Heat Transfer Measurements for a Film Cooled Turbine 

Vane Cascade 

by Douglas R. Thurman, Philip E. Poinsatte, and James D. Heidmann 

ARL-MR-0697 May 2008 

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Army Research Laboratory 

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ARL-MR-0697 May 2008 

Heat Transfer Measurements for a Film Cooled Turbine 

Vane Cascade 

Douglas R. Thurman 

Vehicle Technology Directorate, ARL 

Philip E. Poinsatte and James D. Heidmann 

National Aeronautics and Space Administration

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Heat Transfer Measurements for a Film Cooled Turbine Vane Cascade 

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Douglas R. Thurman, Philip E. Poinsatte, and James D. Heidmann 

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14. ABSTRACT 

Experimental heat transfer and pressure measurements were obtained on a large scale film cooled turbine vane cascade. The 

objective was to investigate heat transfer on a commercial high pressure first stage turbine vane at near engine Mach and 

Reynolds number conditions. Additionally blowing ratios and coolant density were also matched. Numerical computations 

were made with the Glenn-HT code of the same geometry and compared with the experimental results. We used a transient 

thermochromic liquid crystal technique to obtain steady state heat transfer data on the mid-span geometry of an instrumented 

vane with 12 rows of circular and shaped film cooling holes. A mixture of SF6 and Argon gases was used for film coolant to 

match the coolant-to-gas density ratio of a real engine. The exit Mach number and Reynolds number were 0.725 and 2.7 

million respectively. Trends from the experimental heat transfer data matched well with the computational prediction, 

particularly for the film cooled case. 

15. SUBJECT TERMS

Heat transfer, liquid crystal, film cooling, turbine vane 

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Douglas R. Thurman 

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iii 

Contents 

List of Figures iv

List of Tables iv

Introduction 1

Nomenclature 2

Subscripts 3

Apparatus 3

Procedure 7

Glenn-HT Computations 9

Results and Discussion 12

Conclusion 16

References 17 

Distribution List 19 
iv 

List of Figures 

Figure 1. Test facility......................................................................................................................3 

Figure 2. Film cooled vane cascade................................................................................................4 

Figure 3. Film cooled vane. ............................................................................................................5 

Figure 4. Vane Cross section and film hole numbers. ....................................................................5 

Figure 5. Film hole angles. .............................................................................................................6 

Figure 6. Leading edge region of computational grid...................................................................11 

Figure 7. Stagnation temperature contours on plane of fixed span...............................................12 

Figure 8. Mid-span flow pressure survey. ....................................................................................13

Figure 9. Vane loading comparison..............................................................................................13 

Figure 10. Stanton number for no film cooling case, Re = 2.7E6, Mach = 0.725........................14 

Figure 11. Stanton number for film cooled case, Re = 2.7E6, Mach = 0.725, density ratio = 

1.92...........................................................................................................................................15 

List of Tables 

Table 1. Film cooling hole parameters. ..........................................................................................6 

Table 2. Test conditions..................................................................................................................8 
1 

Introduction 

With turbine engine temperatures exceeding material properties, one way to protect the engine 

components exposed to the hot freestream gas is to use film cooling to provide a layer of cooler 

air over the material surface. This is especially useful for first stage high pressure turbine vanes 

which experience the hottest temperatures in the engine. Various studies have looked at the 

characteristics and effectiveness of film cooling. Early research by Goldstein (1971) involved a 

flat plate to simulate the environment of a film cooled turbine blade. Saumweber (2004) and 

Gritsch, et al., (1998) have used flat plates and heated air to match coolant-to-air density ratios of 

real engines. When reproducing the temperature extremes seen in a real engine is not possible, 

foreign gas can be used for the coolant in experiments in order to match the coolant-to-air 

density ratios of real engines. Ekkad, et al., (1997) have used flat plates and foreign gases to 

match the density ratio. Film cooling studies with more realistic geometries have also been 

investigated. Colban, et al., (2006), Cutbirth and Bogard (2002), Ethridge, et al., (2001), and 

Ames (1996) have used scaled blade geometries to study various aspects of film cooling. These 

experimental studies, however, often do not match typical real engine flow conditions, such as 

Mach number and Reynolds number. 

