Turbomachinery asmedigitalcollection

Akhilesh Rallabandi1Turbine Heat Transfer Laboratory,

Mechanical Engineering Dept.

Texas A&M University,

College Station, TX 77843-3123

Jiang Lei2Turbine Heat Transfer Laboratory,

Mechanical Engineering Department,



Je-Chin HanTurbine Heat Transfer Laboratory,

Mechanical Engineering Department,



e-mail: [email protected]

Salam AzadSiemens Energy, Inc.,

4400 Alafaya Trail,

Orlando, FL 32826

Ching-Pang LeeSiemens Energy, Inc.,

4400 Alafaya Trail,

Orlando, FL 32826

Heat Transfer Measurementsin Rotating BladeShapeSerpentine Coolant PassageWith Ribbed Walls at HighReynolds NumbersFlow in the internal three-pass serpentine rib turbulated passages of an advanced highpressure rotor blade is simulated on a 1:1 scale in the laboratory. Tests to measure theeffect of rotation on the Nusselt number are conducted at rotation numbers up to 0.4 andReynolds numbers from 75,000 to 165,000. To achieve this similitude, pressurized FreonR134a vapor is utilized as the working fluid. Experimental heat transfer coefficient meas-urements are made using the copper-plate regional average method. Regional heat trans-fer coefficients are correlated with rotation numbers. An increase in heat transfer ratesdue to rotation is observed in radially outward passes; a reduction in heat transfer rate isobserved in the radially inward pass. Strikingly, a significant deterioration in heat trans-fer is noticed in the hub regionbetween the radially inward second pass and the radi-ally outward third pass. This heat transfer reduction is critical for turbine coolingdesigns. [DOI: 10.1115/1.4026945]

Introduction

Gas turbine rotor blades and stator vanes use rib-turbulatedinternal convective cooling for thermal management. Newer land-based gas turbine engines (with a higher power-rating) are physi-cally larger in size than aviation turbine blades and require higherhydraulic diameter internal passages. This results in a higher inter-nal Reynolds number (of up to 500,000). The goal of this work isto acquire heat-transfer coefficient data for a full-scale rotatingturbine blade (Fig. 1) with high internal flow Reynolds numbers.A significant amount of open literature has been generated for

rib-roughened internal channels over the years. Han [1] conductedtests on channels with various aspect ratios using orthogonal lowblockage ratio ribs (0.02< e/D< 0.08) for spacing to height (p/e)ratios ranging from 10 to 20. It was concluded that, for the spacingvalues studied, taller ribs placed closer together (higher e/D andlower p/e) performed the best. Han and Park [2] and Park et al. [3]studied heat transfer with nonorthogonal ribs and concluded thatthe secondary flows induced by the ribs increase the heat transfercoefficient on the ribbed surface. They concluded that using ribsangled at 45 deg60 deg was the most beneficial from a thermalperformance point of view.The effect of channel aspect ratio has been studied by Han and

Park [2] and Park et al. [3]. Though the ribbed side heat transferaugmentation is of the same order in all cases, the friction factor

is much higher for channels with wider aspect ratios. Studies byTaslim et al. [4] have focused on cooling passages embedded inthe leading edge of the blade. Zhang et al. [5] performed studieson a ribbed triangular channel. A flow visualization of the second-ary flows is presented in Ref. [6], and a computational picture ofsecondary flows is discussed in Ref. [7]. Su et al. [8] performedcomputations on a rotating channel with inclined ribs and pre-sented predictions of secondary flows in the first channel. Highblockage channels have been investigated by Taslim and Leng-kong [9] and more recently by Bunker and Bailey [10] and Ralla-bandi et al. [11]. They found that friction factors increased withreduced rib spacing; the Normalized Nusselt number ratio (Nu/Nu0) was reduced by increasing the Reynolds number. They alsofound a monotonic dependence of Nu on e/D.Further derivatives of the inclined rib concept (crossed ribs,

v-shaped ribs, inverted v-shaped ribs [12], broken parallel ribs,v-shaped broken ribs, d, and wedge shaped ribs [13], etc.) havebeen studied. V-shaped ribs are shown to enhance heat transferover parallel ribs at lower friction factors than comparable parallelribs. The broken v-shaped ribs are found to enhance heat transferfurther at a comparable pressure drop. Studies have shown a dete-rioration in heat transfer enhancement due to the filleting, thoughcompensated by a reduction in corresponding friction factor[10,1416]. This deterioration in heat transfer coefficient wasfound to be less at higher Reynolds numbers [11]. The reductionin friction factor was found to be larger at a higher e/D ratio[10,11,14], indicating that this effect could actually be beneficial.Dippery and Sabersky [17] detailed a method to analyze

the pressure drop and heat transfer in a rough duct using nondi-mensional parameters based on the turbulent boundary layerlaw-of-the-wall. Han et al. [2,15,18] adapted these parameters torib-roughened channels. This correlation has been applied to vari-ous rib configurations (v-shaped, v-shaped broken, etc.) [12] andaspect ratios [3]. At lower Reynolds numbers in the ribbed chan-nel, both the friction factor and average fully developed Nusseltnumber are found to be relatively low. The typical friction factor

