CHAPTER 1 Impingement Cooling in Gas Turbines: Design ......6 IMPINGEMENT JET COOLING IN GAS...

CHAPTER 1

Impingement Cooling in Gas Turbines: Design, Applications, and Limitations

Ronald S. Bunker1, Jason E. Dees2 & Pepe Palafox1

1GE Aviation, Cincinnati, OH, USA.2GE Global Research, Niskayuna, NY, USA.

Abstract

The use of impingement cooling to maintain acceptable material temperatures in cooled gas turbine components is one of the handful of very robust and widely employed cooling methods throughout propulsion and power-generating engines. Impingement cooling utilizes an available pressure differential and one or many appropriate nozzle confi gurations to generate high-speed jets that are directed at component interior surfaces for the purpose of cooling the walls. The potential design space for impingement cooling in these components is immense, yet the actual design space in use tends to be well characterized by a relatively few estab-lished correlations. Impingement cooling is used in virtually all of the hot gas path components of the engine, including the combustor, turbine vanes, blades, and shrouds, and even the rotating disks. The applications span jet Reynolds numbers from 10,000 to 500,000. The complexity of impingement cooling comes with the vast numbers of specifi c jet and surface geometry combinations that are involved with the actual components. Impingement cooling is a very fl exible method, and for this very reason must be made to conform to applications that are dictated by other design requirements, such as those from aerodynamics, mechanical struc-ture, vibrations, low cycle fatigue, and creep rupture limits. This chapter presents the basics of gas turbine impingement cooling design, the various applications within the engine, and the typical limitations imposed upon its use. Generic use and application examples for single impingement jets, in-line rows of jets, and arrays of jets are described for the hot gas fl ow path components of gas turbines, including the combustor system and high-pressure turbine. Specifi c applications to airfoil aerodynamic leading and trailing edge regions, combustor liners, and rotat-ing disks are provided, as well as emerging applications within confi ned channels, blade tips, and fi lm cooling. Signifi cant effects infl uencing impingement cooling

www.witpress.com, ISSN 1755-8336 (on-line) WIT Transactions on State of the Art in Science and Engineering, Vol 76, © 2014 WIT Press

doi:10.2495/978-1-84564-907-4/001

2 IMPINGEMENT JET COOLING IN GAS TURBINES

are discussed including surface roughness, rotation, and accumulated cooling cross-fl ow degradation. The chapter wraps up with a discussion of the main limi-tations placed on impingement cooling due to available pressure, manufacturing, turbine operation, and system thermal management.

Keywords: Gas turbine, hot gas path, impingement cooling, thermal design.

1 Introduction

The use of fl uids within a gas turbine engine defi nes the very function of the engine, both the motive power derived from the chemical energy of the fuel and the means of containing and controlling that power. Relatively few regions of the engine utilize purely static fl uids, instead the engine is dominated by convec-tion of fl uids, mainly forced but also natural. The fl ow path of the main working fl uid through the fan, compressor, turbine, and exhaust entails stagnation regions, boundary layers, separation regions, core fl ows, leakages, injections, and even shocks. Impingement as a fl ow mechanism in this fl ow path plays an important role, particularly in determining aerodynamic losses throughout and in setting sometimes limiting conditions on turbine heat loads. The use of impingement, however, fi nds its primary useful function in the cooling and sealing of the engine hot gas path. Figure 1 shows the layout of a typical aviation gas turbine engine and the main uses of impingement to maintain the engine function and operability. While the use of impingement for cooling of the combustor, high-pressure turbine (HPT) vanes, blades, and shrouds is most notable, impingement is also used for active clearance control of the turbine casing, sealing of the turbine secondary systems, pre-swirl cooling of the turbine disks, and in a heating mode to provide de-icing in the fan-compressor system. Since most gas turbines are self-contained systems, all of these impingement fl ows usually use air bled from appropriate stages of the compressor.

Figure 1: General uses of impingement in the gas turbine engine.


IMPINGEMENT COOLING IN GAS TURBINES 3

The technology of cooling gas turbine components, primarily via internal con-vective fl ows of single-phase gases and external surface fi lm cooling with air, has developed over the years into very complex geometries involving many dif-fering surfaces, architectures, and fl uid–surface interactions. The fundamental aim of this technology area is to obtain the highest overall cooling effectiveness with the lowest possible penalty on the thermodynamic cycle performance. Why use impingement? Simply stated, it is because impingement provides the highest possible convective heat transfer coeffi cient augmentation factors compared with all other modes of single-phase cooling. Almost all highly cooled regions of the HPT components involve the use of turbulent convective fl ows and heat transfer. Very few if any cooling fl ows within the primary hot section are laminar or transitional. Moreover, the typical range of Reynolds numbers for cooling techniques, using traditional characteristic lengths and velocities, is from 10,000 to 60,000. This is true for both stationary and rotating components. The enhance-ment of heat transfer coeffi cients for turbine cooling makes the complete use of the turbulent fl ow nature by seeking to generate mixing mechanisms in the cool-ant fl ows that actively exchange cooler fl uid for the heated fl uid near the walls. These mechanisms include various forms of shear layers, boundary layer disrup-tion, and vortex generation. Impingement cooling literally bypasses these mech-anisms by impacting the coolant directly on the surface, thereby creating a new and continually refreshed local boundary layer. In a marked difference from con-ventional heat exchangers, most turbine cooling methods do not rely on an increase in cooling surface area, since the available surface area to volume ratios

Figure 2: Relationship of heat transfer coeffi cient augmentation to friction coeffi -cient augmentation for basic gas turbine cooling enhancement methods.



are very small. Surface area increases are benefi cial but are not the primary objective of turbine cooling enhancements. The use of various enhancement techniques typically results in at least 1.5× and as much as 5× increase in local heat transfer coeffi cients over that associated with fully developed turbulent fl ow in a smooth duct. As depicted in Fig. 2, these enhancements typically occur at the price of much higher friction coeffi cient augmentations. This is very impor-tant to note because most cooling applications within the engine are highly restricted in the available pressure loss allowed, which directly affects the effi -ciency through the reduction of useful work extraction and/or thrust. Of all the cooling methods, impingement alone operates in the region around the Reynolds analogy where the augmentation in heat transfer is equivalent to that of friction. Impingement cooling requires the pressure driver to accelerate the fl ow as jets but suffers negligible pressure loss due to the actual act of impingement on the surface.

Since the cooling effectiveness of impingement jets is very high, this method of cooling provides an effi cient means of component heat load management, given suffi cient available pressure head and geometrical volume for implementa-tion. Regular arrays of impingement jets are used within turbine airfoils and end-walls to provide relatively uniform and controlled cooling of fairly open internal surface regions. Such regular impingement arrays are generally directed against the target surfaces by the use of sheet metal baffl e plates, inserts, or covers that are fi xed in position relative to the target surface. Figure 3 shows a conventionally

Figure 3: Turbine inlet guide vane’s typical use of impingement cooling.



cooled turbine inlet guide vane that utilizes two shaped impingement baffl es to create jet arrays within the airfoil. The inner and outer endwalls of the vane may also employ impingement baffl es, each with varying regional differences in the arrays. These arrangements allow for the design of a wide range of impingement geometries, including in-line, staggered, or arbitrary patterns of jets. In more confi ned regions of airfoils such as the leading edge or trailing edge, spanwise lines of impingement jets are sometimes used to focus cooling on one primary location of high external heat load like the airfoil aerodynamic stagnation region. Figure 4 shows an example of an HPT blade cooling design that highlights such lines of impingement jets for cooling both the leading and trailing edges. There also exist many other applications for individual impingement jets on selected stationary and rotating surfaces. Endwalls and platforms, unattached shrouds, and combustor liners may all have specifi c local cooling requirements well suited to the use of individual jet cooling. Summaries of applicable fundamental impingement heat transfer research may be found in Martin [1] and Han and Goldstein [2].

