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CHAPTER 4

Impingement Cooling for Combustor Liner Backside Cooling

Srinath V. EkkadMechanical Engineering, Virginia Tech, Blacksburg, VA, USA.

Abstract

Impingement cooling is one of the best heat transfer mechanisms. It is used widely in many industries for heating and cooling products. Jet impingement cooling is signifi cantly in gas turbine systems for cooling hot gas path components, primar-ily in vanes and combustor liners. In modern low NOx gas turbine combustors, cooling of the combustor liner is achieved from the backside through innova-tive enhanced heat transfer augmentation methods. As NOx, CO, and unburned hydrocarbon emissions from gas turbines are continually regulated, methods other than fi lm cooling of the combustor liner are implemented to comply. Jet impingement arrays are widely used to effectively cool the liner from the backside, although there are problems with this approach. In this chapter, an overview of jet impingement-related heat transfer research and design background is presented. Additionally, some specifi c studies that focus on innovative cooling methods for combustor liners are presented with highlighted results. Additional comments on possible future research directions and topics under consideration are presented.

Keywords: Impingement, combustors, cooling, heat transfer.

1 Introduction

Gas turbine engines have proven to be an effective means of converting fuel into usable power either through direct shaft power, or thrust produced from the high momentum exhaust gasses. Prior to World War I, gas turbine engines were nothing more than a concept conceived by physicists and engineers of the era. After the war, governments started to pursue the idea, looking for effective high output propulsion systems to power the greatest addition to the military arsenal, the air-plane. The fi rst gas turbine engines were aimed at performance at the expense of

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104 IMPINGEMENT JET COOLING IN GAS TURBINES

effi ciency and emissions. Modern gas turbine military fi ghters took the world by surprise during World War II, and led to new markets for the device, commer-cial air travel and industrial power generation. Both markets needed gas turbine engines that were designed for reliable and effi cient operation instead of overall performance. With the formation of the Environmental Protection Agency in 1970, emission levels from gas turbines were scrutinized.

The fundamental thermodynamic principles that make gas turbine engines fea-sible are based on the properties of the working fl uid, mainly air. As air is compressed and heated with the addition and burning of fuel, the energy released from the air expanding through the turbine is greater than the energy required to compress the air. This action is magnifi ed as the fl uid is compressed to higher pres-sures and raised to higher temperatures. Seeking an increase in thermodynamic effi ciency, engineers designed engines with higher compressor pressure ratios and turbine inlet temperatures. Temperatures were over or near the failure point of the materials used to build the turbines, requiring extensive cooling of the internal parts exposed to the fl ow. Coolant air is pulled from the compressor section prior to the combustion chamber and routed throughout the engine to cool all parts in the hot gas path, including the combustor liner and turbine section.

2 Background

Combustor liners required signifi cant amounts of cooling air since they are exposed to both convective and radiative heating from the combustion process. Turbine engineers incorporated various methods to cool the liner including fi lm cooling and jet impingement cooling. The combustion chamber is comprised of an internal liner exposed to the hot gas, and an outer shell used to separate the hot liner from other engine parts and to create a passage for the coolant air. Film cooling absorbs the heat load on the liner by allowing coolant air to pass along the backside of the liner, and then enter the combustion chamber through holes and fl ow close to the inner walls. This coolant mixes with the hot combustion gasses reducing the near wall temperature of the fl ow and therefore the convective heat fl ux into the combustor liner. Figure 1 shows a typical fi lm cooled combustor liner.

Film cooling was the preferred way of controlling temperatures since substantial heat loads could be dissipated with minimal coolant air. Although benefi cial for cooling the combustor liner, fi lm cooling does adversely affect the combustion pro-cess leading to increased emissions. The air/fuel mixture entering the combustor is initially rich since additional air needed for combustion comes from the coolant air entering through the liner walls, which mixes with the mainstream during combus-tion. The coolant air entering the combustion chamber leads to nonuniform tempera-ture distributions causing incomplete combustion and production of nitrogen oxide (NOx), carbon monoxide (CO), and unburned hydrocarbon emissions (Fig. 2).