1Current address: Test R&D Engineer, Intel Corporation, 5000 West ChandlerBoulevard, Chandler, AZ 85226.

2Current address: Lecturer, SKL SVMS, Xian Jiaotong University, Xian,Shaanxi 710049, China.

Contributed by the International Gas Turbine Institute (IGTI) of ASME forpublication in the JOURNAL OF TURBOMACHINERY. Manuscript received January 8,2014; final manuscript received February 18, 2014; published online March 17,2014. Editor: Ronald Bunker.

The content of this paper is copyrighted by Siemens Energy, Inc. and is licensedto ASME for publication and distribution only. Any inquiries regarding permissionto use the content of this paper, in whole or in part, for any purpose must beaddressed to Siemens Energy, Inc. directly.

Journal of Turbomachinery SEPTEMBER 2014, Vol. 136 / 091004-1CopyrightVC 2014 by ASME

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tends to attenuate to a constant value as the Reynolds number isincreased [9,14]. The Nusselt number shows a monotonicallyincreasing trend with Reynolds number.The effect of rotation on the heat transfer in a rotating configu-

ration has been studied extensively in literature. Of special signifi-cance are the works of Wagner and Johnson [1922]. The fluiddynamics in rotating turbine blades with heated walls is a complexinterplay of various forces: internal, viscous, Coriolis, and centrif-ugal buoyancy. In radially outward flow configurations, the Corio-lis force induces secondary circulations within the flow field,which results in the accumulation of the warm fluid near the lead-ing surface, and cool fluid near the trailing surface. This results ina reduction in heat transfer along the leading surface and anincrease on the trailing surface. Furthermore, the warm fluid nearthe leading surface experiences a Buoyancy force radially inward,creating a recirculation tendency in the flow field.Various studies [1928] show an increase in heat transfer on

the trailing surface on radially outward passes and deteriorationon the leading surface. This occurs even when roughness elementsare provided in the channel. For instance, Wagner and Johnson[21] report that using skewed turbulators undergoes a smallerdeterioration in heat transfer on the leading surface (in the firstpass) in comparison with smooth surfaces.From the brief review of open literature, the majority heat trans-

fer studies are simulated by physically scaled up flow channelmodels. Most of studies are based on Reynolds numbers from10,000 to 30,000 with rotation numbers around 0.20.4. Theresults can be applied to simulate internal coolant passage designsfor aircraft engines. However, for land-based gas turbine blades,Reynolds numbers can be as high as 100,000 to 200,000, withrotation numbers around 0.20.4. Since it is not easy to reach highrotation numbers at very high Reynolds numbers in laboratoryconditions, open literature does not provide heat transfer informa-

tion in real 1:1 scale blade-shaped coolant passage at simulatedengine cooling flow rotation conditions. The objective of the cur-rent study is to provide heat transfer enhancement under stationaryand rotating condition for real serpentine blade-shaped ribbedchannels in a real size land-based gas turbine blade. This studyfocuses on the rotation effect on heat transfer enhancement onchannel midspan as well as before and after the tip and hub turnregions at high Reynolds number and high rotation numberconditions.

Experimental Details. A three-pass serpentine geometry(Figs. 2 and 3) has been used for the current study. The test sec-tion is composed of 15 sequential regions, each region composedof either 3 (regions 5, 6, 10, 11, and 15) or 4 (all other regions)copper plates. Each region has a trapezoidal shape (which changesin hydraulic diameter from the first to the third passes). Withineach region, the plates correspond to leading (suction), trailing(pressure), and divider walls.Copper plates in each region are thermally insulated from each

other; each region is also thermally insulated from other regions.This thermal insulation is facilitated by the use of a garolite insertwhich is designed to hold the plates onto the test section surfacewithout requiring nuts and bolts. Regions 15 comprise the firstpass, 610 the second, and 1115 comprise the third pass.Four silicone heaters are used for each passlabeled suction