In practice, arrays of impinging jets are the most common form of this cooling technique. Figure 5 shows the heat transfer coeffi cient distribution that results from a typical impingement array in which the jets do not have any strong inter-action with other, as well as that of a single impinging jet. In both cases, it is noteworthy that the ratio of maximum-to-minimum heat transfer coeffi cient can be as high as 10. The main geometric parameters include the jet diameter D, the distance to the target surface Z, and the array jet-to-jet spacing X and Y. Most

Figure 4: High-pressure turbine blade’s typical use of impingement cooling.



applications employ target spacing of 1 < z/D < 5, array spacing of 3 <x/D and y/D < 12, and jet Reynolds numbers ReD of 20,000 to 70,000. For designing purposes, impingement jet array heat transfer coeffi cients may be obtained from the correlation of Florschuetz et al. [3]. This correlation includes another main parameter Gj/Gc that accounts for the relative strength (mass fl ux) of the local impinging jet to that of the accumulated post-impingement cross-fl ow, which determines the degree of impinging jet defl ection and/or degradation due to cross-fl ow.

Nu = A Rem {1 – B[(z/D)(Gc/Gj)]n } Pr1/3,

where the coeffi cients A, B, m, and n are each of the form

A, B, m, and n = C (x/D)q (y/D)r (z/D)s.

The full parameter space for generic impingement jets and arrays of jets includes the jet shape, jet length-to-diameter ratio, jet angle (origin and defl ected angles), target surface geometry (e.g. roughness or pattern), non-uniform jet sizing, non-uniform array spacing, geometry confi nements, fl uid entrainment, interactions with coolant extractions, target rotation, Coriolis and centrifugal forces, surface

Figure 5: Basic jet array impingement geometry (below) and array heat transfer coeffi cient distribution (above left) with single-jet relative magnitude range stagnation point to far fi eld (above right).



curvature, and more. Many of these parameters and devices will be discussed with respect to their applications in gas turbines in the following sections.

2 Applications

2.1 Single-jet impingement cooling

Heat transfer from a single impinging jet on a surface represents the most basic impingement confi guration. As mentioned previously, the full parameter space affecting single impinging jets is large. The major parameters include jet shape, jet target distance to diameter ratio (z/d), impingement angle, and jet Reynolds num-ber (Rej). Thorough summaries of single-jet impingement heat transfer research can be found in Martin [1]. As described by Martin [1], the fl ow fi eld of an imping-ing jet can be divided into three regions. These regions consist of the free jet region, the stagnation zone, and a post-impingement region of wall fl ow. Figure 6 shows a schematic representation of an impinging jet fl ow. A more thorough fundamental description of the impinging jet fl ow fi eld can also be found in reference [1].

Two major factors in the cooling effectiveness of impingement heat transfer on a fl at plate, Rej and z/d, were examined for a circular jet issuing into free atmo-sphere by Gardon and Cobonpue [4]. This study examined the round jets with Rej ranging from 7,000 to 112,000 and varied z/d from 0.5 to 50. The study revealed that a maximum stagnation heat transfer coeffi cient (or Nusselt number) exists for a single impinging jet at about z/d = 6–8, which is important in the context that most gas turbine applications are limited to target distances of z/d < 5. As will become apparent though, consideration of the stagnation heat transfer magnitude alone is rarely of primary interest in gas turbine applications. The average stagna-tion regional heat transfer coeffi cient over some annular surface radius from the

Figure 6: Circular impinging jet fl ow morphology (after Martin [1]).



stagnation point is generally more important. The study of Perry [5] showed that the average heat transfer coeffi cient drops off very quickly as the distance from the centerline of the jet increases relative to the jet-to-target spacing Z. By a surface radial distance of 20% of Z from the jet centerline, the heat transfer coeffi cient is reduced to <60% of the peak value. By 80% of Z, the average heat transfer coef-fi cient decreased to <10% of the peak value. A companion study, by Gardon and Akfi rat [6], examined impingement heat transfer from a single, two-dimensional slot jet over a similar parameter space. Qualitatively, the dependence of stagnation heat transfer coeffi cient on Rej and z/d for a round and slot jet was similar. How-ever, quantitative values of stagnation heat transfer due to a slot jet were notably lower than the corresponding round jet values.

The effect of the angle of an impinging jet relative to the target surface on the impingement heat transfer was studied by Goldstein and Franchett [7]. In addition to the impingement angle, the effects of Rej and z/d were also varied. Two- dimensional plots of the impingement heat transfer coeffi cient profi le are elliptical in nature when the angle of the jet relative to the surface is decreased from the typical 90° case. Peak values of heat transfer coeffi cient were not signifi cantly affected by decreasing the angle relative to the surface until the angle was less than about 30°, which resulted in about a 20% decrease in values for all cases. Due to the limited space available in assembled gas turbines and gas turbine components, the case of impingement heat transfer in a confi ned space, where entrainment of surrounding gas is limited, is an important factor to consider. A study on impinge-ment angle in a wall-confi ned fl ow confi guration was performed by Ichimiya [8]. In this case, the confi ning wall prevents the impingement jet from entraining any surrounding gas and limits the area that the post impingement wall fl ow can occupy. Consistent with the results of Gardon and Akfi rat [6], a little decrease in peak heat transfer coeffi cient was seen when decreasing the impingement angle from 90° to 45°. However, the confi ned nature of the fl ow fi eld did cause some qualitative differences in the shape of the two-dimensional heat transfer coeffi cient distribution when compared with the unconfi ned case.

In gas turbine components, impingement jets are often directed toward the target surface by thin metal inserts with holes machined into them. The work of Brignoni and Garimella [9] examined the effect of various chamfer sizes on the resulting impingement heat transfer. Generally, the presence of a chamfer was shown to have little effect on the impingement heat transfer coeffi cient. However, an inlet cham-fer did result in a lower pressure drop across the impingement nozzle in-line with the basic fl uid mechanics of fl ow in a sudden contraction. This implies that similar heat transfer performance can be achieved with a lower coolant supply pressure if a cohesive jet can be formed, and if manufacturing constraints will allow.

2.2 Impingement from in-line jet rows

The arrays of impingement holes are often used in gas turbine component cooling to increase the area being cooled. An inline row of impingement holes is par-ticularly useful in regions such as blade leading edges, where high external heat



transfer coeffi cients along the stagnation line must be balanced by high internal heat transfer coeffi cients to maintain the metal temperature and achieve the desired part life. In addition to the previously discussed parameters relevant to a single impinging jet, the distance between adjacent jets in an in-line row (x/d) is an addi-tional parameter of interest. The work of Metzger and Korstad [10] examined the effects of a single row of impinging jets issuing onto a fl at plate with a cross fl owing gas. In turbine applications, impinging gas is often confi ned and forced to exhaust in one direction. This can lead to signifi cant cross-fl ow effects due to accumulating mass velocity of the spent impingement gas. The results of Metzger and Korstad [10] demonstrate that the interaction of the impinging gas jets with the cross fl ow degrades impingement performance relative to a case with no cross-fl ow, as shown in the example of Fig. 7. As x/d increases, the performance relative to the case with no cross-fl ow degrades less. This is due to the larger jet spacing providing less of a blockage to the cross-fl ow, resulting in less interaction between the cross-fl ow and impinging jets.

2.3 Leading edge cooling

Due to the high heat loads present along the stagnation line of gas turbine airfoils, a spanwise array of impingement jets is often directed against the internal surface of the airfoil leading edge. In turbine vanes, the impinging air is directed against the internal leading edge surface through metal impingement plates as illustrated in Fig. 3. In blade leading edges, the fl ow is often directed against the leading edge through an array of ‘crossover’ impingement holes fed by a coolant cavity that may be an integral part of the serpentine cooling of the blade depending upon the blade cooling design. Figure 8a shows an enlarged view of this example leading edge impingement region. In addition to the major variables discussed previously, the geometry of the concave target surface has an effect on the resultant impinge-ment heat transfer coeffi cient.

