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1 THE AMERCAN SOCETY OF MECHANCAL ENGNEERS 845 E. 47th St., New York, N.Y The Society shall not be responsible for statements or opinions advanced in papers or 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 15 months after the meeting. Printed in U.S.A. Copyright 1994 by ASME 94-GT-65 A NEW UPRATED TURBNE FOR THE GT8 AND GT8C GAS TURBNE FAMLY S. P. Harasgama and F. Kreitmeier Gas Turbine Department ABB Power Generation Limited Baden, Switzerland !) Abstract This paper describes the design of an uprated turbine for the GT8 gas turbine. The new turbine is installed in the GT8C gas turbine unit. The entire aero-thermal performance of the old turbine has been re-assessed and the new design improves upon the old unit by a 3.5% point increase in gas turbine efficiency. Additionally, the heat transfer and cooling of the turbine has been improved with the blading having more uniform thermal loading resulting in improved thermo-mechanical Me-time. This has been achieved with better cooling technology whilst still maintaining the same cooling air mass flow. The paper describes the steps involved in the uprating..ntroductiog The new uprated turbine requirement was that it would operate within the same confines of the present GT8 annulus, be entirely retro-fittable, have improved durability (with the same amount of cooling air) and deliver a higher power output with an improved efficiency. The target values for the turbine are as follows: Pressure Ratio a 14.2 Mean Stage Loading = 1.4 Mean Reaction = 0.47 Flow Coefficient ti 0.57 The turbine was to be designed to be a three-stage unit with two cooled stages. The resulting design was to satisfy the full GT8C technical requirements as given in Table-1. As the table shows, the changes to the turbine were to be based on well proven technology in the existing 6T8. The main thrust of the work was in the improvement in turbine efficiency and power output, with a moderate increase in Turbine nlet Temperature. Turbine Aerodynamics The annulus line and aerofoil aerodynamics were both changed to obtain a more efficient turbine module. The existing 018 Table-1, Specification for GT8C Turbine DT8 GT8C Gross Electrical Output MW 48.2 '50 Gross Elect. Efficiency % Exhaust Mass Flow kg's 183.c 183 Pressure Ratio nlet Temp. (SO Detn.) cc 1085 a 1100 Firing Temp. (ANS Del) 0D 1162 e 1172 Exhaust Gas Temp. 0C turbine is a unit dating back to Since that time A1613 has gained experience with the engine and has carried out full engine testing on the GT8 gas turbine (Graf, 1985). Turbine Conceptual Des:gn The existing turbine design formed the starting point for the uprated unit and Figure-1 shows the overall design of the old & new turbines. The turbine of the GT8 was reshaped and an optimised version was used for the GT8C turbine flow path. This was done using an in-house developed Streamline Curvature ($2) method. The aerodynamic analyses utilise a semi-coupled calculation procedure. The very detailed $2 calculations resulting in the final annulus line are shown on Figure-2. The old and new annuli are compared in Figure 3. The stator aerofoils for the new design have had some Three-dimensional treatment whereby the leading edges of the vanes are leaned forward to operate with near orthogonal ruling to minimise losses. This is required due to the high flare of the outer casing of the turbine. Some trailing edge lean was also required for the improved design. Presented at the nternational Gas Turbine and Aeroengine Congress and Exposition The Hague, Netherlands June 13-16, 1994