or leading; pressure or trailing; divider and leadingedge or trailing edge. Each silicon heater contacts either 4 or 5copper plates. The use of the garolite insert to hold the plates inplace (described earlier) ensures that nuts and bolts are notrequired to hold the copper-plate-heater in place. This allows theuse of heaters without perforations for this purpose.The test section is made of green garolite due to its structural

strength and its resistance to test temperatures (80 C). The testsection is housed inside an aluminum pressure vessel. It is held inplace by a compression plate secured by large 1 in. diameter(2.54 cm) stainless steel bolts. The pressure vessel is attached toan aluminum rotating arm (100 cm in radius). A similar pressurevessel is located on another rotating arm housed 180 deg from theabove arm. The second pressure vessel is a dummy pressure ves-sel installed for rotor-dynamic stability (Fig. 4).The three-pass serpentine resembles the internal cooling chan-

nel of an actual gas turbine blade. As shown in Fig. 1, the centralplains of the first and second pass (as well as the second and thirdpasses) are at an angle. The U-turn also incorporates a change inhydraulic diameterthe first pass has a hydraulic diameter of0.75 in. (9.05mm), the second pass 0.67 in. (17.02mm), and thethird pass has a hydraulic diameter of 0.46 in. (11.68mm). A

Fig. 2 Internal view of test section showing various regions ofmeasurement

Fig. 1 (a) Rotor blade cross-sectional (b) test sectionpassages

091004-2 / Vol. 136, SEPTEMBER 2014 Transactions of the ASME


plenum (4 times in cross section area to the first pass) is providedupstream of the first pass to simulate developing thermal and mo-mentum boundary layers in the first pass. A ribbed geometry with ane/D of 0.081 for the first pass, 0.091 for the second pass, and 0.133for the third pass is studied. Parallel 45 deg angled ribs are placed onboth pressure and suction walls with a staggered pattern. The p/e(pitch to height ratio) is 10, which is a typical value used by gas tur-bine designers. All reported results correspond to this geometry.T-type (copper-constantan) thermocouples are embedded in

each copper plate. For the heat transfer coefficients encounteredin these tests, the Biot number for each copper plate works out toless than 0.1, ensuring that the plates are isothermal for experi-mental considerations. These thermocouples are soldered to feedthrough wires which are routed to the stationary frame through a200 channel slip-ring assembly which is located along the axis ofrotation. The same slip ring/feed through wire assembly is usedfor routing the electric power necessary for energizing the differ-ent heaters present in the test section. The voltage to each heateris controlled through unique external variable transformers. A NISCXI chassis/terminal block assembly and Labview based pro-gram is used to acquire and record the various temperaturesrecorded by the thermocouples.The working fluid for these tests is Freon R134a vapor. R134a

vapor(flow loop details in Fig. 5) is chosen due to its relativelyhigher densityat 4.5 atmospheres (test conditions), its density is18 kg/m3, similar to that of air at gas turbine operating conditions.The molecular weight of Freon R-134 a is almost four times largerthan air, and it can serve purpose better when compressed to samepressure. Its Prandtl number is 0.8 at test conditions, which is sim-ilar to that of air. Mach numbers expected (corresponding with themaximum measured Reynolds numbers of 165,000) are< 0.1, sothe flow regime is incompressible. R134a does not have issueswith safety as it is inert and noncombustible at room temperature.OSHA regulations were followed while handling the refrigerant.R134a vapor flow is established using two flow loops (Fig. 5), a

primary boiler loop which supplies the R134a vapor for the tests.This vapor then passes through the condenser where heat isextracted from the vapor through a noncontact heat exchanger.The secondary working fluid (which absorbs heat from the pri-mary R134a, which condenses in the condenser) is also R134a.

Fig. 3 Internal passage view45 deg rib arrangement

Fig. 4 Rotating rig assemblyCAD and photographic views

Journal of Turbomachinery SEPTEMBER 2014, Vol. 136 / 091004-3


This secondary R134a is a working fluid for a conventional vaporcompression refrigeration system.Flow rate though the test section is metered by a Coriolis flow

meter. Flow rates reported by the Coriolis flow meter have alsobeen cross-checked by a rotameter. Flow is routed to the rotatingframe of reference from the stationary frame of upstream anddownstream of the test section (along the axis of rotation). Vaporpressure of the refrigerant is controlled to 4.5 bar (abs) immedi-ately upstream of the upstream rotary union using a PID control-ler. Vapor temperature is controlled to 15 C to minimize ambientexternal water-vapor condensation on fluid pipe-related issues,using a PID controlled super-heater downstream of the boiler.Flow rate is controlled though a PID activated needle valve torange from 1.65 kg/min to 2.4 kg/min, resulting in a range ofReynolds numbers from 75,000 to 165,000.Rotating speeds of up to 600 rpm are achieved, yielding rotation

numbers of up to 0.4. The axis of rotation is horizontal. Specialefforts on balancing the assembly were necessary to ensure thesafety of the test section, the signal (TC) and power transmittingcables, the housing, and operating personnel.