The relevant geometric variables in leading edge impingement consist of the leading radius as well as leading edge ‘sharpness’ created by elongation of the blade leading edge due to airfoil aerodynamic performance considerations. As explained in Metzger et al. [11], increasing airfoil sharpness can result in z/d

Figure 7: Cross-fl ow degradation effect on heat transfer coeffi cients for in-line impinging jets (reproduced from [49] with the permission of ASME).



values much larger than the optimal jet to target spacing. This results in signifi -cantly reduced heat transfer coeffi cients relative to the optimal z/d spacing. While the previous statement is generally true, the work of Bunker and Metzger [12], which obtained full surface distributions of heat transfer coeffi cients in leading edge models varying the concave relative radius, target distance and jet size, dem-onstrated that impingement heat transfer in the leading edge region is dependent on complex interactions between several parameters and is diffi cult to correlate succinctly. Yet, another important aspect of leading edge impingement cooling is the interaction of the impinging jets with fi lm cooling extraction. As shown by the work of Metzger and Bunker [13], the interaction of fi lm extraction in leading edge impingement is also complex with fi lm extraction causing an increase or decrease in impingement heat transfer coeffi cients depending on the precise geo-metric confi guration. In general, though the presence of fi lm extraction holes serves to eliminate any internal cross-fl ow effects and also breaks down fl ow recir-culation zones that otherwise would degrade cooling effectiveness.

2.4 Trailing edge cooling

Trailing edges of turbine airfoils are characterized by narrow, low aspect ratio cooling channels that typically exhaust through holes in the trailing edge or pres-sure side slots (e.g. Figs 3 and 4). Due to these geometric constraints, common cooling arrangements in the trailing edge are coolant fl ow across a pin bank or protruding turbulator ribs. The work of Taslim and Nongsaeng [14] provides an example of trailing edge impingement fed from a row of elliptically shaped cross-over holes onto a smooth surface. Due to the low aspect ratio of the cooling edge channel, impingement is limited to only one of the internal channel surfaces at a very shallow impingement angle, as shown in the enlarged example of Fig. 8b. This study examined the cases with shallow angle impingement on one of the

Figure 8: (a) Leading edge and (b) Trailing edge impingement (after US Patent 6273682).



channel sidewalls as well as no impingement and only convective heat transfer on the channel sidewalls. The results of the study show modest heat transfer improve-ments on the impinged sidewall with no detrimental effects on the opposite side-wall, relative to the case with no impingement. Due to the shallow impingement angle required by the channel geometry, the full effectiveness of impingement heat transfer cannot be realized. However, augmentation of the impingement heat transfer may be obtained by means of angled turbulators. The work of Coletti et al. [15] investigated impingement to a rib roughened trailing edge surface. Only the trailing edge wall that was impinged upon was roughened; the other wall was left smooth. The presence of the ribs caused enhanced heat transfer on the ribbed wall. Enhanced heat transfer was also observed on the smooth wall due to the ribs defl ecting the post impingement fl ow such that secondary impingement occurred on the smooth wall as well.

2.5 Surface jet array impingement

One of the primary applications of jet impingement cooling in gas turbines comes in the form of two-dimensional arrays of impingement jets. As mentioned previ-ously, arrays of impinging jets allow for highly effective cooling of large areas such as airfoil surfaces and endwalls. In addition to the jet-to-jet spacing in two dimensions (x/d and y/d), the effect of accumulating cross-fl ow is a signifi cant factor in the performance of two-dimensional surface arrays of impinging jets. In most turbine cooling applications, the post-impingement fl uid is forced to fl ow away from the cooled region in a single direction. For a large array of impingement jets, the accumulation of spent impingement fl uid can establish a signifi cant cross-fl ow that interacts with impinging jets in a detrimental way. Figure 9, from Bailey and Bunker [16], shows impingement heat transfer distributions from arrays that

Figure 9: Array effects on heat transfer coeffi cients (a) without and (b) with cross-fl ow buildup (reproduced from [16] with the permission of ASME).



either are not (9a) or are (9b) affected by an accumulating cross-fl ow (no initial cross-fl ow was present). In general, the accumulating cross-fl ow decreases the peak heat transfer coeffi cient from the impinging jets. However, the accumulat-ing cross-fl ow can also serve to increase the heat transfer coeffi cients in regions between the jets due to increased convective heat transfer. In design practice, the correlation presented by Florschuetz et al. [3] serves as a valuable tool for esti-mating the average heat transfer coeffi cients for a given confi guration. Depend-ing on the exact geometric confi guration and fl ow rate, accumulating cross-fl ow effects can decrease the row averaged heat transfer coeffi cients by more than 20%. The work of Kercher and Tabakoff [17] also examined an array of impingement jets and confi rmed the detrimental effects of the accumulating cross-fl ow on the impingement heat transfer coeffi cient.

The work of Florschuetz et al. [3] and Kercher and Tabakoff [17] provide the average heat transfer coeffi cient values on a row-by-row basis, taking into account the major variables of jet-to-jet spacing, jet-to-target distance, and the ratio of cross-fl ow to impingement jet mass velocity. The rows of impingement jets are averaged in the direction perpendicular to the major cross-fl ow direction. In gen-eral, the jet-to-jet spacing has a signifi cant effect on the row averaged values, with larger jet-to-jet spacing resulting in lower averaged heat transfer values. In gen-eral, the heat transfer coeffi cients decrease with increasing cross-fl ow. A primary reason for the decreasing Nu with increasing cross-fl ow is the variation of Rej along the streamwise direction of the impingement array. For an array with con-stant impingement supply pressure, the cross-fl owing gas will cause a streamwise pressure gradient in the exhaust channel. This effect leads to decreasing Rej values as spent impingement fl ow accumulates in the exhaust fl ow. The interaction between the heat transfer coeffi cients, jet-to-jet spacing, jet-to-target distance and the effect of cross-fl ow is complicated, and in some cases the increasing cross-fl ow can overcome the effect of decreasing Rej, causing a minimum average heat trans-fer coeffi cient, with increasing values of heat transfer coeffi cient with further increases in gas cross-fl ow.

2.6 Inner and outer flow path cooling

Impingement cooling is the dominant means of cooling employed for the inner and outer endwalls of the vane, though fi lm cooling is usually also used here, and the stationary shrouds bounding the blade tips. Impingement is also sometimes used to provide localized cooling on the underside of the blade platforms. In all of these uses, the geometric restrictions on the surfaces lead to some regions that can utilize impingement jet arrays, and other regions that must use a hybrid or cus-tom impingement arrangement of individual jets, irregular groups of jets, angled jets, and variable spacings. A common feature found in all these components is a perimeter fi llet where the edges must either accommodate seals, or where the endwall transits into the airfoil. These situations tend to be highly unique to each design and so there is exceedingly little open literature on the impingement cool-ing aspects. Some research on the effects of irregular impingement jet arrays is



obtained in the studies of Gao et al. [18] and Hebert et al. [19]. Both the stretching of array spacings and the variable sizing of impingement jets row-to-row can be used effectively to compensate for cross-fl ow effects and custom geometries as depicted in the example of Fig. 10.