2 Typical results of the 52 calculation are shown on Figures 4a to 4f. These computations specified the final velocity triangles used for the turbine and the overall loss and deviation of each blade row. These data were then used as boundary conditions for the blade-to-blade design of the turbine. Turbine Blade-to-Blade Design The nitial Turbine The mproved Turbine The blade-to-blade design utilised Two-Dimensional (20) Euler ( Denton, 1982) and Navier-Stokes (Dawes, 1986) methods. For the final aerofoil lift coefficients and loadings a Three-Dimensional (3D) Euler method (Denton, 1984) was also used. The new aerofoil geometries were optimised to reduce the effects of poor surface curvature and the resulting aerofoils were, usually, operated with peak suction surface Mach numbers which were subsonic or only marginally transonic. A typical comparison between the old and new aerofoil Mach number distributions are shown on Figure-5. t can be seen that the peak suction Mach number on the new stator#1 aerofoil is reduced even though the lift coefficient of the two design are almost identical. Consequently, the shock loss of the new aerofoils are considerably reduced. n addition the surface curvatures of the new aerofoils show a much smoother progression which leads to a smoother Math number distribution. This results in a more uniform growth of the aerofoil boundary layers and reduces the overall profile loss. 20 design Conventional blade cooling Labyrinth sealings GT8C 3D-design (CAD) State of the art blade cooling mproved labyrinth sealings. Figure-1 Comparison of Old & New Turbines ' Comparison of the Old & New 0T8 Tu bine Anullus Line Figure -3 n order to establish the aerodynamic quality of the turbine and the resulting efficiency, boundary layer calculations were carried out on the aerofoils. The procedure utilises a calibrated integral boundary layer method developed at ABB and also a 20-Finite difference scheme as reported by Harasgama, et al (1993). The 2D/13L program is also used for the analysis of the external heat loads to the turbine aerofoils. Figure - Ba & 8b show the skin friction coefficient (CO for the vane and blade of the 1st turbine stage. t can be seen that the Cf values indicate that the boundary layer is well away from separation as indeed do the shape factors (H12) as shown on Figure 9a & 9b. The overall aerodynamic performance of the turbine depends not only on the profile losses but on the secondary losses such as endwall boundary layer, tip leakage flow, secondary flows, film cooling mixing and base pressure (trailing edge losses). The data for such losses have been generated within ABS over many years and the resulting correlations have been utilised to evaluate the final Total-to-Total losses in the machine. Table-2 shows the comparison of the old and new turbine aerodynamics. Table -2 Aerodynamic Performance of GT8 Vs GT8C Turbine Total-to-Total Pressure Loss Flow Vane-1 Blade- Vane-2 Blade-2 Vane-3 Blade-3 Number Old 4, ,4 GT8 New GTSC As can be seen the main reductions in loss have been in the ro or blading. This comes partly from better aerofoil aerodynamics and partly from better leakage control in the whole turbine. The detail of the leakage control can be seen in Figure 10. Efficient secondary air sealing has been achieved by using honeycomb material and, wherever possible, double sealing strips mounted upon the turbine rotor blading shroud segments. Labyrinth seals have been used on the hub sections of all components. Turbine Heat Transfer and Cooling Figure 2 Streamline Curvature (52) Calculation Grld (143 re 12) As stated above, the initial design of the turbine was carried out with 2D methods. The aerofoil shapes were optimised to achieve the correct exit angles (in Si mode) as established with the S2 calculations. The predicted Mach numbers of the 2D calculations for the aerofoils were checked with the 3D-Euler method and some of the resulting Math number contours are shown on Figure - 6a & 6b - Vane# 2 & Rotor#2. The corresponding aerofoils surface Mach numbers are shown on Figure 7a & 7b. These analyses were performed for all six (6) blade rows. One of the most important design requirements of the industrial gas turbine is the life time requirement. Aerofoil life must be guaranteed to 48,000 hours life (5 years running time) for such machines. This necessitates careful calculation of external heat loads and internal cooling systems to satisfy the customer requirement. External Heat Transfer & Film Cooling A feature of industrial turbines is that the external heat load with time increases rapidly and can reach twice that of initial operation. 2

3 es r, et: a: ED i ( \\*. ass es , clod deg Fig-4a, Flow & Profile Angles, 'Nadel nlet Fig-SP, Flow S Prof i 0 Angles, Vanern nlet 5. nen 775.6,,._.. 4: / deg eleg ' F gac, Flow & Pro le Angles, Bladert2 nlet Fig-4d, Flow & Pro le Angles, Venet$3 nlet This is brought about by the deposition of combustion particles leading to an increased aerofoil surface roughness with time. Other features such as pitting and foreign object damage also contribute to high gas path heat loads. Figure-11 shows the influence of such phenomena on surface roughness as measured on typical ABB gas turbine blading. t can be seen that the roughness can build up to around 25microns over the life time of the blading. ABB has built up calibrated design programs (20-Boundary Layer) which take into account the influence of roughness on hot gas path heat transfer (Harasgama et al and Tarada & Suzuki-1993). These programs have been validated against cascade testing of the GT8 Vane#1 and Blade#1 gas side heat transfer. Figure-12a shows the external heat transfer coefficient, measured Vs predicted. t can be seen that the agreement is indeed very good. Also shown (Figure 12b) is the predicted heat transfer coefficients for the present vane at design operating conditions with and without roughness. Another feature of turbine blade heat transfer is the significant non-uniformity of endwall (Platform & Shroud) heat transfer. The secondary flows set up within the endwall boundary layers and the horse-shoe vortex interact in a complex fluid dynamic manner which generates this non-uniformity ( Harasgama et al , Blair , Harvey ). Examination of the reported experimental data has resulted in physical understanding of the processes governing endwall heat transfer. Recent progress has led to the development of an analytical method coupled with experimental correlations which successfully predicts endwall heat transfer. Figure-13 shows the predicted Nusselt number distribution over the GT8Nane#1 inner platform. These data were used to determine the level of cooling required to operate the GT8C blading at a safe level to obtain the required life-time. 3

