2013-j c han gt heat transfer review

8/17/2019 2013-J C Han GT Heat Transfer Review

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Je-Chin HanTurbine Heat Transfer Laboratory,

Department of Mechanical Engineering,

Texas A&M University,

College Station, TX 77843-3123

e-mail: [email protected]

Fundamental Gas Turbine Heat

Transfer

Gas turbines are used for aircraft propulsion and land-based power generation or industrial applications. Thermal efficiency and power output of gas turbines increasewith increasing turbine rotor inlet temperatures (RIT). Current advanced gas turbine

engines operate at turbine RIT (1700

C) far higher than the melting point of the bladematerial (1000 C); therefore, turbine blades are cooled by compressor discharge air (700 C). To design an efficient cooling system, it is a great need to increase the under-standing of gas turbine heat transfer behaviors within complex 3D high-turbulenceunsteady engine-flow environments. Moreover, recent research focuses on aircraft gasturbines operating at even higher RIT with limited cooling air and land-based gas tur-bines burn coal-gasified fuels with a higher heat load. It is important to understand and solve gas turbine heat transfer problems under new harsh working environments. Theadvanced cooling technology and durable thermal barrier coatings play critical roles for the development of advanced gas turbines with near zero emissions for safe and long-lifeoperation. This paper reviews fundamental gas turbine heat transfer research topics and documents important relevant papers for future research. [DOI: 10.1115/1.4023826]

Introduction

High Temperature Gas Turbines. Developments in turbinecooling technology play a critical role in increasing the thermalefficiency and power output of advanced gas turbines. To doublethe engine power in aircraft gas turbines, the rotor inlet tempera-ture should increase from today’s 1700 C to 2000 C (3100 F to3600 F) using the similar amount of cooling air (3–5% of com-pressor discharge air). For land-based power generation gasturbines, including power generation (300–400 MW combinedcycles), marine propulsion, and industrial applications such aspumping and cogeneration (less than 30 MW), the rotor inlet tem-perature will increase, but at a rate determined by NOx con-straints; with the emphasis on NOx reduction, efficient use of cooling air becomes more important in order to achieve cycle effi-ciency gains. Therefore, high-temperature material developmentsuch as thermal barrier coating (TBC) and highly sophisticatedadvanced cooling are two important issues that need to beaddressed to ensure high-performance, high-power gas turbinesfor both aircraft and land-based applications.

Gas Turbine Heat Transfer. Figure 1 shows a typical heatflux distribution on the surfaces of a turbine vane and blade andthe associated internal and external cooling schemes [1]. As theturbine inlet temperature increases, the heat transferred to the tur-bine blade also increases. The level and variation in the tempera-ture within the blade material, which cause thermal stresses andblade failures, must be limited to achieve reasonable durabilitygoals. Note that the blade life may be reduced by half if the blademetal temperature prediction is off by only 30 C (50 F). There-fore, it is critical to accurately predict the local heat transfer coef-ficient as well as the local blade temperature in order to preventlocal hot spots and increase turbine blade life. Meanwhile, there isa critical need to cool the blades for safe and long-life operation.

Gas Turbines Cooling System. The turbine blades are cooledwith extracted air from the compressor of the gas turbine engine.Since this extraction incurs a penalty on the thermal efficiencyand power output of the gas turbine engine, it is important to fullyunderstand and optimize the cooling technology for a given

turbine blade geometry under engine operating conditions. Gasturbine blades are cooled both internally and externally as shownin Fig. 2 [2]. Internal cooling is achieved by passing the coolantthrough several rib-turbulated serpentine passages inside of theblade. Both jet impingement and pin-fin cooling are also used as amethod of internal cooling. External cooling is also called filmcooling. Internal coolant air is ejected out through discrete holesto provide a coolant film to protect the outside surface of the bladefrom hot combustion gases. The engine cooling system must bedesigned to ensure that the maximum blade surface temperaturesand temperature gradients during operation are compatible withthe allowable blade thermal stress for the life of the design. Toolittle coolant flow results in hotter blade temperatures and reducedcomponent life. Similarly, too much coolant flow results in

reduced engine performance. The turbine cooling system must bedesigned to minimize the use of compressor discharge air for cooling purposes to achieve maximum benefits of the high turbineinlet gas temperature.

Advanced Hydrogen Gas Turbines. In addition to conven-tional natural gas, syngas or hydrogen (produced from coal gasifi-cation) are viable alternative fuels. These coal gasified fuelsproduce a high percentage of water vapor (steam) and increaseheavy heat load to the turbine components. These coal-based fuelsmay introduce impurities into the mainstream gas which can causecorrosion and erosion on the surface of the airfoils, or the impur-ities may be deposited on the components. All of these scenariosincrease the surface roughness of the components, increase heattransfer, increase aerodynamic losses, increase clogging of film

cooling holes, decrease film-cooling effectiveness, and decreasethe performance of the turbine. Recent research suggests that theadvanced cooling technology and durable thermal barrier coatingsstill play critical roles for the development of future coal-basedgas turbines with near zero emissions (for example, integratedgasification combined cycles (IGCC), thermal efficiency greater 60%).

Important Literature Survey. Research activities in turbineheat transfer and cooling began in the early 1970s; since then,many research papers, state-of-the-art review articles, and bookchapters have been published. Several publications are availablethat address state-of-the-art reviews of turbine blade cooling and

Manuscript received October 15, 2012; final manuscript received February 9,2013; published online May 17, 2013. Assoc. Editor: Srinath V. Ekkad.

Journal of Thermal Science and Engineering Applications JUNE 2013, Vol. 5 / 021007-1CopyrightVC 2013 by ASME

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heat transfer. These include publications by “Film Cooling” byGoldstein [3], “Turbine Cooling” by Suo [4], “Cooling Techni-ques for Gas Turbine Airfoils” by Metzger [5], “Some Considera-tions in the Thermal Design of Turbine Airfoil Cooling Systems”by Elovic and Koffel [6], and “Turbine Cooling and Heat Trans-fer” by Lakshminarayana [7]. Several review articles related togas turbine heat-transfer problems by Graham [8] and Simoneauand Simon [9] are also available.

Several Books Have Been Published Since 2000: Gas Turbine Heat Transfer and Cooling Technology by Han et al. [10], Heat Transfer in Gas Turbines edited by Sunden and Faghri [11], Heat Transfer in Gas Turbine Systems edited by Goldstein [12],Special Section: Turbine Science and Technology (included 10

review papers) edited by Shih [13], Heat Transfer in Gas TurbineSystems (included 10 keynote papers) edited by Simon and Gold-stein [14]. Meanwhile, many review papers related to gas turbineheat transfer and cooling problems are available: “ConvectionHeat Transfer and Aerodynamics in Axial Flow Turbines” byDunn [15], “Gas Turbine Heat Transfer: 10 Remaining Hot GasPath Challenges” by Bunker [16], “Gas Turbine Film Cooling” byBogard and Thole [17], “Turbine Blade Cooling Studies” at TexasA&M 1980–2004 by Han [18], “Turbine Cooling System Design-Past, Present and Future” by Downs and Landis [19], and“Aerothermal Challenges in Syngas Hydrogen-Fired and OxyfuelTurbines” by Chyu et al. [20].

The ASME Turbo Expo (IGTI International Gas Turbine Insti-tute) has made conference CDs available to every year’s attendees

Fig. 1 Cross-sectional view and heat flux distribution of a cooled vane and blade [1]

021007-2 / Vol. 5, JUNE 2013 Transactions of the ASME



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since 2000. These conference CDs contain all gas turbine heattransfer papers presented in each year’s IGTI conference. Thenumbers of heat transfer related conference papers have increasedfrom about 100 in the year 2000 to about 200 in the year 2010.Approximately 25–30% of each year’s conference heat transfer papers have been subsequently published in the ASME Journal of Turbomachinery. These tremendous amounts of conference and journal papers are the main research sources of gas turbine heattransfer and cooling technology for interested readers.

Gas Turbine Heat Transfer

Fundamentals of Gas Turbine Heat Transfer. Current tur-bine designs are characterized by an inability to accurately predictheat-transfer coefficient distributions under turbomachinery flowconditions. This results in a nonoptimized design using inordinateamounts of cooling air, which ultimately causes significant penal-ties to the cycle in terms of thrust and specific fuel consumption.Hot-gas path components include turbine stator vanes and turbinerotor blades. Turbine first-stage vanes (so called nozzle guidevane, NGV) are exposed to high-temperature, high turbulence hotgases from the exit of the combustor as sketched in Fig. 1 [1]. It isimportant to determine the heat load distributions on the first-stage vanes under engine flow conditions for a typical gas turbineengine. An accurate estimate of the heat-transfer distributions canhelp in designing an efficient cooling system and prevent local

hot-spot overheating. Gas to airfoil heat transfer can be affectedby airfoil shape (surface curvature and pressure gradient), bound-ary layer transitional behavior, free-stream turbulence, airfoil sur-face roughness, film coolant injection location, flow separationand reattachment, and shock/boundary layer interaction, exitMach number and Reynolds number.

