The CW251B12 Gas Turbine Engine

Transcription

1 THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS 345 E. 47 St., New York, N.Y GT-119 The Society shall not be responsible for statements or opinions advanced in papers or in discussion 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. Papers are available from ASME for fifteen months after the meeting. Printed in USA. Copyright 1989 by ASME The CW251B12 Gas Turbine Engine IHOR S. DIAKUNCHAK Gas and Steam Turbine Engineering Turbine & Generator Division Westinghouse Canada Inc. Hamilton, Ontario, Canada ABSTRACT The latest variant of the W251 family of engines, the CW251612, rated at 48 MW, is described in this paper. The improved performance, compared to the CW251B1 model, is achieved by the redesign of the compressor for increased mass flow and efficiency and by hot end modifications, which allow a modest increase in firing temperature. Some of the details of the compressor redesign and turbine modifications are outlined. The uprated engine performance information is provided, as well as plans for the full load factory Prototype tests for engine performance and mechanical integrity verification. INTRODUCTION The CW251B12 engine is the latest variant of the successful line of heavy duty, single shaft, industrial gas turbines, designed for 5 and 6 Hz utility and industrial service in simple cycle, combined cycle and cogeneration applications. The engine is capable of operation on both the conventional gas turbine fuels and the majority of current alternative fuels, dual fuel operation, and operation with steam or water injection for NO emissions reduction to the desired levels. The output power target for this engine is 48 MW, with the initial version of the engine producing 45 MW. The uprated performance will be achieved by compressor redesign and hot end modifications, and not a more substantial redesign in order to reduce the time required to incorporate the changes into the engine, to keep the number of new parts to a minimum, and to make the changes readily retrofitable into existing CW251B1 engines. The planned changes will not detract from the availability and reliability records established by the CW251B1 engines to date. This paper provides some details of the changes incorporated into the engine to achieve the performance uprating objectives. The compressor was redesigned for a four per cent increase in the inlet air flow and about one percent increase in overall efficiency, without reducing the compressor operational stability over a wide range of ambient temperatures. New airfoils were designed for the inlet guide vanes and the first two compressor stages. Most of the remaining downstream stators were restaggered and the airfoil number, chord and spacing of the outlet guide vanes were changed. The only hot end modifications required to accommodate the planned firing temperature increase to 115 C (21 F) were the incorporation of shroud cooling on the first stage turbine stator and the redesign of the first stage turbine blade cooling. The factory test of the CW251B12 engine Prototype is planned for APPLICATIONS The CW251B12 engine is intended for 6 Hz power generation, with the 5 Hz version designated as CW251B11. It is a heavy duty design suited for a wide variety of applications and a wide range of ambient temperatures. It can be used for base load power production and for peaking service. The engine is designed for operation on a broad range of gaseous and liquid fuels. The fuel metering and combustion systems can accommodate low BTU gas fuels, residual liquid fuels and most crude oil fuels without requiring the derating of the engine output. Existing CW251 engines operate on such fuels as refinery gas, hydrogen rich gas, butane, and Diesel oil. The engine is ideally suited for combined cycle and cogeneration applications because of its high exhaust temperature and flow, its cold end drive and straight through exhaust system, and its variable inlet guide vanes. When combined with a suitable steam turbine the combined plant output power will be increased by about 45% and the plant thermal efficiency will be approximately 48%. The engine can also be used in the process industry applications where large amounts of controlled exhaust heat are required. The proven, reliable plant control system can be integrated into any process controller for remote monitoring of operation and performance. The engine maintainability and reliability characteristics make it a good choice for industrial and utility applications where continuous, trouble free service is of utmost importance. ENGINE DESCRIPTION Figure 1 illustrates the longitudinal cross-section of the CW251B12 engine. It differs from its predecessor, the CW251B1 engine (Kuly, 1986), in that it incorporates the design changes described in this paper. The engine weight is about 113,4 kg (25, Ib) and its dimensions are: 3.34 m width, 4 m height and 14.7 m length (131.5 in. by 157 in. by 579 in.). The engine incorporates many long established and well proven design concepts. It has two bearing single shaft construction, cold end Presented at the Gas Turbine and Aeroengine Congress and Exposition June 4-8, 1989 Toronto, Ontario, Canada