There have been several computational studies that have focused on film cooling with realistic 

blade geometries and plenum effects. Garg and Rigby (1998) modeled a film cooled blade with 

angled showerhead holes as well as with plenums. Bohn, et al., (1997) also modeled a film 

cooled blade with normal holes. Heidmann, et al., (2000) modeled a realistic Honeywell vane 

geometry with film cooling and near-real engine flow conditions. This geometry consisted of 

showerhead compound-angled holes and pressure and suction side holes with shaped holes on 

the pressure side. 

Little experimental data has been published at realistic engine flow conditions. Abuaf, et al., 

(1998) studied roughness effects on heat transfer of an uncooled turbine vane at several flow 

conditions typical of real engines. Haldeman et al., (2006) published unsteady aerodynamic and 

heat transfer results for a Honeywell film cooled rotating turbine stage in a blow-down facility. 

This was the first attempt to measure film cooled turbine heat transfer in such a realistic 

environment with research-quality instrumentation. The present geometry nominally represents 

the extruded vane mid-span profile and film cooling pattern from the Haldeman et al., (2006) 

study, and matches the geometry investigated by Heidmann, et al., (2000). 

The present study provides experimental data at near-real engine flow conditions for the 

Honeywell vane, and offers the opportunity to generate high-resolution vane heat transfer data in 

a steady-flow cascade environment. The data will also help validate the computational heat 

transfer analysis performed by NASA’s Glenn-HT code on this vane geometry. We used a 

transient liquid crystal technique to obtain the heat transfer data. Mach number and Reynolds 
2 

number nearly match real engine conditions at the exit of a first stage turbine vane. Coolant-togas density ratio is also matched using a combination of Argon and SF6 gases for film coolant. 

The matching of these key flow parameters for a film cooled turbine with detailed local 

temperature measurement is a key element of this study. In the current test, the vane model is 

scaled large enough to yield adequate special resolution of the surface heat transfer. This NASA 

wind tunnel facility has enough flow throughput to nearly match real engine conditions of not 

only Reynolds number, but also exit Mach number at generally atmospheric tunnel conditions. 

Nomenclature 

c vane true chord 

cp specific heat of test vane material 

d film hole diameter 

h heat transfer coefficient 

k thermal conductivity 

p film hole spanwise pitch 

Re Reynolds number based on true chord and exit conditions 

s streamwise distance from leading edge 

St Stanton number 

t time 

T temperature 

V velocity 

α streamwise film hole angle 

β spanwise film hole angle 

βh non-dimensional time 

ρ density of test vane material 

θ non-dimensional temperature
3 

Subscripts 

i initial 

s surface 

t total 

Apparatus 

We used a four blade, five passage linear cascade for the experimental tests. The facility is 

shown in figure 1, and the cascade is shown in figure 2. The simulated turbine vanes were based 

on the mid-span coordinates of a first stage high pressure turbine of a Honeywell engine. Room 

temperature air was drawn into the tunnel by a vacuum exhaust system. We attached an 

elliptical one-dimensional aluminum bellmouth to the inlet of the cascade. The walls of the 

cascade were made of 0.75 in (1.91 cm) thick clear acrylic. The test vanes were made by rapid 

prototyping technique of a Duraform polyamide sintering material (nylon), and had a chord of 

8.09 in (20.5 cm), span of 5 inches (12.7 cm), a 76 degree turning angle, and were scaled 2.9:1 to 

match the engine exit Mach number and exit Reynolds number. The passage width between the 

vanes was 6 inches (15.2 cm). We used adjustable tailboards to ensure periodicity of the flow 

between the passages of the cascade. 

Figure 1. Test facility. 
4 

Figure 2. Film cooled vane cascade. 