2.7 Rotating disk impingement

While the majority of impingement applications are to be found in the turbine hot gas path components, the secondary fl ow systems also contain uses for impinge-ment. The principal direct application of impingement is in the wheelspaces of the turbine, commonly referred to as rotor-stator or rotor disk systems. All rotors pump fl uid radially outward simply due to the centrifugal force acting upon the adjacent fl uid. Depending on the specifi c design need, some rotor disks require additional cooling purge fl ows to both maintain the disk temperature and prevent the ingestion of hot gases. These purge fl ows are sometimes also combined with coolant that is being delivered to the blades. Some systems are supplied near the rotor hub or shaft, in which case the fl ow can act as an impinging fl ow in that region and transitions to a boundary layer fl ow radially. Systems may also be supplied by pre-swirl devices to reduce losses in bringing the coolant up to rotational speed, in which case these pre-swirl jets can also be used to impingement cool the rotor face in more outward locations. Figure 11 shows the notional locations for such jet fl ows.

Purge fl ows introduced nearer the hub tend to behave more as free jets, but the resulting rotor face Nusselt numbers depend very strongly on the relative strengths of the rotor pumping and the supplied impinging fl ow. Metzger and Grochowsky [20] and Popiel and Boguslawski [21] studied the fl ow interaction and heat trans-fer between a single jet and a free rotating disk for a combination of jets and disk sizes. These studies looked at a range of impingement fl ow rates, impingement radial locations from the disk center line and disk rotational speeds. They found that higher rotational speed, larger impingement location radius, and smaller jet fl ow rates favored a rotationally dominated fl ow interaction, where heat transfer rates were essentially independent of jet fl ow rate. Lower rotational speed, smaller impingement location radius, and larger jet fl ow rates favored impingement domi-nated interaction, where heat transfer rates increased with increasing jet fl ow rates. In Metzger et al. [22], the same single jet with free rotating disk setup was used to investigate four different disk contoured shapes to understand the effect of the resulting heat transfer. The contour shape was found to have a very little effect on

Figure 10: Example of tailoring an impingement heat transfer coeffi cient distribu-tion via stretched array spacing and varied jet sizing.



the heat transfer, with and without impingement. The impact from impingement jet was found to be negligible in the cases where the impingement fl ow rate from a single jet was less than one-tenth the fl ow rate due to disk pumping. With jet impingement fl ow rate levels comparable to the disk pumping fl ow rates, the heat transfer rates were seen to increase to the values two to three times the values for the non-impingement cases.

In regions closer to and including the rim seal, rotor-stator spacing is reduced and the presence of the stator can have a signifi cant infl uence on the fl ows and the rotor heat transfer. Metzger et al. [23] measured local rotor heat transfer coeffi -cients in a shrouded disk cavity with impinging circular nozzle jet. The local heat transfer distribution was found to be controlled by varying the radial location of the impinging nozzle on the rotating disk. Interestingly, the maximum rotor- averaged heat transfer was found to be obtained when the impingement jet was located at the center of the rotating disk. At higher radial locations of the impingement jet, the local heat transfer coeffi cient in the vicinity of the impingement zone was seen to increase signifi cantly, by as much 80%, when approaching the outer rim. This work was carried further by Bunker et al. [24, 25] to examine not only the rotor heat transfer but also the associated stator heat transfer coeffi cient distributions, the latter due to secondary impingement on the stator face. While the local stator heat transfer was not as signifi cant as that on the rotor, the effects of impingement-rotational regimes and location still were prevalent.

2.8 Impingement in rotating cooling passages

As previously described, impingement cooling is typically utilized to provide effec-tive heat transfer in the leading and trailing edges of turbine blades. The cooling

Figure 11: The use of impingement (as shown by arrows) in disk cavities.



fl uid within rotating blade passages experience both Coriolis and centrifugal forces that may change the characteristics of the impinging jets. The blade lead-ing and trailing edge have the jets angled to target the high heat load regions, which can result in impingement jet angles of some relative orientation to the direction of rotation. Impingement may also be utilized within the mid-chord or central fl ow passages of the rotating blades, though such practice is as yet not a commonplace. The use of internal sheet metal impingement inserts like those in vanes has been attempted in the past, but vibrational issues have usually made this option unfeasible. Current advanced designs as may be found in high numbers within patent literature use the so-called near-wall cooling schemes involving confi ned impingement jets, such as that highlighted in Fig. 12. This topic will be covered in the next section.

There are relatively few studies on the impact of rotation on the impinging fl ow and the resulting heat transfer within rotating passages. Hsieh et al. [26] performed measurements on a single round jet within a square channel with impingement in-line with the axis of rotation (e.g. a blade leading edge model). They developed correlations for average Nusselt numbers for inward radial fl ow and outward radial fl ow based on jet Reynolds number and rotational Reynolds numbers. Reduction in averaged Nusselt number of 15% and 25% was observed for the outward wall and the inward wall, respectively, as rotational effects

Figure 12: An example of confi ned near-wall impingement (from EU Patent-1022432).



decreased the effectiveness of the impinging jets. In a much more relevant work, Mattern and Hennecke [27] studied the effect of rotation on an in-line row of impinging jets to a concave surface where the angle between the jet axis and rotational axis was varied. In this study, the rotational effect was seen to reduce the impingement heat transfer coeffi cients by as much as 40% locally as com-pared with a channel with no rotation. Strong interaction effects with the Coriolis force could be discerned here as heat transfer for both in-line and transverse oriented jets was most severely reduced, while that for 45° oriented impingement was least affected.

Parsons et al. [28] performed heat transfer measurements on two impinged surfaces, representative of the internal pressure side and suction side walls in a near-wall cooled blade mid-chord passage. Two rows of in-line impingement holes were used on each surface of the rotating channel model with coolant cross-fl ow and exhaust radially outward. Impingement was orthogonal to the axis of rotation. Rotation resulted in a 20% reduction in surface average heat transfer coeffi cients on all channel walls as compared with the results for similar channels with no rota-tion. Similar to non-impingement cooled rotating passages, there were also differ-ences between the rotationally leading (suction side) and trailing wall passage (pressure side) heat transfer. The trailing wall heat transfer coeffi cients were as much as 15% lower than those on the leading wall. The heat transfer on the impingement target surfaces and the jet issuing surfaces tended to be roughly equal in a given passage.

2.9 Confined channel impingement

While all internal cooling passages are spatially confi ned within vanes and blades, the design of airfoil cooling has recently been making advances toward moving the cooled surfaces closer to the heat source. One way to accomplish this is by casting smaller cooling passages inside the walls of the airfoils. Such designs are known as double-walled cooling, or near-wall cooling, and take advantage of the full inte-rior passage surface as signifi cant to the thermal management of the airfoil. An example of cast wall channels with impingement is shown in Fig. 12. Gillespie et al. [29] looked at the heat transfer from a double row of staggered impinge-ment jets in a channel, with fi lm cooling extraction holes, typical of an integrally cast impingement channel in a turbine airfoil. The resulting confi ned impinge-ment channel had a width of just seven and a half impingement hole diameters. Impingement in the confi ned channel, combined with the locally accelerating fl ow due to the fi lm holes, resulted in heat transfer enhancement on the surrounding walls of the channel. The recirculating fl ow generated in the tight passage resulted in heat transfer coeffi cient levels on the impingement jet wall as high as 50% of the levels measured on the target surface. This is sometimes referred to as second-ary impingement and in such confi ned confi gurations can be quite signifi cant. Son et al. [30] studied the impingement fl ow in a confi ned channel with a staggered array of jets alternating in the streamwise direction and accumulating in a channel



cross-fl ow (no fi lm extraction). The heat transfer performance between the jets had values in the order of 30%–50% compared with the stagnation regions, attributed to the mixing interaction of the adjacent jets and the inhibition of spanwise fl ow dispersion. In a companion work, Son et al. [31] measured the heat transfer coef-fi cients due to the secondary impingement to be as much as 50% of the average heat transfer on the target surface.