Data Reduction. Silicone rubber heaters are used to generate heatwhich is dissipated within the test section. The resistance of eachheater, Res is measured before each run. The voltage, Vo volts sup-plied across each heater is measured by a multimeter. The thermalpower generated by the resistance heater is given by Vo2/Res watts.Each heater provides constant heat flux along its length. Chang-

ing the voltage across the heater changes the steady state tempera-ture measured on each copper plate. Since heat transfer coefficientsvary from region to region, it is not possible to control the tempera-ture recorded for each plate to a constant value. The voltages acrossthe heaters are actively controlled to ensure that the temperatures ata central region for each pass correspond to 65 C.An energy budget for each plate is performed to obtain the heat

transfer coefficient.

h Q=A q00loss

Tw Tb (1)

Here, q00loss is the heat loss, which is determined experimentally,based on special test. A low conductivity material is placed inside

the test section inhibiting heat transfer within the test section. Theheaters are energized with zero bulk flow. The voltage across theheaters is adjusted to control the steady state temperature at region4. Two steady state temperatures are studiedone lower than therange encountered in the test, and one higher. Based on these twotemperatures, a heat loss characteristic is obtained for each of thecopper plates. Heat loss tests are run at all RPM values.The local bulk mean temperature Tb is linearly interpolated

from the value measured at the exit of the third pass. A is the pro-jected (smooth) surface area which corresponds to the area of thecopper plate. The Nusselt number for the current study is definedas Nu hD/k, where k is the thermal conductivity of Freon R134avapor. A nondimensional analysis similar to that provided byWagner et al. [19] identifies the following parameters that governthe heat transfer phenomena under rotating conditions:

The rotation number

Ro XDV

(2)

Density ratio

DR Dqq

x

Tw TBTw TB =2

x

(3)

These two parameters can be combined together with the modelradius to diameter ratio to yield the Buoyancy number, Bo

Bo Ro2 DR R=D (4)The mean radius (R) corresponds to the distance between the axisof rotation and the center of the test section (Regions 3, 8, and13); the mean density ratio corresponds to the bulk mean tempera-ture at the center of the test section. The mean Reynolds numberis based on the mass flow rate at region 4.In order to isolate the effect of rotation and to eliminate the

Reynolds number effect, data for rotating conditions is presentedas a ratio, Nu/NuS, where NuS is the measured stationary Nusseltnumber data acquired for different runs (of Re and RPM) is corre-lated against Ro, at the corresponding buoyancy parameter asshown in Table 2. Since Bo is directly proportional to the square

Fig. 5 Refrigerant R134a vapor working loop schematic



of Ro (Eq. 4), dependency of Nu/NuS is not shown in this work.Thus, the obtained results are scalable to engine conditions.

Uncertainties. The main source of error in determining theReynolds number is the resolution of the Coriolis flow meter tomeasure the mass-flow rate (0.01 lb/min 0.00 45 kg/min) andtemperature measurements to estimate the properties of R134aaccurately. Reynolds numbers are thus estimated with an error ofless than 1%.The primary contribution to the error in Nusselt number mea-

surement is the thermocouple accuracy. Assuming a 0.1V voltageerror and a 1 C temperature measurement uncertainty (per infor-mation supplied by DAQ supplier), for a typical bulk temperaturedifference (30 C), the error in Nusselt number measurement(based on the KlineMcClintock [29] error propagation scheme)is less than 4%. The error estimate for the Nusselt numberenhancement due to rotation (NuNuS) is 8%. This value ishigher because of errors in the rotating Nusselt number as well asthe stationary Nusselt numbers.

The measured resistance of the heaters is found to change lessthan 0.5% for the operating range of temperatures. Additionalresistance contribution of the slip rings is measured to be negligible.