In the 1970s, restrictions began to be placed on the emission levels of gas tur-bines and have only grown stronger in the passing years. Initial methods to reduce emissions relied on lean combustion and steam injection. Lean combustion engines use low fuel to air ratio combustors to decrease the combustion temperature and


IMPINGEMENT COOLING FOR COMBUSTOR LINER BACKSIDE COOLING 105

reduce the coolant required for fi lm cooling the liner. This technique results in more complete combustion and lower unburned hydrocarbon emissions. The use of steam injection into the combustor further reduces the combustion temperature and controlled NOx levels but requires the addition of steam to the engine. Unfor-tunately, neither of these two methods have a signifi cant effect on limiting the levels of CO. Modern industrial gas turbine engines use dry, no steam injection, low emission (DLE) combustors to overcome emission restrictions while main-taining effi ciency. DLE combustors use lean, premixed technology previously discussed and further limit emissions by reducing or eliminating the introduction of cool air used for fi lm cooling on the liner. The heat load on the liner must be dealt with in other environmentally friendly ways.

Figure 1: Combustor liner fi lm cooling illustration.

Figure 2: Emission distribution with respect to relative combustion temperature [16].



The principal disadvantage of a pure fi lm cooling approach as stated by Nealy et al. [1] is that the heat sink potential of the cooling air is not effectively utilized as 50%–60% of the total combustor fl ow is coming from the coolant through slots and fi lm holes. According to their calculation, the current burner outlet tempera-tures will affect the percentage air fl ow that is required to support combustion and effectively control burner pattern.

3 Jet Impingement Cooling

Jet impingement cooling is an enhanced heat transfer method capable of cooling a combustor liner without injecting cool air directly into the combustion chamber. Cooling the liner from the backside enables engineers to dissipate the heat load and maintain more uniform temperatures in the combustion region needed for effi cient combustion. Figure 3 shows a typical combustor liner using backside jet impinge-ment cooling. An impingement array is comprised of a jet plate typically having round holes, which produces the impinging jets. The jets strike the surface to be cooled, referred to as the target plate. Traditionally, the structure of an impinging jet is broken down into three parts, the potential core, shear layer, and the wall jet. Figure 4 shows the impinging jet structure and the associated regions.

At the discharge of the jet plate, the velocity profi le of the jet is relatively uni-form. As the jet discharges, viscous forces acting in the shear layer cause the jet velocity profi le to develop and expand. The potential core of the jet is defi ned as the region where the viscous forces have little or no effect on the velocity profi le. Once the jet strikes the target plate, the wall jet is formed as the fl uid travels along the wall. Again viscous forces act on the fl uid decreasing the peak velocity and causing the wall jet to thicken as it moves away from the stagnation point. In jet impingement arrays consisting of multiple jets, the ideal jet structure is signifi -cantly altered due to jet-to-jet interactions. The wall jet region is forced upwards as it collides with the adjacent wall jet in the array. This upward movement of the

Figure 3: Jet impingement cooling of combustor liner.



fl uid amplifi es the strength of the ring vortex near the jet exit. As the vortex strengthens, the mixing in the shear layer of hot spent air and coolant air decreases the effectiveness of the jet. Jet arrays, which are generally confi ned on three sides, contend with crossfl ow caused from the spent air of upstream jets intersecting downstream jets before exiting. The crossfl ow of hot spent air degrades the heat transfer characteristics of downstream jets and limits the practical size of jet impingement arrays. Since most experiments are conducted at conditions suitable for accurate testing and not actual engine conditions, dimensionless parameters must be used to scale results. Impinging jets are characterized by the jet Reynolds and Mach numbers, while the heat transfer is characterized by the Nusselt number. The Reynolds number, ReD, is defi ned in eqn (1):

(1)

where U is the mean jet velocity at the discharge, d is the jet hole diameter, and ν is the kinematic viscosity of air. The Mach number, M, is a dimensionless velocity term defi ned in eqn (2) as:

(2)

where a is the local speed of sound at the jet discharge. The heat transfer coef-fi cient, h, is nondimensionalized by relating it to the hole diameter, d, and the thermal conductivity, k, of the fl uid. Equation (3) defi nes the Nusselt number, NuD.

