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1HE AMERICAN SOCIETY OF MECHANICAL ENGINEERS 345 E. 47th St., New York, N.Y. 10017 96-GT-150 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. Authorization to photocopy material for internal or personal use under circumstance not falling 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 SOIVICO 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 shout be addressed to the ASME Technical Publishing Department. All Rights Reserved Copyright C 1996 by ASME Printed in U.SA. Dieter Bohn*, Karsten Kusterern and Harald Schonenborn" Institute of Steam and Gas Turbines Aachen University of Technology Germany ABSTRACT NOMENCLATURE High process efficiencies and high power-weight ratios are two major requirements for the economic operation of present day gas turbines. This development leads to extremely high turbine inlet temperatures and adjusted pressure ratios. The permissible hot gas temperature is limited by the material temperature of the blade. Intensive cooling is required to guarantee an economically acceptable life of the components which are in contact with the hot gas. Although film-cooling has been successfully in use for a chord length blowing ratio Ma Mach number passage vortex pressure side PS Re Reynolds number shed vortex SS suction side temperature heat transfer coefficient pressure velocity cartesian coordinate couple of years along the suction side and pressure side, problems occur in the vicinity of the stagnation point due to high stagnation pressures and opposed momentum fluxes. In this area basic investigations are necessary to achieve a reliable design of the cooled blade. In the present calculations, a code for the coupled simulation of fluid flow and heat transfer in solid bodies is employed. The numerical scheme works on the basis of an implicit finite volume method combined with a multi-block technique. The full, compressible 3-1) Navier-Stokes equations are solved within the fluid region and the Fourier equation for heat conduction is solved within the solid body region. An elliptic grid generator is used for the generation of the structured computational grid, which is a ;.‘ combination of various C-type and H-type grids. Results of a 3-D numerical simulation of the flow through a turbine blade cascade with and without cooling ejection at the leading edge through two slots are presented. The results are compared with 2-1) numerical simulations and experimental results. It is shown that the distribution of the coolant on the blade surface is influenced by secondary flow phenomena which can not be taken into account by the 2-D simulations. Further coupled simulations with non-adiabatic walls in the leading edge region are performed with realistic temperature ratios and compared to the same case with adiabatic walls. It is shown that in the case of 111111111111, [11111111111 Greek letters density Indices is 1 2 coolant conditions isenuopic total at inlet at exit INTRODUCTION In present day gas turbines, film-cooling is a very efficient cooling method in order to protect the blade material from direct contact with the hot gas. Although film-cooling has been in use for over ten years along the suction side and the pressure side (e.g. (Schonung, Rodi, 1987), (Bassi et al., 1992), (Bohn, Bonhoff, 1994)), problems occur in the vicinity of the stagnation point due to high stagnation pressures and opposed momentum fluxes. But non-adiabatic walls the temperature on the blade wall is the leading edge in particular has to be protected efficiently because thermal loading is at its highest in this region. Thus, basic investigations are necessary in this area to achieve a reliable significantly lower than in the case of adiabatic walls. design for cooled blades. Full professor, director " Research engineer Presented at the International Gas Turbine and Aeroengine Congress & Exhibition Birmingham, UK — June 10-13, 1996 Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/GT1996/78750/V004T09A002/4215782/v004t09a002-96-gt-150.pdf by guest on 17 November 2020 3-D NUMERICAL SIMULATION OF THE FLOW THROUGH A TURBINE BLADE CASCADE WITH COOLING INJECTION AT THE LEADING EDGE found in (Bohn et al., 1995). To demonstrate the performance of the coupling procedure, results from Bohn et al. (1995) are shown in Fig. 1. The calculated and experimentally (Hylton et al., 1983) determined temperatures along the blade