For the film cooled heat transfer tests, we attached the vane to a cut-out section of the top lid of 

the cascade as shown in figure 2. Thin-film thermocouples were attached to the inside and 

outside surfaces of the vane. We sprayed the outside surface of the vane with black paint, then 

with narrow-band microencapsulated chiral nematic liquid crystal paint, which displays 

distinctive colors at particular temperatures. This test vane, shown in figure 3, had two plenums 

to feed 12 rows of film cooling holes. Figure 4 shows the rows of holes that were fed by each 

plenum. Table 1 summarizes the hole geometry, and figure 5 illustrates the definitions of 

streamwise and spanwise angles α and β. The first four rows of holes on the pressure side were 

shaped and staggered, with an expansion angle of 10 degrees on each side of the holes and on the 

downstream edge. Rows 5 through 10 had round, staggered showerhead holes. Rows 11 and 12 

had round, aligned holes on the suction side. Inside the center of each plenum was a pressure tap 

and an open ball thermocouple. An open ball thermocouple was also located at the opening of 

each plenum. The no-hole heat transfer vane was instrumented with liquid crystals in a similar 

manner. 
5 

Figure 3. Film cooled vane. 

Figure 4. Vane Cross section and film hole numbers. 
6 

Table 1. Film cooling hole parameters. 

Row s/c No. 

holes Type p/d α β

1 –0.447 10 shaped, staggered 

pressure side 4.06 65.5 0 

2 –0.385 10 shaped, staggered 

pressure side 4.06 62.9 0 

3 –0.182 10 shaped, staggered 

pressure side 4.06 63.8 0 

4 –0.142 10 shaped, staggered 

pressure side 4.06 61.4 0 

5 –0.053 11 round, staggered 

showerhead 4.06 6.2 60 

6 –0.032 11 round, staggered 

showerhead 4.06 2.7 60 

7 –0.011 11 round, staggered 

showerhead 4.06 0.9 60 

8 0.011 11 round, staggered 

showerhead 4.06 0.9 60 

9 0.033 11 round, staggered 

showerhead 4.06 2.9 60 

10 0.054 11 round, staggered 

showerhead 4.06 4.4 60 

11 0.116 15 round, aligned 

suction side 2.71 44.5 0 

12 0.157 15 round, aligned 

suction side 2.71 47.0 0 

Figure 5. Film hole angles. 

We used a heater oven, covered with an automatic temperature-controlled heater blanket, to heat 

the heat transfer vanes. Low heat emitting lights were used to illuminate the model. We 

recorded the video data with cameras and digital recorders. A time-date generator was used to 

superimpose the time onto the video data. A commercial data acquisition system recorded 

pressure and temperature data. 

Since the temperature difference that exists in a real engine could not be reproduced in this 

facility, we used a combination of Argon and Sulfur Hexafluoride (SF6) for the film cooling gas 

which provided a coolant-to-freestream density ratio of 1.92. Coolant flow to the instrumented 

heat transfer vane was supplied by Argon and SF6 tanks. To achieve the correct density ratio, 
7 

four mass flow controllers were used to control the gas flow rates, with each tank flow being 

split into two mass flow controllers. We then mixed the gases from one SF6 controller and one 

Argon controller together prior to entering each vane plenum. 

To obtain static surface pressure data, we used a vane with no film cooling holes with static 

pressure taps located along the mid-span height of the vane. Static pressure taps were also 

located along the mid-span height of the adjacent vanes, as well as on the lower endwall of the 

tunnel at discreet locations upstream and downstream of the vanes. Total pressure survey data 

was taken at mid-height upstream and downstream of the vanes with a 3-hole total pressure 

probe. 

Procedure 

The tunnel flow conditions were set prior to the test by a downstream flow control valve. We 

conducted heat transfer tests using a transient liquid crystal technique similar to that described in 

Poinsatte and Thurman (2007) to measure heat transfer coefficients. We placed the test vane in 

the heater oven and heated it to a uniform temperature of nominally 150 °F (65.6 °C). When the 

model reached the desired temperature, the data acquisition system and video recorders were 

started. We opened the oven and placed the heated vane in the tunnel. The time to transfer the 

vane to the tunnel was less than two seconds. The surface temperature decreased less than 1 

degree and was included in our uncertainty calculations. Room temperature air was then 

suddenly drawn through the cascade with a fast-opening valve, which produced a near step 

change in flow startup. The airflow cooled the previously heated model and the model surface 

temperature was indicated by the liquid crystal color. We recorded the resulting liquid crystal 

color patterns showing continuous surface temperatures with time. Knowing this time 

temperature history allows calculation of heat transfer coefficients, as described below. For the 

present study, we examined only the mid-span region of the vane. The yellow color band of the 

liquid crystal was used to calculate the heat transfer coefficients and was calibrated to be 100 °F 

(37.8 °C). Test conditions are summarized in table 2. 
8 

Table 2. Test conditions. 