2 FIGURE 1 - LONGITUDINAL CROSS-SECTION OF CW251B12 ENGINE power drive, and axial exhaust. The two engine bearings are of the inherently stable tilting-pad type. The front and rear journal bearings are equipped with thermocouples and X-Y vibration probes. Accelerometers are used on the inlet and exhaust, as well as a keyphaser. The engine cylinders are horizontally split to facilitate field maintenance without rotor removal. The inlet casing incorporates the fixed front end engine supports. The front part of the compressor forms a separate cylinder. The high pressure end of the compressor, the combustor and the turbine are in a single cylinder. Two cooled engine support trunnions are located near the downstream flange of this cylinder. The exhaust cylinder carries the hot end journal bearing, which is contained in a casing supported by six tangential struts. This strut system allows for thermal expansion of struts without changing bearing centreline location. The struts are protected from the hot gases by airfoil shielding. Numerous borescopic inspection ports permit internal inspection of critical components. Two interstage bleeds, in conjunction with the variable inlet guide vanes, are employed during startup and shutdown to avoid compressor surge and rotating stall. The compressor discs are shrunk on to a hollow forged shaft. The three turbine discs are connected through CURVIC clutches, which are made up of toothed connection arms that extend from adjacent discs and interlock when the discs are bolted together. The compressor and turbine rotors are connected by a torque tube and a centre coupling. The compressor is a 19 stage design derived from the successful W51 compressor, with redesigned front end. The stators are made of stainless steel and fabricated into diaphragms. This allows changeout by rolling out half diaphragms without rotor removal. The first two rows of blades are precision forged and the rest are machined from envelope forgings. They are secured in the discs using pinned dovetail roots. All the compressor blades can be replaced in situ. The stators are coated with SERMETEL 538 DP to provide improved surface finish and environmental protection. The coating of blades is optional. The combustion system consists of eight axial baskets and is capable of operation on a wide range of conventional gas turbine fuels as well as on the majority of alternative fuels. It has provision for dual fuel operation and either steam or water injection for NO suppression. The cooled, goose-neck style transition ducts direct the combustor exit gases into the turbine. The three stage turbine is a high efficiency design based on the latest design tools. All turbine stators and the first two stages of blades are cast. The first stage stator is cast in single segments in ECY 768 material and is coated. The single segment design has been shown to be the least susceptible to thermal cracking and allows the removal of the first stator through the elongated combustor cover openings. The first stage stator is cooled by compressor delivery air and its cooling scheme consists of two full impingement inserts, three rows of surface ejection holes, pin fins in the trailing edge region and trailing edge ejection holes. The second stator is cast in X-45 material in segments of two airfoils and is coated. Its cooling scheme is similar to that of the first stator, but without the surface cooling air ejection holes and using lower pressure air. The third stator is cast in X-45 in segments of three airfoils, but is not coated. Although it is not cooled, it is hollow and carries air from the outer cavity into the second and third turbine disc cavities. The stators are supported by two blade rings, which are designed so as to allow rolling out of the stator segments with the rotor in situ. The first and second stage turbine blades are cast in IN 738, HIP'd and coated. Both blades are cooled with spanwise holes and use cooled cooling air. Figure 2 shows the details of the engine cooling system. The third stage blade is forged in UDIMET 52 and is uncooled. The turbine blades are secured to the discs with four land fir tree roots and all can be individually replaced with the rotor in place. DESIGN CHANGES Compressor The compressor redesign objectives were as follows: maximum mass flow increase, maximum efficiency increase, no reduction in surge margin, no change in mechanical integrity, minimum number of changes, retrofitability, and minimum risk. The redesign included the modification of the inlet bellmouth, new inlet guide vanes, new airfoils in the first two compressor stages, and the restagger of the downstream stators. Retrofitability of the new compressor into existing CW251B1 engines and the minimum number of required replacement parts were taken into serious consideration during the design. The original compressor was based on the modified NACA 65 Series airfoil profile design technology. Although this compressor design achieved excellent performance, it was decided to employ the Double Circular Arc airfoil design technology in the redesign of the first two stages. This airfoil profile, in conjunction