(3)

Due to the complexity of impinging jet structures in actual arrays, researchers have systematically studied the effects of geometrical parameters on the heat transfer characteristics of impinging jets. Dano et al. [2] researched the effects of nozzle geometry on the fl ow characteristics and heat transfer performance. San

Figure 4: Impinging jet structure.



and Lai [3] studied the effect of jet-to-jet spacing on heat transfer in staggered arrays. Cheong and Ireland [4] experimentally measured local heat transfer coef-fi cients under an impinging jet with low nozzle-to-plate, z/d, spacings. Several others have studied the effect of crossfl ow on jet structure and heat transfer includ-ing Kercher and Tabakoff [5]. Both Florschuetz [6, 7] a nd Kercher and Tabakoff [5] developed correlations to predict the effect of crossfl ow on jet impingement heat transfer for inline and staggered arrays, which are still used today in jet impingement research. Bailey and Bunker [8] studied the effect of sparse and dense arrays for large numbers of jets. Hebert and Ekkad [9] investigated the effect of a streamwise pressure gradient for an inline array of sparse and dense confi gura-tions. As more information on geometrical parameters and their effect on imping-ing jets became available, others studied ways of increasing the jet effectiveness through target surface modifi cation. Surface geometries such as trip strips, protru-sions, or dimples can signifi cantly alter the jet structure and potentially provide enhanced heat transfer.

Kercher and Tabakoff [5] studied the impingement heat transfer in the perspec-tive of turbine blade designers. Figures 5 and 6 show their correlation for Nusselt number distributions for multiple jet impingements on a fl at plate. Two correlating parameters that are plotted are developed to correlate the Nusselt number distribu-tion in the presence of crossfl ow. The two parameters, φ1 and φ2, are defi ned as:

(4)

Figure 5: Heat transfer coeffi cient correlation with a constant defi ned as φ1 [5].



where, NuD is the average Nusselt number based on the jet diameter in the absence of crossfl ow and ReD is the jet Reynolds number. This parameter, φ1, is evaluated at jet-to-target-plate spacing of one jet diameter. The other parameter, φ2 is defi ned as:

(5)

Kercher and Tabakoff [5] have shown that heat transfer prediction based on graphical values of these two correlation parameters gave excellent prediction of the heat transfer coeffi cient. The impingement Nusselt number can be calculated from these two functions and a correction factor from the nondimensional jet to target plate spacing (Z/D) as:

(6)

It is interesting to note that unlike a single jet, the Nusselt number increases with an increase in the jet-to-target plate distance. With an increase in this distance, the

Figure 6: Impingement heat transfer coeffi cient correlation based on degradation coeffi cient φ2 from Ref. [5].



crossfl ow gets more space to develop and therefore, the defl ection of the impinge-ment jet is less.

Florschuetz et al. [6] indicated that inline and staggered arrays are similar to pin-fi n array arrangements, but the heat transfer characteristics of these two arrangements are entirely different. The fundamental difference in crossfl ow effects between pin-fi n heat transfer and that with impingement is, in a pin-fi n arrangement crossfl ow can increase heat transfer from the pins by impinging on them. Whereas, in a jet impingement, jets get defl ected away from the target sur-face by the crossfl ow and therefore, crossfl ow effects are mostly detrimental for impingement cooling. The impingement heat transfer correlation for both inline and staggered arrays is given as:

(7)

where the normalizing Nusselt number is given as:

(8)

The values of constants in the equation are dependent on the array confi guration and are given as shown in Table 1.