4 ConECTED SP7 SsOis Cuts mat i L. deg Fig-4e, FlowS Profile Angles Blarlee3 nlet Mach Mumbo Mach Nu mber as 4.1. as e.., s at tit s ss is Oki G11(3) la Stagy Bledng Dscontntsata Curraluns Leads to Pow votary OssinbuSorts Sheck WaseS and Agit Scot;! a 3 Blade Cooling Design The temperature levels of the GT8C turbine dictated that film cooling was required on the 1st stage nozzle guide vane. Leading edge heat transfer coefficients also indicated that impingement cooling was required. The film cooling of Vane#1 was established by careful consideration of the external aerodynamics and heat transfer. t was decided to use films on the suction side close to the crown of the aerofoil. The trailing edge region on the pressure side was also filmed. Calculations were performed with a 30/Navier- Stokes code (Dawes-1993). The procedure utilised an unstructured adaptive grid and ABB have developed the process of including film cooling with this program ( Weigand & Ha 071(C) S stage lading Sseettli awns,. sesong to OCCel Velcan, thstistudans and by/ Loss Figure 5 mprovements in Veneer Aerodynamics. Harasgama -1994).Figure-14 shows the computations carried out on the GT8 aerofoil with film cooling. t was established that the stagnation point was at 13 0 around on the pressure surface (positive incidence). Subsequent experimental data indicated an identical result. t was therefore with some confidence that the film cooling locations and internal impingement locations could be placed with accuracy. This vane has a separate leading edge compartment for better control of the impingement region. Figure- 15 shows the final cooling design of the new GT8C vane, also shown is the old GT8. Mean Section Vane#2 Figure 6 sentropic Mach Number Distributions Blade#2 4

5 MEAN / + ME / ri , AN PRPHRY Mach Nu mber \\, ',.///,, Vane#2 Blade#2 Figure 7 Surface lsentropic Mach numbers Figure 8a Skin Friction Coefficient Vane#1 Figure 8b Skin Friction Coefficient Eilade# r P essuro Sid 0075 Pressure Side i o o. 0015!, 1, i i ir 1 A! i i Evaluation of current experience on platforms of the GT8 vane indicated high heat loads. This is now predictable and the new GT8C blading have improved platform cooling, Figure-16. The stage-1 rotor blade cooling schemes for the old and new designs are compared on Figure-17. The new blade has "Chevron" shaped ribs placed within its internal cooling passages to produce a highly augmented internal cooling system. The leading edge has a separate cooling channel for better control of cooling air distribution. Coolant is discharged over the tip of the o ,2) blade to prevent tip problems and the balance is exhausted through the trailing edge passage. Thermal Analysis of the Blading The resulting internal and external heat transfer rates were used in a conduction solver to establish the temperature distributions on the aerofoil surfaces and within the material of the aerofoils. A typical result is shown on Figure-18. The finite element grid used for the calculations is also shown on the same figure. 3 5

6 Figure 9a Shape Factor Vane#1 Figure 9b Shape Factor Blade#1 H12 4, , ---\ Pressure S -de Pressure Side _ a!nur. 11 Measured Surface ReuchnesS on Opmmtkmal Take. Figure 10 Overview of Turbine & Leakage Control n order to more easily compare the improvement in cooling design, Figures 19 & 20 indicate the surface temperatures on the old and new Stagelt1 vane and blade respectively. t can be seen that Vane#1 has a reduction in surface temperature of approximately 75 0C whilst the blade indicates a reduction of 4000, both around the leading edge area. t is also dear that the temperature gradients for both aerofoils is much more uniform and this will result in a longer thermo-mechanical life-time. Conclusions Table-3: Final Performance Parameters for the GT8C Turbine GT8 GT8C Gross Electrical Output MW Gross Elect. Efficiency % Exhaust Mass Flow kg/s Pressure Ratio nlet Temp. (SO Defn.) CC Firing Temp. (ANS Dot) PC Exhaust Gas Temp. oc The original targets laid out for the turbine improvements have been achieved. Table-3, below indicates the final performance of the Uprated turbine unit. 6