2.10 Impingement onto randomly rough and textured surfaces

Although it has been noted that very high convective heat transfer coeffi cients can be achieved through the use of impinging jets, further augmentation of the heat transfer coeffi cient can be achieved via the use of a roughened target surface. Chakroun et al. [32] examined the effect of a single-jet impinging onto a surface with small cube-shaped elements intended to model roughness. The roughened surface caused increases in heat transfer coeffi cient levels of 8%–28% relative to a smooth impingement surface. El-Gabry and Kaminski [33] studied the heat trans-fer augmentation of an array of jets impinging on a roughened surface formed by brazing small metal particles in a close-packed format. The surface-averaged Nu values show that a roughened surface produced signifi cantly higher impingement heat transfer coeffi cients relative to a smooth surface at all Reynolds numbers and impingement angles tested. Similar results were noted by Bunker and Bailey [34] for an impinging jet on a randomly roughened surface applied to the underside of a blade platform. An example of each roughened surface is shown in Fig. 13. It is evident that the addition of surface roughness to the target surface can be an invaluable tool for the gas turbine designer looking for increased internal heat transfer coeffi cients, most specially for localized effects.

Several forms of patterned surface augmentation, more amenable to investment casting, have also received increasing attention for use under impingement jets.

Figure 13: Rough surfaces for impingement augmentation (a) close-packed [33] and (b) random particles [34]; roughness average (RA) level of about 30 µm (reproduced with the permission of ASME).



These include pin arrays in the form of discrete-shaped bumps, regularly spaced turbulators, and even arrays of concavities (also known as dimples). The intent of these surface augmentation methods is to derive a greater benefi t than simple wet-ted surface area increase by also creating additional vortices and turbulence. Gau and Lee [35] studied various forms of two-dimensional jet impingement on rib-roughened surfaces showing that with proper relative geometry control, heat fl ux can be signifi cantly increased by as much as 100%. The addition of rib features signifi cantly altered the fl ow structure and resulting heat transfer compared with a smooth target plate. However, the addition of ribs also promoted the formation of fl ow recirculation zones, which under some conditions prevented the direct impingement onto the target surface and thus resulted in lower heat transfer levels. Haiping et al. [36] also carried out an experimental study of heat transfer from an array of impinging jets on a rib-roughened surface concluding that other general fi ndings for jet array heat transfer behavior with respect to target distance, array spacing, and Reynolds number applied to the rib-roughened surface. Son et al. [37] measured the heat transfer on a surface populated with protruding elements. They showed a higher row-averaged heat transfer for the protruding elements with lower pressure drop as compared with conventional fl ow over turbulators in a duct. These tests were carried out in a confi ned channel with alternating impingement rows of two and three impingement jets. The high heat transfer impingement zones seen on a typical smooth surface were preserved with 40% blockage of cylindrical protruding elements with a pressure penalty of only 2% over the smooth case. Taslim et al. [38] investigated conical bumps to augment airfoil leading edge impingement cooling. Both Son et al. [37] and Taslim et al. [38] found heat trans-fer to be augmented by roughly an amount equal to the increased surface area ratio factor. Annerfeldt et al. [39] examined the heat transfer enhancement from four different enlarging geometries (triangles, wings, cylinders, and dashed ribs) in an impinging fl ow. Some geometries shows enhancement levels of up to 30% com-pared with impingement on a fl at plate. Kanokjaruvijit and Martinez-Botas [40] tested several confi gurations of impingement jet array parameters with dimpled target surfaces of various forms. They found that shallow dimples could provide as much as 50% improved heat fl ux (including area factor), whereas deeper dimples could actually serve to create poor recirculation zones. Of great signifi cance throughout these studies was the fi nding that pressure losses were negligibly affected by the use of such surface methods with impingement fl ows.

2.11 Blade tip internal cooling

A very common form of blade tip internal cooling is that of serpentine passages creating a 180° turn region under the tip cap. The present state of the art for cool-ing of blade internal tip turn regions, and the tip cap surface of the 180° turn spe-cifi cally, consists of the use of smooth (cast or machined) internal surfaces that are naturally augmented by the enhanced heat transfer coeffi cients due to the three-dimensional fl ow turning and pseudo-impingement present. The effect of turning



fl ow-induced secondary fl ows in the tip turn regions serves to somewhat lessen the natural cooling augmentation, due to the radial infl ow motion of the secondary fl ow. These secondary fl ows were detailed through particle imaging velocimetry measurements in the study of Schabacker et al. [41]. Their measurements at Reyn-olds number of 50,000 in a sharp 180° turn showed weak recirculating fl ows in the upstream corners of the turn, strong impingement fl ows over the initial and mid-turn regions, followed by a very dominant dual vortex pair generated in the downstream portion of the turn. The tip cap heat transfer can be expected to follow these fl ow features in its distribution.

In serpentine blades, the 180° turns under the blade tip are typically treated as surfaces of relatively constant heat transfer coeffi cient of the order of 2 to 3 times that of fully developed turbulent duct fl ow at the same Reynolds number. In the study of Han et al. [42], mass transfer data were obtained for smooth surface tip turn surfaces with both smooth and turbulated inlet/exit channels. They found that the all-smooth case with a Re number of 30,000 to have an average tip heat trans-fer augmentation factor of 1.8. A factor of 2.5 was found when the inlet/exit chan-nels were turbulated, thereby raising the entire level of augmentation for the tip surface internal heat transfer. Strong centrifugal forces, in addition to the generally high passage Reynolds number, overcome the secondary fl ow effects and lead to heat transfer coeffi cients well above those of fully developed, stationary smooth duct fl ow. As such, rotational effects are expected to have little or no detrimental infl uence on the tip cap surface heat transfer. Mochizuki et al. [43] showed that rotational effects do not diminish the heat transfer coeffi cient enhancements on the internal tip cap surface for all smooth conditions. Wagner et al. [44] tested a multi-pass serpentine channel with entirely smooth walls, fi nding that the tip turn surface average heat transfer coeffi cient was augmented by about 50%. However, Wagner et al. [45] performed a similar study with repeated normal turbulators prior to the tip turn and found the average tip turn surface heat transfer coeffi cient augmenta-tion to be only about 10%. One may conclude that rotational effects will not degrade the tip surface heat transfer compared with stationary data, and in fact might further augment cooling.

Heat transfer coeffi cient augmentation for tip turn internal surfaces depends on other factors as well, including the channel aspect ratio, the transition of channel shape through the turn, Reynolds number, internal divider spacing to the tip sur-face, the specifi c turbulation form of the inlet/exit channels, and also any other surface augmentation used through the turn region. One such tip cap augmentation proposed by Bunker [46] consists of arrays of discrete-shaped pins on the internal tip cap surface designed to accommodate a mixture of impingement-like fl ow, channel fl ow, and strong secondary fl ows. The physics of the augmentation method combine two mechanisms. First, the short pin height-to-diameter ratio of two or less assures that the majority of the pin and fi llet surface areas are effective as heat transfer wetted area. Secondly, the placement of such pins on the turning surface represented by the tip cap allows a combination of impingement and cross-fl ow convection over the pins, which generate fl ow mixing and turbulence on the local



level and as interactions in the array format. This specifi c fl ow–surface interaction serves to at least partially disrupt the secondary fl ows that otherwise might decrease heat transfer on this surface. As shown in Fig. 14, effective heat transfer coeffi -cients based on the original smooth surface area were increased by up to a factor of 2.5. Most of this increase is due to the added surface area of the pin array. Factoring out this surface area shows that the heat transfer coeffi cient has been increased by about 20%–30%, primarily over the base region of the tip cap itself. This augmentation method resulted in negligible increase in tip turn pressure drop over that of a smooth surface.

