1HE AMERICAN SOCIETY OF MECHANICAL ENGINEERS
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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
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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
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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
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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
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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
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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
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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
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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
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