After accelerating from the first-stage vanes, hot gases moveinto the first stage rotor blades to produce turbine power assketched in Fig. 1 [1]. At the inlet of the first-stage rotor blade,both the temperature and turbulence levels are lower compared tothe inlet of the first-stage vane. However, the inlet velocity couldbe two to three times higher. Besides, the blade receives unsteadywake flows from the upstream vane trailing edge (turbulent inten-sity up to 20%). More importantly, blade rotation causes hot gases

to leak from the pressure side through a tip gap to the suction side.This often causes damage on the blade tip near the pressure sidetrailing-edge region. It is important to understand the complex 3Dflow physics and associated heat-transfer distributions on the rotor blade, particularly near the tip region, under typical engine flowconditions. It is important to note that rotation causes the peak gastemperature to shift from the blade pitch line toward the tipregion. It is also important to correctly predict the RIT profile aswell as the associated unsteady velocity profile and turbulencelevels. Many papers were reviewed and cited in Chapter 2 of Hanet al. [10].

Heat Transfer Through Turbine Stator Vanes. For typicalNGV designs, heat-transfer coefficients on the pressure surfacedecrease rapidly from the leading edge to about 20% surface dis-tance and then gradually increase toward the trailing edge. Theheat-transfer distribution on the pressure surface is not affected bychanging the exit Mach number (Mach¼ 0.75–1.05). On the suc-tion surface, heat-transfer coefficient distributions show laminar boundary layer separation, transition, and turbulent reattachmentat about 25% surface distance. The location of the laminar bound-ary layer transition seems to move upstream with decreasing exitMach number, and downstream of that location, heat-transfer coefficients are higher with decreasing exit Mach numbers. Inregions where the boundary layer remained attached (before tran-sition), there is no apparent effect of the exit Mach number. More-

over, the transition location on the suction surface moves closer tothe leading edge with an increase in exit Reynolds number (Re¼ 1–1.5 106). Overall, heat-transfer coefficients over theentire airfoil surface showed significant increases with an increasein exit Reynolds number reported by Nealy et al. [21].

Turbulence and Roughness Effects. Combustor generated tur-bulence (high turbulence intensity up to 20% and large scale) con-tributes to significant heat transfer enhancement. Turbulence canstrongly affect laminar heat transfer to the stagnation region, pres-sure surface, transition, and turbulent boundary layer heat transfer by Ames [22]. Another severe heat-transfer enhancement factor for NGV heat transfer is the surface roughness effect. Combustiondeposits may make the vane surface rough after several hours of

Fig. 2 Gas turbine blade cooling schematic: (a ) film cooling, (b ) internal cooling [2]

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operation, and this roughness could be detrimental to the life of thevane due to increased heat-transfer levels that are much higher than design conditions, as well as decreased aerodynamicperformance cited by Abuaf et al. [23]. The already enhancedheat-transfer coefficients on the pressure surface due to high free-stream turbulence (Tu¼ 10%) are unaffected by the surface rough-ness. However, the effect on suction surface is significant. Acombination of surface roughness with high free-stream turbulencecauses the boundary layer to undergo transition more rapidly thanfor the high free-stream turbulence case only by Hoffs et al. [24].

Heat Transfer Through Turbine Rotor Blades. Figure 3(a)depicts a complex flow phenomenon in a turbine rotor hot gaspassage including secondary flows, horseshoe vortex, end wallpassage vortex, film cooling, tip flows, wakes, and rotation flows[25]. The heat-transfer distributions for not film-cooled blades arehigher than those for film-cooled blades at the same engine flowconditions. These heat-transfer distributions could differ for var-ied engine flow conditions; therefore, it is critical for a designer tobe able to accurately predict these distributions for film-cooled or no film-cooled blades in order to design an efficient coolingscheme. Most of the primary results on real rotor/stator heat trans-fer have been provided by Calspan Advanced Technology Center by Dunn et al. [26,27]. They used a full-stage rotating turbine of the Garrett TFE 731-2 engine with an aspect ratio of around 1.5,

as well as a full-stage rotating turbine of an Air Force/Garrettlow-aspect-ratio turbine with an aspect ratio of approximately 1.0.They reported heat flux measurements on the NGV airfoil and endwall, the rotor blade, blade tip, and shroud of the turbine. Ashock-tunnel facility was intended to provide well-defined flowconditions and duplicate sufficient number of parameters to

validate and improve confidence in design data and predictivetechniques under development.

Blair et al. [28,29] conducted experiments on a large-scale am-bient temperature, turbine-stage model. The turbine-stage modelconsisted of a stator, a rotor, and an additional stator behind therotor (11=2 stage). They also studied the effects of inlet turbulence,stator-rotor axial spacing, and relative circumferential spacing of the first and second stators on turbine airfoil heat transfer. Thistest facility was designed for conducting detailed experimentalinvestigations for flow around turbine blading. Guenette et al. [30]

presented local heat-transfer measurements for a fully scaled tran-sonic turbine blade. The measurements were performed in theMIT blow down turbine tunnel. The facility has been designed tosimulate the flow Reynolds number, Mach number, Prandtl num-ber, corrected speed and weight flow, and gas-to-metal tempera-ture ratios associated with turbine fluid mechanics and heattransfer. They used thin-film heat flux gauges to measure theupstream NGV trailing edge unsteady wake effect on the down-stream rotor blade surface time-dependent heat-transfer coeffi-cients. They found that, on the suction surface, the blade-passingeffect is stronger at the leading edge and attenuates toward thetrailing edge. The steep variations of the heat transfer enhance-ment on the suction surface indicate strong wake propagation to-ward the suction surface near the leading edge and then movingtoward the pressure surface near the trailing edge.

Effect of Upstream Unsteady Wake. The upstream vane-generated unsteady wake impinging on the downstream rotor issimulated using a stationary blade cascade and an upstream wakegenerator. Wake-simulation experiments typically used a rotatingspoked-wheel wake generator by Dullenkopf et al. [31] upstreamof the stationary blade cascade to simulate the relative motion of vane trailing edges. The relative motion of the rods on the wakegenerator creates unsteady wakes that impinge on the downstreamblade cascade. Guenette et al. [30] confirmed the validity of usingrotating-bar simulation. In general, unsteady wake promotes earlyboundary layer transition and increases heat load to suction sur-face as well as pressure surface of the turbine blade. Mayle [32]developed an intermittency model to predict laminar-turbulentboundary layer transition due to turbulence spots production ratefrom various experiments. The intermittency model was able to

predict the turbine blade suction-surface heat transfer enhance-ment due to unsteady wake flow conditions.

Heat Transfer Through Turbine Blade Tip. Turbine bladetip and near-tip regions are typically difficult to cool, and are sub- jected to potential damage due to high thermal loads (blade tipheat transfer coefficients are same order as leading edge stagna-tion region). Unshrouded blades have a gap existing between theblade tip and the shroud surface, which is known as tip gap. Theleakage flow accelerates due to a pressure difference betweenboth the pressure and suction sides of the blade, causing thinboundary layers, high heat transfer rates, and low turbine effi-ciency. It has been recognized that the blade tip geometry andsubsequent tip leakage flows have a significant effect on the aero-dynamic efficiency of turbines. A common technique to reduce

the tip leakage flow is to use a recessed tip, which is known as asquealer tip as shown in Fig. 4 [33]. A squealer tip allows asmaller tip clearance, without the risk of a catastrophic failure, incase the tip rubs against the shroud during turbine operation. Thetip recess also acts as a labyrinth seal to increase flow resistanceand reduce the tip leakage flow. Thus, it is important to under-stand both the flow and heat transfer behavior on the squealer tipof a gas turbine blade. Reliable experimental data are also impor-tant to develop and validate computational codes to predict flowand heat transfer distributions on turbine blades. Several paperspresented heat transfer results under engine representative main-stream flow conditions by Ameri et al. [34], Bunker et al. [35],Azad et al. [36] and Dunn and Haldeman [37]. Results show vari-ous regions of high and low heat transfer coefficient on the tip

Fig. 3 (a ) Typical film cooled airfoil [25] and (b ) end wall vorti-ces [39]




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surface; heat transfer coefficient increases with tip clearance andturbulence intensity. Azad et al. [36] compared squealer tip andplane tip geometry and concluded that the overall heat transfer coefficients were lower for squealer tip case. Kwak et al. [38]investigated the heat transfer on several different squealer geome-tries. They found that a suction side squealer tip gave the lowestheat transfer coefficient among all cases studied.

