David L. Rigby

QSS Group, Inc., Cleveland, Ohio

Ronald S. Bunker

General Electric R&D Center, Schenectady, New York

Heat Transfer in a Complex Trailing Edge

Passage for a High Pressure Turbine Blade

Part 2: Simulation Results

NASA/CR—2002-211701

August 2002

2002–GT–30213
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David L. Rigby

QSS Group, Inc., Cleveland, Ohio

Ronald S. Bunker

General Electric R&D Center, Schenectady, New York

Heat Transfer in a Complex Trailing Edge

Passage for a High Pressure Turbine Blade

Part 2: Simulation Results

NASA/CR—2002-211701

August 2002

National Aeronautics and

Space Administration

Glenn Research Center

Prepared under Contract NAS3–00145

Prepared for the

Turbo Expo 2002

cosponsored by the American Society of Mechanical Engineers

and the International Gas Turbine Institute

Amsterdam, The Netherlands, June 3–6, 2002

2002–GT–30213
Acknowledgments

The present work was supported by the NASA Glenn Research Center as part of the Coolant Flow Management

Program. The authors wish to thank Mr. Steven Hippensteele and Dr. Raymond Gaugler, Chief of the Turbine

Branch, for their support and encouragement of this work. Calculations were done on the NAS SGI clusters at the

NASA Ames Research Center.

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Available electronically at http://gltrs.grc.nasa.gov/GLTRS

The Aerospace Propulsion and Power Program at

NASA Glenn Research Center sponsored this work.
ABSTRACT

A combined experimental and numerical study to investigate

the heat transfer distribution in a complex blade trailing edge

passage was conducted. The geometry consists of a two pass

serpentine passage with taper toward the trailing edge, as well as

from hub to tip. The upflow channel has an average aspect ratio

of roughly 14:1, while the exit passage aspect ratio is about 5:1.

The upflow channel is split in an interrupted way and is smooth

on the trailing edge side of the split and turbulated on the other

side. A turning vane is placed near the tip of the upflow channel.

Reynolds numbers in the range of 31,000 to 61,000, based on

inlet conditions were simulated numerically. The simulation was

performed using the Glenn-HT code, a full three-dimensional

Navier-Stokes solver using the Wilcox k-ω turbulence model. A

structured multi-block grid is used with approximately 4.5

million cells, and average y+ values on the order of unity.

Pressure and heat transfer distributions are presented with

comparison to the experimental data. While there are some

regions with discrepancies, in general the agreement is very

good for both pressure and heat transfer.

NOMENCLATURE

A Area of inlet

friction coefficient,

D inlet hydraulic diameter

h heat transfer coefficient

k thermal conductivity

mass flow rate

Nu Nusselt number, hD/k

P pressure

Pr Prandtl number

C f 2τw/ρV2

m ̇

Wall heat flux

Re Reynolds number, VD/

T Temperature

Tt,in Total temperature at inlet

V characteristic velocity, /( A)

y+ dimensionless distance from the wall,

distance normal to wall

viscosity

density

kinematic viscosity

Subscripts

air bulk condition

t total condition

w wall value

INTRODUCTION

The present work is motivated by the need to accurately

predict heat transfer in turbomachinery. For efficient gas turbine

operation, flow temperatures in the hot gas path exceed

acceptable metal temperatures in many regions of the engine. To

maintain the integrity of the parts for an acceptable engine life,

they must be cooled. Efficient cooling schemes require accurate

heat transfer prediction to minimize regions which are overcooled, and even more importantly to ensure adequate cooling

in high heat flux regions. With the advent of more powerful

computers, three dimensional simulations of the internal coolant

flow are becoming more common. The Glenn-HT code is a

Reynolds averaged Navier-Stokes solver which has proven to be

qw′′

ν

m ̇ ρ

η

D

--- Re

C f

2

------

η

μ

ρ

ν

Heat Transfer in a Complex Trailing Edge Passage for a High Pressure Turbine Blade

 Part 2: Simulation Results

David L. Rigby

QSS Group Inc.

