimization process to provide maximum competitiveness of new generating installations. This counts specifically for those designs created for combined cycle applications. Gas turbine The onset of deregulation will change fundamental business practice within the electric power industry. This change requires electric power producers to find increasingly innovative ways to provide competitive power. These ways must provide an edge performance and its associated exhaust temperature has been increasing at a rapid pace over recent years. The conventional method of selecting a GT based upon price and performance, over the competition and improve profit levels at the same time. Previously, a common solution was for the gas turbine manufacturer to increase engine firing temperatures. This and then designing a complex bottoming cycle does not provide sufficient solutions for power generation in an open access marketplace. The optimal solution takes into account the interrelation between the CT and WS cycle, leading to a more efficient, simplified and flexible power plant. This analysis shows how different levels of GT exhaust energy lead to different optimum cycle solutions. It shows, as postulated above, that considering the WS cycle demands in gas turbine design leads to a simpler cycle with inherent advantages in efficiency, reliability and flexibility. increase provided for both simple cycle and combined cycle improvements. However, the ability to finance and insure these large machines has recently become an issue, especially as a deregulated marketplace eliminates the financial security of a long term power purchase agreement. Nomenclature 3PRH Triple Pressure Reheat 2PRH Dual Pressure Reheat CC Combined Cycle FOR Forced outage rate EFOR Equivalent Forced Outage Rate GT Gas Turbine HP High Pressure LP Low Pressure NPV Net present value RH Relative humidity SP Superheater ST Steam Turbine WS Water Steam VP Evaporator LO Loss of Output Specific exergy Specific enthalpy Specific entropy amb ambient Efficiency Traditionally, gas turbines provided emergency peaking power. With the growing importance of energy efficiency, combined cycle concepts based on such gas turbines, designed primarily for simple cycle operation, became economical. However, to utilize the full potential of the combined cycle power plant it is not enough to optimize either the gas turbine or the bottoming water steam (WS) cycle independently. The most economical and hence most competitive solution must consider a plant approach that integrates the CT and WS cycle in a manner most beneficial to the owner of the plant. The CT exhaust energy is the link between CT and WS cycle. Therefore, the design and operating concept of the CT must consider this link and provide exhaust energy at a level leading to a simple, reliable and efficient WS cycle concept. To ensure high reliability and financeability, the plants design criteria should be based upon proven concepts, rather than new technology that consistently 'pushes the envelope'. Gas turbine exhaust energy - link between topping and bottoming cycle Over the last 15 years, gas turbine exhaust temperature has risen from below 480°C (900 °F) to almost 650°C (1200 °F) [1, 2]. Due to this fact, CC performance was marginal in the early years (<50% gross thermal efficiency). With CT exhaust temperatures of 480 °C, the efficiency of the bottoming steam cycle was much below that of conventional rankine cycle plants. Presented at the International Gas Turbine Sr Aeroengine Congress & Exhibition Orlando, Florida — JuneTerms 2—June 5,1997 Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 12/29/2014 of Use: http://asme.org/terms Significant improvements (ft 56% gross thermal efficiency) began only a few years ago when GT exhaust temperatures rose to 540°C (1000 °F) and beyond. This allowed live steam temperatures of CT based combined cycles to match that of steam boiler plants (540°C - 565°C (1000 °F - 1050°F)), increasing bottoming cycle efficiency and improving overall CC plant performance. Property ST exhaust mass flow GT exhaust temperature ST cooling energy' Value/ Range of Values 540 Ir4/s (4,285,786 Ibihr) Variable (550°C / 600"C / 650°C) (1022° F/1112° F/1202°F) 19 MW (64.8 MMlittuihr) 70 ban (1015 psia) Minimum live steam generating pressure The heat recovery curves associated with CT exhaust temperatures between 550°C and 600°C (1000°F - 1100°F) (Figure 1) favor the introduction of multi-pressure level cycles for improved cycle efficiency, leading to today's 'industry standard' of a triple pressure reheat cycle (31 Figure 1 also shows the reason for these multi-pressure cycles: the heat recovery steam generator (HRSG) stack temperature for a single pressure level is very high and leaves room for economical intermediate and low pressure steam production. Reheat pressure 1/5 of live steam pressure Steam turbine