Heat Transfer Through Turbine End-Wall. Several studieshave shown how a horseshoe vortex develops at the leading edgeof the turbine vanes and blades, as shown in Fig. 3(b) [39]. Elimi-nating the formation of the horseshoe vortex at the leading edgeof the turbine blade will positively impact the performance of theengine. Adding a fillet at the junction of the airfoil and platformhas been shown to eliminate the vortex formed at the leading edgeof the blade by Sauer et al. [40]. Not only does the elimination of the horseshoe vortex decrease aerodynamic losses, but it also hasa positive impact on the end wall heat transfer and film cooling.Shih and Lin [41] predicted fillets not only reduce aerodynamiclosses, but the surface heat transfer is reduced by more than 10%on the airfoil surface and more than 30% on the vane end wall. Asecond method to mitigate the secondary flow along the end wallis to implement end wall contouring. Kopper et al. [42] deter-mined the secondary losses are reduced by up to 17% through apassage with a contoured end wall (compared to a flat end wall).Saha and Acharya [43] applied nonaxisymmetric profiling to theend wall of rotor blade cascade. Schobeiri and Lu [44] recentlyreported the efficiency of a three-stage rotating turbine can be

greatly improved by applying a physic-based diffuser-flow con-cept for optimizing nonaxisymmetric end wall contouring.

Thermal Barrier Coating and Spallation Effects. For safer operation, the turbine blades in current engines use nickel-based super alloys at metal temperatures well below 1000C(2000 F). For higher RIT, the advanced casting techniques,such as directionally solidified and single crystal blades with TBCcoating, are used for advanced gas turbines. TBC coating servesas insulation for the turbine airfoils and allows a 100–150C(200 F–300 F) higher RIT, thereby enhancing turbine effi-ciency. There are two types of coating techniques: (1) air plasmaspray with plate structure/porosity/low thermal conductivity and

(2) electron beam physical vapor deposition with column struc-ture/dense/high thermal conductivity by Nelson et al. [45]. Theperformance of TBC coatings, the zirconia-based ceramics,depends on the aforementioned coating techniques and the coatingthickness (5–50 mil). The United States government laboratories,gas turbine manufacturers, and university researchers have con-ducted research to identify better coating materials, better coatingtechniques, controllable coating thicknesses, good bonding coats,and hot corrosion tests for TBC life prediction. It is important todetermine the effects of TBC roughness and the potential TBC

spallation on turbine aerodynamic and heat-transfer performance.Ekkad and Han [46,47] studied the effect of simulated TBC spal-lation shape, size, and depth on heat transfer enhancement over aflat surface as well as on a cylindrical leading-edge model. Theyfound that the spallation can enhance the local heat transfer coeffi-cients up to two times as compared to that with the smoothsurface.

Deposition and Roughness Effects. Recent experimentalwork in measuring the formation of deposits has been done under the UTSR program by Bons et al. [48]. In a series of experimentsin an accelerated test facility, Wammack et al. [49] investigatedthe physical characteristics and evolution of surface deposition onbare polished metal, polished TBC with bond coat (initial averageroughness was less than 0.6 micrometers) and unpolished oxida-

tion resistant bond coat (initial average roughness around 16micrometers). Based on these results, they inferred that the initialsurface preparation has a significant effect on deposit growth, thatthermal cycling combined with particle deposition caused exten-sive TBC spallation while thermal cycling alone caused none, andfinally that the deposit penetration into the TBC is a significantcontributor to spallation. Subsequently, Bons et al. [48] made con-vective heat transfer measurements using scaled models of the de-posited roughness and found that the Stanton number wasaugmented by between 15 and 30% over a smooth surface. Theyconcluded that deposition increased by a factor of two as the massmean diameter of the particle was increased from 3–16 micro-meters. Second, particle deposition decreased with decreasing gastemperature and with increased backside cooling. They found athreshold gas temperature for deposition to occur at 960 C.

Gas Turbine Film Cooling

Fundamentals of Film Cooling. In turbine blade film cooling,as sketched in Fig. 2, relatively cool air is injected from the insideof the blade to the outside surface, which forms a protective layer between the blade surface and hot mainstream. Film coolingdepends primarily on the coolant-to-hot-mainstream pressure ratio( pc/pt), temperature ratio (Tc / Tg), and the film-cooling-hole loca-tion, configuration, and distribution on a film-cooled airfoil.The coolant-to mainstream pressure ratio can be related to thecoolant-to-mainstream mass flux ratio (blowing ratio), while thecoolant-to-mainstream temperature ratio can be related to the cool-ant-to-mainstream density ratio. In a typical gas turbine airfoil, the pc/pt ratios vary from 1.02 to 1.10, while the corresponding blowing

ratios vary approximately from 0.5 to 2.0. Whereas the Tc / Tg val-ues vary from 0.5 to 0.85, the corresponding density ratios varyapproximately from 2.0 to 1.5. In general, the higher the pressureratio, the better the film-cooling protection (i.e., reduced heattransfer to the airfoil) at a given temperature ratio, while the lower the temperature ratio, the better the film-cooling protection at agiven pressure ratio. However, a too high pressure ratio (i.e.,blowing too much) may reduce the film-cooling protectionbecause of jet penetration into the mainstream (jet liftoff from thesurface). Data from numerous studies available in open literaturesuggest a blowing ratio near unity is optimum, with severe penal-ties at either side. As mentioned earlier, turbine-cooling systemdesigners need to know where the heat is transferred from thehot mainstream to the airfoil (Fig. 1) in order to design better

Fig. 4 Typical gas turbine blade squealer tip cooling configu-

ration [33]




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film-cooling patterns for airfoils. These film-hole patterns (i.e.,film-hole location, distribution, angle, and shape) affect film-cooling performance. The best film cooling design is to reduce theheat load to the airfoils using a minimum amount of cooling air from compressors. Many papers were reviewed and cited in Chap-ter 3 of Han et al. [10].

Flat Plate Film Cooling. It is common in literature to use aflat plate to perform fundamental studies on various parametriceffects on film cooling such as the study done by Goldstein [3].

Moreover, results on flat plates have been used to calibrate andstandardize various experimental techniques to measure film cool-ing effectiveness and heat transfer coefficients. While the bestfilm cooling coverage can be obtained by injecting the fluid paral-lel to the mainstream, manufacturing constraints dictate that holesbe angled. Using film cooling holes perpendicular (90 deg) to themainstream, results in very low film cooling effectiveness. Theuse of holes inclined at 35 deg typically gives a balance betweenfilm cooling performance and manufacturing ease.

Effect of Coolant-Mainstream Blowing Ratio and Density Ratio. Blowing ratio ( M ) is defined as the ratio of the coolantmass flux to that of the mainstream. In general, regardless of hole-shape and angle, film cooling effectiveness is found to increasewith blowing ratio at low blowing ratios (less than 0.5). However,beyond a critical blowing ratio, film cooling effectiveness is found

to decline. This decline can be attributed to the phenomenon of film-cooling lift-off from boundary layers, wherein the high mo-mentum film-cooling jet fails to attach with the plate surface andpenetrates into the mainstream reported by Goldstein et al. [50]and Pedersen et al. [51]. The coolant to mainstream density ratio(DR) in modern gas-turbine engines is typically around 2.0 due tothe coolant temperature being significantly lower than hot main-stream. Scaled down laboratory tests (to simulate engine DR con-ditions) usually involve chilling the coolant to very lowtemperatures by Sinha et al. [52] or using a foreign gas with ahigher density by Goldstein et al. [50], Pedersen et al. [51], andEkkad et al. [53]. In general, increasing DR at a given M results ina higher effectiveness, especially at higher blowing ratios, sincethe momentum of a high density coolant is lower at a given M ,there is a lower tendency to lift-off.

Effect of Hole Exit Shape and Geometry. Injecting the filmcoolant at an angle to the mainstream (a compound angle), resultsin higher film cooling effectiveness due to greater lateral diffusionof the coolant. Compound angled configurations are also found toresist to lift-off more than simple angled configurations by Ekkadet al. [53]. Using shaped film cooling holes (with a fan shapeddiffuser on the blade surface) results in a lower tendency to liftoff due to the reduction in momentum due to the increase cross-sectional area for the coolant by Goldstein et al. [50], Schmidtet al. [54], and Gritsch et al. [55]. The converging slot-hole (or console) provides the same level of cooling effectiveness as thatof the slot or the fan shaped-hole by Sargison et al. [56]. Embed-ding film cooling holes in slots by Bunker [57], trenches by Wayeand Bogard [58], and craters by Lu et al. [59] (to simulate thermalbarrier coating sprays) has been found to increase film cooling

effectiveness in the proximity of the hole.