2.12 Combustor cooling

The cooling of combustion systems is generally divided into two distinct formats. Aviation engine combustors rely heavily on fi lm cooling due to the very high fl ame temperatures and the compact design of the systems. Power- generating gas turbine engines must meet far more stringent emissions requirements (e.g. <25 ppm NOx) and so the cooling of these combustors is achieved primarily through the use of convective backside heat transfer, with little or no injection of coolant into the hot gas path as fi lm cooling. Figure 15 shows a typical can-annular type combustor system for an industrial gas turbine. Cooling is through a reverse fl ow design in which the compressor discharge air, minus the cooling air used in the turbine, is split into two paths, one to impingement cool the transition piece and the other to cool the combustor liner. Given the high levels of fl ow required to perform this cooling, the pressure drop allocated to the combustion system is an important

Figure 14: Nusselt number impingement augmentation of a blade tip cap (repro-duced from [46] with the permission of ASME).



factor. A typical combustion system may use up to 7% of the available pressure from the compressor. In round numbers, about two-thirds of this pressure drop is used in the cooling of the liner and transition piece, and the remaining one-third is used in the combustion mixing and reaction processes.

The cooling capability for the combustor liners impacts the liner life, both in terms of bulk temperatures and thermal gradient induced stresses. Methods for the cooling of combustor liners are of as many forms as there are liner designs. An excellent summary of combustor liner cooling via fi lm cooling methods, as well as combined fi lm and convective cooling methods, is presented by Schulz [47]. In the example of Fig. 15, roughly 50% of the air cools the transition piece by arrays of impingement jets. Since the accumulated fl ow rate at downstream locations is very high, these jet arrays employ both jet size and spacing varia-tions to maintain effective impingement without large cross-fl ow degradation as depicted in the lower right inset fi gure from Bunker et al. [48]. The jet Reynolds numbers may range from 50,000 to 150,000, which is considerably higher than most impingement cooling applications found in the turbine. The combustor liner is cooled initially by a combination of the spent cooling fl ow from the transition piece and the other 50% of air introduced as more impingement jets. These rows of impinging jets must penetrate an even higher cross-fl ow and so they have Reynolds numbers ranging up to as much as 300,000. The remainder of the combustor liner is cooled by convection over augmented surfaces (e.g. tur-bulated) to maintain system pressure losses within limits. The study of Bailey et al. [49] explored heat transfer distributions for a combustor liner model of this type, resulting in the heat transfer coeffi cient distributions shown in Fig. 15

Figure 15: Can-annular combustor system impingement cooling and typical axial heat transfer coeffi cient distribution.



(lab conditions). The liner impingement cooling is clearly evident, though the overall lower minimum-to-maximum variation than many turbine impingement cooling scenarios.

2.13 Closed-circuit impingement cooling

In a manner of speaking, the transition piece and combustor liner cooling system displayed is a localized closed circuit in that no coolant is released along the cooling path; it all fl ows to the combustor head end for use in pre-mixed combus-tion. The impingement jet cooling must be designed to accommodate all local effects of accumulated cross-fl ow, both in terms of jet defl ection and mixing. Such closed-circuit cooling can also be found in the turbine components, both the vanes and the blades, of some engine designs. In the most extreme case, closed-circuit cooling may employ a cooling fl uid that is contained in a system entirely separate from the gas turbine. An example of this is the H-System™ [50] combined cycle power plant that utilizes super-heated steam from the bottoming cycle to cool the gas turbine. In this case, impingement cooling is designed with not only variably sized jet diameters and spacing but also uses shaped impingement inserts. The airfoil shown in Fig. 16 from Cunha [51] has inserts that are shaped to allow the spent impingement cooling fl uid to accumulate in side channels for return. This mechanism avoids cross-fl ow degradation effects, excessive pressure loss, and shields the return coolant from further heat pick up. The example impinge-ment heat transfer coeffi cient distribution on a smooth surface shows a designed impingement array that maintains a fairly even average distribution along the

Figure 16: Use of impingement within closed-circuit cooling [48, 51]; color heat transfer coeffi cient distributions show the effect of texturing the surface.



mean fl ow direction but has the typical minimums and maximums of impinge-ment. The other distribution shown here uses impingement onto a textured surface (discussed further in a later section), which aids in smoothing out the heat transfer coeffi cient gradients. The plot of heat transfer coeffi cient augmentation factors and min-to-max ratios versus cross-fl ow mass velocity ratio from [48] shows that very high impingement augmentation (relative to fully developed turbulent duct fl ow) can be maintained over a wide range of cross-fl ow strength and resulting local gradients.

2.14 Impingement in film cooling

Conventionally, the only connection between impingement cooling and fi lm cool-ing comes in the manner of how post-impingement internal fl ow enters fi lm holes, thereby infl uencing the development of fi lm cooling at its initiation. Another means by design is to utilize impingement cooling to deliberately reduce the cool-ant pressure prior to entering fi lm holes, thereby decreasing the fi lm blowing ratio to a regime of improved behavior. Impingement is not normally associated directly with the external fi lm cooling confi gurations. However, many recent investigations have been conducted concerning the fi lm effectiveness for various forms of shal-low trenches. Results have been varied, but generally show positive conclusions with at least equivalent effectiveness to that of diffuser shaped holes. Figure 17 shows one variation on such a shallow trench as described by Bunker [52]. In this geometry, the shallow trench of Bunker [53] is provided diffuser-like reliefs in the regions between the fi lm hole exits. The concept is to allow each fi lm jet to impinge on the downstream edge of the trench, forcing lateral coolant fl ow inside the trench, which then exits gracefully onto the fl ow path surface. As reported by Bunker [54], this concept was tested in a high-speed vane cascade and compared against diffuser-shaped fi lm holes of the same size, number, fl ow rate, location, blowing ratio, and density ratio. The infrared imaging data in Fig. 17 show an

Figure 17: Use of impingement in fi lm cooling to affect lateral spreading of cool-ant [52, 54]; raw infrared image (right) shows surface temperature dis-tribution, for example, of same geometry.



example of surface temperature distribution indicating a very uniform cooling condition along the entire fi lm row. This geometry also results in about a 10% increase in the adiabatic fi lm effectiveness for the entire downstream surface of the airfoil.

3 Limitations

In practice, the design and application of impingement cooling within a gas turbine encounters many factors that limit or constrain its full potential. Broadly speaking, these limitations fall into the three main categories of manufacturing, operation, and thermal management. Constraints, however, should not always be viewed in a negative connotation. For example, while certain factors may restrict an applica-tion to the use of a fairly narrow range of jet Reynolds numbers, the result may be benefi cial from the perspective of decreased variability in operation. What is most important in all elements of constraint is that designers have a thorough knowl-edge of the causes, effects, accuracy, and variability.

Manufacturing limitations encompass both engineering and non-engineering factors. On the surface of the issue, either a desired impingement hole or array of holes can be manufactured or it cannot. In truth, with respect to conventional and foreseen impingement cooling methods, virtually any desired arrangement can be manufactured, but with what cost and with what resulting component integ-rity. Looking at the example of a double-walled cooled HPT blade of Lee et al. [55] shown in Fig. 18, there are several possible ways to manufacture the mid-chord impingement chambers including (1) investment casting using tiny quartz rods held between ceramic core elements to maintain the impingement holes during metal pour, (2) electro-discharge machining of the internal holes using a tool entry from the blade tip with later welding of the tip cap, and (3) casting the blade without external walls in the desired regions such that the impingement holes can be machined easily, followed by bonding (e.g. braze) of the external wall sections. These approaches are all feasible, but they will not result in the same yield of fi nal acceptable blades (cost), nor will the resulting blades have the same mechanical strength or temperature capability (integrity), nor will the resulting impingement holes all conform to the same required geometry for cool-ing purposes.