backpressure 42°C (110°F) Minimum HRSG stack temperature 80°C (175°F) Drum approach temperature Minimum superheater approach gig Maximum steam temperature lailM --tingmtm . e 10 K (18W) 5 K (9°F) 251< (45°F) 565"C (1050°F) Slagle p eeeeee tint • ■•• ■• Tab. 1.: Parameters used for cycle investigations. Exhaust mass flow is approximately equal to the exhaust mass flow of the largest industrial gas turbines available today. Live steam pressure values and maximum live steam temperature cover the well proven experience range of combined cycle power plants and conventional steam plants (most conventional steam plants commissioned in the 1980s run with live steam pressures of 170 bar (2470psia) and five steam temperatures of 540°C to 565°C and use single or double reheat at 35 bar (500psia) 540°C to 565°C (2)). Reheat pressures at 1/5 of live steam pressure ensure acceptable temperature differences at the steam turbine (cold reheat/ hot reheat) and in the HRSG (parallel superheater and reheater). steam turbine backpressure of 0.065 bar is economically achievable with a wet cooling tower at 15 °C (59°F) ambient at 60% RH. FIRSG feedwater at 42°C is appropriate for combined cycles firing natural gas with low sulfur content (the pipe metal temperature of the feedwater preheating economizer remains above the water dewpoint). Values for pinch points and approach temperatures follow the recommendations of the international standard ISO 3977 Annex F. IME ! =1419111 • : '414 C •. ■■■■■■ 0.065 bara (7 Hg) HRSG feedwater temperature Evaporator pinch point C IMO; 190 bare (2755 psia) Maximum live steam generating pressure .• ■.1 •• • ■ • ■■ •■ • ■ •.■, •••• LDS MSS Heat transferred Fig. 1.: Heat recovery profile of a triple pressure reheat cycle at CT exhaust temperatures of 550°C (1020 °F). The dashed line shows the heat recovery profile of a single pressure reheat cycle. The high stack temperature leaves room for economical production of low- and intermediate pressure steam The following section investigates the effect of increasing CT exhaust temperatures on heat recovery cycle design. It answers the question if and how optimum cycle design will change with increasing gas turbine exhaust temperatures. Table 1 shows the parameters used in this investigation. Heat Recovery & Exergy Transfer This section compares the suitability of three cycle designs for given levels of gas turbine exhaust energy: the single pressure reheat, dual pressure reheat and triple pressure reheat cycles. Figure 2 shows flow diagrams of the HRSGs of the compared cycles. 'Most high efficiency gas turbines require some amount of external cooling of high temperature parts, either by direct steam cooling or external cooling of compressor air. This investigation assumes that 20% of compressor massflow are used as cooling air and cooled down by 200°C. This heat is Introduced into the cycle as high pressure steam. 2 Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 12/29/2014 Terms of Use: http://asme.org/terms Figures 4 and 5 show block diagrams of the HRSGs used for the reliability calculations. vr "Tr HP HP 13P HP - E m P4 HP - GP HP Er LP VP LP DIA LP "Tr I tea CP LP 1 650 tea • 600 X xmoo VV CT CITM CP LP tea • SSO MO x Fig. 4.: Block diagram for reliability analysis of the dual pressure reheat HRSG. EC ...economizer, VP...evaporator, SP...superheater, CP...circulation pump, VV. valves, CT...controls, OTH...other MOO Met SO I. •• 126 III 1111 Evaporation pressure [bare] ••■■•■ EC VP HP Fig 5 SP EC IP VP IP DM IP EC LP VP LP DM LP I. Fig. 6.: Heat exchange surface difference between dual pressure reheat and triple pressure reheat cycles as a function of live steam pressure with gas turbine exhaust temperature as parameter. This picture shows that the narrowing gap between dual pressure reheat and triple pressure reheat cycle is not 'bought' with a more elaborate cycle design. This improvement is - as also shown in the previous section - a result of the better 'fit' of the dual pressure reheat cycle for higher gas turbine exhaust temperatures. VV - CT OTM CP LP •■■• •••■ Block diagram for reliability analys's of the triple pressure reheat HRSG. Symbols a e explained in figure 4 above. The results presented in the previous two sections lead to the following conclusions: Table 5 below shows the reliability comparison of dual and triple pressure reheat HRSG. The comparison shows an advantage of about 0.1% points in equivalent forced outage rate for the dual pressure reheat HRSG. System Dual Pressure FOR % High pressure Intermediate • ressure Low pressure General (Valves, control etc. Total 0.22 Tnple Pressure LO % EFOR % FOR % 100 0.22 LO % EFOR % 0.21 100 0.21 0.19 100 0.19 0.20 100 0.20 0.13 70 0.09 0.09 100 0.09 0.10 100 0.10 0.51 0.63 0.51 (1) The thermal advantage of a triple- over a dual pressure reheat cycle sharply diminishes with increasing OT exhaust temperatures. (2) With increasing GT exhaust temperatures, higher HP steam pressures show an increasing output benefit for the dual pressure reheat cycle. (3) The simpler design of the dual pressure reheat HRSG results in an advantage of about 0.1% points in equivalent forced outage rate (EFOR). (4) The dual pressure reheat cycle requires consistently less HRSG surface than the triple pressure reheat cycle irrespective of GT exhaust temperature. 