The confi dence level for this correlation is 95% and the advantage is that it is easy to compute. An inline arrangement shows a higher heat transfer coeffi cient than that with staggered arrangement. Inline array is less sensitive to jet-to-jet spacing (nx and ny are smaller) and more sensitive to jet-to-target-plate spacing (nz is greater). A more detailed correlation developed earlier by Florschuetz et al. [6] was:

(9)

where the constants are defi ned by geometrical parameters as:A, m, B, or n = C (xn/d)nx (yn/d)ny (z/d)nz

The constants to be used for this correlation are given in Table 2.This correlation is based on the results presented in Figure 7 that shows the ratio

of Nusselt numbers with staggered jet pattern to that with inline jet pattern. As z/d ratio increases, the staggered jet pattern shows a reduction in the heat transfer coeffi cient relative to the inline hole pattern. This is due to the associated spanwise distribution of crossfl ow. The tendency of the crossfl ow to be channeled between adjacent streamwise rows in the inline array reduces the impact of the crossfl ow on

Table 1: Constants obtained for eqn (7) for different conditions.

Jet array C nx ny nz n

Inline 0.596 −0.103 −0.38 0.803 0.561Staggered 1.07 −0.198 −0.406 0.788 0.660



the impinging jets. In contrast, the crossfl ow distribution in staggered array is more uniformly distributed and affects impinging jets more signifi cantly. Note that a pin-fi n array behaves exactly the opposite way. The heat transfer coeffi cient in a pin-fi n array improves in the staggered confi guration compared with an inline arrangement. Since the crossfl ow is detrimental in a jet impingement, an inline array that has less jet defl ection from crossfl ow performs better.

3.1 Effect of initial crossflow

In the previous subsection, we discussed the effect of crossfl ow developed by the spent jets. Therefore, the fi rst row of jets did not have any crossfl ow effects. However, the crossfl ow may develop from the upstream cooling conditions. For example, the crossfl ow developed by impingement cooling in the leading edge can affect the mid-chord jet impingement. Presence of initial crossfl ow can signifi cantly alter the aerodynamic boundary layer and mass distribution in jet rows; and therefore, heat transfer patterns are different from that with spent jet

Table 2: Correlation constants for different jet arrangements [6].

Inline pattern Staggered pattern

C nx ny nz C nx ny nz

A 1.18 −0.944 −0.642 0.169 1.87 −0.771 −0.999 −0.257M 0.612 0.059 0.032 −0.022 0.571 0.028 0.092 0.039B 0.437 −0.095 −0.219 0.275 1.03 −0.243 −0.307 0.059N 0.092 −0.005 0.599 1.04 0.442 0.098 −0.003 0.304

Figure 7: Nusselt number ratio of staggered arrangement and inline arrangement for different crossfl ow conditions [6].



crossfl ows. This study [7] is different from their previous study [6] due to the initial crossfl ow provided in the left side of the confi guration. Figure 8 shows the Nusselt number and jet impingement effectiveness for different initial crossfl ow conditions. The effectiveness is defi ned as:

(10)

This effectiveness defi nition is analogous to the effectiveness used for fi lm cool-ing analysis. However, unlike fi lm cooling, for jet impingement, the jet fl ow is the primary fl ow and the crossfl ow is the secondary fl ow. Since the dominance of jet fl ow is desirable in jet impingement, lower effectiveness indicates a better impinge-ment performance.

Figure 8: Effect of initial crossfl ow on impingement.



The cooling effectiveness of jets decreases with the axial distance measured in the direction of the crossfl ow. The effect of initial crossfl ow can clearly be seen in the heat transfer pattern. A higher mass fl ow from a jet hole increases the related heat transfer. This fi gure compares heat transfer coeffi cients for staggered and inline jet array confi gurations. Nusselt number is signifi cantly reduced in the presence of crossfl ow for both inline and staggered formations. The cooling effec-tiveness of inline formation is lower and Nusselt number is higher than corre-sponding staggered jet array formation.

3.2 Impingement cooling for combustor liners

Nealy et al. [1] proposed a multi-layer fi lm convection scheme that was called Lamilloy©. Figure 9 shows that a combination of multi-jet impingement with roughened wall was the best option for low pressure drop systems such as combus-tors. The multi-jet system offered more fl ow control and ability to deliver cooling fl ow more evenly and maintain temperature uniformity.