Cleveland, Ohio 44135

Phone: 216-433-5965, Fax: 216-433-5802, Email: rigby@grc.nasa.gov

Ronald S. Bunker

General Electric R&D Center

Schenectady, New York 12301

NASA/CR—2002-211701 1
an accurate and robust simulation tool for flowfield and heat

transfer prediction. Many papers have been published

demonstrating the use of the code on many configurations. Some

recent configurations include turbine rotors with clearance

(Ameri and Bunker [1], Ameri and Rigby [2]), internal/external

flow coupled through cooling holes (Heidemann et al. [3], Garg

and Rigby [4]), and rotating internal passages with ribs (Rigby

[5]). Additionally, simulations have been performed which

include internal and external flows coupled through a conjugate

solution of the solid region (Rigby and Lepicovsky [6])

Traditional cooling passage analysis relies heavily on empirical

data bases and quasi-one-dimensional techniques. Higher fidelity

prediction capabilities are needed to allow attempts at more

advanced and novel configurations. The capabilities of the

current computers make it possible to simulate reasonably

complex geometries, such as the one in the present study. Beyond

obtaining the required computing resources, there are two

additional major issues. As we turn our attention to more

complex geometries, the two additional issues are grid generation

and turbulence modeling.

The present work was undertaken in an attempt to assess our

abilities in grid generation and to test the turbulence model on

increasingly complex geometries. The present geometry includes

variation in channel shape and size in the flow direction. Also,

turbulated regions are combined with smooth regions such that

there are several areas where turbulators do not end at a side wall.

A final complication arises from the interrupted nature of the split

upflow channel, which has turbulators to one side and smooth

region on the other side. Three basics questions were to be

answered in this undertaking: First could a structured multi-block

grid be generated, second would the code run to convergence on

what had to be a less than ideal grid, and third would the

simulation agree with the experimental data? The answer on all

three counts is yes. The commercially available grid generation

software, ICEMCFD, proved to be quite capable of generating a

reasonable grid. The Glenn-HT code also proved to be robust and

produced results which agree very well with the data.

NUMERICAL PROCEDURE

Brief description of Glenn-HT code

The simulations performed in this study were done using a

computer code currently called Glenn-HT. In the past the code

was named TRAF3D.MB (Steinthorsson et al. [7]). This code is a

general purpose flow solver, designed for simulations of flows in

complicated geometries. The code is based on the TRAF3D code,

an efficient computer code designed for simulations of flows in

turbine cascades (Arnone et al. [8]). The Glenn-HT code employs

the full compressible Navier-Stokes equations. It uses a multistage Runge-Kutta scheme to march in pseudo time. For

convergence acceleration the code uses, among other techniques,

local time stepping. The spatial variation in the time step takes

into account both convective and diffusive stability limits. In

addition, the code utilizes multi-grid and implicit residual

smoothing to accelerate convergence to steady state. Convective

and diffusive fluxes are computed using central differencing.

Artificial dissipation is added to prevent odd-even decoupling.

The discretization is formally second order accurate. To handle

complex geometries, the code uses contiguous structured multiblock grid systems but has the added capability of handling grids

with non-contiguous grid lines across branch cuts. For

contiguous systems, all internal boundaries are conservative. The

TRAF3D.MB code was described in detail by Steinthorsson et al.

[7]. Some aspects of the formulation used in the code are the

same as those described by Arnone et al. [8]. The code is fitted

with the low Reynolds number k-ω turbulence model of Wilcox

[9].

DESCRIPTION OF GEOMETRY AND CONDITIONS

Figures 1 and 2 show a schematic of the test model. Figure 1

shows the bottom surface which is flat. The two main passages

are seperated by a continuous 3.18 mm rib. The inlet passage is

divided by an interupted 3.18 mm rib, allowing flow to pass

between 1A and 1B. A turning vane is located in the trailing edge

tip corner. As shown in figure 2, the upper surface slopes from

passage 2 to the trailing edge. In addition, the upper surface

slopes from the root to the tip. All turbulators are square in crosssection and angled at 45 degrees. In passage 1B the turbulators

near the inlet are 2.03 mm high, while the seven nearsest the tip

are 1.52 mm high. In passage 2, the six turbulators nearest the tip

are 1.52 mm high, with the remainder being 2.54 mm high. In

general, the ratio of turbulators spacing to height is 10. On the

upper surface, the turbulators are staggered relative to the lower

surface.

In the experiment, an inlet plenum is used for passage 1. For

the numerical calculation an additional 7.62 cm was added so

there would be no turbulators at the inlet or exit. Without this

extension, it is expected that there would have been problems

with boundary conditions and robustness. Since heat transfer data

is not collected until well into the test section, an acceptable

match in the measurement region was expected.