0.59 Combining these facts and carrying out the net present value (NPV) calculation with the parameters from Table 4 results in Figure 7 below. Figure 7 shows the difference in NPV between dual pressure reheat and triple pressure reheat cycles with gas turbine exhaust temperature as parameter. The figure shows how the facts work in combination to result in an increasing NPV advantage of the dual pressure reheat cycle over the triple pressure reheat cycle with increasing GT exhaust temperature. Tab. 5.: Reliabi ity values for dual pressure reheat and triple pressu e reheat HRSGs [4, 51. FOR ... forced outage rate; LO ... loss of output; EFOR ... equivalent forced outage rate Cycle investment cost: The primary difference between the two cycles compared is the design of the HRSG. The dual pressure reheat cycle is a simpler design requiring less heat exchange surface. Figure 6 shows the heat exchange surface difference between dual and triple pressure reheat cycles with gas turbine exhaust temperature as parameter. The figure shows how the dual pressure reheat cycle maintains an almost constant HRSG surface advantage over the triple pressure reheat cycle regardless of OT exhaust temperature. 5 Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 12/29/2014 Terms of Use: http://asme.org/terms One important conclusion is that an increase in gas turbine exhaust temperature significantly increases the effectiveness and efficiency of heat recovery irrespective of the cycle configuration employed. For an optimum bottoming cycle, a gas turbine should, therefore, deliver the highest exhaust temperature possible with the chosen gas turbine cycle concept tutu • G50 C OST, 2PRII as V. OfPSTO PRH ...dr. The most relevant fact for further investigations, however, is that the advantage of a triple pressure reheat cycle sharply diminishes with increasing gas turbine exhaust temperatures. The next section will, therefore, answer the question whether the triple pressure reheat cycle remains the economical cycle at higher gas turbine exhaust temperatures? Dual Pressure Reheat versus Triple Pressure Reheat Cycle: Economic Comparison nem 13 texts • GOO C is.., 117.501.--x 'i.e.., texh • ssoc MIS II DM SO BS WO 120 Evaporation pressure (bare] This section compares the dual pressure reheat and triple pressure reheat cycles in respect of economical properties over the range of exhaust temperatures and live steam pressures specified in Table 1. The economical comparison uses the parameters listed in Table 4 Property Value Financial assessment period 15 years Discount Factor for NPV Analysis Plant operating hours Electricity price 10% 7900 hours/year 3.5 centsAd/Vh Escalation Electricity Revenue Fuel Price (LHV) 1.5% Fixed O&M Costs Cycle availability: Dual / and triple pressure systems differ through the additional components required for intermediate pressure steam generation by the triple pressure reheat cycle. This comparison considers only the HRSG components. The comparison assumes that failures of the high pressure part (dual and triple pressure cycle) or of the intermediate pressure part (triple pressure cycle) of the HRSG result in a plant shutdown. Failure of the low pressure part results in a plant shutdown for the dual pressure cycle and in a load reduction to 70% for the triple pressure cycle. Scheduled HRSG outages (maintenance) coincide with the gas turbine scheduled outages, differences in availability between the cycle concepts are, therefore, equal to differences in reliability. 3 2.25 USS/MMBTU Escalation Fuel Cost Variable O&M Costs Fig. 3. Steam turbine output of the dual pressure reheat cycle as a percentage of the steam turbine output of the triple pressure reheat cycle with gas turbine exhaust temperature as parameter. This picture confirms the conclusion of the previous section that with moderate exhaust temperatures between 550 °C and 600°C the triple pressure reheat cycle with a five steam pressure of 110 bar (1550psia) was a good choice. With higher exhaust temperatures it gets significantly more interesting to increase pressure and switch to dual pressure reheat. 