Andrews et al. [10] proposed a full coverage impingement system for combus-tion chamber wall cooling. They provided detailed design data for different pitch-to-diameter ratios with a constant gap with just multi-jet impingement but also a combination of jet impingement with effusion cooling. They concluded that impingement cooling combined with effusion cooling provided a signifi cant increase in cooling effectiveness. They found that their simple combination design performed as well as the complex Lamilloy© wall design.

Schulz [11] showed that new cooling schemes with ribbed channels, pin-fi ns, and/or impingement cooling could be more effective in modern low emission combustors. Figure 10 shows the concepts and heat transfer results with ribs, pins, and impingement cooling. They indicated that large impingement holes are not as effective as ribs or pins in such narrow gap channels. However, impingement effectiveness can be greatly enhanced with smaller holes but will result in greater pressure penalty.

Bailey et al. [12] explored combustor liner models utilizing both impingement jet cooling, high Reynolds number turbulated fl ow between the liner and fl ow sleeve, and variable passage geometry. The impingement jet diameters are not uniform as each row has different jet size, hence there is a broad range of jet Reyn-olds number conditions and crossfl ow ratios. There were six spanwise rows of impingement jets with average target distance-to-diameter ratio of 2.3 for these jets. The average jet array spacing streamwise and spanwise was 4. Sharp, square turbulators with full fi llet radius were machined in the liner surface over the latter 50% of the fl ow path. The turbulators were placed transverse to the fl ow, with height of 0.76 mm, a pitch-to-height ratio of 10, and an average height-to-channel height ratio of 0.022. Figure 11 shows the test confi guration with initial crossfl ow entering the channel and then the jet impingement array and fi nally the turbulated channel after the impingement. The study compared heat transfer behavior in the channel with only initial crossfl ow with downstream rib turbulators, initial crossfl ow with impinging jets without downstream turbulators, and fi nally initial crossfl ow with impingement and downstream rib turbulators. Figure 12 shows the



heat transfer measurements in the channel for all three conditions. The fi rst case with pure convection shows entrance effect with high heat transfer at the entrance and decreases as the channel boundary layer develops. There is a signifi cant increase when the channel counters the ribs downstream and there is a large enhancement in the downstream region. With the jet impingement only case, the peaks and valleys caused by jet impingement are clearly visible in the upstream region and there is a strong decrease downstream as the jet effect is dissipated. The fl ow develops into a channel fl ow downstream as there are no ribs to enhance the

Figure 9: (a) Film convection cooling schemes (b) Comparison of different con-vective cooling schemes [1].



heat transfer. With the combined system of impingement and trips, heat transfer peaks and valleys are not as signifi cant as in the pure impingement case. However, the downstream region with ribs is greatly enhanced due to the increased turbu-lence produced by the jets upstream. Overall, the combination of jet impingement and rib turbulators seems to perform the best in terms of heat transfer removal from the liner wall.

Figure 10: Effect of heat transfer augmentation on the outer surface of liner walls [11].

Figure 11: Flow crosssection and detailed surface construction for multi-scheme cooling employed by Bailey et al. [12].



Gao et al. [13] studied i mpingement geometries where the spacing between the holes increased in both the streamwise and spanwise direction simulating the stretching of the hole arrays downstream. Two different arrays were investigated with the fi rst array having uniform diameter holes through the array placed in a stretched format. The second array has holes placed in the same locations with increasing diameter along the streamwise direction. Figure 13a shows the uniform diameter jet plate. All the holes are of a uniform diameter with holes of diameter 0.635 cm. The fi rst row of holes are placed 2 hole diameters apart from each other in the spanwise direction. The second row of holes are placed 3 hole diameters apart and the second row is placed 2 hole diameters downstream of the fi rst row. Similarly, the downstream rows are placed in an increasing distance both in span-wise and streamwise direction simulating a stretched array of holes. Figure 13b shows the varying diameter plate. In this case, the hole diameters are increasing from the fi rst row with hole diameters of 0.3175 cm to the last row with 0.635 cm. The locations of the holes are identical on the plate to the uniform diameter holes resulting in varying spanwise and streamwise normalized distances. Table 3 shows the normalized distances comparing the uniform and varying diameter plates. The spacing between the jet plate can be varied by changing the wall spacers along the three closed sides and changing the wall distance-to-jet diameter ratio (Z/D). These values are also presented in Table 4.