Effect of Multiple Rows. Multiple rows of film cooling holesare conventionally used in turbine blade designs. Ligrani et al.[60] studied typical distributions with both simple and compoundangles. At lower blowing ratios (less than 0.5), the effects of thenumbers of rows is fairly insignificant. However, on increasingthe blowing ratio, the double jet row showed a higher effective-ness. More recently, Kusterer et al. [61] studied two rows of filmcooling holes with opposite orientation and internal supply geo-metries. These holes resulted in higher film cooling effectivenessby canceling out the counter-rotating ‘kidney’ vortices (whichare induced by the interaction of the inclined jet with themainstream). Dhungel et al. [62] presented measurements of film

cooling effectiveness using film cooling holes supplemented withspecial anti-vortex holes to increase the effectiveness.

Turbine Vane Film Cooling. It is well known that nozzleguide vanes, being just downstream of the combustor exit, experi-ence the hottest gas path temperatures. The vanes also experiencehigh free-stream turbulence caused by combustor mixing flows.Depending on the requirements, vanes are cooled internally andsome coolant is ejected out as film cooling. A typical film cooledvane is shown in Fig. 1. Nirmalan and Hylton [63] studied the

effects of parameters such as Mach number, Reynolds number,coolant-to-gas temperature ratio, and coolant-to-gas pressure ratioon the C3X vane film cooling. The leading edge has a showerheadarray of five equally spaced rows with the central row located atthe aerodynamic stagnation point. Two rows each on the pressureand suction surfaces are located downstream. With increasingblowing strength, the effect on the pressure surface increases far-ther downstream and the suction surface shows higher effective-ness due to favorable curvature. Ames [64] studied film coolingon a similar C3X vane. Turbulence (Tu ¼ 1–12%) was found tohave a dramatic influence on pressure surface film cooling effec-tiveness, particularly at the lower blowing ratios. Turbulence wasfound to substantially reduce film cooling effectiveness levels pro-duced by showerhead film cooling. Drost and Bolcs [65] foundthat mainstream turbulence (Tu¼ 5.5–10%) had only a weakinfluence on suction surface film cooling. Higher film effective-ness was noted on the pressure surface at high turbulence due toincreased lateral spreading of the coolant. Ethridge et al. [66]studied the effect of coolant-to-mainstream density ratio on a vanewith high curvature. Dittmar et al. [67] studied different film cool-ing hole configurations on the suction (convex) surface and con-cluded that shaped holes provide better coverage at higher blowing ratios by resisting jet penetration into the mainstream.

Turbine Blade Film Cooling. A typical film cooled blade isshown in Fig. 1. Most experimental results for turbine blades areobtained on simulated cascades under simulated engine condi-tions. Ito et al. [68] studied the effect of surface curvature andfound that film cooling effectiveness is relatively subdued on theconcave (pressure) side in comparison with the convex (suction)side, with the flat plate effectiveness values lying in between.Lift-off occurs at a lower blowing ratio on the concave side. How-ever, the curvature of the concave surface results in a reattach-ment of the lifted-off coolant on the pressure side, resulting inhigher downstream effectiveness.

Effect of Unsteady Wake and Secondary Flow. A rotatingspoke-wheel wake generator installed upstream of a typical highpressure film cooled model turbine blade has been to simulate theeffect of an upstream wake by Mehandale et al. [69]. A reductionin film cooling effectiveness due to the unsteady wake wasobserved across the board. The effect of coolant to mainstreamdensity ratio and an unsteady stator wake was studied by Ralla-bandi et al. [70] using the pressure sensitive paint method. Foreigngases with variable density (Nitrogen for DR¼ 1.0, CO2 for DR¼ 1.5 and a mixture of Ar þSF6 for DR¼ 2.5) were used to

simulate realistic engine density ratios. Results show a longer coolant trace on the suction side compared with the pressure side.Due to the concave geometry of the pressure side, at higher blow-ing ratios, a reattachment of the lifted off jet is observed. Anincrease in effectiveness at higher density ratios for a given blow-ing ratio is observed as well as deterioration in film cooling effec-tiveness due to the average effect of the unsteady wake. Gao et al.[71] used the pressure sensitive paint method to characterize fullcoverage film cooling on the blade surface equipped with axiallaid-back fan-shaped holes (expansion and diffusion angles of 10 deg) and another with compound angled laid-back fan-shapedholes. The compound angled holes, in general, resulted in a higher film cooling effectiveness than the axial holes. The effects of tipleakage vortices and horseshoe vortices on the film coolant




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flow-path can be seen. These vortices result in an expansion of thefilm coolant on the pressure side, and a contraction of the coolanttrace on the suction side.

Leading-Edge Film Cooling. A large semicylinder is conven-tionally used as a good approximation to the stagnation region ona turbine vane. Initial studies into film cooling effectiveness near the leading edge were performed by Luckey et al. [ 72]. Ekkadet al. [73] presented the effect of coolant density and free-streamturbulence on a cylindrical leading edge model using a transient

liquid crystal technique to obtain the detailed film effectivenessdistributions. They showed that the film cooling effectiveness val-ues for air as the coolant are highest at a low blowing ratio of 0.4and decrease with an increase in blowing ratio up to 1.2. In themeantime, for CO2 as the coolant, the highest film cooling effec-tiveness is obtained at a blowing ratio greater than 0.8. Gao andHan [74] reported showerhead film cooling effectiveness measure-ments using pressure sensitive paint method for a cylindrical lead-ing edge model. Leading edges with up to seven rows of radialand compound angled shaped and cylindrical holes were studiedfor blowing ratios studied range from 0.5 to 2.0 with DR¼ 1.0,Tu¼ 7%. Results showed that radial angles performed better thancompound angles; shaped holes performed better than cylindricalholes for the range studied.

End-Wall Film Cooling. Due to the large difference in pres-sure between the pressure and suction side of the blade, secondaryvortices are formed in the hub end wall region by Langston [ 39]as shown in Fig. 3(b). These vortices increase heat transfer, neces-sitating provisions for aggressive film-cooling of the end wall. Toachieve this end, film coolant is typically ejected from thecombustor-vane gap and the stator-rotor gap to cool the NGV endwall and rotor blade platform, respectively. Besides this, enginedesigns also incorporate discrete film cooling holes along the endwall and platform. Friedrichs et al. [75] used an innovativeammonia-diazo mass transfer analogy to measure film coolingeffectiveness on the end wall due to discrete holes. The effective-ness corresponding to the film coolant discharged through thecombustor-stator gap has been studied by Oke et al. [76] andZhang and Jaiswal [77]. The additional momentum introduced in

the near wall region by the slot coolant tends to reduce thestrength of the secondary flows. More recently, the improvementin film cooling due to the usage of shaped holes in the end wallhas been studied by Colban et al. [78] and Gao et al. [79]. Resultsshow that shaped holes offer significantly better coverage than cy-lindrical holes. The effect of hub secondary flows (horseshoe vor-tices, etc.) on film cooling is evident from the film coolingeffectiveness contours. The effect of coolant density ratio on filmcooling effectiveness was studied by Narzary et al. [80], with theconclusion that higher density coolants are more resilient to lift-off and result in higher film cooling effectiveness.

Blade-Tip Film Cooling. Film cooling on the blade tip has adual purpose-to protect the tip by forming an insulating film, andto reduce hot-gas tip leakage from pressure side to the suction

side, reducing heat transfer coefficients on the tip. A review of thework done on tip-gap film cooling by Metzger’s group is availablein Kim et al. [81]. More recently, Kwak and Han [82] used liquidcrystal imaging technique to measure detailed film cooling effec-tiveness contours on the squealer tip with tip hole cooling. In gen-eral, the literature agrees that a higher blowing ratio and a smaller tip clearance result in better film cooling performance. Mhetraset al. [33] used a model rotor blade with a cut-back squealer tip toallow the film coolant accumulated on the tip to discharge, in theprocess cooling the trailing edge region of the tip, as shown inFig. 4 [33]. Results (using the pressure sensitive paint method)show that for the tip film cooling holes, the effectiveness increaseswith blowing ratio. The cutback allows the coolant to flowover the trailing edge region, resulting in higher effectiveness.

A blowing ratio of 1.0 seems optimal for the holes on the near-tipregion of the pressure side, indicating that liftoff occurs at higher blowing ratios.

Trailing-Edge Film Cooling. A comprehensive survey of filmcooling investigations prior to 1971 was done by Goldstein [3]and included data for slots as well as discrete holes. Emphasis wason two-dimensional slots. Taslim et al. [83] found that the lip-to-slot height ratio has a strong impact on film cooling effectiveness.Martini et al. [84] measured the film cooling effectiveness and

heat transfer on the trailing edge cutback of gas turbine airfoilswith different internal cooling structures using the IR thermogra-phy method, showing the strong impact of internal design on thefilm cooling performance downstream of the ejection slot. Thefast decay in film cooling effectiveness was attributed to vortexshedding from the pressure side lip. Recently, Cakan and Taslim[85] measured the mass/heat transfer coefficients on the trailingedge slot floor, slot sidewalls and lands using naphthalene subli-mation method. They found that averaged mass-transfer on theland sidewalls are higher than that on the slot floor surface. Choiet al. [86] measured film cooling effectiveness values for differentinternal cooling configurations on a cut-back trailing edge usingthe transient liquid crystal method.