Total Pressure 14.3 psi (98.6 kPa) 

Initial Model Temperature 150 °F (65.6 °C) 

Liquid Crystal temperature 100°F (37.8 °C) 

Inlet Airstream Temperature 75 °F (23.9 °C) 

Reynolds number 2.7E6 

Inlet Mach number 0.125 

Exit Mach number 0.725 

(ρck)0.5 of test model 1.486 Btu/hr0.5/ft2

/°R 

 (506 W sec0.5/m2/°K) 

Coolant Temperature 75 °F (23.9 °C) 

Mass flow rate coolant 0.065 lbm/sec (.029 kg/sec) 

Nominal coolant mixture 

composition by weight 60% Ar, 40% SF6

Blowing ratio 

(upstream plenum) 2.8 

Blowing ratio 

(downstream plenum) 1.4 

Density ratio 1.92 

Assuming one-dimensional conduction out of a semi-infinite wall, the heat transfer coefficient h 

can be determined from the solution of the conduction equation for a step change in flow: 

1 ( ) 2

h e erfc h θ β β = − (1) 

where 

i t

i s

T T

T T

−

− θ = (2) 

c k

h t

p

h

ρ

β = (3) 

Ti is the initial surface temperature, Ts is the wall surface temperature, which is the liquid crystal 

calibrated temperature, Tt is the freestream total air temperature, which was assumed to be room 

temperature, and t is the time from the start of the test. Note that the experimental results using 

this one-dimensional conduction assumption are not actually valid near the cooling holes, as 

these regions are no longer one-dimensional and are also affected by the film coolant flowing 
9 

through the holes. This will result in higher experimental uncertainties near the holes. When 

film coolant was used, the coolant was at the same room temperature as the freestream air 

temperature. Heat transfer magnitudes and locations were determined by the location of the 

yellow isotherms produced by the liquid crystals on the surface. We then calculated Stanton 

number based on inlet conditions using the following definition: 

Vc p

h St

ρ = (4) 

We acquired pressure survey measurements under steady state conditions. Total pressure and 

flow angle measurements were taken upstream and downstream of the vanes at mid-span height 

to verify that the flow between the passages was periodic. We also took endwall pressure 

measurements on the lower surface of the cascade. 

An uncertainty analysis was performed based on the method of Kline and McClintock (1953). 

Overall uncertainties in heat transfer ranged from 5% to 8% for the no-hole vane, and from 8% 

to 15% for the film cooled vane with most points around 10 percent. The relatively high 

uncertainty for the film cooled cases is due to the relative uncertainty regarding the coolant 

temperature at startup. For the first couple of seconds the coolant picks up the residual heat from 

the plenum of the preheated model. The coolant quickly comes down to the room temperature 

which is the same as the inlet freestream. This startup transient can bias the measured heat 

transfer data, but the effect is lessened when the time from startup is sufficiently large, which is 

where most of the data was taken. 

Glenn-HT Computations 

The following is a brief summary of the computational methods employed in previously reported 

calculations in Heidmann et al., (2000). The simulations were performed using a multiblock 

computer code called Glenn-HT, previously known as TRAF3D.MB (Steinthorsson et al., 1993) 

which is based on a single block code designed by Arnone et al., (1991). This code is a general 

purpose flow solver designed for simulations of flows in complicated geometries. The code 

solves the full compressible Reynolds-averaged Navier-Stokes equations using a multi-stage 

Runge-Kutta-based multigrid method. It uses the finite volume method to discretize the 

equations. The code uses central differencing together with artificial dissipation to discretize the 

convective terms. The overall accuracy of the code is second order. The code employs the k-ω

turbulence model developed by Wilcox (1994a, 1994b), with subsequent modifications by 

Menter (1993) as implemented by Chima (1996). Accurate heat transfer predictions are possible 

with the code because the model integrates to the walls and no wall functions are used. Rather, 

the computational grid is generated to be sufficiently fine near walls to produce a y+

 value of less 

than 1.0 at the first grid point away from the wall. For heat transfer a constant value of 0.9 for 
10 

turbulent Prandtl number, Prt, was used. A constant value of Pr = 0.72 was used. Laminar 

viscosity was a function of temperature through a 0.7 power law (Schlichting, 1979) and cp was 

taken to be a constant. 