Figure 14 presents results from Gao et al. [13] for the regional average heat transfer coeffi cient comparisons with a Re = 6,000 for jet height to diameter ratio of Z/D = 3 with the predictions correlations provided by Kercher and Tabakoff [5] and Florschuetz et al. [6]. Both the correlations predicted very high heat transfer coeffi cients for the fi rst row and show immediate degradation downstream for the

Figure 12: Comparison of circumferentially averaged test data for different geom-etries as presented by Bailey et al. [12].



Table 3: Jet plate geometry for stretched impingement arrays presented by Gao et al. [13].

Row # Hole diameter (cm)

Normalized spanwise spacing based on local hole diameter (Y/D)

Normalized streamwise spacing based on upstream hole diameter (X/D)

U.D. V.D. U.D. V.D. U.D. V.D.

1 0.635 0.3175 2 4 – –2 0.635 0.3630 3 5.25 2 43 0.635 0.4082 4 6.22 3 5.254 0.635 0.4765 5 6.66 4 6.225 0.635 0.4989 6 7.63 5 6.666 0.635 0.5443 7 8.17 6 7.637 0.635 0.5895 8 8.62 7 8.178 0.635 0.635 9 9 8 8.62

rest of the rows. The degradation drops past the fi fth row because the degradation constant (φ2) is held constant beyond this point. The fi rst row is severely over-predicted as the correlation uses no crossfl ow at this point. In the present experi-ment, the fi rst two rows produce identical levels of heat transfer coeffi cients for all three jet heights and degrade downstream but not as rapidly as the correlation predicted values. Far downstream, it appears that the correlations and the experi-ments are in good agreement. Kercher and Tabakoff [5] correlation consistently predicts values lower than that from Florschuetz et al. [6] correlation. In fact, the zero-crossfl ow fi rst row has Kercher and Tabakoff [5] predicting closer values to

Figure 13: Linearly stretched arrays studied by Gao et al. [13] (a) uniform hole diameter (b) varying hole diameter.



the experiment. The crossfl ow effect is over-predicted for both the correlations resulting in lower heat transfer coeffi cients than measured. Also, the lack of deg-radation values for large crossfl ow-to-jet fl ow ratios resulted in extrapolation, which may be causing the discrepancies in the predictions. It is clearly evident that more number of downstream rows will show increased effectiveness of stretched arrays and also reduce overall coolant usage for cooling combustor liners.

Table 4: Jet plate-to-target wall dimensions for stretched arrays [13].

Jet-to-wall spacing →

H = 0.635 cm H = 1.905 cm H = 3.175 cm

Row number↓

U.D. V.D. U.D. V.D. U.D. V.D.

1 1 2 3 6 5 102 1 1.75 3 5.25 5 8.253 1 1.56 3 4.67 5 7.784 1 1.33 3 3.99 5 6.655 1 1.27 3 3.82 5 6.366 1 1.17 3 3.51 5 5.837 1 1.08 3 3.24 5 5.408 1 1 3 3 5 5

Figure 14: Comparing performance of linearly stretched arrays to standard impingement correlations [13].



Facchini and Surace [14] evaluated the heat transfer performance of sparse arrays, to increase our knowledge of such confi gurations and also provided basis for impingement to be used in combustor liner confi gurations. Effectiveness is a com-mon measure of how close the heat-transfer-driving temperature (adiabatic wall temperature) is to the jet or crossfl ow temperature; this is commonly expressed as:

(11)

where Tc and Tj, respectively, indicate crossfl ow and jet temperature. According to the defi nition, the adiabatic wall temperature, Taw, is the surface temperature of a perfectly insulated wall. As a consequence, heat fl ux can be computed here by use of the equation:

(12)

They found that streamwise heat transfer coeffi cient comparison with correla-tions showed minor deviations but the lack of a similar correlation for effective-ness is emphasized to effectively calculate local heat fl ux.