surface of a turbine nozzle guide vane with ten cooling channels, admitted with air, are shown. Only the average coolant temperature and heat transfer coefficient in the channels were prescribed according to the experiments. The surface temperature was determined fully coupled. The maximum local differences are smaller than 2% over the whole surface. The advantage of this procedure is that fluid flow and heat transfer can be determined in one single code without any further external iteration processes. The governing equations for the conservative variables are formulated in arbitrary, body-fitted coordinates in order to allow the simulation of complex geometries. The conservation equations are discretized implicitly first order in time making use of the Newton method (Schmatz, 1988). Upwind discretization is used for the inviscid fluxes With respect to numerical diffusion. Godunov type flux-differencing (Eberle, 1987) is employed. In order to achieve third order accuracy, van Leer's MUSCL-technique (Anderson et al., 1985).is used. Since this Godunov flux is not sufficiently diffusive to guarantee stability in regions with high gradients (Schmatz, 1989), it is combined with a hyperdiffusive modified Steger-Warming flux (Eberle et al., 1990). The viscous fluxes are approximated using central differences. The resulting system of linear equations is solved by a GaussSeidel point iteration scheme allowing high vectorization on present day computers. The closure of the conservation equations is provided by the algebraic eddy-viscosity turbulence model by Baldwin and Lomax (1978). A grid consisting of six blocks was generated for the 2-D simulations using an elliptic grid generation procedure. Since the blade under investigation is a plain cascade, the 3-1) grid was obtained by stacking the 2-D grid in radial direction, taking into MATHEMATICAL MODEL The numerical scheme for the coupled simulation of the fluid flow and heat transfer works on the basis of an implicit finite volume method combined with a multiblock technique (Bohn et al., 1995). The physical domain is divided into separate blocks for the fluid and solid body regions. The full, compressible, two- or three-dimensional Mater-Stokes equations are solved in the fluid blocks. The Fourier equation is solved in the solid body blocks. The coupling of fluid blocks and solid body blocks is achieved via a common wall temperature for each cell, resulting from the equality of the local heat fluxes passing through the contacting cell faces. This common wall temperature serves as a boundary condition for the present time step in the grid zone under consideration. A more detailed description of the coupling can be pt i=3.34 bar 0.9 T1 1=788 K Taref Tref = 811 K o Experiment (Hylton et al.) cr i =900 —411• 0.8 (Mal= 0.19) — Calculation (Bohn et al.) ' fr. 0.7 AL• If 0.6 pressure side suction side PP. • P2 /Pt1=0.5 (Ma2is =1.04) --i. 03 1.0 03 0.0 0.5 Fie. 1: TEMPERATURE DISTRIBUTION ALONG THE BLADE SURFACE OF THE CONVECTION COOLED N0771 F GUIDE VANE MARK 11 2 -.1 Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/GT1996/78750/V004T09A002/4215782/v004t09a002-96-gt-150.pdf by guest on 17 November 2020 By means of experiments, Beeck (1992) investigated a gas turbine blade with coolant ejection at the leading edge through two slots. These experiments were mainly focused on the aerodynamic behaviour of the ejection; thus the temperature ratio of coolant and main flow was nearly equal to unity. This configuration was investigated by Beeck et al. (1992) and Irmisch (1995) numerically, using a 2-D code. The distribution of the coolant over the blade surface and thus the cooling efficiency are influenced significantly by secondary flow in the blade passage and further flow phenomena induced by the inhomogeneous slot geometry. This can only be taken into account by 3-D simulations. Vogel (1994) performed 2-D and 3-D calculations of the aerodynamics of this case. But, to be able to draw conclusions about the thermal effects of the film-cooling, real temperature ratios between coolant and main stream have to be considered. Furthermore, non-adiabatic walls have to be taken into account to determine the thermal load in the vicinity of the stagnation point. These investigations are presented here, following the presentation of the numerical model and the validation of the code against the experimental data. mid-plane, symmetry condition BC Exit: P2 Fig. 2: COMPUTATIONAL DOMAIN AND GRID account the two interruptions of the slots. These are responsible for interesting 3-D effects, which are explained later. Figure 2 shows the computational domain and the grid employed for the 3-1) investigations for one half of the passage. The grid consists of a total of 233,826 grid points. Symmetry conditions are used in the midplane. The detail shows the discretization of the leading edge. Because the flow development in the inner geometry significantly influences the flow behaviour at the exit of the ejection slots, it is important to include the inner computational results on the pressure side is very good. On the suction side between x/L=03 and x/1:---0.6, the pressure is predicted a few percent higher than the experimental data. Further downstream, the agreement between experiments and calculations is again satisfactory. A 2-0 grid independence study with 60 additional grid points in flow direction for the C-type grid, which encloses the blade, showed that the pressure distribution does not depend on the grid. Therefore, the smaller grid was chosen as a geometry in the simulations. Total pressure, total temperature and incidence at the inlet of the cascade and the coolant inlet are prescribed as boundary conditions. The static pressure is fixed at the exit No inlet profile in radial direction was prescribed because the boundary layer was sucked off in the experiments. and running time. As an example of the secondary flow occuring in the passage, the flow vectors calculated in a plane downstream of the trailing edge normal to the main flow direction are displayed in Figure 4. This figure represents a qualitative sketch of the flow phenomena RESULTS which can be observed in this plane, looking upstream. The shed vortex S I can clearly be discerned. Due to the position of the basis for the 3-0 calculations in order to save computer memory Aerodynamic Investigations Table 1 shows the conditions at the passage inlet and exit of the two cases investigated with (M=0.47) and without ejection (M=0.0). The incidence is equal to at=43 ° for both M=0.0 M=0.47 Ma inlet 0.375 0.38 Ttot inlet [IC] 312.3 3124 ptot inlet [iPa] 201.3 201.7 p exit BiPal 138.9 147.4 Re inlet 370000 370000 0 experimental data — 3-D calculation -- 2-D calculation 0.90 cases. Tab, 1 . TEST CASE CONDITIONS Figure 3 presents a comparison between the numerically and experimentally (Beeck, 1992) determined pressure distribution for the case without ejection along the blade surface in the mid-plane. In addition, the results for the 2-D simulations are displayed. Distribution for the 2-0 and 3-1D simulations is nearly identical. Thus, the 3-0 effects have very little influence on the pressure distribution in the midplane. The coincidence between the experimental data and the 0.80 0.70 0.60 0.50 00 0.2 0.4 0.6 0.8 x/L Fig 3. PRESSURE DISTRIBUTION WITHOUT EJECTION (M).0) 10 Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/GT1996/78750/V004T09A002/4215782/v004t09a002-96-gt-150.pdf by guest on 17 November 2020 Cutting plane in Fig. 4 S shed vortex P passage vortex • • • SI S2 ty, :1111 114.F.-i zi :111 ••••••••••0111•• dries. P2 __swat' 1 Fig. 4: SECONDARY FLOW VECTORS IN A PLANE DOWNSTREAM OF THE TRAILING EDGE cutting plane it is possible to observe another shed vortex S 2 from the neighbouring blade, which is already further developed. The comparison of the two vortices shows that the center of the vortex is displaced toward the midspan and widens in tangential direction due to the dissipation. The comparison of the passage vortices P 1 and P 2 shows that the passage vortex does not change its position. The passage vortex, too, loses some of its strength. In comparison to the shed vortex, the passage vortex is quite small. This is due to the fact that much of the energy of the passage vortex is lost to the shed vortex. Furthermore, secondary flow from the pressure side to the suction side along the lower side wall can be discerned. The same observations were also made by Bassi and Savini (1992) in their numerical simulations of a different case. The investigations of the blade with ejection at the leading edge were performed with a blowing ratio of M=(Pw)c/(Pw)i= 0.47 . Figure 5 shows the corresponding pressure Experiment (Beeck, 1992) experimental data 3-D calculation 2-D calculation 1.00 P/N Fig. 6: STREAMLINES AT THE LEADING EDGE From the observations made in