A structured multi-block grid approach was adopted to model this complex geometry. The grid 

was composed of 1.2 x 106

 computational cells. A y+ value of less than 1.0 at the first grid point 

from the wall was employed at all locations. Several calculations were performed for varying 

wall spacings, and it was found that further reductions produced little change in the solution. 

The grid consisted of 20 cells across both the inlet and outlet boundaries, 60 cells on the periodic 

boundary, over 200 cells around the vane, and 44 cells from the vane to the periodic boundary. 

These values are consistent with good computational practice. 

Figure 6 shows the grid in the leading edge region of the vane. All film cooling holes were 

discretized explicitly, including the fan-shaped holes with expanded exits. In addition, the 

clustered viscous grids remain attached to all wall boundaries without extending into the flow 

field. The freestream inlet flow to the vane was at an angle of 0 degrees to the axial direction, 

with all temperatures and pressures normalized by the inlet stagnation values of 294 K and 

101 kPa, respectively. Inlet turbulence intensity of 8.0% and turbulence length scale of 15.0% of 

vane true chord were applied. The vane downstream exit flow was defined by imposing a 

constant normalized static pressure which was empirically determined to yield an exit Mach 

number. Periodicity was enforced in both the blade-to-blade and spanwise directions based on 

vane and film hole pitches, respectively. We gave all solid walls a no-slip boundary condition in 

addition to a fixed temperature. The isothermal wall boundary condition extended to all wall 

surfaces, including the film hole surfaces and plenum surfaces, so heat transfer in the plena and 

film holes occurred and provided a stagnation temperature profile in the jet exiting each hole. 

Plenum inflows were modeled as if flowing from the plenum side wall to preserve periodicity in 

the spanwise direction.
11 

Figure 6. Leading edge region of computational grid. 

We ran the code for three cases: a solid isothermal blade with no film cooling holes and a wall 

temperature of 0.7 times the freestream temperature as a baseline, and two film cooled 

isothermal-wall cases with wall temperatures of 0.7 and 0.8 times the freestream temperature, 

respectively. These two cases were then used to generate independent distributions of film 

cooling effectiveness and heat transfer coefficient. The geometry consisted of a spanwiseperiodic pitch of the film cooling hole pattern, so no endwall effects were considered. The film 

cooling holes were fed by two plenums as in the experiment, and the coolant mass flow was set 

to the engine design overall blowing ratio for each plenum. We set the nominal design density 

ratio of 2.0 by imposing a film coolant temperature of 0.5 times the freestream inlet temperature. 

Exit Mach and Reynolds numbers of 0.876 and 2.9 x 106

 were employed, respectively. 

Figure 7 shows a result of the Glenn-HT simulation, illustrating the stagnation temperature on a 

fixed-span plane through the centerline of the central suction side film holes. It is immediately 

evident that the stagnation line intersects this plane between rows 6 and 7 (the second and third 

showerhead rows from the pressure side) since the high temperature isotherms approach the 

vane. The freestream impinges at this point, diverting the cooling flow from these rows to 

opposite sides of the vane. A thinning of the thermal boundary layer is also evident on the 

pressure side, just upstream of the shaped holes. This will manifest itself as an increase in heat 

flux at this location. A thermal boundary layer is present inside the plenum and film holes due to 

the isothermal boundary condition there. The resultant stagnation temperature profile in the film 
12 

jets is of great interest, as is the velocity profile. Complete results and discussion of the GlennHT simulation can be found in Heidmann et al., (2000). 

Figure 7. Stagnation temperature contours on plane of fixed span. 