As previous studies have shown, the crossfl ow induced by upstream jets tend to push the downstream jets away from the surface resulting in strong degradation of impingement heat transfer. The jet has to make strong impact on the heat transfer surface to provide higher heat transfer coeffi cients. As the number of jet rows increase, the degradation is signifi cant and results in almost channel like fl ow. Esposito et al. [15] presented low crossfl ow designs for jet impingement for arrays with ten rows. An extended port design was studied for the jet plate and compared with a corrugated wall design and baseline design to redirect spent air away from downstream jets in an effort to reduce the detrimental effects of crossfl ow. Corru-gated wall design traps the spent air in the corrugations between impingement jets to reduce crossfl ow effects on downstream jets. Extended ports provide directed impingement on the test surface with suffi cient space for spent fl ow to expand and not push the downstream jets.

Table 5 summarizes the test geometries. The baseline impingement plate consisted of a fl at jet plate with circular holes that were 1 hole diameter in length (L/d = 1). The spacing of the holes in the spanwise, y/d, and streamwise, x/d,

Table 5: Geometrical parameters for all cases studied by Esposito et al. [16].

Baseline Corrugated Extended ports Variable extended

z/d 3 Vary from 6 to 3H/d 6x/d and Y/d 5L/d 1 4 Vary from 1 to 4Si/d 3.35Si/d 1.65



directions was 5. The fi rst new geometry tested was a corrugated wall design. The corrugations in the wall allow spent air from upstream jets to exit the impingement array without interfering with the downstream jets. Figure 15a shows the corru-gated wall design and how the spent air is expected to exit the impingement array. In addition to routing the spent air around the downstream jets, the increase in crosssectional area of the section due to the corrugations also decreases the overall crossfl ow velocity.

Figure 15b shows the uniform extended port design. This design offers a higher crosssectional area for crossfl ow than the baseline and corrugated wall, therefore reducing the overall crossfl ow velocity. Also the length of the impingement tube is an additional benefi t that allows a more developed jet fl ow. This increases the peak jet velocity and further reduces crossfl ow effects by increasing the core jet to crossfl ow velocity ratio. A variation to the extended port design was also tested with variable extended port lengths, shown in Fig. 15c. This further increases the crossfl ow area especially for the front nine rows. The length of the extended ports was linearly varied from the fi rst to the last rows of jets. All ports were of uniform length in the spanwise direction. A total of four cases are investigated. Case 1 is the baseline case. Case 2 is the corrugated wall. Cases 3 and 4 are the uniform extended port and variable extended port designs. Three jet Reynolds numbers, Red, were tested for 20,000, 40,000, and 60,000.

Figure 16 presents the detailed Nusselt number distributions for Re = 60,000. The baseline case shows the effect of crossfl ow-related heat transfer degradation beyond the 5th row. The corrugated wall does not show signifi cant enhancement in the core jet region but produces almost uniform heat transfer characteristics for all 10 rows. The uniform extended port case shows strong core impingement for the

Figure 15: Different crossfl ow effect reducing impingement schemes tested by Esposito et al. [15]; (a) Corrugated wall (b) Extended port – uniform (c) Extended port – varying.



Figure 16: Detailed heat transfer coeffi cient distributions for different geometries tested by Esposito et al. [15] at a jet diameter-based Reynolds number of 60,000.

fi rst 9 rows and shows slightly lower values for the 10th row. The variable extended port shows lower Nusselt numbers for the upstream rows and higher for the downstream rows due to the longer tube lengths for the downstream rows. However, the variable extended ports show some mixing between adjacent jets compared with corrugated wall and uniform extended port cases.