3 TO EXTERNAL AIR COOLER r- lluv Flu_ G - CJVl71INC i!u JLIIM7 O r 1 Cwl with thinner airfoil cross-sections, would result in reduced surface Mach Numbers, improved airfoil choke margin, and hence in improved performance. Since this design technology was used in previous successful compressor designs (Marson and Nevin, 1987), there is considerable confidence of achieving the performance improvement goals with the redesigned CW251B12 compressor. Table 1 shows the CW251B12 compressor design point conditions and compares them with those of the original design. The design point pressure ratio selected was for engine application with steam injection for NO, control. This explains why the new compressor design pressure ratio is much higher than that of the original one, which is quoted "dry". 2.A larger inlet area increase by increasing the outer radius could not be accommodated within the existing compressor cylinder casing, and the reduction of the inner radius was impossible because of geometric constraints caused by the front journal bearing. The increased inlet area was achieved by a 12.7 mm (.5 in.) increase in compressor inlet outer radius (see Figure 3). This radius increase necessitated the modification of the inlet bellmouth outer surface profile. The increased radius was kept constant in the downstream direction until it intersected the existing compressor outer blade path slope. TABLE 1 CW251 B12 COMPRESSOR DESIGN POINT CONDITIONS INLET TEMPERATURE = 15 C 59 F) INLET PRESSURE = kpa (14.63 psia) CW251B12 COMPRESSOR CW251 B1 COMPRESSOR Inlet Flow, kg/sec (Ib/sec) (365.) (351.) Rotational Speed, Rev/sec (Rev/min.) (54.) (54.) Pressure Ratio Overall Efficiency The compressor was redesigned for about 4% increase in mass flow and 1% increase in efficiency (Marson, 1989). A larger mass flow increase was considered, but was rejected due to the following considerations: 1. A higher increase in mass flow would have resulted in adverse flow conditions in the compressor back end and a possible reduction in surge margin CW251B12 CW2511 FIGURE 3 - CW251B12 COMPRESSOR INLET MODIFICATIONS In order to achieve a satisfactory inlet bellmouth surface velocity distribution, the outer profile modification extended a considerable distance upstream of the inlet guide vanes. A Streamline Curvature Axisymmetric Flow Analysis Program was used to analyze the inlet duct surface velocities and generate the flow conditions for the design of the inlet guide vanes. The results of this analysis showed that the inlet bellmouth outer surface velocity diffusion was similar to that in the original design and was less than 14%. The inlet guide vanes were designed using the latest turbine aerodynamic design technology. A three section design was employed and the shape of each design section was modified till the desired surface

4 velocity distribution and the required airfoil thickness were achieved. A Time Marching Channel Analysis Program was used in the estimation of surface velocities (Denton and Singh, 1979). The final design achieved satisfactory surface velocity distributions, the required radial distribution of outlet flow angle, and satisfactory bending stresses and vibration characteristics. Table 2 summarises some of the more important design parameters of the new inlet guide vane. The surface diffusions on all three design sections were substantially lower than in the original design. TABLE 2 CW251 B12 INLET GUIDE VANE DESIGN DETAILS HUB SECTION MEAN SECTION TIP SECTION Inlet Mach No Outlet Mach No Incidence, degrees Outlet Flow Angle, degrees Pitch/Chord Max. Airfoil Thickness Chord Suction Surface Diffusion, % The compressor design was based on aerodynamic flow conditions generated by the Streamline Curvature Axisymmetric Flow Analysis program. The compressor was analyzed at the design point conditions, as well as over a wide ambient temperature range, to ensure acceptable compressor flow conditions over the complete operating range of the engine. For optimum performance, the stage loading was kept fairly constant, so that the front stages would not be excessively loaded on extremely hot days and the rear stages on extremely cold days. Close attention was paid to the airfoil incidence variation over the whole ambient temperature range to ensure safe compressor operation at all ambient conditions. The diffusion factor was kept below.4 on all airfoils, except for some airfoils in the middle stages on extremely hot ambient temperatures, where it was marginally above.4. Table 3 gives some of the first stage design parameters and compares them to those of the original design. TABLE 3 CW251 B12 COMPRESSOR FIRST STAGE DESIGN POINT S CW251B12 CW251B1 Blade Inlet Relative Tip Mach Number Midspan Meridional Velocity, m/sec (ft/sec) (517.) (528.) Blade Tip Speed, rn/sec (ft/sec) (1182.) (1158.) Blade Diffusion Factor HUB TIP Stage Work, KJ/kg (Btu/lb) (8.61) (8.63) Based on the design flow conditions obtained from the through flow solution, the new airfoils were designed and the downstream stators were restaggered. Double circular arc airfoil cross-sections and multisection design were used in the design of the four new airfoils. In addition, the airfoil thicknesses were