Results and Discussion 

Figure 8 shows inlet and exit total pressure survey data for the cascade test section taken with a 

3-hole pressure probe. The surveys confirm that the inlet flow was smooth and that the flow 

downstream of the vanes was periodic. Figure 9 shows the experimentally measured Mach 

number distribution on the mid-span of the tested vane, as well as the computational prediction 

for this vane in the cascade. The loading is typical of a vane of this type and the measured 

values agree with the prediction fairly well. Flow angle and endwall pressure measurements also 

were consistent with a well-conditioned periodic cascade. No inlet turbulence measurements 

were made during this test, however, many previous hot wire measurements were made in this 

facility with very similar inlet conditions. Based on these results, the inlet turbulence intensity 

was expected to be between 1 and 3 percent. 
13 

0.94

0.95

0.96

0.97

0.98

0.99

1

1.01

1.02

1.03

1.04

0 5 10 15 20 25 30

N

o

r

m

a

l

i

z

e

d

t

o

t

a

l

p

r

e

s

s

u

r

e

,

P

/

P

b

a

r

distance from sidewall (in)

upstream

downstream

Figure 8. Mid-span flow pressure survey. 

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

0 0.1 0.2 0.3 0.4 0.5 0.6

axial chord distance/true chord

Mach

CFD Prediction

instrumented vane

adjacent vanes

Figure 9. Vane loading comparison. 
14 

Figure 10 shows the measured heat transfer for the vane with no film cooling holes along with 

the Glenn-HT prediction. The graph shows Stanton number versus normalized surface distance 

for a Reynolds number based on true chord and exit conditions of 2.7 x 106

 and an exit Mach 

number of 0.725. The experimental data shows a local heat transfer maximum at the stagnation 

point with a nominal St value of 0.005. The heat transfer initially decreases on both the suction 

and pressure sides. On the suction side, the Stanton number initially decreases to a value of 

0.004 at s/c of 0.055 and then quickly increases to a value around 0.01, showing a rather quick 

transition to a turbulent boundary layer. The heat transfer is highest on the suction side at an s/c 

of 0.2 to 0.4 as the flow accelerates around the vane curvature, and then is fairly constant and at a 

high level for most of the suction surface. The downstream suction surface heat transfer is 

nearly twice that of the stagnation point. On the pressure surface the Stanton number drops to a 

value about half that of the stagnation region, indicating a laminar boundary layer, and then 

monotonically rises gently moving further downstream. Note that the surface roughness of this 

vane was measured with a profilometer and the arithmetic mean peak height was roughly 

350 microinches (8.89 μm), which could be considered fairly rough. The film cooled vane 

roughness was also measured and had a similar value. 

0.000

0.005

0.010

0.015

0.020

‐1.0 ‐0.8 ‐0.6 ‐0.4 ‐0.2 0.0 0.2 0.4 0.6 0.8 1.0 1.2

S

t

a

n

t

o

n

N

u

m

b

e

r

surface distance, s/c

no‐hole data

GlennHT prediction

pressure suction

Figure 10. Stanton number for no film cooling case, Re = 2.7E6, Mach = 0.725. 

The Glenn-HT heat transfer prediction matched the experimental values fairly well on the 

suction side but over-predicts on the pressure side. While the heat transfer trends are generally 

similar, the prediction also shows an earlier boundary layer transition to turbulent. The predicted 

stagnation Stanton number was roughly 60% higher than the experimental value. Generally the 

suction side heat transfer prediction matched the experiment within 10%, while on the pressure 

side the prediction over-predicted the measured values by roughly 100%. These results indicate 

that the k-ω turbulence model employed in Glenn-HT does a good job of predicting heat transfer 

when the boundary layer is turbulent but less so when the flow is laminar. There may be some 
15 

sensitivity of the surface heat transfer to the inlet turbulence intensity and length scale. The 

computational work was done with 8% turbulence intensity and 15% of vane true chord length 

scale that are larger than the experimental case. This, in concert with the use of the k-ω

turbulence model which is essentially fully turbulent, may contribute to the over-prediction of 

Stanton number by the Glenn-HT code shown in figure 10. 