Results and Discussions

Stationary Channel Results. The Reynolds numbers studiedare depicted in Figs. 68. Table 1 shows that Reynolds numbergradually increases from the first passage to the third passage.Due to a reduction in hydraulic diameter going from the first passto the third (the channels have the largest cross section in theupstream portion of the serpentine channel, closer to the leadingedge portion of the blade), the Reynolds number graduallyincreases from the first pass to the third. A large volume of litera-ture exists on the subject of roughened channel heat transfer.Acquired results for stationary channels (reported for the leadingand trailing surfaces of the first passage, in open literature the45 deg inclined rib turbulators increase the heat transfer to around23 times of the DittusBoelter correlation (Eq. 5)) depending onReynolds numbers. For comparison with simple shaped channels,stationary NuS/Nu0 data with p/e 10, e/D 0.1, 45 deg rib anglefor Re 60,000200,000 from Rallabandi et al. [30] wereincluded in Fig. 6 at region 8. The current data show slightlyhigher than those of Ref. [30] due to complicated 180 deg tip turn-ing effect. The comparison of data trend is considered asacceptable.

Nu0 0:023Re0:8Pr0:4 (5)

Fig. 6 Variation of internal NuS/Nu0 for representative regionsin the first pass (Region 3), second pass (Region 8), and thirdpass (Region 13). Results from region 8 are compared withRef. [31].

Fig. 7 Variation of internal NuS/Nu0 in regions around the firstturn of the serpentine passage



It is noticed that heat transfer in the first passage is higher thanthe second and third passagesbefore the turn is higher than afterthe turn. This is due to the combinational effects of 45 deg ribangle and 180 deg sharp turn, as well as channel cross section andorientation between passages.Since the channel in consideration is a ribbed channel (and the

Reynolds numbers are very high), the flow in the channel is turbu-lent. This results in a relatively short development length and arelatively constant heat transfer coefficient throughout the chan-nel. Local increases in the heat transfer coefficient occur inregions 5, 6 and 10, 11, since these regions correspond with U-turns. This is consistent with results in the open literature with sta-tionary channels with U-turns.Asymmetry in the Nu/Nu0 results between the pressure and suc-

tion sides can be attributed to three factors: (a) geometrythetrapezoidal cross section of the channel results in a deviation insymmetry from the square/rectangular channel case, (b) staggeredrib turbulators, (c) vorticity triggered by relative angularity andchange in cross section between the first, second, and third passes.

Rotating Channel Results. Similar data has been obtained atthe three rpm values (0, 300, and 500) for Reynolds numbers

Fig. 8 Variation of internal NuS/Nu0 is regions around thesecond turn of the serpentine passage

Table 1 Nominal and actual Reynolds numbers in study

Nominal case 75 k 100 k 125 k 150 k

1st passage 72,696 94,820 118,526 142,8632nd passage 79,932 104,259 130,323 157,0833rd passage 84,142 109,751 137,188 165,358

Fig. 9 Nondimensional parameter range studied

Fig. 10 Effect of rotation: variation of internal Nu/NuS versusrotation number (Ro) for regions 3 (first pass), 8 (second pass),and 13 (third pass). Results from region 8 are compared withresults from Lei et al. [33].



ranging from 75,000 to 165,000. Rotation numbers of up to 0.4are obtained in the first pass. A detailed test parameter range ispresented in Fig. 9. The effect of rotation is best visualized byusing Nu/NuS as the ordinate and Ro as the abscissa. At Ro 0,the case in consideration is stationary and the corresponding Nu/NuS 1. Increasing Ro (simulated by increasing the rpm) effectsthe Nusselt number enhancement. This effect on the enhancementis dependent on the location under consideration.Figures 10, 11, and 12 show the effect of increasing the rotation

number on the measured heat transfer coefficient. For the parameterrange studied (for Re 100,000200,000; Ro 0.20.4), the effectof the Coriolis flow on the flow field seems reduced; rotationinduced asymmetry in heat transfer coefficient between the leadingand trailing surfaces is not very obvious, as compared with those inopen literature (for Re 10,00030,000, Ro 0.20.4).Table 2 shows the relations between Reynolds number, rotation

number, and local buoyancy parameter in the first, second, andthird passage, respectively. The effect of rotation in the first passis to increase the average heat transfer coefficient (as seen inFig. 10, Pass 1: Region 3). The flow in the third pass (region 13)is also radially outward, but rotation is observed to have a moremuted effect.This can be attributed to the lower rotation numbers in the third