Figure 5 and 6 it can be deduced that the code predicts the pressure distribution in the case of filmcooling with ejection at the leading edge with a sufficient level of accuracy, even with 2-D simulations. But the results presented further on demonstrate that the coolant is distributed unevenly on the blade surface, something that has severe consequences for the thermal aspects if real temperature levels are used. 3-D investigations are necessary in order to take these effects into account. The 3-D pressure distribution calculated on the suction side and on the pressure side is displayed in Figure 7. It can be seen that on the suction side the pressure decreases until the minimum is reached and increases again towards the trailing edge. There is almost no pressure gradient in radial direction except close to the lower side wall of the blade. The locally higher pressure in this 0.90 0.80 0.70 0.60 0.50 00 0.2 0.4 0.6 Calculation 0.8 x/L 10 Fig, 5: PRESSURE DISTRIBUTION WITH EJECTION 4 Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/GT1996/78750/V004T09A002/4215782/v004t09a002-96-gt-150.pdf by guest on 17 November 2020 distributions for the experiments, the 2-0 and the 3-0 calculations in the mid-plane. The broader pressure minimum on the pressure side immediately downstream of the leading edge, as compared to the case without ejection (Fig. 3), is due to a separation bubble. This region is predicted somewhat better by the 3-D calculations than by the 2-0 calculations. This is due to secondary flow phenomena which influence the size of the separation bubble as explained later. Overall, coincidence between experiments and calculations is good, except on the suction side close to the trailing edge. The differences between the 2-D and 3-D simulations are again rather small. In Figure 6, streamlines in the region of the leading edge are presented. The experimental results are based on Laser-2-Focus measurements (Beeck, 1992). The corresponding 2-D numerical results are presented on the right hand side. It can be seen that both the position of the stagnation point and the size of the separation bubble on the pressure side show good qualitative agreement with the experimental results. As already observed in the experiments, and as documented by Irmisch (1995), the mass flow rates are distributed unevenly between the two slots. This can be seen by the distribution of the streamlines in the inner region. The asymmetric arrangement of the two slots and the different curvatures on the pressure side and the suction side lead to a higher mass flow rate through the slot on the pressure side (PS) than through the slot on the suction side (SS). A very small separation bubble downstream of the slot on the suction side can be recognized. suction s pressure side mid-plane 0.1883 0.1202 0.1856 \•-\\\_,LN,■....■___,2.2.2j y i dip = 0.0032 bar [p] = bar CTh lower side wal lower side wall leading edge 4 = 0.0027 bar Fit 7. CALCULATED ISOLINES OF PRESSURE ON SUCTION SIDE AND PRESSURE SIDE region is due to the secondary flow directed towards the blade wall. The influence of the paths on pressure distribution on the suction side is negligible. Contrary to this, locally higher pressure can be observed on the pressure side downstream of the paths. In the region downstream of the slots, where the separation bubble is located, a pressure minimum can be seen. The influence of this inhomogeneous pressure field on the distribution of the coolant is explained in the following figures. Figure 8 shows the distribution of streamlines coming out of the middle of the slots on the pressure side. It can be seen that a large region of the blade surface is not covered at all by the coolant. This is due to the fact that local pressure maxima are formed in the region between the two slots and close to the lower blade wall. This leads to a displacement of the coolant from the upper slot towards the mid-plane and to a contraction of the cooling film from the lower slot. Material from the main stream is conducted under the cooling film and this leads to displacement of the film. At the end of the slots, it can be seen that streamlines are sucked into the separation bubble due to the pressure field and are lost for cooling purposes. Due to the symmetry conditions in the mid-plane, the above-mentioned effect does not take place in this plane. Because of the displacement effect mentioned previously, the separation