2% 0.4 cents/IcWh 3.8 MM US$/year Escalation O&M 2% Tab. 4.: Parameters used for economical comparison Cycle characteristics governing the economical comparison are: • cycle efficiency (steam turbine output differential) • cycle availability • cycle investment cost Cycle efficiency: Figure 3 shows the steam turbine output of the dual pressure reheat cycle as a percentage of the steam turbine output of the triple pressure reheat cycle with gas turbine exhaust temperature as parameter. The shown percentages reflect the trend explained in the previous section. The relative advantage of the triple pressure reheat cycle decreases significantly with increasing GT exhaust temperature. The figure also shows another interesting trend, the benefit of higher pressures for the dual pressure reheat cycle with increasing exhaust temperatures. In combined cycle power plants scheduled outages (maintenance) are coordinated with the gas turbine as lead machine. As the duration of HRSG maintenance is significantly shorter than gas turbine maintenance irrespective of cycle concept, differences in HRSG maintenance duration do not influence overall availability. 3 4 Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 12/29/2014 Terms of Use: http://asme.org/terms Figures 4 and 5 show block diagrams of the HRSGs used for the reliability calculations. vr "Tr HP HP 13P HP - E m P4 HP - GP HP Er LP VP LP DIA LP "Tr I tea CP LP 1 650 tea • 600 X xmoo VV CT CITM CP LP tea • SSO MO x Fig. 4.: Block diagram for reliability analysis of the dual pressure reheat HRSG. EC ...economizer, VP...evaporator, SP...superheater, CP...circulation pump, VV. valves, CT...controls, OTH...other MOO Met SO I. •• 126 III 1111 Evaporation pressure [bare] ••■■•■ EC VP HP Fig 5 SP EC IP VP IP DM IP EC LP VP LP DM LP I. Fig. 6.: Heat exchange surface difference between dual pressure reheat and triple pressure reheat cycles as a function of live steam pressure with gas turbine exhaust temperature as parameter. This picture shows that the narrowing gap between dual pressure reheat and triple pressure reheat cycle is not 'bought' with a more elaborate cycle design. This improvement is - as also shown in the previous section - a result of the better 'fit' of the dual pressure reheat cycle for higher gas turbine exhaust temperatures. VV - CT OTM CP LP •■■• •••■ Block diagram for reliability analys's of the triple pressure reheat HRSG. Symbols a e explained in figure 4 above. The results presented in the previous two sections lead to the following conclusions: Table 5 below shows the reliability comparison of dual and triple pressure reheat HRSG. The comparison shows an advantage of about 0.1% points in equivalent forced outage rate for the dual pressure reheat HRSG. System Dual Pressure FOR % High pressure Intermediate • ressure Low pressure General (Valves, control etc. Total 0.22 Tnple Pressure LO % EFOR % FOR % 100 0.22 LO % EFOR % 0.21 100 0.21 0.19 100 0.19 0.20 100 0.20 0.13 70 0.09 0.09 100 0.09 0.10 100 0.10 0.51 0.63 0.51 (1) The thermal advantage of a triple- over a dual pressure reheat cycle sharply diminishes with increasing OT exhaust temperatures. (2) With increasing GT exhaust temperatures, higher HP steam pressures show an increasing output benefit for the dual pressure reheat cycle. (3) The simpler design of the dual pressure reheat HRSG results in an advantage of about 0.1% points in equivalent forced outage rate (EFOR). (4) The dual pressure reheat cycle requires consistently less HRSG surface than the triple pressure reheat cycle irrespective of GT exhaust temperature. 0.59 Combining these facts and carrying out the net present value (NPV) calculation with the parameters from Table 4 results in Figure 7 below. Figure 7 shows the difference in NPV between dual pressure reheat and triple pressure reheat cycles with gas turbine exhaust temperature as parameter. The figure shows how the facts work in combination to result in an increasing NPV advantage of the dual pressure reheat cycle over the triple pressure reheat cycle with increasing GT exhaust temperature. Tab. 5.: Reliabi ity values for dual pressure reheat and triple pressu e reheat HRSGs [4, 51. FOR ... forced outage rate; LO ... loss of output; EFOR ... equivalent forced outage rate Cycle investment cost: The primary difference between the two cycles compared is the design of the HRSG. The dual pressure reheat cycle is a simpler design requiring less heat exchange surface. Figure 6 shows the heat exchange surface difference between dual and triple pressure reheat cycles with gas turbine exhaust temperature as parameter. The figure shows how the dual pressure reheat cycle maintains an almost constant HRSG surface advantage over the triple pressure reheat cycle regardless of OT exhaust temperature. 