Figure 3 shows a perspective view of the geometry. The bottom

flat wall is shown in gray and the black lines show the outlines of

grid blocks. Flow enters through passages 1A and 1B, and exits

through passage 2. From figure 3 it can be seen that the passages

are trapezoidal in cross section and are thickest near passage 2

and thinnest near passage 1A, the trailing edge. In addition, the

passages thin as the tip is approached. In figure 3, the location

x=0 is where the experimental test section began.

Figure 4 shows the bottom wall grid in the region which is

heated. The interrupted nature of the split between 1A and 1B,

and the turning vane are visible in figure 4. Also, it can be seen

that passage 1A is smooth, while 1B and 2 have turbulators at a

45 degree angle. The upper wall looks much the same as the

bottom, with the turbulators at the same angle and staggered

NASA/CR—2002-211701 2
relative to those on the bottom wall. Additional details on the

geometry are provided in Part 1 of this work (Bunker et al.

[10]).

Three flow conditions were run numerically. In each case the

exit pressure was the same and the wall temperature in the

heated region was maintained at 1.1 times the inlet total

temperature. The average inlet total pressure for each of the

three cases was 1.28, 1.46, and 1.81 times the exit static

pressure. The total pressure imposed on passage 1A was slightly

less than that on passage 1B as was observed in the experiment.

The resulting Reynolds numbers based on inlet conditions for

the Low, Medium and High pressure ratios were 31,000, 43,000,

and 61,000, respectively. The inlet hydraulic diameter is based

on the combination of both 1A and 1B.

Each case took approximately 1000 cpu hours on a SGI Origin

cluster. Most runs were done with 32 processors in parallel

resulting in running times on the order of 30 hours. This case ran

effeciently even at 95 processors, the largest number attempted.

However, due to the shared nature of the machine, a request for

95 processors resulted in a very long wait in the queue.

Heated Unheated

45.72

30.48

7.62

1B 1A 2

x

y

Trailing Edge Apex

1A

1B

2

x

y

z

x=0

x=15.24 cm x=45.72 

x=38.1 cm

7.14 12.64

1.92 2 0.26 1B 1A

Figure 1. Schematic of flat lower surface.

(dimensions in centimeters)

Figure 3. Perspective view of geometry including

block boundaries.

Figure 4. Grid on heated region.

Figure 2. Schematic of Inlet Cross-Section.

(dimensions in centimeters)

NASA/CR—2002-211701 3
RESULTS

Pressure

Figure 5 shows the pressure distribution down the center of

each of the passages, for the three different pressure ratios. The

results for channel 1A are provided above the figure for clarity,

since they are nearly equal to those in channel 1B. The first

thing apparent from figure 5 is that the pressure distributions are

very similar for each case. The pressures in channel 1A fall

smoothly in the flow direction closely matching the drop in

channel 1B. In channel 1B the pressure is fairly uniform

between turbulators with jumps across the turbulator. The rapid

drop in pressure near the end of channel 1B occurs because of

the rapid acceleration of the flow through the reduced area.

channel 2 shows similar jumps in pressure across the turbulators

to those in channel 1B, along with significant adverse pressure

gradient between the turbulators.

Figure 6 is similar to figure 5, except the experimental results

are included and only the high pressure ratio for the Glenn-HT

case is shown. It is apparent that comparison with the

experiment is very good with the exception of the region of

channel 2 near the tip. In that region, which is the turn and

recovery region, the experimental pressures fall well below the

prediction as well as below the exit pressure. As shown in the

next section, the turn and recovery region is associated with a

discrepency in heat transfer results between CFD and the

experiment. The pressures in channel 1A and 1B compare quite

well with the simulation matching the rate of drop as well as the

rapid drop in channel 1B near the tip. Pressures in the middle of

channel 2 also match well.

In Part 1 of this work the results of a one-dimensional, multielement flow model were used to estimate the flow split between

passages 1A and 1B. Those results are shown in figure 7 along

with the results from the Glenn-HT simulation. The agreement

of the Glenn-HT simulation with the Network Model is very

good. These results also indicate that the flow split is essentially

independent of Reynolds number.

0 5 10 15 20 25 30 35 40

Distance from Inlet [cm]

−0.2

0.0

0.2

0.4

0.6

0.8

1.0

(P−Pout)/(Ptin−Pout)

Lo, Glenn−HT

Medium, Glenn−HT

High, Glenn−HT

Leg 1A

Leg 1B

Leg 2

0 5 10 15 20 25 30 35 40

0.6

0.8

1.0

Figure 5. Normalized pressure down center of each

channel.