pass and due to increased local Reynolds number. In the secondpass (Region 8), however, the effect of rotation is to suppress theheat transfer coefficient up to 15%. This can be attributed to rota-tional (centrifugal) buoyancy. Since the centrifugal force actsradially outward (i.e., heavier fluid particles are pushed outwards)in a rotating frame of reference, the buoyant force has an oppositedirection. Warmer (i.e., lighter) fluid pockets are subject to a netforce which acts radially inward in all the three passes. In the firstand third passes, the buoyant force (directed radially inward) is

opposing the bulk fluid flow. In the second pass, the buoyant forcealigns itself with the bulk fluid flow.A survey of literature [31,32] indicates that a reduction in heat

transfer coefficient is indeed expected in turbulent internal flowfields when buoyancy is aligned with bulk flow. This explains thereduction in heat transfer in the second (radially inward) pass.This phenomenon is attributed to the suppression of turbulencegeneration due to increase in momentum aligned to the bulk flow

Fig. 11 Effect of rotation: variation of internal Nu/NuS versusrotation number (Ro) along turn between the radially outwardfirst pass and the radially inward second pass

Fig. 12 Effect of rotation: variation of internal Nu/NuS versusrotation number (Ro) along the turn between the radially inwardsecond pass and the radially outward third pass

Table 2 Rotation numbers and local buoyancy parameters

Rotate speed 300 rpm 500 rpm

Region Re 75 k 100 k 125 k 150 k 75 k 100 k 125 k

R3 Ro 0.2388 0.1831 0.1464 0.1215 0.3995 0.3063 0.2450Bo 0.3401 0.2062 0.1986 0.1427 0.9358 0.5828 0.3875

R5 Ro 0.2388 0.1831 0.1646 0.1215 0.3995 0.3063 0.2450Bo 0.2576 0.1585 0.1083 0.0794 0.6719 0.3922 0.2823

R6 Ro 0.1745 0.1338 0.1070 0.0888 0.2919 0.2238 0.1791Bo 0.1131 0.0748 0.0519 0.0379 0.3003 0.1758 0.1258

R8 Ro 0.1745 0.1338 0.1070 0.0888 0.2919 0.2238 0.1791Bo 0.1733 0.1061 0.0731 0.0528 0.4831 0.2789 0.1946

R10 Ro 0.1745 0.1338 0.1070 0.0888 0.2919 0.2238 0.1791Bo 0.2968 0.1785 0.1154 0.0813 0.8463 0.4969 0.3249

R11 Ro 0.0770 0.0591 0.0472 0.0392 0.1289 0.0988 0.0791Bo 0.0285 0.0177 0.0117 0.0087 0.0837 0.0505 0.0346

R13 Ro 0.0770 0.0591 0.0472 0.0392 0.1289 0.0988 0.0791Bo 0.0387 0.0239 0.0162 0.0119 0.1001 0.0613 0.0426



direction in warm fluid pocket. The effect is opposite in radiallyoutward passesresulting buoyant force in warm fluid pockets(close to the wall) opposes the bulk flow and therefore increasesshear and generates turbulence increasing heat transfer coefficients.Similar measurements have been made by Wagner and Johnson [19].For comparison with simple shaped channels, rotating Nu/NuS

data with p/e 8, e/D 0.1, 45 deg rib angle for Re 10,00040,000 from Lei et al. [33] were included in Fig. 10 atregion 8. The current data show much lower Nu/NuS values thanthose of Ref. [33]. This might be due to high Reynolds numbersand complicate 180 deg tip turning effect. This implies rotatingdata from the simple multipass channels with same channel angleto rotation direction, such as rectangular-to-rectangular multipasschannels at Reynolds numbers from 10,000 to 40,000 and rotationnumbers up to 0.3 [33] might be different from the complicatereal multipass channels with different channel angle to rotationdirection, such as trapezoidal-to-trapezoidal channels at Reynoldsnumbers from 75,000 to 165,000 and rotation numbers up to 0.3.Comparing Figs. 11 and 12, a qualitative reduction in heat

transfer rates due to rotation is observed in the turn correspondingwith the hub region (regions 10 and 11). The trend in Fig. 11(the turn corresponding with the tip region) is reverse; rotationincreases the heat transfer in the internal passage near the tipregion. This difference can be attributed to the arrangement of thebuoyancy driven secondary flows, which, as discussed in the pre-vious paragraph, are always directed towards the hub. The warmerfluid in the buoyancy driven boundary layer impinges on thehub region, suppressing the internal cooling ability at the hub.