reduced where possible without compromising the mechanical integrity. The blade tip section thickness was reduced on both stages to improve the choke margin. The tip chord of the first and second stage blades was increased to improve the stall margin. Table 4 summarises the design parameters of stage 1 and 2 airfoils. To achieve the maximum performance improvement, the downstream stators on all stages, except the sixteenth and seventeenth, were restaggered open by up to four degrees. To reduce the last stage stator and outlet guide vane losses the solidity of the latter was reduced and the two airfoils were made into a tandem cascade. TABLE 4 MIDSPAN DESIGN POINT S OF STAGE 1 AND 2 COMPRESSOR AIRFOILS STAGE 1 STAGE 2 ROTOR STATOR ROTOR STATOR PITCH/CHORD ASPECT RATIO, HEIGHT/CHORD MAX. AIRFOIL THICKNESS CHORD INLET ANGLE, DEGREES CAMBER ANGLE, DEGREES INCIDENCE, DEGREES DIFFUSION FACTOR (NACA DEFINITION) % CHOKE MARGIN Stress and vibrations calculations were carried out on all the new airfoils and most of the remaining stators. This included the use of WECAN Finite Element Analysis Program (see Figure 4). The results showed that the stresses and vibration characteristics of the stators were the same or better than in the original design. In order to achieve the desired mechanical integrity goals for the new first and second stage blades, a better material (modified AISI 63, precipitation hardened stainless steel) was selected for their manufacture. Based on the extensive aerodynamic and mechanical analyses carried out on this compressor design and past experience on previous compressor designs, it can be safely concluded that this redesign will achieve its performance and mechanical integrity goals. Turbine The increase in firing temperature from 1113 C (235 F) on current CW251B1 engines to 115 C (21 F) on the developed CW251B12 engine required the incorporation of shroud cooling on the first stage turbine stator and improved cooling of the first stage turbine blade. Based upon two fully loaded factory tests of the CW251B1 engine at 115 C (21 F) firing temperature, a thorough reassessment of the rest of the

5 hot end components for acceptability at the increased temperature level concluded that no additional modifications were required. The firing temperature increase was considered easily achievable and not requiring a major redesign of the turbine cooling, without compromising the engine reliability and mechanical integrity, because of the following: 5 HOLES D 1. The original CW251 B1 engine design was based on a firing temperature of 113 C (265 F). 2.Two fully loaded factory tests and the field operational experience with the current CW251B1 engines. VIEW OF INNER PLATFORM 3 FROM SECTION BB. ^^ \SECTION DD. In order to provide adequate service life of the first stage turbine stator in the CW251B12 application, shroud film cooling, using compressor delivery air, will be incorporated. Two rows of film cooling holes on the outer platform and one row on the inner platform will be electrochemically machined at 3 to the surface. The hole location is as shown on Figure 5. About.9 kg/sec (2 Ib/sec) of cooling air will be ejected into the gas path to provide a cooling film in the critical regions of the stator end walls. 5 HOLES A ES 11 6 HOLES VIEW OF OUTER PLATFORM FROM SECTION AA. FIGURE 5 - FIRST STAGE STATOR PLATFORM COOLING MAJOR DIAMETER 2.3 mm (.9 in.) FIGURE 4- CW251B12 COMPRESSOR FIRST STAGE BLADE FINITE ELEMENT MODEL To improve the first stage turbine blade life at the increased firing temperature the first stage turbine blade cooling will be enhanced with turbulators. The turbulators, or "ribs", incorporated in the blade cooling holes will promote turbulence in the cooling flow and therefore increase heat transfer. The cooling holes will be cast with a marginally larger diameter than in the original blade and, subsequently, the ribs will be electrochemically machined. Figure 6 shows the schematic view of the turbulators. The CW251B12 blade with turbulators will have a longer service life (in excess of 5, hours), using the same amount of cooling air, than the original blade in the CW251B1 application. To verify the theoretically estimated cooling enhancement with turbulators, a rig was designed, built, and tests were carried out. The test specimen was a smooth stainless steel tube with brazed in rings, which simulated the effect of the turbulators. A completely smooth tube was tested to obtain reference base data. The test results demonstrated that the cooling effectiveness was approximately doubled with the turbulators. Since a lower cooling effectiveness increase was used in the original calculation, the estimated first stage blade life will have a considerable margin. FILLET RAD CORNER RAD RIB HEIGHT.15 mm.6 in.) )RAW ANGLE TCH a 2.5 mm (.1 in.) -IICKNESS.15 mm (.6 in.) FIGURE 6 - SCHEMATIC VIEW OF TURBULATED COOLING SYSTEM PERFORMANCE The CW251B12 engine performance was based on the following: 1. Site performance acceptance test results on CW251B1 engines (Diakunchak and Nevin, 1989). 2. Improvements incorporated into the CW251B1 engines. 3. Estimated improvements due to the design changes described in this paper.