Figure 11 shows the measured heat transfer for a vane with film cooling along with the GlennHT prediction. The graph shows Stanton number versus normalized surface distance at a 

Reynolds number based on true chord and exit conditions of 2.7 x 106

, exit Mach number of 

0.725, and a density ratio of 1.92. Nominal average film cooling blowing ratios of 2.5 based on 

freestream conditions were employed. The positions of the rows of film cooling holes are 

indicated by the dashed lines. The addition of film cooling generally increases the heat transfer 

over the no-hole vane. This is especially seen in the showerhead film cooling region where the 

experimental Stanton number has increased roughly 150% to 200%. The Stanton number at the 

stagnation region for the film cooled case has a value of roughly 0.012. The trends for the film 

cooled case are similar to the no-hole case, except that each row of film cooling holes causes a 

local increase in heat transfer. Suction side heat transfer is highest as the flow accelerates around 

the vane curvature and then drops and levels off to a Stanton number value just below 0.01. The 

pressure side Stanton number drops down to a minimum value of 0.004 and then increases 

toward the trailing edge again with local maxima at the location of film cooling jets. 

0.000

0.005

0.010

0.015

0.020

‐1.0 ‐0.8 ‐0.6 ‐0.4 ‐0.2 0.0 0.2 0.4 0.6 0.8 1.0 1.2

S

t

a

n

t

o

n

N

u

m

b

e

r

surface distance, s/c

film‐cooled data

Glenn‐HT prediction

Filmhole locations

pressure suction

Figure 11. Stanton number for film cooled case, Re = 2.7E6, Mach = 0.725, 

density ratio = 1.92. 

The Glenn-HT heat transfer prediction for the film cooled case matches the experimental data 

extremely well. This is especially true on the suction side where the agreement is right on at 

most locations. On the pressure side, the prediction is again rather good. The prediction shows a 

slightly earlier boundary layer transition and thus over-predicts the pressure side heat transfer, 
16 

but for most points the predicted Stanton numbers match the experimental values within 20%. It 

should be noted that the sharp peaks around the cooling holes seen in the prediction cannot be 

captured well with this experiment. 

Conclusion 

Heat transfer measurements were made on a commercially designed film cooled turbine vane 

employing a transient liquid crystal technique at near realistic engine conditions. We made 

comparisons with a prediction of flow and heat transfer of the same vane previously made using 

the Glenn-HT computation code. 

1. The no-hole case shows high heat transfer on the suction side and much lower heat transfer 

on the pressure side, and indicates quick boundary layer transition on the suction side and 

mostly laminar flow on the pressure side. 

2. Addition of film cooling enhances heat transfer relative to the no-hole vane baseline case. 

Film cooling jets cause a local increase in the heat transfer; this is especially true in the 

showerhead region. 

3. Glenn-HT does a very good job predicting the heat transfer on the film cooled vane. The 

predicted Stanton number essentially matches much of the data well within the 

experimental uncertainty. The Glenn-HT code also does a fairly good job predicting the 

heat transfer on the no-hole vane. Comparison with the experimental data shows that the 

data trends are similar, and while the suction side shows adequate agreement, the pressure 

side is over-predicted by the computational model. 
17 

References 

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Aerodynamic Performance of Turbine Airfoils. Journal of Turbomachinery 1998, 120, 522–

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Arnone, A.; Liou, M.-S.; Povinelli, L. A. Multigrid Calculation of Three-Dimensional Viscous 

Cascade Flows; AIAA Paper 91-3238; 1991. 

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Bohn, D. E.; Becker, V. J.; Rungen, A. U. Experimental and Numerical Conjugate Flow and 

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Holes With Expanded Exits. ASME Journal of Turbomachinery 1998, 120, 549–556. 
18 

Haldeman, C. W.; Mathison, R. M.; Dunn, M. G.; Southworth, S.; Harral, J. W.; Heitland, G. 

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348-359. Also NASA TM 1999-209078. 

Kline, S. J.; McClintock, F. A. Describing Uncertainties in Single-Sample Experiments. 

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1994b, 32 (2), 247–255. 
19 

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