Conclusions

A significant amount of internal channel heat transfer coeffi-cient data is collected on a 1:1 scaled turbine blade replica in thelab at simulated engine conditions for high Reynolds numbersfrom 75,000 to 165,000 and high rotation numbers up to 0.4. Theeffects of rotation (i.e., the composite of the Coriolis and theBuoyancy effect) are studied. Results obtained reflect the compli-cated nature of flow field (the effect of rib turbulators, channelshape, sharp turns, rotation etc.).

(1) Using R134a as a working fluid, prominent trends in openliterature (obtained using air as a working fluid) can bereproduced. Nondimensional numbers corresponding toactual engine conditions (high Reynolds numbers and highrotation numbers) can be more easily simulated with R134adue to its higher density.

(2) Rib turbulators are found to increase the heat transfercoefficient to around 23 times the value predicted by theDittusBoelter correlation. The NuS/Nu0 values decreasewith Reynolds numbers, a trend widely documented inopen literature.

(3) For the high Reynolds numbers studied, the effect of theCoriolis force in generating asymmetry between the leadingand trailing surfaces is reduced.

(4) Heat transfer is found to be relatively suppressed in the sec-ond pass up to 15%, which has radially inward flow. Thiseffect is attributed to the suppression of turbulence genera-tion due to the alignment of the buoyancy force acting onthe warmer fluid pockets and the bulk flow.

(5) Heat transfer is found to be suppressed in the hub regionpressure side (the region between the second and the thirdpasses) up to 25%. This is also attributed to buoyancyeffects. This heat transfer reduction is critical for turbinecooling designs.

Acknowledgment

This project was sponsored by Siemens Energy Company (theproject initiated by Sanjay Chopra in 2004). The authors from

Texas A&M University would like to acknowledge and extendtheir gratitude to Randy Tucker, Lesley Wright, Mike Huh, andYao-Hsien Liu who have made contributions to the project duringtheir study in TAMU-Mechanical Engineering. The project wasalso partially supported by TAMU-Marcus Easterling endowmentfund.

Nomenclature

D hydraulic diameter of channelh heat transfer coefficient, W/m2 Kk thermal conductivity of Freon R134a vapor

Nu Nusselt numberNuS measured Nusselt number for stationary case at same

Reynolds numberNu0 number for smooth duct, based on the DittusBoelter

correlation (Eq. 5)Pr Prandtl numberPS trailing side of internal passages, corresponding to blade

external pressure sideQ heat dissipated by heater, W

q00loss local heat loss, W/m2R mean radiusRe Reynolds number in channel (for pass consideration)Res electrical resistance of heater, ohmsRo rotation numberSS leading side of internal passages, corresponding to blade

external suction sideTb local bulk mean temperatureTi test section inlet temperature, 22 CTw local wall temperatureV fluid velocity, m/s

Vo voltage, voltsl dynamic viscosityq densityX angular velocity, rad/s

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[8] Su, G., Chen, H., Han, J., and Heidmann, J., 2004, Computation of Flow andHeat Transfer in Rotating Two-Pass Rectangular Channels (AR 1:1, 1:2, and1:4) With Smooth Walls by a Reynolds Stress Turbulence Model, Int. J. HeatMass Transfer, 47(26), pp. 56655683.

[9] Taslim, M., and Lengkong, A., 1998, 45 Deg Staggered Rib Heat TransferCoefficient Measurements in a Square Channel, ASME J. Turbomach., 120(3),pp. 571580.

[10] Bunker, R., and Bailey, J., 2003, Heat Transfer and Friction in Channels WithVery High Blockages 45 Degree Staggered Turbulators, ASME Paper No.GT2003-38611.

[11] Rallabandi, A. P., Alkhamis, N., and Han, J. C., 2011, Heat Transfer andPressure Drop Measurements for a Square Channel With 45 Deg RoundEdged Ribs at High Reynolds Numbers, ASME J. Turbomach., 133(3),p. 031019.

[12] Han, J., Zhang, Y., and Lee, C., 1991, Augmented Heat Transfer in SquareChannels With Parallel, Crossed, and V-Shaped Angled Ribs, ASME J. HeatTransfer, 113(3), pp. 590596.



[13] Han, J., and Zhang, Y., 1992, High Performance Heat Transfer Ducts WithParallel Broken and V-Shaped Broken Ribs, Int. J. Heat Mass Transfer, 35(2),pp. 513523.

[14] Taslim, M., and Spring, S., 1994, Effects of Turbulator Profile and Spacing onHeat Transfer and Friction in a Channel, J. Thermophys. Heat Transfer, 8(3),pp. 555562.