0 5 10 15 20 25 30 35 40

Distance from Inlet [cm]

−0.2

0.0

0.2

0.4

0.6

0.8

1.0

(P−Pout)/(Ptin−Pout)

Low, Exp.

Medium, Exp.

High, Exp.

High Glenn−HT, Leg 1

High Glenn−HT, Leg 2

Leg 1A

Leg 1B

Leg 2

0 5 10 15 20 25 30 35 40

0.6

0.8

1.0

Figure 6. Normalized pressure down center of each

channel with experimental data.

0 10 20 30 40

Distance from Inlet [cm]

0.0

0.2

0.4

0.6

0.8

1.0

Fraction of Passage 1 Flow in 1B

 Network Model Prediction

 Lo, Glenn−HT

 Medium, Glenn−HT

 High, Glenn−HT

Figure 7. Fraction of passage 1 mass flow in 1B.

NASA/CR—2002-211701 4
Heat Transfer

Heat transfer results are presented in terms of the heat transfer

coefficient defined by

(1)

where Tair is the average between inlet and exit total

h qw′′

Tw Tair ( ) – ---------------------------=

temperatures, which should produce essentially the same value for

Tair as used in the experiment. For turbulent flows, it is generally

expected that the heat transfer coefficient will vary nearly like the

Reynolds number to the 0.8 power. In light of that expectation the

results for the different cases are presented in normalized form.

The heat transfer coefficient for each case is normalized by the

value predicted by the Dittus-Boelter correlation based on the

Reynolds number at the inlet to passage 1. The Dittus-Boelter

correlation

(2)

predicts the heat transfer coefficient for a smooth pipe with fully

developed turbulent flow.

Figure 8 shows the normalized heat transfer coefficient for the

Low, Medium and High pressure ratio simulations. The purpose of

showing this figure first is to simply show that, when normalized,

the cases are nearly identical. The cases differ significantly in

Reynolds number, doubling between the Low and High pressure

ratios, and yet the match is nearly perfect. So the code is definitely

producing a dependence on Reynolds number to the 0.8 power as

was observed in the experimental results presented in Part 1 of this

work.

Figure 9 shows the normalized heat transfer coefficient for the

experiment at the high pressure ratio. The color bar is labeled for

both normalized and dimensional [W/m2K] heat transfer

coefficient. Since the colors are chosen to match between the CFD

and experimental figures, one can now compare the experimental

results in figure 9 to the numerical results in figure 8. The white

h0 0.023 k

D

---     ReD

0.8Pr0.4 =

1.75 3.752.75 4.75 5.75 h/h0

a.)

b.)

c.)

1.75 3.752.75 4.75 5.75 h/h0

h

[W/m2/K]

Figure 9. Normalized heat transfer coefficient for

High Pressure Ratio experiment. (color bar also

shows dimensional value.

Figure 8. Normalized heat transfer coefficient for

Low(a), Medium(b), and High(c) pressure ratio

simulations.

NASA/CR—2002-211701 5
regions in the experimental figure are where data was not

obtained, due to temperature limitations of the model materials.

Qualitatively, the comparison is very good. High heat transfer is

observed between the turbulators, as expected. In addition, the

low streak emanating from the tip of the split between passages

1A and 1B is closely matched. Very high values after the final

rib in passage 1B, as well as the low region around the tip of the

split between passages 1B and 2, are visible in both experiment

and simulation. Although, the simulation produces a much

larger low heat transfer region at the entrance to passage 2. The

rise in heat transfer on the outer wall of passage 2 shown in the

experiment is also predicted by the simulation. The level of

increase in heat transfer between the turbulators is generally

underpredicted by the simulation. The level in the smooth

passage, 1A, is predicted closely by the simulation.

In figure 10, the normalized heat transfer coefficient on a line

down the middle of passage 1B is shown for the Low, Medium

and High pressure ratio cases. One can now see, in a

quantitative way, the similarity between the solutions. The

average of each of the lines in figure 10 are each within ±5%

and the maximum diference is <10%.

Figure 11 shows the same result as figures 8c and 9, except

the scaling is changed to highlight the features in the tip region.

In this figure, once again the features are quite similar. The low

streak emanating from the splitter between 1A and 1B, along

with the high streak wrapping around the tip of the splitter

between 1B and 2. A high heat transfer region exists at the end

of the turning vane for both experiment and simulation.

Although, in the experiment the high region seems to be aligned

with the turning vane, while in the simulation it occurs more

toward the inside of the turn. One additional discrepancy is

noted in the region to the outside of the turning vane, the very

low experimental values are not present in the simulation.