Effect of Thermal Barrier Coating Spallation. Thermal bar-

rier coatings (TBC) are often used to protect turbine componentmetal surfaces from high temperature gases. The spallation canoccur at random; that there is no defined shape or size of the spallmakes it difficult to analyze the actual spallation phenomenaoccurring on a real turbine blade. Thus, it needs to be modeledwith predefined shape, size, and location to understand its effecton local heat transfer coefficients and film cooling effectiveness.Ekkad and Han [87] studied the detailed heat transfer coefficientand film cooling effectiveness distributions on a cylindricalleading-edge model with simulated TBC spallation using a tran-sient liquid crystal technique. The two rows of film cooling holeslocated atþ 15 deg and 15 deg from stagnation. The simulatedspallation cavities were rectangular in shape and had roundededges and are similar to the spallation that typically occur on theturbine blade. In general, presence of spallation enhances heattransfer coefficients and causes variation in film cooling effective-ness distributions.

Effect of Deposition and Blockage on Hole Exits. Bunker [57] presented an experimental study to determine the effects of typical turbine airfoil protective coatings on film cooling effec-tiveness due to the partial blockage of film-hole exits by the TBCcoatings. The measurements indicated significant degradation tofilm performance can result from coatings which are deposited inthe hole-exit regions, or inside the holes themselves, during thespray application process. Results also show that shaped filmholes are generally very tolerant of coatings and do not show thedegradation shown for cylindrical holes. Sundaram and Thole[88] used a large-scale turbine vane cascade to study endwall filmcooling. They showed that partially blocked holes had the greatest

detrimental effect on degrading film-cooling effectiveness down-stream of a film-cooling row. Somawardhana and Bogard [89]indicated that as much as 50% degradation occurred withupstream obstructions, but downstream obstructions actuallyenhanced film cooling effectiveness. The transverse trench config-uration performed significantly better than the traditional cylindri-cal holes, both with and without obstructions and almosteliminated the effects of both surface roughness and obstructions.Ai et al. [90] performed particulate deposition experiments in aturbine accelerated deposition facility to examine the effects of flyash particle size and trench configuration on deposits near filmcooling holes. Deposits that accumulated on the downstream sideof the trench between cooling holes eventually changed the geom-etry of the trench and clogging cooling holes.




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Film Cooling Under Rotating Conditions. Due to the diffi-culty of acquiring data on a rotating blade, literature studying theeffect of rotation is very scarce. Dring et al. [91] reported filmcooling effectiveness in a rotating configuration in a low speedtunnel. Takeishi et al. [92] also studied film cooling effectivenesson a stator-rotor stage, simulating a heavy duty gas turbine.Measured effectiveness values on the suction side for the rotatingturbine blade seemed to match the data from the stationary cas-cade whereas the rotating effectiveness on the pressure sideseemed to be significantly lower than the nonrotating case. Effects

of rotation are attributed to the deflection of the film cooling jetdue to centrifugal forces. Abhari and Epstein [93] reported filmcooling heat transfer coefficients by the superposition method onthe short-duration MIT blowdown turbine facility using heat fluxgauges. Time resolved heat transfer coefficient data was obtained-and the benefit of using film cooling on the blade surface isevident.

More recently, using the PSP method, film cooling effective-ness values under rotating conditions were measured on the lead-ing edge by Ahn et al. [94], and on the rotor platform bySuryanarayanan et al. [95], using a three stage multipurposeresearch turbine at the Turbomachinery Performance and FlowResearch Laboratory at Texas A&M University. Ahn et al. [94]used two rows of showerhead holes, one on the suction side andthe other on the pressure side. Results (using pressure sensitivepaint method) showed that the film cooling effectiveness was sen-

sitive to the location of the stagnation line. When running atdesign condition, the stagnation line was such that coolant wouldbe uniformly dispersed onto the suction and pressure sides. Theeffect of rotation on the film cooling effectiveness on the end wallwas studied by Suryanarayanan et al. [95] due to coolant dis-charged rotor-stator purge slot and discrete holes under rotatingconditions. Results indicated that a blowing ratio of around 1.0was optimal. Film cooling coverage was also found to be optimalwhen running at design condition. Also evident are the effects of the passage vortex, as can be seen by the angle of the film coolanttraces.

Gas Turbine Internal Cooling

Fundamentals of Internal Cooling. The gas turbine bladesare convectively cooled with compressor bled air passing throughthe complex shaped internal cooling channels. These channels arespecifically designed to fit the blade profile and have irregular cross sections (Figs. 1 and 2). Since the design of these channelsvaries from blade to blade, and increased complexities of the flowfield are introduced by irregular cross sectional shapes, research-ers have mostly used square and rectangular channels as modelsin the study of heat transfer. The square and rectangular channelsare categorized by aspect ratio, as seen in Fig. 5 [96]. In thisreview paper, the channel aspect ratio (AR) is defined as the ratioof the channel width (W) to the channel height (H) or AR ¼W/H.Furthermore, the channel height is the distance from the suctionsurface to pressure surface as seen in Fig. 5. The channel width isthe dimension of the surface on which the rib turbulators are cast.

Another point of clarification is in regard to the distinctionbetween “leading edge,” “leading surface,” “trailing edge,” and“trailing surface.” Commonly, the phrase leading surface has beenused interchangeably with suction side/surface. Likewise, trailingsurface is interchangeable with the pressure side/surface.

The internal cooling channels near the blade leading edge havebeen modeled as narrow rectangular channels with AR¼1:4 and1:2. The cross section of the cooling channels changes along thecord length of the blade due to the blade profile. In the middle of the blade, the channels are square in shape. Towards the trailingedge, the channels have wider aspect ratios of AR¼ 2:1 and 4:1.An experimental study on the effects of the buoyancy parameter in various aspect ratio channels was performed by Fu et al. [96].The study considered five different aspect ratio channels

(AR¼

1:4, 1:2, 1:1, 2:1, and 4:1) with a fully developed flow inletcondition. The results showed that the overall levels of heat trans-fer enhancement (Nu/Nuo) for all the ribbed channels were com-parable. However, significant differences arose in the pressurelosses incurred in each of the channels. The 1:4 channel incurredthe lowest pressure penalty; therefore, the thermal performance(TP) of the 1:4 channel was superior to the 1:2, 1:1, and 2:1channels. It is worth noting that the thermal performance takesinto account the pressure penalty (f/f o) and the heat transfer enhancement, and for a constant pumping power, TP¼ (Nu/Nuo)/ (f/f o)

1/3. Many papers reviewed and cited in chapters 4 and 5 of Han et al. [10].

Mid-Chord Rib Turbulated Cooling. In advanced gas turbine

blades, rib turbulators are often cast on two opposite walls of in-ternal coolant passages to augment heat transfer as seen in Fig.2(b) [2]. Rib turbulators are also widely known as “trip strips” asthey simply trip the boundary layer in the internal cooling chan-nel. The heat transfer augmentation in rectangular coolant pas-sages with rib turbulators primarily depends upon the ribturbulators’ geometry, such as rib size, shape, distribution, flow-attack-angle, and the flow Reynolds number. There have beenmany basic studies by Han et al. [97 – 99] to understand the heattransfer augmentation versus the pressure drop penalty by the flowseparation caused by rib-turbulators. The Reynolds numbers basedon coolant channel hydraulic diameter vary from 10,000 to80,000. However, the Reynolds numbers can be up to 500,000 for the coolant passages in large power generation turbine blades. Ingeneral, repeated ribs, used for coolant passages, are nearly square

in cross section with a typical relative rib height of 5–10% of thecoolant channel hydraulic diameter (e/D), a rib spacing-to-heightratio (p/e) varying from 5 to 15, and a rib flow-attack-anglearound 30 deg to 60 deg.

In general, smaller rib height is more efficient for higher Reyn-olds number flows, and the heat transfer enhancement decreasesbut pressure drop penalty increases with the Reynolds number.For example, the heat transfer can be enhanced about three timeswith five times the pressure drop penalty in a square channel withtypical rib geometry (e/D¼ 0.06, p/e¼ 10, and 45 deg rib flow-attack-angle) at a Reynolds number around 30,000. Han andZhang [100] showed that the V-shaped ribs provide better heattransfer performance than the typical angled rib geometry for agiven pressure drop penalty. Smaller gas turbine blades have

Fig. 5 Typical turbine blade internal cooling channel withrotation-induced vortices [96]




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larger blockage ribs with e/D¼ 0.1 0.2 at closer spacing withp/e¼ 3 5 reported by Taslim and Lengkong [101].