7 Heat Trcnsfer Coefficient (W/m 2/K) Oisionce from.e. (in) Succor He. Figure 13 nner Platform Nusult Numbers Rough Diatonce from i.e. (rn) Figure 12b PM Span Heat Transfer Coefficients mprovements in turbine aerodynamics were identified and put into practice for the new turbine. Annulus geometry was improved and so was the aerofoil aerodynamics. n particular, it was found beneficial to lean the leading and trailing edges of the aeroloils for loss reduction. The turbine heat transfer and cooling has been the subject of an intensive experimental development study. The cooling design has been radically improved. This has been brought about by a more accurate prediction of external & internal heat transfer coupled with more modern cooling design philosophy. Finally, the temperature distributions in the cooled blading have been made significantly more uniform due to the better cooling design. The stage#1 vane has a temperature reduction of around 750C, whilst the corresponding blade has a reduction of 40 0C. Based on the materials used in the new design, it is expected that the uprated turbine will have a much longer thermo-mechanical life-time than its predecessor. Linear Cascade Test Facility (EPFL, Lausanne) OD sa. Figure 12a Linear cascade tots facility and measured heat transfer coefficients

8 Figure 14 Computations ot the GT8 film Cooled VaneM1 Old Vane Old Made an Mae Figure 15 VaneS1 Cooling Schemes Figura 17 Bladial Cooling SchemeS REDUCED MPNGEMENT GAP TOR MPROVED COOLNG NNER PLATFORM COOLED OLD VANE NEW VANE Figure 15 mprovements n Platform Coolng 8

9 (References; Denton, J.D.(1982) An improved time marching method for turbomachinery flow applications. AMSE Paper No: 82-61%129 Denton, J.D. (1976) Extension of the finite area time marching method to three-dimensions. VK Lecture series VK-LS-84 Dawes W.N. A numerical method for the analysis of threedimensional viscous compressible flows in turbine cascades: Application to secondary flow development in a cascade with & without dihedral. ASME Paper No: 86-GT-145, 1986 Nine Flii110 Cenral, calculations 01 Madan temperatures al m10 span MPROVEMENT TO 1st VANE FOR CT 8C Dawes, W.N. Simulating unsteady turbomachinery flows on unstructured meshes which adapt both in time and space. ASME Paper No: 93-GT-1 04, 1993 Harasgama, S.P. Crawford, ME., Tarada, F. Neelakantan & Baumann, R. (1993) Calculation of heat transfer to turbine blading using two-dimensional boundary layer methods. ASME Paper No: 93-GT-79 Tarada, F & Suzuki, M. (1993) External heat transfer enhancement to turbine blading due to roughness. ASME Paper No: 93-GT-74 Harasgama, S.P. & Wedlake, ET. (1991) Heat transfer and aerodynamics of a high rim speed turbine nozzle guide vane as tested in the RAE sentropic Light Piston Cascade. ASME, Jr11. Turbomachinery, Vol.-113, pp la SO 60 SURFACE LENGTH - % Harasgama, S.P. & Burton, C41992) Film cooling research on the endwalls of a turbine nozzle guide vane in a short duration annular cascade. Parts - 1 & 2, ASME, Jn of Turbomachinery, Vol. 114 Flgyra 19 liudspen temperature reduction,. Ventre' Blair, M.F. (1974) An experimental study of heat transfer & film cooling on a large scale turbine endwall. ASME Paper No: 74-CT- 33 MPROVEMENT TO 1st BLADE FOR CT 8C GT 813 GT SC Harvey, NM. (1991) Turbine endwall heat transfer Oxford University, PhD Dissertation Weigand, B & Harasgama, S.P. (1994) Computations of a film cooled turbine rotor blade with non-uniform nlet temperature distribution using a three-dimensional viscous procedure. lobe published, ASME Gas Turbine Conference, Hague, PSSUM FAS SURFACE LENGTH - % Flours 213 anaemia teinporatur0 reductions. Btadeel 9