6 The starting point for the performance calculation was the current CW251B1 packaged unit performance, with base load output power in excess of about 42.5 MW at ISO conditions, on standard natural gas, with no injection. The performance benefits expected from the redesigned compressor and the firing temperature increase were then incorporated to the base performance standard. Since the estimated compressor mass flow and overall efficiency increases are considered realistic, and likely to be exceeded, there is considerable confidence that the engine performance goals will be achieved with considerable margin. In fact, it is expected that the initial CW251B12 engine target output power of 45 MW will be achieved without any firing temperature increase over the current level. In the estimation of the performance improvement resulting from the firing temperature increase, the increase in cooling air flow (mostly for stator shroud cooling) was taken into account. The resulting developed CW251B12 engine base load target output power of 48 MW is considered a conservative estimate. The estimated target performance data of the initial and the developed CW251B12 engine, as well as the original CW251B1 engine, are shown on Table 5. As was mentioned previously, the target performance levels should be achieved on the first engine build. TABLE 5 CW251B ECONOPAC PERFORMANCE FOR BASE RATING AT ISO CONDITIONS ON NATURAL GAS FUEL WITH NO INJECTION ORIGINAL INITIAL DEVELOPED CW251 BI O CW251B12 CW251 B12 PLANT NET POWER, KW 41,2 45, 48, PLANT HEAT RATE, KJ/KW.HR, LHV 11,445 11,29 11,8 (Btu/KW. HR, LHV) (1,845) (1,7) (1,5) EXHAUST FLOW, kg/sec (Ib/sec) (35) (368) (372) EXHAUST TEMP., C ( F) (941) (95) (99) NO, emissions have posed a significant environmental concern in many parts of the world. Therefore, the emissions characteristics of industrial gas turbines are becoming very important considerations in the selection of new units. In order to meet the stringent emissions limits imposed in many installations, provision will be made in the CW251B12 engine for steam or water injection through the fuel nozzles to reduce NO emissions. To achieve the NO emission level of 42 PPMV EPA Corrected (Environmental Protection Agency, 1979), at ISO conditions, base load, on natural gas fuel, the estimated steam injection rate of about 1.2 times the fuel flow, or water injection rate of about.8 times the fuel flow, will be employed. The developed CW251B12 engine estimated performance with steam injection for NO reduction to 42 PPMV EPA Corrected is shown on Table 6. For comparison, similar performance information for the current CW251 B1 engine is included. It should be noted that, although a lower amount of water injection would be required for the same NO reduction than steam, steam injection is preferable in many installations because of availability of inexpensive low pressure steam and because of better engine heat rate with steam injection. The CW251B12 engine will be used in many combined cycle and cogeneration applications where steam production will be of considerable or even primary importance. Figure 7 shows the steam production capability of the developed CW251B12 engine operating at base load, at ISO conditions, on natural gas, and without injection. It shows the amount of steam, at varying steam conditions, produced by an unfired boiler, which utilizes the engine exhaust heat. The one condition steam production of the original CW251B1 engine is shown for comparison. As can be seen, the CW251B12 engine steam production is about 17%/ higher. Figure 8 depicts the steam production variation with ambient temperature. = TABLE 6 CW251 B ECONOPAC PERFORMANCE FOR BASE RATING AT ISO CONDITIONS ON NATURAL GAS FUEL WITH STEAM INJECTION FOR 42PPMV NOx EPA CORRECTED ORIGINAL CW251 B1 DEVELOPED CW251B12 PLANT NET POWER, KW 44,94 52,8 PLANT HEAT RATE, KJ/KW.HR, LHV 11,2 1,67 (Btu/KW.HR, LHV) (1,44) (1,11) 82 NOTES: ISO CONDITIONS BASE RATING (18) NATURAL GAS NO INJECTION UNFIRED BOILER 78 17) \\ CW251B12 J o STEAM PRESSURE KPa (45 psig) (16) = 458 KPa (65 psig) 7 (15) p 66 J LL CW251B1 Q (14) \ KPa (15 psig) 32 KPa (45 psig) (76) (8) (84) (88) (92) STEAM TEMPERATURE, C ( F) FIGURE 7- CW251B12 STEAM GENERATION Although the simple cycle gas turbine engine performance is of considerable interest and importance, in many applications the economic/ performance considerations of the combined cycle or cogeneration mode of engine application play the decisive role in engine selection. Combining three CW251B12 engines with a dual pressure unfired boiler and a Westinghouse SC23 condensing steam turbine, will result in a net combined cycle thermal efficiency of about 47%. Using a higher pressure boiler and reheat steam cycle with a more sophisticated and costly steam turbine, net plant thermal efficiency of about 48% could be achieved. In installations where both additional electricity and medium to low pressure steam are required, the CW251B12 engine could be used in conjunction with an extraction/back pressure steam turbine to provide the required performance.