Figure 12 shows results for the low pressure ratio case. All of

the features and the level of comparison are the same as for the

high pressure ratio.

An observant reader may have noticed that block boundaries

are visible in some regions of the heat transfer figures. For

a.)

Figure 11. Normalized heat transfer coefficient for

high pressure ratio simulation(a) and experiment(b)

scaled to show turn.

1.75 2.75 3.75 h/h0

h

[W/m2/K]

Figure 10. Normalized heat transfer coefficient on a

line down the middle of passage 1B.

15 20 25 30 35 40

x [cm]

0.0

2.0

4.0

6.0

h/h0

Lo

Medium

High

b.)

NASA/CR—2002-211701 6
instance, figure 11a near the turning vane. These are merely

artifacts of less than perfect post-processing. The solution is

produced on cell centers. The heat transfer is then calculated on

face centers and interpolated to the grid nodes. Because of these

manipulations, a little bit of “noise” is to be expected. Figure 13

is offered as further evidence of the high quality of the codes

block handling ability. Figure 13 shows the temperature

distribution one grid space off the wall. From figure 13 it can be

seen that the temperature is continuous and smooth across block

boundaries.

One benefit of running a simulation is the ability to

investigate the flow field everywhere in the domain, which can

in turn allow further insight into why things appear as they do.

Figure 14 shows the velocity component normal to the heated

wall. The plane shown is located above the wall at a height

equal to the height of the ribs in the tip region. No scale is

shown for figure 14, since its purpose is to highlight the sign of

the normal velocity component. Figure 14 bears a striking

resemblance to figure 8, with the colors reversed. In general,

regions of high heat transfer coincide with regions of flow

toward the surface. Notice that between the turbulators in

passage 1B there is a region of down flow just after the

turbulator, followed by a region of upflow just upstream of the

turbulator. From figure 8 we see that the highest heat transfer

between turbulators in passage 1B tends to be just downstream

of the rib where the flow at rib height is toward the surface. In

passage 2, which has significantly less blockage compared to

passage 1B, the normal velocity at rib height is toward the wall

over much of the region between turbulators. Referring to figure

8 again we see that the predominantly downward flow between

the passage 2 turbulators leads to a heat transfer distribution

which is not skewed toward the upstream turbulator. In the turn

region we see a strong upwelling of flow associated with the low

1 23456 h/h0

h

[W/m2/K]

Figure 12. Normalized heat transfer coefficient for

low pressure ratio simulation(a) and experiment(b)

scaled to show turn.

Figure 13. Temperature normalized by inlet total

temperature. Cut plane is located one grid space off

wall.

b.)

a.)

NASA/CR—2002-211701 7
heat transfer streak emanating from the splitter between

passages 1A and 1B. Also, a downward flow is associated with

the high heat transfer on the outer wall of passage 2. One final

observation on figure 14 is that flow away from the surface does

not always imply low heat transfer. Looking at the end of

passage 1B, where the area becomes quite small, figure 14

shows a strong upwelling of flow. However, figure 8 shows high

heat transfer in this area. This reversal of the general trend is of

course due to the tremendous streamwise acceleration occurring

as the flow exits passage 2.

CONCLUSIONS

A simulation of flow and heat transfer in a complex trailing

edge passage for a high pressure turbine blade was done using

the Glenn-HT Navier-Stokes code. The geometry attempted

represents the most complex internal passage attempted with

this code to date. The code proved to be very robust even though

the grid was highly nonorthogonal and somewhat nonsmooth in

places. Results of the pressure and heat transfer distributions

compare very well with the available experimental data, with

the exception of some localized discrepenancies. The

dependence of heat transfer on Reynolds number to the 0.8

power observed in the experiment, was reproduced by the

simulation very closely. The availability of the simulation

allows in depth investigation of flow features and their effect on

heat transfer.

REFERENCES

[1] Ameri, Ali and Bunker, R.S., “Heat Transfer and Flow on

the First-Stage Blade of a Power Generation Gas Turbine: Part

2-Simulation Results,” ASME J. Turbomach., 122 pp 272-277,

2000.

[2] Ameri, Ali and Rigby, D.L., “A Numerical Analysis of the

Heat Transfer and Effectiveness on the Film Cooled Turbine

Blade Tip Models,” 4th Int. Sym. on Air Breathing Engines,

Florence, Italy, 1999.