[15] Han, J., Glicksman, L., and Rohsenow, W., 1978, An Investigation of HeatTransfer and Friction Factor for Rib-Roughened Surfaces, Int. J. Heat MassTransfer, 21(3), pp. 555562.

[16] Chandra, P., Fontenot, M., and Han, J., 1998, Effect of Rib Profiles on Turbu-lent Channel Flow Heat Transfer, J. Thermophys. Heat Transfer, 12(1), pp.116118.

[17] Dipperey, D. and Sabersky, R. H., 1963, Heat and Momentum Transfer inSmooth and Rough Tube at Various Prandtl Numbers, Int. J. Heat Mass Trans-fer, 6(5), pp. 329353.

[18] Han, J., 1984, Heat Transfer and Friction in Channels With Two OppositeRib-Roughened Walls, ASME J. Heat Transfer, 106(4), pp. 774781.

[19] Wagner, J., Johnson, B., and Kopper, F., 1991, Heat Transfer in Rotating Ser-pentine Passages With Smooth Walls, ASME J. Turbomach., 113(3), pp.321330.

[20] Wagner, J., Johnson, B., Graziani, R., and Yeh, F., 1992, Heat Transfer inRotating Serpentine Passages With Trips Normal to the Flow, ASME J. Tur-bomach., 114(4), pp. 847857.

[21] Johnson, B., Wagner, J., Steuber, G., and Yeh, F., 1994, Heat Transfer inRotating Serpentine Passages With Trips Skewed to the Flow, ASME J. Tur-bomach., 116(1), pp. 113123.

[22] Johnson, B., Wagner, J., Steuber, G., and Yeh, F., 1994, Heat Transfer inRotating Serpentine Passages With Selected Model Orientations for Smooth orSkewed Trip Walls, ASME J. Turbomach., 116(4), pp. 738744.

[23] Guidez, J., 1989, Study of the Convective Heat Transfer in a Rotating CoolantChannel, ASME J. Turbomach., 111(1), pp. 4350.

[24] Dutta, S., and Han, J. C., 1996, Local Heat Transfer in Rotating Smooth andRibbed Two-Pass Square Channels With Three Channel Orientations, ASMEJ. Heat Transfer, 118(3), pp. 578584.

[25] Azad, G. S., Uddin, M. J., Han, J. C., Moon, H. K., Glezer, B., 2002, HeatTransfer in a Two-Pass Rectangular Rotating Channel With 45-Deg Angled RibTurbulators, ASME J. Turbomach., 124(2), pp. 251259.

[26] Fu, W., Wright, L., and Han, J.-C., 2006, Rotational Buoyancy Effects on HeatTransfer in Five Different Aspect Ratio Rectangular Channels With SmoothWalls and 45 Degree Ribbed Walls, ASME J. Heat Transfer, 128(11), pp.11301141.

[27] Liu, Y., Huh, M., Han, J., and Chopra, S., 2008, Heat Transfer in a Two-PassRectangular Channel (AR 1:4) Under High Rotation Numbers, ASME J.Heat Transfer, 130(8), p. 081701.

[28] Huh, M., Lei, J., and Han, J., 2012, Influence of Channel Orientation on HeatTransfer in a Two-Pass Smooth and Ribbed Rectangular Channel (AR 2:1)Under Large Rotation Numbers, ASME J. Turbomach., 134(1), p. 011022.

[29] Kline, S. J., and McClintock, F. A., 1953, Describing Uncertainty in Single-Sample Experiments, ASME J. Mech. Eng., 75(1), pp. 38.

[30] Rallabandi, A. P., Yang, H., and Han, J. C., 2009, Heat Transfer and PressureDrop Measurements for a Square Channel With 45 Deg Ribs at High ReynoldsNumbers, ASME J. Heat Transfer, 131(7), p. 071701.

[31] Mceligot, D., and Jackson, J., 2004, Deterioration Criteria for Convective HeatTransfer in Gas Flow Through Non-Circular Ducts, Nucl. Eng. Des., 232(3),pp. 327333.

[32] Lee, J., Hejzlar, P., Saha, P., Stahle, P., Kazimi, M., and Mceligot, D.,2008, Deteriorated Turbulent Heat Transfer (DTHT) of Gas Up-Flow in aCircular Tube: Experimental Data, Int. J. Heat Mass Transfer, 51(13), pp.32593266.

[33] Lei, J., Li, S. J., Han J. C., Zhang, L., Moon, H. K., 2013, Heat Transfer inRotating Multi-Pass Rectangular Ribbed Channel With and Without a TurningVane, ASME J. Heat Transfer, 135(4), p. 041903.



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