7 U The developed CW251B12 engine will provide excellent performance in simple cycle, combined cycle, and cogeneration applications. The CW251B engine development will continue in the future. The main objective of this effort will be the further improvement of both the simple cycle and combined cycle plant efficiencies. This objective will be achieved by a more extensive compressor redesign, redesign of the turbine first stage, and a further increase in firing temperature. NOTES: ISO CONDITIONS BASE RATING NATURAL GAS NO INJECTION UNFIRED BOILER -'I STEAM CONDITIONS: 32 KPa, 4 C 86 ( 19 ) (45 psig, 75 F) cc I Y 82 (18) 3 78 J LL 2 a 74 (17) () (2) (4) (6) (8) (1) AMBIENT TEMPERATURE, C ( F) FIGURE 8 - CW251B12 STEAM GENERATION VS AMBIENT TEMPERATURE TEST PROGRAM To verify the performance and the mechanical integrity of the CW251B12 engine, especially of the new components, extensive factory tests are planned for The factory tests will be carried out on Diesel oil fuel. The main objectives of these tests are the confirmation of the following: 1. Engine starting and acceleration characteristics. 2.Compressor inlet air flow and overall efficiency over the entire inlet guide vane setting range. 3.Optimum inlet guide vane setting. 4. Compressor surge margin. 5. Turbine performance. 6. Engine performance. 7. Engine emissions, both dry and with steam injection. 8. Mechanical and thermal performance of the engine, by measurement of gas and metal temperatures and pressures. 9. Vibration characteristics of the engine and its components, including telemetry tests on the new first stage compressor blade. The engine will be heavily instrumented to obtain the required performance and mechanical information. The redesigned compressor will be fully instrumented to measure interstage temperatures and pressures. A traverse will be carried out at the inlet guide vane exit to measure the flow conditions into the first stage compressor blade. The first stage compressor blade vibration patterns and stresses will be measured with the aid of telemetry equipment. The vibratory stresses in the diaphragms of the first few compressor stages will be measured with the aid of strain gauges. The above information will be analyzed to obtain stage by stage performance, as well as overall compressor performance, and to confirm its acceptability from mechanical and vibration considerations. In the hot end a multitude of metal temperatures will be measured to assess the performance of the components at the increased firing temperature, to verify the mechanical design safety factors, and to ensure the required service life of the components. The test data will be automatically recorded, reduced, and the pertinent information will be available within minutes of the completion of the actual test. The detail analysis of the test results will help in confirming that performance and mechanical integrity goals of the CW251 B12 engine have been achieved. SUMMARY This paper has described the main design features of the CW251B12 engine, especially the redesigned compressor and the hot end modifications. The engine performance in both simple cycle and combined cycle applications was summarized. Based on past experience, there is great confidence in achieving these performance goals. Extensive factory test plans for the engine performance and mechanical integrity verification were outlined. The scope of the engineering design effort and the planned test program confirm Westinghouse Canada Inc. commitment to the critical evaluation of the new design features and to the continuing endeavor to improve the performance and mechanical integrity of the CW251 line of gas turbine engines. REFERENCES Kuly, P., 1986, "The CW251B1 Gas Turbine Engine", ASME Paper No. 86-GT-82. Marson, E., and Nevin, D. R., 1987, "Compressor Redesign: An Economical Approach to Near Term Combined Plant Repowering", International Symposium on Turbomachinery, Combined-Cycle Technologies and Cogeneration- IGTI Vol. 1, Book No Marson, E., 1989, "Compressor Uprating for CW251B12 Gas Turbine", Paper to be presented at the 1989 ASME IGTI Conference. Denton, J. D., and Singh, U. K., 1979, "Time Marching Methods for Turbomachinery Flow Calculation", VKI Lecture Series: Transonic Flows in Turbomachinery. Diakunchak, I. S., and Nevin, D. R., 1989, "Site Performance Testing of CW251B1 Gas Turbines", Paper to be presented at the 1989 ASME IGTI Conference. Environmental Protection Agency, Standards of Performance for New Stationary Sources, 4 CFR Part 6, Subpart GG, Published 1979.