[3] Heidmann, J.D., Rigby, D.L., and Ameri, Ali, “A ThreeDimensional Coupled Internal/External Simulation of a FilmCooled Turbine Vane,” IGTI paper 99-GT-, 1999.

[4] Garg, V.K. and Rigby, D.L., “Heat Transfer on a FilmCooled Blade - Effect of Hole Physics,” Int. J. of Heat and Fluid

Flow, Vol. 20, pp. 10-25, 1999.

[5] Rigby, D.L., “Prediction of Heat and Mass Transfer in a

Rotating Ribbed Coolant Passage with a 180 Degree Turn,”

IGTI paper 98-GT-329, 1998.

[6] Rigby, D.L. and Lepicovsky, J., “Conjugate Heat Transfer

Analysis of Internally Cooled Configurations,” IGTI Paper

2001-GT-405, 2001.

[7] Steinthorsson, E., Liou, M.S., Povinelli, L.A.,

“Development of an Explicit Multiblock/Multigrid Flow Solver

for Viscous Flows in Complex Geometries,” AIAA-93-2380,

1993.

[8] Arnone, A., Liou, M.-S., and Povinelli, L. A.,1991,

“Multigrid Calculation of Three-Dimensional Viscous Cascade

Flows,” AIAA-91-3238.

[9] Wilcox, D.C., 1994b,“Simulation of Transition with a TwoEquation Turbulence Model,” AIAA Journal, Vol. 32, No.2, pp.

247-255.

[10] Bunker, R.S., Wetzel, T.G. and Rigby, D.L., “Heat Transfer

in a Complex Trailing Edge Passage for a High Pressure

Turbine Blade- Part 1: Experimental Measurements,” IGTI

Paper 2002-GT-30212, 2002.

Figure 14. Velocity component normal to surface. Cut

plane is located at the height of the turbulators near

the tip. (Red is away and blue is toward surface)

NASA/CR—2002-211701 8
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15. NUMBER OF PAGES

20. LIMITATION OF ABSTRACT

Unclassified Unclassified

Final Contractor Report

Unclassified

1. AGENCY USE ONLY (Leave blank)

10. SPONSORING/MONITORING

 AGENCY REPORT NUMBER

9. SPONSORING/MONITORING AGENCY NAME(S) AND ADDRESS(ES)

National Aeronautics and Space Administration

Washington, DC 20546–0001

Available electronically at http://gltrs.grc.nasa.gov/GLTRS

August 2002

NASA CR—2002-211701

2002–GT–30213

E–13430

WU–708–28–13–00

NAS3–00145

14

Heat Transfer in a Complex Trailing Edge Passage for a High Pressure

Turbine Blade

Part 2: Simulation Results

David L. Rigby and Ronald S. Bunker

Heat transfer; Turbine; Cooling; CFD

Unclassified -Unlimited

Subject Categories: 07 and 34 Distribution: Nonstandard

QSS Group, Inc.

21000 Brookpark Road

Cleveland, Ohio 44135

Prepared for the Turbo Expo 2002 cosponsored by the American Society of Mechanical Engineers and the International

Gas Turbine Institute, Amsterdam, The Netherlands, June 3–6, 2002. David L. Rigby, QSS Group, Inc., Cleveland, Ohio

44135; and Ronald S. Bunker, General Electric R&D Center, Schenectady, New York 12301. Project Manager, Stephen

Hippensteele, Turbomachinery and Propulsion Systems Division, NASA Glenn Research Center, organization code 5820,

216–433–5897.

A combined experimental and numerical study to investigate the heat transfer distribution in a complex blade trailing edge

passage was conducted. The geometry consists of a two pass serpentine passage with taper toward the trailing edge, as

well as from hub to tip. The upflow channel has an average aspect ratio of roughly 14:1, while the exit passage aspect

ratio is about 5:1. The upflow channel is split in an interrupted way and is smooth on the trailing edge side of the split and

turbulated on the other side. A turning vane is placed near the tip of the upflow channel. Reynolds numbers in the range of

31,000 to 61,000, based on inlet conditions, were simulated numerically. The simulation was performed using the

Glenn-HT code, a full three-dimensional Navier-Stokes solver using the Wilcox k-ω turbulence model. A structured

multi-block grid is used with approximately 4.5 million cells and average y+ values on the order of unity. Pressure and

heat transfer distributions are presented with comparison to the experimental data. While there are some regions with

discrepancies, in general the agreement is very good for both pressure and heat transfer.