5 Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 12/29/2014 Terms of Use: http://asme.org/terms increased bottoming cycle efficiency as these elevated pressures define the top of the ST expansion curve and an overall increase in the power density of the cycle (power per unit mass flow of working fluid). Y01/4 2111011 texh•650C .00.0"4" • _ LI I zioll<1 a 1091000 —X irrn ri: In C 11 6) X X x x In considering higher live steam pressures, the next logical step is evaluating the use of a high speed, high efficiency geared HP turbine section. With higher live steam pressures - which are favored by sequential combustion and a dual pressure reheat cycle, as shown in the previous sections - the value (kW/kg) of live steam increases significantly. Hence higher HP steam turbine efficiency pays off disproportionately. The geared ST technology, with many years of proven experience, allows for complete optimization of the steam flow profile for maximum efficiency resulting in increasing benefits with increasing steam pressures (Figure 8). Through their unique design, high speed turbines are more efficient that 3600 rpm turbines. The reason is that the circumference (blade-hub-blade) of the stages is smaller than.a comparable 3600 rpm turbine. This in turn yields longer, more efficient blades since although the circumference is smaller, the steam path area is constant (the tip speeds of both turbine types are identical). Since tip losses and root losses are basically constant, longer blades (high speed) are always more efficient than shorter blades. Therefore, the resulting HP turbine section can achieve higher efficiency with fewer stages than a comparable direct drive unit. 1 X texh•550C 1601090 Evaporation pressure [bare] Fig. 7.: Difference in NPV between dual pressure reheat and triple pressure reheat cycle as a function of live steam pressure with gas turbine exhaust temperature as parameter. This picture shows that for CT exhaust temperatures above approximately 630 °C (1150°F) the dual pressure reheat cycle is the most economical solution regardless of live steam pressure. For a CT exhaust temperature of 650 °C (1200°F) the picture also shows an increasing advantage of the dual pressure reheat cycle with increasing evaporation pressure. This NPV comparison shows that the NPV advantage of the dual pressure reheat cycle increases with increasing live steam pressures. Figure 7 compares dual pressure reheat and, triple pressure reheat cycles. Therefore, it does not give any information as to where the optimum live steam pressure for the dual pressure reheat cycle lies. The following section answers that question and draws some conclusions as to which components shall be applied to make this cycle work most effectively. Figure 8 shows the power output difference between a geared ST at 170 bar (2465psia) evaporation pressure and a nongeared ST at 110 bar (1600psia) evaporation pressure as a function of CT exhaust temperature with the cycle type as parameter. The figure shows the increasing rewards of using higher steam pressures and a geared steam turbine with increasing CT exhaust temperatures. With increasing exhaust temperatures the geared steam turbine becomes ever more favorable for the dual pressure reheat cycle. Specification of the dual pressure reheat cycle t it 2 0 This section optimizes the live steam pressure of the dual pressure reheat cycle based on the economical parameters shown in Table 4. It also shows which components the optimized cycle requires to work most effectively. z .6 IA Tr1919 I• ***** • F 2.6 2 2 o. < .2 21 0 The previous section has shown that above CT exhaust temperatures of approximately 630 °C the dual pressure reheat cycle is the most economical cycle regardless of live steam pressure (Figure 7). However, it has also shown that increasing live steam pressures yield increasing benefits for the dual pressure reheat cycle with increasing CT exhaust temperatures (Figure 3). Conventional plants, where heat input at high temperature levels has always been available, have used live steam parameters of 170 bar (2465psia) and 565°C (1050psia) for more than two decades. This was not possible with combined cycle plants until recently when high GT exhaust temperatures have become available. 2, to uti "0 III In 710 OT exhaust temperature [C] Fig. 8.: Difference in power output between a geared ST at 170 bar evaporation pressure and a non-geared ST at 110 bar evaporation pressure The CT exhaust energy determines the amount of live steam that can be produced at a given pressure level. Increasing live steam pressures produce a smaller mass- and volume flows of higher energy steam. In principle, this provides two benefits: Evaluating the dual pressure reheat cycle at an elevated gas turbine exhaust temperature with a geared steam turbine results in the curve shown in Figure 9 below. Figure 9 shows the NPV 6 Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 12/29/2014 Terms of Use: http://asme.org/terms of the dual pressure reheat cycle as a function of evaporation pressure with the cycle at 70 bar (1000psia) as reference. Figure 9 shows that the economical evaporation pressure lies between 150 and 170 bar (2200 and 2500psia) for an optimized dual pressure reheat cycle with a geared steam turbine at