This example raises other manufacturing issues as well. A selected manufactur-ing approach to producing impingement holes or orifi ces will result in a specifi c characteristic cross-section, internal surface condition, and entry and exit effects. Orifi ces machined by laser will be very different from those made by electro- discharge methods, and so the impingement holes discharge coeffi cients, or fl ow coeffi cients, will also differ. This directly impacts the mass fl ow rate through each hole, as well as the cooling fl ow balance within a component. Furthermore, for impingement applications where the jet holes are manufactured by investment casting (e.g. leading and trailing edge crossover holes), the effective length-to-diameter ratio plays a role in determining the resultant directionality of the jets. If



the metering length is very short, and the inlet and exit rounded as demanded by casting, then the impingement jet may have a severely reduced ability to form a jet core with defi nite direction. The fact that a hole exists is not suffi cient to produce the desired impingement cooling.

The discussion above provides only a small sample of perhaps the largest man-ufacturing limitation that being the resulting variability in impingement geometry. Variability affects dimensions and placement within the component. But with dif-fering manufacturing methods, variability affects not only the impingement jet orifi ces but also the coolant supply chambers and the impingement targets. These sources of variability must then be characterized in two respects, the tolerances allowed by the manufacturing and design specifi cations, and the effect of these tolerances on coolant fl ow and ultimately the thermal–mechanical integrity of the component in operation. A summary of how these manufacturing tolerances may affect cooling can be found in Bunker [56]. Typical impingement applications are shown to result in ± 20°C variability in blade bulk metal temperature, which in

Figure 18: Manufacturing impingement in a fabricated double-walled cooled high-pressure turbine blade [55].



some cases could be half the life of the component. Fortunately, such large effects are generally unlikely to result from a single non-conforming impingement jet unless it is in a sensitive region such as the leading edge. Such effects take on even more prominence when the part-to-part and engine-to-engine variations are also considered, or when the effects of impingement variability in the combustion sec-tion are translated into the HPT section through, for example, alteration of the combustor exit gas temperature profi les.

While manufacturing limitations impact what can be designed and fabricated, operational limitations impact how and where impingement cooling can be uti-lized. The HPT vane and blade provide ample examples of the operational con-straints. One of the most restrictive requirements placed on the HPT airfoils is that of back-fl ow margin (BFM). BFM refers to the condition that no hot gases shall be allowed to fl ow into the airfoils through any surface holes, or even through any location that is postulated to experience a through-wall crack. Margin means that an excess of available pressure exists at each point in the cooling circuit to assure that coolant always fl ows outward while accounting for the aforementioned toler-ances and variability. Since impingement typically utilizes pressure ratios between 1.005 and 1.05 to drive the required jet Reynolds numbers, impingement cooling can become infeasible in some circumstances because it will violate BFM. This is not the only concern over pressure loss, since any reduction in pressure means a loss of available power output from the engine. Hence, impingement cooling can become expensive with respect to engine effi ciency. Along similar lines of reason-ing, impingement cooling demands a certain amount of airfoil volume to be geo-metrically feasible. Altering the vane or blade aerodynamic shape to accommodate impingement cooling could lead to a non-advantageous loss of turbine effi ciency despite the cooling advantages.

The integrity of the components comes into play here as well. Consider a rotat-ing blade that uses internal impingement realized with an inserted sheet metal impingement baffl e. Under vibrational conditions, while the blade may be well cooled, such a baffl e could experience high cycle fatigue failure, or worse yet induce failure of the structural airfoil holding it. Whether in rotating or stationary applications, impingement orifi ces are also subject to potential partial or total plugging due to dust, metal oxides shed within the engine, environmental constitu-ents such as salt over oceans or volcanic ash, and even foreign object ingestion to the engine such as ground up bird bones. Impingement holes are, therefore, usu-ally required to be of a certain minimum effective diameter or greater to avoid plugging, including an allowance for manufacturing tolerances.

Since the purpose of impingement cooling is to manage the thermal conditions and responses of the engine, it is important to note that greater impingement cool-ing magnitude is frequently not the correct design choice. The two thermal condi-tions that most affect the life of a HGP component are the bulk metal temperature and the local thermal stresses. A high bulk temperature will usually lead to early failure of the part, whereas high thermal stress will typically lead to crack initia-tion, propagation, and failure over time (cycles). Increasing cooling fl ow may reduce a high bulk temperature, but may actually exacerbate a thermal stress issue.



Impingement by its very nature can create very high local thermal gradients in the material as implied by the heat transfer coeffi cient distributions of Fig. 5. In extreme cases, the footprint of internal impingement jets can be seen in the exter-nal surface of a part due to thermal discoloration and even variations in the chem-istry of debris that adheres to the surface. Closer jet-to-jet array spacing can reduce the local thermal gradients, but at a cost of higher cooling fl ow rate. Other mitiga-tion strategies might be employed, as described in earlier sections, to reduce local gradients, such as the use of surface roughness to augment heat transfer coeffi -cients about a lesser initial magnitude, or angling of the impingement jets to spread out the coolant.

Thermal gradients are of concern not only from the in-plane stresses but also due to the through-wall stress. Impingement cooling can also be damaging simply by creating a very high local temperature gradient through the materials leading to low cycle fatigue issues, or even strain concerns in the protective coatings. Both in-plane and through-wall thermal gradients become increased concern when local conditions interact with other thermal sinks or sources within the components. A prime example is the interaction of the cold internal structural rib where it con-nects to the airfoil outer wall with the thermal gradients caused by impingement cooling. The combination can lead to magnifi ed stresses in regions where other stress concentrations exist, in this case a fi llet, and again cracking. Another example that is quite common is the interaction of impingement cooling with high internal cooling inside fi lm holes, which also have local geometric stress concen-trations, leading to fi lm hole cracking. As a consequence, both direct and indirect impingement cooling effects must be considered, requiring the knowledge of the full interior heat transfer coeffi cient distribution.

Lastly, the effects of coolant temperature rise due to impingement must be con-sidered as a real limitation. Impingement can result in very high heat transfer coef-fi cients and this also means the potential for very substantial relative increases to the coolant temperature depending on the coolant-to-wall temperature ratio condi-tions present. Impingement arrays must be designed to account for coolant heat up in the post-impingement regions, both locally as well as in later portions of the component cooling circuit, or as fi lm cooling. In this respect, there can be a syner-gistic relationship between internal impingement cooling and external fi lm cool-ing, where the latter can help to decrease the heat fl ux burden over post-impingement regions that have reduced cooling potential due to coolant temperature rise. In many applications, coolant temperature rise is a deliberate part of the design, for example, in the use of either ‘cold bridge’ or ‘warm bridge’ leading and/or trailing edge cooling in airfoils.

4 Summary

This chapter has presented the basics of gas turbine impingement cooling design, the various applications within the engine, and the typical limitations imposed upon its use. The discussions and design examples are by no means a compre-hensive review, but do provide the generic uses and applications impingement



jets for the hot gas fl ow path components of gas turbines. Specifi c targeted appli-cations have been presented for airfoils and combustor systems, as well as several emerging applications. Several important effects infl uencing impingement cooling have been discussed including surface roughness, rotation, and accumulated cool-ing cross-fl ow degradation. The potential design space for impingement cooling in gas turbines is shown to be immense, with the complexity of impingement cooling coming from the vast number of specifi c jet and surface geometry combinations that are involved with the actual components. Impingement cooling is a very fl ex-ible method capable of both localized and distributed cooling effectiveness, yet must also be made to conform to applications that are dictated by other design requirements, such as those from aerodynamics, mechanical structure, vibra-tions, low cycle fatigue, and creep rupture limits. The main limitations placed on impingement cooling are those due to available coolant pressure, component manufacturing methods, turbine operation, and system thermal management. In the end, impingement cooling must reach compromises that balance the desires of effective cooling with the turbine engine cost and life management.