Heat Transfer Correlation. More recently, Rallabandi et al.[102] performed systematic experiments to measure heat transfer and pressure losses in a stationary square channel with 45 deground/sharp edged ribs at a wide range of Reynolds numbers rang-ing from 30,000 to very high flows of Re¼ 400,000. These highReynolds are typical of land based turbines. The correlations of Han and Park [99] were modified to fit into the new extended

parameter range. This work has extended the eþ

(a nondimen-sional roughness Reynolds number) range of previous work fromeþ¼ 1,000 (Re¼ 70,000, e/D¼0.078) to eþ¼ 18,000(Re¼ 400,000, e/D¼ 0.18). With round edged ribs, the frictionwas lower, resulting in a smaller pressure drop. The heat transfer coefficients for the round ribs, on the other hand, were similar tosharp edged ribs.

Rotational Effect on Internal Passage Flow and HeatTransfer. Rotation induces Coriolis and centrifugal forces whichproduce cross-stream secondary flow in the rotating coolantpassages; therefore, heat transfer coefficients in rotor coolant pas-sages are very much different from those in nonrotating frames.One important finding from recent studies is that rotation cangreatly enhance heat transfer on one side of the cooling channel

and reduce heat transfer on the opposite side of the cooling chan-nel due to rotating-induced secondary flow, depending on the ra-dial outflow or inflow of the cooling passages (Fig. 5). Withoutconsidering rotational effect, the coolant passage would beovercooled on one side while overheated on the opposite side.Recent studies focus on the combined effects of rotation, channelshape, orientation, and aspect ratio on rotor coolant passage heattransfer with various high performance rib turbulators. Resultsshow that the channel shape, orientation, and aspect ratio signifi-cantly change local heat transfer coefficient distributions in rotor coolant passages with rib turbulators.

Fluid Flow in Rotating Coolant Passages. Heat transfer is aside effect of the flow field. Flow in a rotating channel is signifi-cantly different from flow in a nonrotating channel. The secondary

flow in rotation redistributes velocity and also alters the randomvelocity fluctuation patterns in turbulent flows. Cheah et al. [103]used the LDA to measure velocity and turbulence quantity in arotating two-pass channel. Bons and Kerrebrock [104] measuredthe internal flow in a rotating straight smooth-wall channel withparticle image velocimetry (PIV) for both heated and nonheatedcases. Liou et al. [105] measured pressure and flow characteristicsin a rotating two-pass square duct with 90-deg ribs by using theLAD. Rotation shifts the bulk flow toward the trailing side, andthe turbulence profile shows a different distribution in rotation.The above-mentioned flow measurements help to understand theflow physics and serve to explain the heat transfer results obtainedin two-pass rotating channels with smooth and ribbed walls.

Square Cross-Section Serpentine Channel. Heat transfer in

rotating multipass coolant passages with square cross section andsmooth walls was reported by Wagner et al. [106]. Results showthat the heat transfer coefficient can enhance 2–3 times on thetrailing surface and reduce up to 50% on the leading surface for the first-pass radial outward flow passage; however, the reverse istrue for the second-pass radial inward flow passage due to theflow direction change. Results also show that the heat transfer dif-ference between leading and trailing surfaces is greater in thefirst-pass than that in the second-pass due to the centrifugal buoy-ancy opposite to the flow direction. Heat transfer in rotating multi-pass coolant passages with square cross section with 45 deg ribturbulated walls was reported by Johnson et al. [107]. Resultsshow that rotation and buoyancy in general have less effect on therib turbulated coolant passage than on the smooth-wall coolant

passage. This is because the heat transfer enhancement in theribbed passages is already up to 3.5 times higher than inthe smooth passages; therefore, the rotational effect is still impor-tant but with a reduced percentage. Results also show that, like anonrotating channel, the 45 deg ribs perform better than 90 degribs and subsequently better than the smooth channel.

Wall Heating Condition Effect. Since the temperature differ-ence between the coolant and the channel walls varies along thecoolant passages, so does the rotation buoyancy. Therefore, it is

expected that the channel wall heating conditions would affectrotor coolant passage heat transfer. Han et al. [108] studied theuneven wall temperature effect on rotating two-pass square chan-nels with smooth walls. They concluded that in the first pass, thelocal uneven wall temperature interacts with the Coriolis force-driven secondary flow and enhances the heat transfer coefficientsin both leading and trailing surfaces as compared with the uniformwall temperature case. Zhang et al. [109] studied the influence of wall heating condition on the local heat transfer coefficient inrotating two-pass square channels with 90 deg ribs and 60 deg ribson the leading and trailing walls, respectively. They concludedthat the uneven wall temperature significantly enhances heat trans-fer coefficients on the first-pass leading and second-pass trailingsurfaces as compared with the uniform wall temperaturecondition.

Channel Orientation Effect. Since the turbine blade iscurved, the rotor blade cooling passage can have different channelorientations with respect to the rotating plane. Johnson et al. [110]studied the effects of rotation on the heat transfer for smooth and45 deg ribbed serpentine channels with channel orientations of 0 deg and 45 deg to the axis of rotation. They found that theeffects of Coriolis and buoyancy forces on heat transfer in therotating channel are decreased with the channel at 45 deg com-pared to the results at 0 deg. This implies that the difference inheat transfer coefficient between leading and trailing surfaces dueto rotation will be reduced when the channel has an angle to theaxis of rotation. Dutta and Han [111] used high performance bro-ken V-shaped ribs in rotating two-pass square channels to studythe effect of channel orientation on heat transfer. The channel ori-

entation with respect to the rotation axis influences the secondaryflow vortices induced by rotation, as shown in Fig. 5. They con-cluded that the broken V-shaped ribs are better than the 60 degangled ribs; the parallel 45 deg angled ribs are better than thecrossed 45 deg angled ribs. In general, the difference betweenleading and trailing wall heat transfer coefficients is reduced for the channel with a 45 deg angle to the axis of rotation.

Rotation Number and Buoyancy Parameter. It is worth-while, then, to develop nondimensional parameters that may beused to correlate rotating effects to heat transfer. The rotationnumber (Ro) has been widely accepted to establish the strength of rotation by considering the relative strength of the Coriolis forcecompared to the bulk inertial force. As such, the rotation number is defined as Ro¼XDh /V. The buoyancy parameter (Bo) is useful

to include the effects of density variation (centrifugal effects) andis defined as the ratio of the Grashoff number to the square of theReynolds number; both of which are based on the channel hydrau-lic diameter. Thus Bo¼ (Dq / q)(Ro2)(R/Dh). Typical rotationnumbers for aircraft engines are near 0.25 with Reynolds numbersin the range of 30,000. One method to achieve conditions similar to a real gas turbine engine in the laboratory is to use air at highpressures. As the pressure of the air increases, so will the density.For a fixed Reynolds number, dynamic viscosity, and hydraulicdiameter, an increase in density will proportionately decrease thebulk velocity. A lower bulk velocity will in turn increase the rota-tion number since the rotation number is the ratio of the Coriolisforce to bulk inertial force. Increasing the range of the rotationnumber and buoyancy parameter is very important since gas




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turbine engineers can utilize these parameters in their analysis of heat transfer under rotating conditions.

Rectangular Cross-Section Two-Pass Channel. Zhou andAcharya [112] studied a 4:1 aspect ratio channel with a rotationnumber of 0.6 at a Reynolds number of 10,000. Huh et al. [ 113]increased the range of the rotation number by a factor of 4 for theAR¼ 2:1 channel. Huh et al. [114] studied heat transfer in a 1:4aspect ratio channel (Fig. 5). Results show that heat transfer onthe trailing surface with radially outward flow does indeed

increase under rotating conditions due to the flow phenomena pre-viously described. Rotation reduces the heat transfer on the lead-ing surface by a very significant 50%. However, due to buoyancyeffects, the leading surface heat transfer trends reverse after a crit-ical rotation number is reached. With radially inward flow, theheat transfer in the smooth channel shows the expected behavior on the leading wall. Surprisingly, however, due to the aspect ratioof the channel, the heat transfer on the trailing wall also increases.In square channels this is not the case.

Blade Tip Internal Cooling. A gas turbine blade experienceshigh heat loads on the tip portion due to high velocity fluid leak-age between the rotating blade and casing. Until recently, most of the studies that have considered heat transfer in multipass internalserpentine channels provided minimal information on heat trans-

fer on the inside of the blade tip. Even fewer studies are availablethat consider the effect of rotation on blade tip cap heat transfer.The effects of rotation on tip cap internal heat transfer in rectangu-lar channels with AR¼2:1 was presented in the study by Huhet al. [113] and Huh et al. [114], which provided heat transfer results on the tip cap of the 1:4 aspect ratio channel. Results revealthat rotation helps to increase cooling of the blade tip internal sur-face. Rotation doubles the heat transfer coefficients on the tip capsurface in both passages.