Figure 17 presents the comparison of span averaged Nusselt number distribu-tions for Re = 60,000. The Nusselt numbers for the fi rst three rows are similar for all four cases except for variable extended ports that shows lightly lower Nusselt numbers. Beyond the fourth row, the Nusselt numbers for the new geometries are signifi cantly higher than the baseline. The high Reynolds number associated mass fl ow rate is clearly causing a severe degradation on baseline heat transfer. The cor-rugated wall is at similar levels to both the extended port confi gurations at this high Reynolds number condition. The variable extended port performs better for the downstream rows compared with the other two geometries. The baseline com-parison with Florschuetz correlation is signifi cantly different at the fi rst row and the last few rows. The correlation predicts a larger crossfl ow effect than the exper-iment.

Overall pressure drop was measured for all cases where the overall pressure drop was the differential pressure before the impingement plate to the ambient. The four geometries tested indicated less than 5% variation in overall pressure drop due to variation of geometry. It was clear that the change in pressure drop compared with heat transfer enhancement was minimal.

4 Conclusions

As environmental restrictions take precedence in next generation turbine engine design, combustor technology improvements will be critical. Also, of importance is the overall effi ciency and choice of fuel. Impingement cooling of combustor

Figure 17: Comparing low crossfl ow effect geometries with baseline and Florschuetz correlation [15].



liners can provide the option of running the combustor in a lean-burn mode and near-stoichiometric conditions to improve effi ciency. Impingement cooling can be used to cool the liners and the heated cooling air can then be bypassed back to mix with the fuel to reduce overall fuel usage to maintain temperatures that produce minimum emissions. Several options of impingement cooling and combined cool-ing systems have been studied in the past decade or so that clearly show the high value of back side cooling geometries. The ability to design, predict, evaluate, implement, and fabricate such complex cooling geometries will require testing, development of correlations to predict heat transfer behavior, and new manufac-turing techniques, and focus on joints, welds, and other components needed to meld new features.

References

[1] Nealy, D.A., Reider, S.B. & Mongia, H.C., Alternate cooling confi gurations for gas turbine combustion systems. AGARD Conference Proceedings No. 390, 1985, Paper 25.

[2] Dano, B.P.E., Liburdy, J.A. & Kanokjaruvijit, K., Flow characteristics and heat transfer performance of a semiconfi ned impinging array of jets: effect of nozzle geometry. International Journal of Heat and Mass Transfer, 48, pp. 691–701, 2005.

[3] San, J.-Y. & Lai, M.-D., Optimum jet-to-jet spacing of heat transfer for stag-gered arrays of impinging air jets. International Journal of Heat and Mass Transfer, 44, pp. 3997–4007, 2001.

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[5] Kercher, D.M. & Tabakoff, W., Heat transfer by a square array of round air jets impinging perpendicular to a fl at surface including the effect of spent air. ASME Journal of Engineering and Power, 92, pp. 73–82, 1970.

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[11] Schulz, A., Combustion liner cooling technologies in scope of reduced pol-lutant formation and rising thermal effi ciencies. Annals New York Academy of Sciences, 934, pp. 135–146, 2000.

[12] Bailey, J.C., Tolpadi, A., Intile, J., Fric, T., Nirmalan, N.V. & Bunker, R.S., Experimental and numerical study of heat transfer in a gas turbine combus-tor liner. ASME Paper GT-2002-30183, 2002 Turbo Expo, Amsterdam, Netherlands.

[13] Gao, L., Ekkad, S.V. & Bunker, R.S., Impingement heat transfer, part I: linearly stretched arrays of holes. AIAA Journal of Thermophysics and Heat Transfer, 19(1), pp. 57–65, 2005.

[14] Facchini, B. & Surace, M., Impingement cooling for modern combustors: experimental analysis of heat transfer and effectiveness. Experiments in Fluids, 40, pp. 601–611, 2006.

[15] Esposito, E., Ekkad, S.V., Kim, Y.W. & Dutta, P., Novel jet impingement cooling geometry for combustor liner backside cooling. ASME Journal of Thermal Science & Engineering Applications, 1, 021001, 2009.

[16] http://www.e-inst.com/combustion/carbon-monoxide


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