a ST exhaust temperature of 650 °C. • a decreasing output difference between dual pressure reheat and triple pressure reheat cycles and an increasing benefit of higher live steam pressures for the dual pressure reheat cycle with increasing ST exhaust temperatures. • an advantage of about 0.1% points in equivalent forced outage rate for the dual pressure reheat HRSG • an almost constant HRSG surface advantage of the dual pressure reheat over the triple pressure reheat cycle regardless of ST exhaust temperature. The comparison of net present value based on these facts shows an increasing NPV advantage of the dual pressure reheat cycle over the triple pressure reheat cycle with increasing GT exhaust temperature. For ST exhaust temperatures above approximately 630°C (1150°F) the dual pressure reheat cycle is the more economical solution regardless of live steam pressure. For a ST exhaust temperature of 650 °C (1200°F) the advantage of the dual pressure reheat cycle continually increases with increasing evaporation pressure. II HI Also demonstrated was the increased potential of higher live steam pressures at high gas turbine exhaust temperatures can best be utilized with a geared high pressure steam turbine. The geared ST technology, with many years of proven experience, allows for complete optimization of the steam flow profile for maximum efficiency resulting in increasing benefits with increasing steam pressures. At a ST exhaust temperature of 650°C the resulting optimum live steam pressure for the dual pressure reheat cycle lies between 150 and 170 bar (2200 and 2500psia). The increase in power output largely outweighs the cost increase for the high pressure parts in HRSG and water steam cycle. Evaporation pressure [bare] Fig. 9.: NPV of the dual pressure reheat cycle a function of evaporation pressure with the cycle at 70 bar (1000psia) as reference. As the evaluation shows the economical evaporation pressure lies between 150 and 170 bar (2200 and 2500psia). The increase in power output largely outweighs the cost increase for the high pressure parts in HRSG and water steam cycle. Summary and Conclusions How these thermodynamic advantages of the dual pressure reheat cycle translate into an optimum power plant concept together with the ABB sequential combustion gas turbines 5T24 and 5126 will be the subject of a follow-up paper. As earlier postulated, the most economical and hence most competitive combined cycle power plant solution must consider a plant approach that integrates the ST and WS cycle in a manner most beneficial to the owner of the plant. As the deregulated electric industry evolves and cost transparency is no longer available, the owners of new generating facilities must seek the true optimal configurations. ST exhaust energy is the link between ST and WS cycle. Therefore, the design and the operating concept of the ST must consider this link and provide the exhaust energy at a level that leads to a simple, reliable and efficient WS cycle concept. Acknowledgements The authors would like to acknowledge the contributions of the following individuals to the publication of this paper Rolf Bachmann, Yves Carels, Sep van der Linden and Melinda Martin. A comparision of heat recovery effectiveness (fit of a cycle with a given level of ST exhaust energy) and heat recovery efficiency for single pressure reheat, dual pressure reheat and triple pressure reheat cycles showed that an increase in gas turbine exhaust temperature significantly increases the effectiveness and efficiency of heat recovery irrespective of the cycle configuration employed. The comparison also showed that the advantage of a triple pressure reheat cycle over a dual pressure reheat cycle sharply diminishes with increasing gas turbine exhaust temperatures. An economic comparison of the dual pressure reheat and triple pressure reheat cycles, considering cycle efficiency, cycle availability and cycle investment costs, showed 7 Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 12/29/2014 Terms of Use: http://asme.org/terms References [1] Jury W. and Luthi H.: 'Advanced GTs call for advanced CC's (or the 4 S's)', ASME 94-GT-303, paper presented at the international gas turbine and aero engine congress and exhibition, the Hague, Netherlands - June 13-16, 1994 [2] Chiesa P. et. al.: 'Predicting the ultimate performance of advanced power cycles based on very high temperature gas turbine engines', paper presented at the 38th ASME international gas turbine and aero engine congress, Cincinnati, May 1993 Lugand P. and Parietti C.: 'Combined cycle plants with [3] frame 9F gas turbines', ASME paper 90-GT-345 [4] NERC statistics 1978-1987 and 1988-1994 VGB analysis of non-availabilities 1988-1994 (5) 8 Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 12/29/2014 Terms of Use: http://asme.org/terms
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