bubble is larger than in the 2-D case. On the right hand side of Figure 8, secondary flow vectors in a cutting plane at 7JL = 0.9 perpendicular to the blade surface are displayed. The approximate location of the plane is indicated on the blade surface. This diagram underlines the previous observations. The distribution of the streamlines on the suction side is presented in Figure 9. It can be observed that the streamlines coming out of the upper slot are distributed almost evenly on the blade surface due to the homogeneous pressure field in this region. The cooling film formed by the fluid coming out of the lower slot is displaced towards the mid-plane, so that the region close to the lower side wall of the suction side is not covered by the coolant This phenomenon can be attributed to the passage vortex. The pressure gradient in the passage leads to a greater diversion of the low momentum fluid in the side wall boundary layer than of the main flow. Thus, it is conducted towards the suction side. Here it changes direction and flows towards the midplane. This can be observed by means of the secondary flow vectors in the cutting plane at x/L = 0.9 perpendicular to the blade wall. It should be noted that the view direction is opposed to the view direction downstream mid-plane 2 4 view direction upstream I mid-plane 3 cutting-plane a x/L=0.9 cutting-plane at x/L=0.9 Fig 9. STREAMLINES ON THE BLADE SURFACE - Fig, 8 . STREAMLINES ON THE BLADE SURFACE PRESSURE SIDE SUCTION SIDE 5 Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/GT1996/78750/V004T09A002/4215782/v004t09a002-96-gt-150.pdf by guest on 17 November 2020 0.117 stream. The interaction between the cooling film and the passage vortex leads to a distortion of the passage vortex. The extension of the passage vortex towards midspan is disturbed by the cooling flow direction of the main 1.1 .0 Figures 8 and 9 only show the qualitative distribution of the coolant. Due to the view direction, the location of the streamlines Tirtl — non-adiabatic ——— -- adiabatic 0.9 0.8 0.7 Thermal Investigations Because the experimental investigations were conducted with the injection of cold gas into a cold mean stream in order to 0.6 investigate the aerodynamic aspects of film-cooling, 2-D numerical investigations with realistic temperature ratios of coolant and main flow were also performed in order to determine the temperature distribution on the blade surface. The temperature 0.5 ,-- — -. -. --. -.... -... .. S. 5. .9 .--- , I \ i .. . .. pressure Side / suction side injection slot injection slot leading edge ratio chosen was Inal=0.5, which corresponds to the ratio of the air temperature at the last compressor stage and the turbine inlet temperature under normal operating conditions for present day gas turbines. The total pressure chosen at the coolant inlet was the same as for the case with the temperature ratio equal to unity. The coolant conditions thus changed result in a blowing factor of Fig. 11 . TEMPERATURE DISTRIBUTION ALONG THE BLADE SURFACE AT THE LEADING EDGE M=0.56, which is still comparable to the previous one. Special attention was paid to the region around the leading edge. Because the usual assumption of adiabatic walls gives temperatures in the vicinity of the stagnation point which are far too high, simulations i/ / with non-adiabatic walls were conducted and compared to the same simulations with adiabatic walls. Figure 10 shows the discretization of the leading edge with the non-adiabatic walls. Five solid body blocks were added to the 2-D grid. This was possible due to the multi-block technique, which allows an arbitrary number of fluid blocks and solid body blocks. The numerical model described above allows the iteration-free coupling of these blocks in just one code and thus allows the _ ki/A.,, 111 determination of the temperature influence in this region. áT=5 .4 K .4, Aillii "4 detail Figure 14 adiabatic calculation / i tb k, k AT=5.4 K -i. SIM detail Figure 14 non-adiabatic calculation Fig. 12z ISOLINES OF TEMPERATURE IN THE LEADING EDGE REGION Fig. 0: LEADING EDGE DISCRETIZATION WITH NON-ADIABATIC WALLS 6 Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/GT1996/78750/V004T09A002/4215782/v004t09a002-96-gt-150.pdf by guest on 17 November 2020 in relation to the blade surface is difficult to discern, but the effect of the pressure field on the coolant can be demonstrated by these figures. fluid solid