Developing Flow Entrance Effect. Some gas turbine bladedesigns provide a developing flow entrance. It is well acceptedthat due to the thin boundary layer, heat transfer with developingflow is markedly different from fully developed flows. Wrightet al. [115] performed experiments in channels with three different

entrance geometries. They concluded that the entrance conditionwill enhance the heat transfer. They also pointed out that theeffect of the entrance weakens as the rotation number increases.The influence of the entrance geometry also is stronger in thesmooth channel when compared to the ribbed channel. Huh et al.[113] studied a sudden expansion from a circular tube to the rec-tangular cross section of the channel. Notable is the lack of degra-dation in heat transfer, until large Bo values, on the leadingsurface for the developing flow cases. Heat transfer is clearlydominated by the entrance.

Leading-Edge Impingement Cooling. Jet impingement cool-ing is most suitable for the leading edge of the blade where thethermal load is highest and a thicker cross section of this portionof the blade can suitably accommodate impingement cooling

(Figs. 1 and 2). There are many studies focused on the effects of jet-hole size and distribution, jet-to-target surface distance, spent-air cross flow, cooling channel cross section, and the targetsurface shape on the heat transfer coefficient distribution (for example, Chupp et al. [116], Metzger et al. [117], etc.). Recentstudies have considered the combined effects of target surfaceroughening coupled with jet impingement for further heat transfer enhancement. Taslim et al. [118] investigated heat transfer on acurved target surface to more realistically simulate the leadingedge of the blade. Three different roughening techniques werestudied: conical bumps, tapered radial ribs, and sand paper typeroughness. Kanokjaruvijit and Martinez-Botas [119] showed thatby impinging on the dimple, higher energetic vortices were gener-ated and thus heat transfer was increased. Since the leading edge

of the gas turbine blade incorporates a showerhead film coolingdesign, Taslim and Khanicheh [120] showed that the heat transfer can be significantly increased by including the film cooling holeson the target plate.

Rotational Effect on Impingement Cooling. All of the stud-ies previously mentioned considered jet impingement heat transfer under stationary conditions. Of course, however, the turbine bladeis rotating. Overall, the effectiveness of the jet is reduced under rotating conditions due to deflection from the target surface.

Epstein et al. [121] studied the effect of rotation on impingementcooling in the leading edge of a blade. They reported that the rota-tion decreases the impingement heat transfer, but the effectiveheat transfer is better than a smooth rotating channel. Mattern andHennecke [122] reported the effect of rotation on the leading edgeimpingement cooling by using the naphthalene sublimation tech-nique. They found that the rotation decreases the impingementheat transfer for all staggered angles. Glezer et al. [123] studiedthe effect of rotation on swirling impingement cooling in the lead-ing edge of a blade. They found that screw-shaped swirl coolingcan significantly improve the heat-transfer coefficient over asmooth channel and the improvement is not significantly depend-ent on the temperature ratio and rotational forces. Parsons et al.[124] studied the effect of rotation on impingement cooling in themid-chord region of the blade. A central chamber serves as thepressure chamber, and jets are released in either direction toimpinge on two heated surfaces. The jet impinging directionshave different orientations with respect to the direction of rotation.They reported that the rotation decreases the impingement heattransfer on both leading and trailing surfaces with more effect onthe trailing side (up to 20% heat transfer reduction).

Trailing-Edge Pin Fins Cooling. Pin-fins are mostly used inthe narrow trailing edge of a turbine blade where impingementand ribbed channels cannot be accommodated due to manufactur-ing constraint (Figs. 1 and 2). Pin-fins commonly used in turbineblade cooling have pin height-to-diameter ratio between 1

2 and 4.

Heat transfer in turbine pin-fin cooling arrays combines the cylinder heat transfer and end wall heat transfer. Due to the turbulenceenhancement caused by pins (wakes and horseshoe vortex), heattransfer from end-walls is higher than smooth wall cases; however,casting pins will cover a considerable end wall area, and that areaneeds to be compensated for by the increased pin surface area for cooling. In addition to flow disturbances, pins conduct thermalenergy away from the end wall surface. Long pins can increase theeffective heat transfer area and perform better than short pins. Therehave been many investigations that studied the effects of pin array(inline or staggered), pin size (length-to-diameter ratio¼ 0.5 to 4),pin distribution (streamwise-and spanwise-to-diameter ratio¼ 2 to4), pin shape (with and without a fillet at the base of the cylindricalpin; oblong, cube, and diamond shaped pins as well as the stepped di-ameter cylindrical pins), partial length pins, flow convergence andturning, and with trailing edge coolant extraction on the heat transfer coefficient and friction factor distributions in pin-fin cooling channels(for example, Metzger et al. [125], Chyu et al. [126], etc.)

Rotational Effect on Pin Fins Cooling. Wright et al. [127]studied the effect of rotation on heat transfer in narrow rectangular channels (AR¼ 4:1 and 8:1) with typical pin-fin array used in tur-bine blade trailing edge design and oriented at 150deg withrespect to the plane of rotation. Results show that turbulent heattransfer in a stationary pin-fin channel can be enhanced up to 3.8times that of a smooth channel; rotation enhances the heat trans-ferred from the pin-fin channels up to 1.5 times that of the station-ary pin-fin channels. Most importantly, for narrow rectangular pin-fin channels oriented at 135 deg with respect to the plane of rotation, heat transfer enhancement on both the leading and trail-ing surfaces increases with rotation. This provides positive infor-mation for the cooling designers.




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Rotational Effect on Wedge-Shaped Cooling Channel. Thetrailing edge cooling passage has been represented with wide as-pect rectangular channels. However, the cross sectional shape isbest represented with a wedge or trapezoid. To enhance heat trans-fer in this region of the blade, the leading and trailing surfaces areroughened with ribs or pin-fins. Further protection is providedwith coolant ejection from the narrow portion of the channel(Fig. 5), and the additional effects of Coriolis induced secondaryflows and centrifugal driven buoyancy alter the heat transfer char-acteristics. Chang et al. [128] studied heat transfer in rib rough-

ened trapezoidal duct with bleed holes. Liu et al. [129] consideredheat transfer in a trailing edge cooling passage with smooth wallswith trailing edge slot ejection. The channel was placed at an angleof 135 deg respective of the direction of rotation. Most notably, for all three surfaces, the Nusselt number ratios increase as the rotationnumber increases. The heat transfer enhancement with slot ejectionis much higher than the cases without slot ejection. Rallabandiet al. [130] studied the effect of full length conducting, partiallength conducting and nonconducting pins in a wedge-shaped chan-nel with trailing edge bleeding. The rotational effects were alteredin different regions by the presence of the pins.

Rotational Effect on Dimples Cooling. Dimples are recentlybeing considered for turbine blade trailing edge cooling designs.Dimples provide reasonable heat transfer enhancement with a rel-

atively low pressure loss penalty as compared with the ribs andpin-fins. The dimple cooling can be a good choice if the pressureloss is the main concern in the cooling design. Due to the disturb-ance enhancement caused by dimples, heat transfer from dimpledsurface is higher than the smooth wall conditions. This is becausedimples induce flow separation and reattachment with pairs of vortices. In addition to flow disturbances, dimples increase heattransfer area. In general, higher heat transfer enhancement occurson the flow reattached regions either at the dimple cavity down-stream or on the dimple downstream flat surface. The heat transfer enhancement is typically around 2–2.5 times that of the smoothwall value with 2–4 times pressure loss penalty and is fairly inde-pendent of Reynolds number and channel height or aspect ratio.There have been a number of studies that evaluated the effects of dimple size, dimple depth (depth-to-print diameter ratio ¼ 0.1 to

0.3), distribution, shape (cylindrical, hemispheric, and teardrop),and channel height on the heat transfer coefficient and friction fac-tor distributions in dimple cooling channels (for example, Mah-mood et al. [131], etc.). However, the majority of investigationsinvolving dimple cooling have been limited to stationary channelsthat are applicable for stator blade trailing edge cooling designs;only a few studies focus on rotor blade dimple cooling. Zhou andAcharya [132] studied heat/mass transfer in a rotating squarechannel with typical dimple array. They found that rotation enhan-ces heat transfer on the trailing dimple surface and reduces heattransfer on the leading dimple surface in a similar manner as therotational effect on the trailing and leading surfaces of the squarechannel with ribs. Griffith et al. [133] studied heat transfer inrotating rectangular channels (AR¼ 4:1) with typical dimplearray. The results show that rotation enhances heat transfer onboth trailing and leading surfaces of the narrow dimpled channelin a similar trend as the rotational effect on the trailing and lead-ing surfaces of the narrow rectangular channel with pins; how-ever, the heat transfer enhancement of the pinned channel exceedsthat of the dimpled channel. Additionally, the dimpled channeloriented at 135 deg with respect to the plane of rotation providesgreater overall heat transfer enhancement than the orthogonaldimpled channel.