Nomenclature

Cf Coeffi cient of frictionCfo Normalizing Cf of turbulent fl ow in a smooth ductD, d Jet diameter at point of issueGc Cross-fl ow mass velocityGj Jet mass velocity at point of issueNu Nusselt numberNuo Normalizing Nu of turbulent fl ow in a smooth ductPr Prandtl numberRe, ReD, Rej Jet Reynolds number based on diameterX, x Axial spacing between jets in arrayY, y Lateral spacing between jets in arrayZ, z Target distance from jet issue to impingement surface

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[24] Bunker, R.S., Metzger, D.E. & Wittig, S., Local heat transfer in turbine disk cavities: part I - rotor and stator cooling with hub injection of coolant. Jour-nal of Turbomachinery, 114(1), pp. 211–220, 1992.

[25] Bunker, R.S., Metzger, D.E. & Wittig, S., Local heat transfer in turbine disk cavities: part II - rotor cooling with radial local injection of coolant. Journal of Turbomachinery, 114(1), pp. 221–228, 1992.

[26] Hsieh, S.S., Huang, J.T. & Liu, C.F., Local heat transfer in a rotating square channel with jet impingement. ASME Journal of Heat Transfer, 121, pp. 811–818, 1999.

[27] Mattern, C. & Hennecke, D.K., The infl uence of rotation on impingement cooling. ASME Turbo Expo Conference Paper 96-GT-161, Birmingham, UK, 1996.

[28] Parsons, J.A., Han, J.C. & Lee, C.P., Rotation effect on jet impingement heat transfer in smooth rectangular channels with four heated walls and radi-ally outward crossfl ow. ASME Turbo Expo Conference Paper 96-GT-387, Birmingham, UK, 1996.

[29] Gillespie, D.R.H., Wang, Z., Ireland, P.T. & Kohler, S.T, Full surface local heat transfer coeffi cient measurements in a model of an integrally case impingement cooling geometry. ASME Paper No. 96-GT-200, 1996.

[30] Son, C., Gillespie, D. & Ireland, P., Heat transfer and fl ow characteristics of an engine representative impingement cooling system. ASME Paper No. 2000-GT-219, 2000.

[31] Son, C., Gillespie, D. & Ireland, P., Heat transfer characteristics of an impingement plate used in a turbine vane cooling system. ASME Paper No. 2001-GT-154, 2000.

[32] Chakroun, W.M., Al-Fahed, S.F. & Abdel-Rehman, A.A., Heat transfer aug-mentation for air jet impinged on rough surface. 97-GT-436, IGTI Turbo Expo, Orlando, 1997.

[33] El-Gabry, L.A. & Kaminski, D.A., Experimental investigation of local heat transfer distribution on smooth and roughened surfaces under an array of angled impinging jets. Journal of Turbomachinery, 127, pp. 532–544, 2004.

[34] Bunker, R.S. & Bailey, J.C., Turbine blade platform impingement cooling with effects of fi lm extraction and roughness. Paper No. GT-2005-68415, IGTI Turbo Expo, Reno-Tahoe, Nevada, USA, 2005.



[35] Gau, C. & Lee, C.C., Impingement cooling fl ow structure and heat transfer along rib-roughened walls. International Journal of Heat and Mass Transfer, 35(11), pp. 3009–3020, 1992.

[36] Haiping, C., Zhang, J. & Huang, T., Experimental investigation on impinge-ment heat transfer from rib roughened surface within arrays of circular jets: effect of geometric parameters. ASME Paper No. 98-GT-208, 1998.

[37] Son, C., Gillespie, D., Ireland, P. & Dailey, G.M., Heat transfer enhancement strategy for an impingement cooling system. Proceedings of the 8th Inter-national Symposium on Transport Phenomena and Dynamics of Rotating Machinery, Hemisphere Pub. Co., Vol. 2, pp. 721–729, 2000.

[38] Taslim, M.E., Setayeshgar, L. & Spring, S.D., An experimental evaluation of advanced leading edge impingement cooling concepts. ASME Journal of Turbomachinery, 123, pp. 147–153, 2002.

[39] Annerfeldt, M.O., Persson, J.L. & Torisson, T., Experimental investigation of impingement cooling with turbulators or surface enlarging elements. ASME Paper No. 2001-GT-0149, 2001.

[40] Kanokjaruvijit, K. & Martinez-Botas, R.F., Heat transfer and pressure inves-tigation of dimple impingement. Paper No. GT-2005-68823, IGTI Turbo Expo, Reno-Tahoe, Nevada, USA, 2005.

[41] Schabacker, J., Boelcs, A. & Johnson, B.V., PIV investigation of the fl ow characteristics in an internal coolant passage with two ducts connected by a sharp 180-deg bend. Paper No. 98-GT-544, IGTI Turbo Expo, Stockholm, Sweden, 1998.

[42] Han, J.C., Chandra, P.R. & Lau, S.C., Local heat/mass transfer distributions around sharp 180-deg turns in two-pass smooth and rib-roughened channels. Journal of Heat Transfer, 110, pp. 91–98, 1988.

[43] Mochizuki, S., Takamura, J., Yamawaki, S. & Yang, W.J., Heat transfer in serpentine fl ow passages with rotation. Journal of Turbomachinery, 116, pp. 133–140, 1994.

[44] Wagner, J.H., Johnson, B.V. & Kopper, F.C., Heat transfer in rotating serpentine passages with smooth walls. Journal of Turbomachinery, 113, pp. 321–330, 1991.

[45] Wagner, J.H., Johnson, B.V., Graziani, R.A. & Yeh, F.C., Heat transfer in rotating serpentine passages with trips normal to the fl ow. Journal of Tur-bomachinery, 114, pp. 847–857, 1992.

[46] Bunker, R.S., The augmentation of internal blade tip-cap cooling by arrays of shaped pins. Journal of Turbomachinery, 130(4), 041007, 2008.

[47] Schulz, A., Combustor liner cooling technology in scope of reduced pollut-ant formation and rising thermal effi ciencies. Heat Transfer in Gas Turbine Systems, ed. R.J. Goldstein, Annals of the New York Academy of Sciences: New York, Vol. 934, pp. 135–146, 2001.

[48] Bunker, R.S., Bailey, J.C., Widener, S., Johnson, T. & Intile, J., Method and apparatus for cooling combustor liner and transition piece of a gas turbine. US Patent 7,493,767, February 24, 2009.



[49] Bailey, J.C., Intile, J., Tolpadi, A., Fric, T., Nirmalan, N.V. & Bunker, R.S., Experimental and numerical study of heat transfer in a gas turbine com-bustor liner. Journal of Engineering for Gas Turbines and Power, 125, pp. 994–1002, 2002.

[50] Layne, A.W. & Hoffman, P.A., Gas turbine systems for the 21st century. Paper No. 99-GT-367, ASME International Gas Turbine Conference, India-napolis, 1999.

[51] Cunha, F.J., Turbine stator vane segments having leading edge impingement cooling circuits. US Patent 5,762,471, 1998.

[52] Bunker, R.S., Film cooled slotted wall and method of making the same. US Patent 7,553,534, June 30, 2009.

[53] Bunker, R.S., Film cooling effectiveness due to discrete holes within trans-verse surface slots. GT-2002-30178, IGTI Turbo Expo, Amsterdam, 2002.

[54] Bunker, R.S., Gas turbine fi lm cooling: breaking the limits of diffusion shaped holes. Heat Transfer Research, 41(6), pp. 627–650, 2010.

[55] Lee, C.P., Isburgh, A.M., La Chapelle, D.G. & Wilson, P.S., Airfoil blade having a serpentine cooling circuit and impingement cooling. US Patent 5,660,524, 1997.

[56] Bunker, R.S., The effects of manufacturing tolerances on gas turbine cooling. ASME Journal of Turbomachinery, 131(4), 041018, 2008.


CHAPTER 1 Impingement Cooling in Gas Turbines: Design ......6 IMPINGEMENT JET COOLING IN GAS...

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Transcript of CHAPTER 1 Impingement Cooling in Gas Turbines: Design ......6 IMPINGEMENT JET COOLING IN GAS...