body fluid 1.10 Tau 1.03 0.80 0.70 0.60 0.50 adiabatic calculation A DES TEMPERATURE DISTRIBUTION ALONG THE CROSS SECTION A-B In Figure 11, a comparison of the temperature distributions along the blade surface in the immediate vicinity of the leading edge with adiabatic and non-adiabatic walls is displayed. At the leading edge itself there is a significant difference in the surface temperature. While in the adiabatic case, the temperature is equal to the stagnation temperature, in the non-adiabatic case, the temperature is only about 60% of said temperature. This difference in the determination of the surface temperature has severe consequences for the predicted life-time of the blade. The temperatures at the exit of the injection slots are almost equal. But then the surface temperature on the suction side as well as on the pressure side increases in the adiabatic case while in the case of non-adiabatic walls, the temperature remains nearly constant due to the heat flux into the blade material. non-adiabatic calculation fly. 14. DETAILS OF FIGURE 12- FLOW VECTORS AND ISOLINES OF TEMPERATURE The previous observations can be explained qualitatively by adiabatic walls demonstrate the ability of the code to predict a means of the isolines of the temperature in the vicinity of the leading edge in Figure 12, Figure 13 shows the temperature along the cross section A-B indicated in Figure 12. It can be seen that there is a large heat flux from the leading edge into the blade material, thereby heating up the coolant in the slots. This is demonstrated by the isolines within the slot geometry. Furthermore, it can be seen that the different surface temperature realistic temperature field in the flow and blade wall. In film-cooling, the coolant is usually ejected through a row of holes. A comparison of slot injection and hole injection with non-adiabatic walls was performed by Bohn et al. (1995). Figure 15 shows a comparison of the temperature distribution on the surface of a blade with injection through a slot and a row of holes on the suction side. The differences are mainly due to the heat conduction in the blade material between the holes. Therefore, it is difficult to transfer results from a case with slot injection to one with hole injection. Further investigations will concentrate on the thermal aspects of 3-D film-cooled turbine blades, taking into leads to different flow fields. In the case with non-adiabatic walls, the separation bubble is slightly larger because the coolant is already warmed up. This means that for the same mass flow, the velocity increases and the coolant jet penetrates a little deeper into the main stream. This effect can also be observed in Figure 14, where a zoomed view at the exit of the pressure side slot is shown with flow vectors and isolines of temperature. In the case of account the heat transfer in the blade walls. CONCLUSIONS adiabatic calculations, the inflow region is larger forming a very hot zone inside the slot wall with large temperature gradients. In the case of the non-adiabatic walls, a more moderate temperature The code presented in this paper predicts the aerodynamic aspects of film-cooling with coolant ejection at the leading edge with a good degree of accuracy. The investigations revealed that some aerodynamic aspects of film-cooling can be taken into level and temperature gradient in this region is predicted, leading to less undesired effects such as thermal stress and corrosion. Considering previous validations of the code, these thermal account by 2-1) simulations, but that the complex 3-1) flow phenomena associated with turbine blades, in general, and the slot investigations of the leading edge region with adiabatic and non- 7 Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/GT1996/78750/V004T09A002/4215782/v004t09a002-96-gt-150.pdf by guest on 17 November 2020 0.90 Mil 1.0 Beeck, A., 1992, "Stremungsfelduntersuchungen zum aerodynamischen Verhalten eines hochbelasteten Turbinengitters mit Killilluftausblasung an der Vorderkante", Ph.D. thesis, pressure side UniBW Munchen 111,11 suction side 0.9 1111111b• - 0.8 _ ---c-hole-injection 0.7 -7-7---=-‘;,s ot-thjection--- -0.6 Heat-Transfer and Cooling in Gas Turbines, Anralaya Irmisch, S., 1995, "Simulation of Film Cooling Aerodynamics with a 2-D-Navier Stokes Solver Using Unstructured Meshes", ASME-paper 95-GT-024 — Vogel, D. T., 1994, "Navier-Stokes Calculation of Turbine Rows