Numerical Modeling

CFD for Turbine Internal Cooling. In recent years, manyresearchers have made computational studies on internal coolingchannels of the rotating blade. Numerical predictions provide the

details that are difficult to obtain by experiments. Moreover, theincrease in computation power in desktop computers has made iteconomical to optimize the design parameters based on numericalanalyses. Most common models are based on a two-equation tur-bulence model; namely, the k– e model, low Reynolds number k– emodel, the two-layer k– e model, and the low Reynolds number k– x model. The Reynolds averaged Navier–Stokes (RANS) andlarge–eddy simulations (LES) are the most commonly used simu-lation methods for turbine blade internal flow and heat transfer predictions. Direct numerical simulation (DNS) is to solve every

flow in detail. The extremely small grid spacing and time incre-ments makes this type of simulations extremely expensive interms of time and computational resources. Jang et al. [134]employed Reynolds stress turbulence model to predict the flowand heat transfer in turbine blade cooling passage with rib turbula-tors. They concluded the second moment solutions display largeanisotropy in turbulent stress and heat flux distributions. Withrotation, the Coriolis and buoyancy forces result in strong noniso-tropic turbulence flows. Viswanathan and Tafti [135] present thelarge eddy simulations (LES) of flow and heat transfer in rotatingsquare duct with 45 deg rib turbulators. The unsteady temperaturefield in periodic domain is computed directly. The authors observethat the large scale vortices play a major role in the mixing of thecore fluid and the near-wall heated fluid, the vortex sheddingbehind the ribs are responsible for the large spike in the energyspectrum, and the time variation of the flow rate is attributed to

the variation dominated by the vortex shedding frequency.

CFD for Turbine Film Cooling. It is difficult to model turbineblade with film cooling due to the complicated flow phenomenacoupling with heat transfer process. Turbulence is resolved by dif-ferent CFD methods, including RANS, URANS, LES, DNS, etc.Although most of the models have good agreements with experi-mental results, the accuracy still needs to be further improved.Generally, unsteady models perform better than steady models. Inthis section, numerical simulations of film cooling from a fewselected papers are mentioned. Voigt et al. [136] performeddetailed comparison and validation of RANS, URANS, and SASSimulations on flat plate film cooling. Five different turbulencemodels are used to simulate the flat plate film-cooling process.

The models include three steady and two unsteady methods. Thesteady RANS methods are the Shear stress transport (SST) modelof Menter, the Reynolds stress model of Speziale, Sarkar, andGatski and a k– e explicit algebraic Reynolds stress model. Theunsteady models are a URANS formulation of the SST model anda scale-adaptive simulation (SAS). The solver used in this study isthe commercial code ANSYS CFX 11.0. The results are compared toavailable experimental data. These data include velocity and tur-bulence intensity fields in several planes. Results show that thesteady RANS approach has difficulties with predicting the flowfield due to the highly three-dimensional unsteadiness. TheURANS and SAS simulations on the other hand show good agree-ments with the experimental data. The deviation from the experi-mental data in velocity values in the steady cases is about 20%whereas the error in the unsteady cases is below 10%. Leedom

and Acharya [137] presented Large Eddy Simulations (LES) of film cooling flow fields from cylindrical, laterally diffused, andconsole shaped holes. The results show that the console’s per-formance is superior to the laterally diffused and cylindrical holesin terms of jet penetration into the cross-flow. The turbulence gen-erated in the near field of film cooling jets was found to be highlyanisotropic. Sreedharan and Tafti [138] performed a numericalstudy to investigate deposition and erosion of Syngas ash in theleading edge region of a turbine vane. Large Eddy Simulation(LES) is used to model the flow field of the coolant jet-mainstream interaction and syngas ash particles are modeled usinga Lagrangian framework. Overall, for particles of size 5 lm, thereis a combined increase in deposition and erosive particles from16% to 24% as the blowing ratio increases from 0.5 to 2.0. The




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7lm particles, on the other hand, decrease from 35% to about30% as the blowing ratio increases from 0.5 to 2.

CFD for Conjugate Heat Transfer and Film Cooling. Turbineinternal cooling and film cooling are the major cooling techniquesthat applied to turbine blades. In both types of cooling, heat isremoved by means of both convection and conduction. Moreover,convection and conduction are effected from each other. Conju-gate heat transfer is basically termed as the interaction betweenthe convection heat transfer from the surrounding fluid and the

conduction heat transfer through the solid body. The conductionheat transfer is affected from the convection heat transfer of thesurrounding fluid. Thus, they should be solved simultaneously.This coupling of solid to fluid is usually done by using the samewall temperature for the adjacent fluid block and solid block.The conjugate CFD methods provide good predictions for heattransfer analysis in turbine blades. Eliminating the heat transfer coefficient calculation by utilizing the relationship between solidand fluid interface, conjugate methods can provide direct solu-tions. There is much commercially available software (i.e., ANSYSFLUENT, ANSYS CFX, STAR CCM, etc.). However, the accuracy of conjugate methods must be compared with experimental dataand still remains to be improved. For example, Shih et al. [ 139]performed an extensive study on the effects of Biot number ontemperature and heat-flux distributions in a TBC-coated flat plate

cooled by rib-enhanced internal cooling. He and Oldfield [ 140]conducted a study on modeling effect of hot streak on TBC-coated turbine vane heat transfer by unsteady conjugate heattransfer.

Concluding Remarks

Turbine Film Cooling Heat Transfer. For turbine blade exter-nal cooling, most available experimental data are for the main bodyof turbine blade heat transfer and film cooling. Recent researchfocuses on unsteady wake, high free-stream turbulence, and surfaceroughness/TBC spallation effects on turbine rotor blade heat transfer with film cooling. To optimize the film cooling performance, effectsof film-hole size, length, spacing, shape, and orientation on turbineblade heat transfer distributions need to be considered. Satisfying

the even higher turbine operating temperature requirement for higher power and efficiency makes turbine blade edge cooling anurgent issue for gas turbine blades. Turbine blade edge cooling andheat transfer includes turbine blade leading-edge, trailing-edge, tipand platform, with and without film cooling, under engine Machand Reynolds number flow conditions. Investigations in the futurecould focus on the effect of velocity, temperature, and turbulenceprofiles exiting the combustion chamber on film cooling of surfaceand end-walls of the first high pressure vane. The effect of rotationand unsteady velocity and temperature profiles and stator wakeson rotor blade film cooling effectiveness and heat transfer coeffi-cients is a relatively unexplored subject. Effects of thermal barrier coating spallation, film cooling-hole blockage and surface rough-ness on the blade surface, end-walls and tip also merit investiga-tion. Gas turbine vane and blade designs are beginning toincorporate contoured end-walls to reduce aerodynamic lossesassociated with secondary vortices. The effect of contoured end-walls on film cooling can also be explored. Highly accurate andhighly detailed local heat transfer and film cooling data in theturbine blade main body as well as the turbine blade edgeregions would be critical in preventing blade failure due to localhot spots. Flow visualizations, measurements, and CFD predic-tions would provide valuable information for designing effectivecooled blade for advanced gas turbines.

Turbine Internal Cooling Heat Transfer. For turbine bladeinternal cooling, most experimental data available to date are for rotating rectangular cooling channels with high performance ribturbulators for Reynolds number up to 50,000, rotation number up

to 0.25, and buoyancy parameter up to 0.5. These parameters areapplicable for aircraft gas turbines. More studies are needed for the blade-shaped coolant passages (realistic cooling passage ge-ometry, shape, and orientation) with high performance turbulatorsand with or without film cooling holes, for rotating impingementcooling with or without film coolant extraction, as well as rotatingpin-fin cooling with or without trailing edge ejection in order toguide the efficient rotor blade internal cooling designs. In addi-tion, for land-based power generation turbines, more studies areneeded for rotor coolant passage heat transfer under higher cool-

ant flow (Reynolds number up to 500,000), thermal (buoyancynumber up to 5), and rotation (rotation number up to 0.5) condi-tions. Highly accurate and highly detailed local heat transfer coef-ficient and pressure drop data under these extreme cooling designconditions would be needed to prevent the blade from failure dueto local hot spots. In addition, study of higher heat transfer enhancement versus lower pressure drop penalty should continueto identify the best heat transfer augmentation technique includingcompound and new cooling techniques. Development of accurateand efficient CFD prediction tools should continue to provide val-uable information for designing effective cooled rotor blades for the new generation of gas turbines. With advancements beingmade on alternative fuel sources for turbines (i.e., hydrogen), astep change in the capability of blades to handle higher heatingloads is a must. Compound and new cooling concepts need to bedeveloped and explored, such as double wall cooling for new tur-

bine blades and vanes [141,142], heat pipe and microchannelapplications for blade tip, leading, and trailing edge cooling. Fun-damental studies need to consider the effects of rotation on thesenew cooling concepts. Development of this technology will ensurethat the blade design is not the limiting factor for increased effi-ciency and the move to other fuel sources.

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