with Film Cooling", 1CAS-paper 94-253 „Bohn, D., Bonhoff, B., SchOnenborn, H., Wilhebni, H., 1995, "Valitlitton of. a Numerical Model for the Coupled Simulation of Fluid Flow anethabatic-Walls with Application to Film-cooled Gas Turbine Blades", VD1-13FriChre-Nr..1186, pp. 259-272 Bohn, D., Bonhoff, B., Lang, G.78.arbitenborn,,H., 1995, "Determination of Thermal Stress and Strain BaTsarz--on- a Combined Aerodynamic and Thermal Analysis For a Turbine Nozzle Guide Vane", ASME-paper 95-CT? -089 Bohn, D., Bonhoff, B., Schrinenbom, H., 1995, "Combined Aerodynamic and Thermal Analysis of a Turbine Nozzle Guide Vane", 1GTC-paper-108, Proceedings of the 1995 Yokohama 0.5 0.4 00 t iniection 0.2 04 06 0.8 ya, 1.0 Fig. 15: TEMPERATURE DISTRIBUTION ALONG THE BLADE SURFACE FOR SLOT INJECTION AND HOLE INJECTION geometry, in this special case, result in uneven distribution of the coolant on the blade surface. The 3-D investigations presented show large regions on both suction side and pressure side which are not covered by the coolant. This is due to the inhomogeneous pressure field on the blade surface resulting from secondary flow and the slot geometry. The slot geometry with the ejection opposed to the main stream leads to a large separation on the pressure side. Thus, there is a high potential for aerodynamic optimization by changing the cooling geometry. In addition, 2-1) numerical investigations were carried out with realistic temperature ratios between coolant and hot gas with adiabatic and non-adiabatic walls. It is shown that the surface temperatures are significantly different in the vicinity of the leading edge. A reduction of 40% in the prediction of the leading edge temperature was observed. Even further downstream of the injection slots, the differences are significant. Considering previous validations of the code, this application demonstrates its ability to predict the temperature fields in the flow and blade wall using numerical simulations in order to be able to optimize the design. International Gas Turbine Congress Hylton, L D., Milhec, M. S., Turner, E. R., Nealy, D. A., York, R. E., 1983, "Analytical and Experimental Evaluation of the Heat Transfer Distribution Over the Surfaces of Turbine V anes" , NASA CR 168015, May. 1983 Schmatz, M. A., 1988. "Three-dimensional viscous flow simulations using an implicit relaxation scheme", Notes on Numerical Fluid-Mechanics (NNFM), Vol. 22, Vieweg, Braunschweig, pp. 226-242 Eberle. A., 1987, "Characteristic Flux Averaging Approach to the Solution of Euler's Equations", Lecture Series 4 'Computational Fluid Dynamics at VET, Part I and II Anderson, W. K., Thomas, J. L., von Leer, B., 1985, "A Comparison of Finite Volume Flux Vector Splittings for the Euler Equations", AIM -paper 85-0122 Schmatz, M. A., 1989, "Hypersonic Three-Dimensional Navier-Stokes Calculations for Equilibrium Gas", AIM-paper 89-2183 Eberle, A., Sclunatz, M. A., Bissinger, N., 1990. "Generalized Flux Vectors for Hypersonic Shock-Capturing." AMA -paper 90- REFERENCES 0390 Schonung, B., Rodi, W., 1987, "Prediction of Film-Cooling by a Row of Holes With a Two-Dimensional Boundary-Layer Procedure", Joum. of Turbornachinery Vol 109, pp. 579-587 Bassi, F., Rebay, S., Savini, M., Colantuoni, S., Santoriello, G., 1992, "A Navier-Stokes Solver with Different Turbulence Models Applied to Film-Cooled Turbine Cascades", AGARD-CP- Baldwin, B. S., Lomax, H., 1978, "Thin Layer Approximation and Algebraic Model for Separated Turbulent Flows", AMA- paper 78-257 Bassi, F., Savini, M., 1992 "Secondary Flows in a Transonic Cascade: Validation of a 3-D Navier-Stokes Code", ASME-paper 92-GT-62 527, pp. 41-1 -41-14 Bohn, D., Bonhoff, B., 1994. "Berechnung der Kuhl- und Storwirkung eines filingektihlten, transsonisch durchstromten Turbinengitters mit diabaten Wanden", VDI-Berichte-Nr. 1109, Bohn, D., Bonhoff, B., Schtinenbom, H., Wilbelmi, H., 1995. "Prediction of the Film-cooling Effectiveness of Gas Turbine Blades Using a Numerical Model for the Coupled Simulation of Fluid Flow and Diabatic Walls", ISABE-paper 95-7105, Sept. pp. 261-275 1995, Melbourne 8 Downloaded from http://asmedigitalcollection.asme.org/GT/proceedings-pdf/GT1996/78750/V004T09A002/4215782/v004t09a002-96-gt-150.pdf by guest on 17 November 2020 Beeck, A.. Fottner, L., Benz, E., Wittig, S., 1992, "The Aerodynamic Effect of Coolant Ejection in the Leading Edge Region of a Film-cooled Turbine Blade". AGARD-Meeting: