Heat transfer engineering an international journal, tập 31, số 7, 2010

[43] investigated experimentally the effects of squealer or winglet-squealer tip and tip clearance on the aver-age and local mass transfer coefficients for a large-scale gas turbine blad

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C Taylor and Francis Group, LLC

ISSN: 0145-7632 print / 1521-0537 online

DOI: 10.1080/01457630903425320

Gas Turbine Blade Tip Heat Transfer and Cooling: A Literature Survey

BENGT SUNDEN and GONGNAN XIE

Department of Energy Sciences, Lund University, Lund, Sweden

Gas turbines are widely used for aircraft propulsion, land-base power generation, and other industrial applications like

trains, marines, automobiles, etc To satisfy the fast development of advanced gas turbines, the operating temperature must

be increased to improve the thermal efficiency and output work of the gas turbine engine However, the heat transferred to the

turbine blade is substantially increased as the turbine inlet temperature is continuously increased Thus, it is very important

to cool the turbine blades for a long durability and safe operation Cooling the blade must include cooling of the key regions

being exposed to the hot gas The blade tip region is such a critical area and is indeed difficult to cool This results from the

tip clearance gap where the complex tip leakage flow occurs and thereby local high heat loads prevail This paper presents a

literature survey of blade tip leakage flow and heat transfer, as well as research of external and internal cooling technologies.

The present paper does not intend to review all published results in this field, nor review all papers from the past to now This

paper is limited to a review of recently available published works by several researchers, especially from 2001 to present,

concerning blade tip leakage flow associated with heat transfer, and external or/and internal tip cooling technologies.

INTRODUCTION

A gas turbine is an engine designed to convert the energy

of a fuel into some form of useful power, such as shaft power

or thrust Today, gas turbines (GTs) are widely used in aircraft

propulsion, land-based power generation, and other industrial

applications For example, GTs are used to power commercial

airplanes, marines, trains, electric power generators,

automo-biles, and gas pipeline compressor drivers Figure 1 illustrates

a commercial gas turbine engine The reasons that gas turbine

engines are widely used for aircraft propulsion include that they

are light, compact, and have a high power-to-weight ratio As

shown in Figure 1, there are three main components of a gas

turbine engine: compressor, combustor, and turbine The

com-pressor is used to compress the intake air to a specific high

pres-sure, the combustor is used to burn the input fuel and produce

the high temperature gas, and the turbine extracts the energy of

the gas and converts it into power work A number of

compo-nents sometimes occurs in the gas turbine system to improve the

The authors acknowledge financial support from the TURBO POWER

con-sortium funded by the Swedish Energy Agency (STEM), SIEMENS Industrial

Turbomachinery, and VOLVO AERO Corporation.

Address correspondence to Professor Bengt Sunden, Division of Heat

Trans-fer, Department of Energy Sciences, Lund University, PO Box 118, S-22100,

Lund, Sweden E-mail: Bengt.Sunden@energy.lth.se

network output or thermodynamic efficiency, e.g., intercooler,recuperator, regenerator, and combustion reheater However, thebalance of additional power, efficiency, cost, complexity, dura-bility, compactness, etc must be carefully evaluated

The temperature–entropy diagram for the basic cycle of agas turbine engine with friction is shown in Figure 2 The idealstandard cycle is assumed to be adiabatic, reversible, and fric-tionless The overall thermodynamic efficiency depends on theefficiencies of all components, such as compressor efficiency,turbine efficiency, and combustion efficiency Clearly, the tur-bine efficiency will affect the cycle efficiency at some degree.Thus, improving the turbine efficiency will help to improve theoverall performance of a gas turbine engine, while losses inturbine efficiency and/or output work will reduce the overallperformance of the system Apart from the component efficien-cies, the operating temperature of gas turbine system affects theoverall performance

It is well recognized that one way to increase the power put and thermodynamic efficiency of a gas turbine engine is toincrease the turbine inlet temperature (TIT) From the principles

out-of engineering thermophysics [1, 2], the reason is that at a fixedpressure ratio the net work output of a gas turbine increases

with increasing turbine blade (also called rotor) inlet

temper-ature Figure 3 shows recent development of TIT from 1950

to 2010 Current advanced gas turbine engines are operating atTIT of about 1200–1500◦C To pursue higher power, the inlet

527

Figure 1 Gas turbine illustration From http://www.epower-propulsion.com/

epower/gallery/GasTurbines.htm.

temperature should be raised increasingly to higher certain

tar-gets For example, to double the power of the aircraft, the TIT

should be increased from 1500 to 2000◦C

However, the heat transferred to the blade increases with the

increase of the blade inlet temperature, and the allowable

melt-ing temperature of materials increases at a slower rate This

means that the turbine blade inlet temperature may exceed the

material melting temperature by more than 500◦C Thus, it is

critical to cool turbine blades for a safe and long-lasting

oper-ation The blades can only survive if effective cooling methods

are used Various internal and external cooling techniques are

employed to decrease the blade material temperature below its

melting point Figure 4 depicts the typical cooling technology

for internal and external zones The leading edge is cooled by jet

impingement with film cooling, the middle portion is cooled by

internal serpentine ribbed-turbulators passages, and the trailing

edge is cooled by pin-fins with ejection In internal cooling, the

relatively cold air, bypassed/discharged from the compressor,

is directed into the hollow coolant passages inside the turbine

blade In external cooling, the bypassed air is ejected through

those small holes, which are located in the turbine blade

dis-cretely The commonly used cooling technique for the

high-pressure turbine blade is a combination of internal and film

cooling Most recent developments in TIT increase have been

Figure 2 Temperature–entropy diagram for a basic gas turbine cycle.

2400

2000

1600

1200 2600

Film,Impingement

New cooling concept Projected trend new material

Figure 3 Developments of gas turbine inlet temperature over recent years Reproduction from Rolls Royce plc.

achieved by better cooling of the turbine blade and have proved the understanding of the heat transfer mechanisms in theturbine passages Several recent publications reviewing the gasturbine heat transfer and cooling technology investigations areavailable These include a relevant book [3], edited volumes [4,5], and journal papers [6–9]

im-Film cooling

Trailing edge ejection

Impingement

Tip cap cooling

Rib turbulated cooling

Hot gas

Cooling air

Internal convective cooling

Figure 4 Typical cooling techniques for a blade.

B SUNDEN AND G XIE 529

Figure 5 Clearance gap and leakage flow of unshrouded turbine blade.

Cooling of the blade should include the cooling of all regions

exposed to high-temperature gas and thermal load Among such

regions, particularly for high-pressure turbines, is the blade tip

area Gas turbine blades usually have a clearance gap between

the blade tip and stationary casing or the shroud (as

schemat-ically shown in Figure 5) The clearance gap is necessary to

allow for the blade rotation and for its mechanical and thermal

expansion However, due to the pressure difference between the

pressure side and suction side, the hot gas leaks through the

gap This is known as the tip leakage flow The leakage flow

is undesirable, because it is associated with the generation of

a secondary flow resulting in reduction of the work done and

hence of the overall efficiency, and results in higher heating at

the high pressure tip corner from mid-chord to trailing edge The

hot leakage flow increases the thermal loads on the blade tip,

leading to high local temperature Thus, it is essential to cool the

turbine blade tip and near the tip regions However, it is difficult

to cool such regions and to seal against the hot leakage flow

The blade tip operates in an environment between the rotating

blade and the stationary casing, and experiences the extremes of

the fluid-thermal conditions within the turbine [10–12] A more

detailed discussion of the blade tip can be found in [13]

Because the blade lifetime may be reduced by a factor of 2

if the blade metal temperature prediction is off by only 30◦C, it

is very critical to predict accurately the local heat transfer and

local blade temperature to prevent hot spots and thus increase

the turbine blade life It is important for the gas turbine designers

to know the effects of increased heat load in the area exposed to

hot gas and be able to design efficient cooling schemes to protect

the blade Therefore, fundamental and detailed studies of heat

transfer and flow relating to the blade tip or near blade tip regions

are needed to provide better understanding and prediction of the

heat loads on such regions accurately

Besides conventional techniques of experimental

measure-ments with advanced apparatus, computational fluid dynamics

(CFD) plays an increasingly important role in design and

re-search studies of gas turbines During the past two decades,

CFD has been developed so rapidly that many advanced

computational codes and commercial softwares have uously appeared for solving the heat transfer and flow field ofcomplex geometries like gas turbine passages By validating thecodes with experimental data, many computational results based

contin-on CFD are accurate and reliable This will ccontin-ontribute to theprediction and design of turbomachinery components, withoutdoubt, including the turbine blade and its tip The highly accuratecomputational results can contribute to the design and manufac-ture of gas turbine blades and improve the durability and safeoperation

This paper does not and cannot review all the interestingand important progress related to gas turbine heat transfer andcooling (some may be found in [1–5]), but tries to summarizethe recently published results in the concerned field of blade tipheat transfer and development of cooling technology The firststudies on blade tip heat transfer were reviewed earlier [11–13]

In this paper, published literature from 1995 to 2008 and on,especially the recent years 2001–2008, are reviewed

This paper is organized as follows Gas turbine heat transferand the need for cooling techniques are introduced first Then theblade tip leakage complicated flow associated with heat transfer

on tips or near the tip regions is reviewed Next the ment of external tip cooling methodology is reviewed, whilethe last section reviews the development of internal tip coolingmethodology A summary is presented in the final section

develop-BLADE TIP LEAKAGE FLOW AND HEAT TRANSFER Generic Flow Pattern Associated With Tip Leakage Flow

The flow field in a turbine is very complex It is stronglythree-dimensional, unsteady, and viscous, with several types ofsecondary flows, endwall flows, and vortices (passage vortex,counter vortex, horseshoe vortex, leakage flow vortex, etc.).Transition flow and high turbulence intensity result in additionalcomplexities Figure 6 depicts the complex flow phenomenaheat transfer engineering vol 31 no 7 2010

Figure 6 Complex flow and heat transfer phenomena in the turbine gas path

[14].

in a turbine blade gas path [14] The understanding of such

complex flow fields and heat transfer characteristics is necessary

to improve the blade design and prediction in terms of efficiency

as well as the evaluation of mechanical and thermal fatigue Tip

leakage flow is a dominant source of unsteadiness and

three-dimensionality of the flow in turbomachineries As depicted in

Figure 7, the tip leakage flow passes through the tip clearance

driven by the pressure gradient between the pressure side and

suction side Also, the leakage flow tends to roll up into a vortex

and interacts with the secondary flow Thus, the leakage flow

and its interaction with other flow features show very complex

phenomena

A perfect blade tip will not allow any leakage flow, and no

secondary flows to reduce stage efficiency will be generated,

nor losses for downstream stages created, and cooling is not

required Thereby no thermodynamic losses occur [11] Thus,

the two main objectives of blade design are to reduce the leakage

Figure 7 Schematics of blade tip leakage flow characteristics [11].

Figure 8 Different kinds of blade tips [11].

flow as much as possible and to cool the blade tips using smallquantities of extracted cooling air However, all the blade tips inmodern gas turbines do allow some leakage flow and secondaryflows are generated Today, there are several major types ofblade tips: (a) flat tip, (b) recessed tip with peripheral squealersealing rims, and (c) attached tip shrouds [11], as shown inFigure 8 Each blade tip has its advantages and disadvantages.Although it is easy to design a flat tip and its cooling scheme,very few turbines use flat tips High leakages lead to bad tip

B SUNDEN AND G XIE 531

Figure 9 Streamlines of blade tip flow pattern [15].

aerodynamics and results in higher heat loads on the tip A

recessed tip with sealing rims is the most common design in

practice today for high-pressure turbine blades A recessed tip

with rim reduces the risk of blade damage if the tip rubs against

the shroud; however, the design of a recessed tip is more complex

because of the cooling of the rim and the need to prevent losses

by oxidation and erosion Blades with attached tip shrouds are

mostly used in low-pressure turbine blades This tip has the

lowest aerodynamic loss when properly installed, but it requires

greater attention to stresses because of the heavier weight and

requires a more complex cooling system

Ameri et al [15] performed calculations on flow and heat

transfer of a GE-E3rotor tip considering three types: plane tip,

2% recess tip, and 3% recess tip A two-dimensional (2D) cavity

flow problem was used to validate the k-ω turbulence model.

These authors found two dominant flow structures in the recess

region, which strongly affect the heat transfer rate, as shown

in Figure 9, but no significant effect on the adiabatic efficiency

was observed for these three tips Also, Ameri et al [16] studied

the effects of tip clearance and casing recess on heat transfer

and stage efficiency in axial turbines Their numerical study

reconfirmed a linear relationship between the efficiency and the

tip gap height Introduction of a recessed casing resulted in a

drop in the rate of heat transfer on the pressure side, and a

marked reduction of the heat load and peak values on the blade

tip Ameri et al observed that the recessed casing has a small

effect on the efficiency but can have a moderating effect on the

flow underturning at smaller tip clearances

Experimental Measurements for Tip Region Flowfield

and Heat Transfer

Bunker et al [17], and Ameri and Bunker [18] reported

results of a combined experimental and simulation study

de-signed to investigate the detailed distribution of the convective

heat transfer coefficient on the flat tip surface with both sharp

and rounded edges for a large power generation turbine This

study showed good agreement between experiments and

com-Figure 10 Sharp and rounded edge tip heat transfer coefficients [18].

putations Figure 10 presents a sample of these experimentaland computed tip heat transfer coefficients for the sharp androunded edge tips Ameri [19] also conducted experimental andnumerical studies of detailed heat transfer coefficient distribu-tion on the rounded blade tip of a gas turbine equipped with

a mean-camberline strip Generally good agreement betweenexperimental data and computations was achieved, as shown

in Figure 11 Results showed that the mean-camberline stripcould reduce the tip leakage flow but the total pressure loss wasnot reduced comparatively, and the sharp edge tip was better inheat transfer engineering vol 31 no 7 2010

Figure 11 Heat transfer coefficient and flow pattern for blade tip with

mean-camberline strip [19].

reduction of the tip leakage flow and tip heat transfer compared

to the rounded edge tip

Thorpe et al [20, 21] reported experimental measurements of

time-mean/time-resolved heat transfer and static pressure on the

over-tip casing of a transonic axial flow turbine They presented

axial and circumferential distributions of the heat transfer rate

as well as adiabatic wall temperature, Nusselt number, and static

pressure They found that the rate of heat transfer to casing wall

and the wall temperature varied strongly with axial position

through the rotor, and the effects of the vane exit flow features

were small Through assessments of the relative importance

of different time varying phenomena to the casing heat load

distribution, they concluded that up to half of the casing heat

load was associated with the tip leakage flow Also, discussion

about shroudless turbine design accounting for the high heat flux

was addressed Thorpe et al [22] also experimentally studied the

blade tip heat transfer and aerodynamics in a transonic turbine

stage They observed high heat transfer rates near the nose of the

blade tip and also in the region of high blade lift near the

mid-axial chord, and proposed three primary mechanisms: vane–

shock interaction, relative total temperature fluctuations, and

fluctuations in tip leakage flow speed and direction driving the

unsteady heat transfer

Chana and Jones [23] presented detailed experimental

mea-surements of heat transfer and static pressure distributions on

the shroudless rotor blade tip and casing with and without inlet

nonuniform temperatures Also, a simple 2D model was

devel-oped to estimate the heat transfer rate to tip and casing as a

function of Reynolds number Results showed that the overall

heat load was reduced with inlet nonuniformity, that the highest

heat transfer rate was on the pressure side of the blade where the

highest random unsteadiness was marked, and that the average

static pressures did not show significant difference between the

two cases Camci et al [24] investigated experimentally

aerody-namic characteristics of full and partial length squealer rims in

an axial turbine Figure 12 shows a schematic picture of partial

Figure 12 Geometries of partial squealer rims [24].

squealer rims studied Results showed that the partial squealerrim could seal the tip effectively, and a mid-size partial rim wasmost effective in reducing the tip leakage flow Compared tothe two studied channel arrangements having partial rims nearthe corners of the suction and pressure sides, the sealing perfor-mance of the mid-size rim on the suction side was even better.This indicated that the partial squealer rims on the suction sidewere capable of reducing the exit total pressure loss by the tipleakage flow to a significant degree Key and Arts [25] studiedthe tip leakage flow characteristics for flat and squealer turbinetips The experiments were conducted at different Reynoldsnumber and Mach number conditions for a fixed value of thetip gap in a nonrotating, linear cascade arrangement Oil flowvisualization was used, as shown in Figure 13, and the staticpressure and aerodynamic loss were measured These authorsfound that the squealer tip showed a significant decrease in ve-locity through the tip gap, and for the flat tip the increase ofReynolds number would cause an increase in the tip velocitylevel whereas for the squealer tip the sensitivity was not much.Their data are valuable for validation of CFD computations,and in turn CFD can provide insight to some details of the flowphysics in the tip region

Azad et al [26, 27] and Teng et al [28] measured the heattransfer coefficient and static pressure distributions on gas tur-bine tips in a five-bladed stationary linear cascade Various re-gions of high and low heat transfer coefficients at the tip surfacewere observed The heat transfer coefficients increase with anincrease of the inlet turbulence intensity Compared to the flat

B SUNDEN AND G XIE 533

Figure 13 Flow visualization of squealer and flat tip [25].

tip, the squealer tip showed a lower overall heat transfer

coef-ficient Also, a reduced tip gap clearance resulted in a weaker

unsteady wake effect on the blade tip heat transfer and a

reduc-tion in the heat transfer coefficient over the blade tip surface

Azad et al [29] also studied the effect of the squealer geometry

arrangement on a blade tip The detailed heat transfer

coeffi-cient distributions of six tip geometry cases were obtained It

was shown that the suction-sided squealer could provide a

bet-ter benefit compared to other cases, and the mid-chamber lined

squealer behaved better than the pressure-sided squealer Also,

a single squealer provided better performance in reducing the

overall heat transfer than a double squealer

Dhadwal and Kurkov [30] used a dual-laser probe integrated

fiber optic system to measure the blade tip clearance in a

ro-tating turbomachinery A symmetric configuration of the probe

installation could offer better resolution The time-of-flight

mea-surements were robust and reliable Saxena et al [31] presented

a comprehensive investigation of the effect of various tip sealing

geometries on the blade tip leakage flow and heat transfer of a

scaled up high-pressure turbine They found that compared to

other geometries, the tripped strips placed against the leakage

flow (as shown in Figure 14a) led to the lowest heat transfer

on the tips with a reduction of 10–15% The use of strips and

pin-fins did not decrease the tip surface heat transfer

coeffi-cients Saxena and Ekkad [32] also experimentally investigated

the effect of squealer tip geometries on the blade tip leakage

and associated heat transfer in the same facility It was found

that the suction-sided squealer rim might be favorable for

re-ducing the heat transfer coefficients on the tip surface, whereas

the pressure-sided squealer did not reduce the heat transfer and

behaved like the plane tip Nasir et al [33] also investigated

the effect of tip gap and squealer geometry on the detailed

Figure 14 Blade tip geometries for test [31].

heat transfer over a high pressure turbine rotor blade tip Thesquealer studied altered the tip gap flow significantly and henceresulted in lower heat transfer coefficient Also, experimentalresults showed that some partial burning of the squealers might

be good for overall reduction in the heat transfer coefficient.Rhee and Cho [34, 35] experimentally measured localheat/mass transfer characteristics on tip, shroud, and near-tipsurface of a rotating blade in a low-speed annular cascade Theeffects of rotation and incoming flow incidence angle were ex-amined Results showed that the heat transfer was complex withcomplicated flow patterns such as flow acceleration, laminariza-tion, transition, separation, and tip leakage flow, and the bladerotation caused increased incoming flow turbulence intensitywhile the tip leakage flow was reduced Also, they found thatthe heat/mass transfer coefficients were about 1.7 times thanthose on the blade surface and shroud, and due to the reducedtip leakage flow under rotation the heat/mass transfer coeffi-cients on the tip slightly decreased while they remained similar

on the shroud With a positive incidence angle, more uniformand higher heat transfer rate were found on the tip because ofthe increased tip gap flow and high flow angle Rhee and Choheat transfer engineering vol 31 no 7 2010

Figure 15 Blade geometries for test [40].

[36, 37] experimentally studied the effect of vane/blade position

on heat transfer in a stationary blade and shroud in a low-speed

wind tunnel They presented detailed mass transfer

measure-ments, and the results showed that the mass transfer coefficients

in the upstream region varied up to 25% due to the blockage

ef-fect as the blade position changed The size and level of the peak

region were affected strongly Also, distinctly different patterns

near the blade tip were observed due to the variation in the tip

leakage flow

Matsunuma [38] observed the effect of Reynolds number

and freestream turbulence on turbine tip clearance flow

Three-dimensional flow fields at the exit of the turbine with and without

tip clearance were measured Results indicated that variations

in Reynolds number and freestream turbulence intensity did notaffect the mass-averaged tip clearance loss Due to the stronginteraction between the leakage vortex and tip-side passage vor-tex, the decrease in flow angle at lower Reynolds numbers waslarger than that at higher Reynolds numbers Kwak and Han[39] and Kwak et al [40] conducted a series of measurements

on the tip and near-tip region heat transfer coefficients of aturbine blade with flat or squealer tip, and the effects of rim lo-cation and height as well as tip clearance on heat transfer weremeasured The geometry is shown in Figure 15 The blade tipclearance was 1.0%, 1.5%, and 2.5% and the rim height was2.1%, 4.2%, and 6.3% of the blade span, as shown in Figure 15.Experimental results showed that the heat transfer coefficients

B SUNDEN AND G XIE 535

on the tip surface were higher than those on the shroud and on

the near-tip region of the pressure and suction sides, and with an

increase of the tip clearance the heat transfer on the tip surface

increased whereas heat transfer on the shroud and the suction

side first increased and then decreased On the blade pressure

side the heat transfer coefficient was kept constant They also

found that higher rims could reduce the heat transfer coefficient

on the tip and shroud, while on the pressure side and suction side

the reduction was not significant The suction-sided rim could

provide lower heat transfer coefficient on the tip and near-tip

region than the double-sided rim case Kwak and co-workers

[41, 42] also performed measurements on detailed heat transfer

coefficients on the squealer tip and near-tip region of a turbine

blade Results showed that the overall heat transfer coefficients

on the squealer tip were higher than those on the shroud

sur-face and the near-tip region of the pressure side and suction

side Near the tip region the heat transfer coefficient showed

no significant reduction Also, the suction-sided squealer tip

re-vealed the lowest heat transfer coefficients on the blade tip and

near tip

Papa et al [43] investigated experimentally the effects of

squealer or winglet-squealer tip and tip clearance on the

aver-age and local mass transfer coefficients for a large-scale gas

turbine blade, and used the heat–mass analogy to obtain heat

transfer coefficients Flow visualization on the tip surface was

presented Compared to the winglet-squealer tip, the squealer

tip provided a higher average mass/heat transfer coefficient

Rehder and Dannhauer [44] studied the effect of the tip leakage

flow on the three-dimensional (3D) flow field and end-wall heat

transfer Results showed that when the leakage mass flow rate

increased from 1% to 2%, significant changes in the secondary

and end-wall heat transfer occurred The secondary flow was

amplified as the leakage flow was ejected perpendicular to the

main flow direction, whereas it was reduced significantly as

the leakage flow was ejected tangentially Govardhan et al [45]

investigated the 3D flow in a large deflection turbine cascade

with tip clearance 0.08%, 1.5%, and 3.0% of the chord They

found that there was a strong horseshoe vortex in front of the

leading edge for 0.08% clearance, while for 3% clearance there

was no vortex A small tip separation vortex was also observed

on the tip surface, which made the flow from the pressure side

to be accelerated The passage vortex did not diminish as the

tip clearance increased Also, Govardhan et al [46] investigated

the effect of endwall and tip clearance on the flow in a

two-dimensional turbine rotor blade cascade Five incidence angles

were chosen:−10, −5, 0, 5, and 10◦ Results showed that as

the tip clearance was increased the adverse pressure gradient

upstream the leading edge was reduced, and with the increase

of incidence angle the blade loading due to the static pressure

gradient also increased

Porreca et al [47] conducted experimental and numerical

in-vestigation on flow dynamics and performance of partially and

fully shrouded axial turbines, as shown in Figure 16

Experi-mental results showed that for the partial shroud case a strong

tip leakage vortex was developed from the first rotor and

trans-Figure 16 Shroud configuration and probe planes [41].

ported through the downstream blade row CFD computationalresults showed a good agreement with the measured data at themidspan for the first stage The overall second stage efficiencyfor the full shroud case could be improved by 1% Newton et al.[48] measured the heat transfer coefficient and pressure coeffi-cient on the tip and near-tip region of a generic turbine blade

in a five-blade linear cascade Two tip clearances of 1.6% and2.8% of chord were considered and three tip geometries werestudied: plane tip, suction-sided squealer, and cavity squealer.They found that the flow separation at the pressure side edgedominated the flow through the plain gap, that the highest heattransfer was located in such a region that the flow reattached

on the tip, and that the suction-sided and cavity squealers couldreduce the heat transfer in the gap The suction-sided squealerprovided an overall net heat flux reduction of 15%, while thecavity squealer revealed no net heat flux reduction Palafox

et al [49] measured new detailed flow fields for a very largelow-speed, high-pressure turbine rotor blade using particle im-age velocimetry (PIV) The interaction between the tip leakagevortex and passage vortex was clearly characterized, and the ef-fect on the tip leakage vortex was examined Results showedthat a separation bubble under the tip significantly affectedthe leakage flow, and the end-wall movement influenced theshape and size of the bubble distinctly, while the relative bladeheat transfer engineering vol 31 no 7 2010

casing movement distorted the shape of the tip leakage vortex

and shifted it closer to the suction side

Jin and Goldstein [50, 51] simulated and measured local

mass and heat transfer on a turbine tip and near-tip regions They

concluded that for the smallest tip clearance the mass transfer on

the tip was significant along the pressure side At the largest tip

clearance the separation bubble on the tip could cover the whole

width of the tip on the second half of the tip surface A high

mainstream turbulence could reduce the average mass transfer

rate on the tip, whereas a higher mainstream Reynolds number

provided higher local and average values on the tip and near-tip

surfaces Stephens et al [52] and Douville et al [53] conducted

experiments to study the effects of thickness-to-gap and

gap-to-chord ratios on the tip-gap flows They also performed surface

flow visualization on the blade tip for better understanding of the

gap flow behavior The partial squealer tip or plasma actuators

were used to control the tip leakage flow Results showed that

the squealer tip could effectively reduce the pressure loss, and

by the use of a plasma actuator the effect depended strongly on

the unsteady frequency Srinivasan and Goldstein [54] measured

the local mass transfer on the tip of a turbine blade in a

five-blade linear cascade with a five-blade-centered configuration, and

used a moving end wall mounted on the top of a wind tunnel

to observe the effect of relative motion between the casing and

the tip Results showed that at a clearance of 0.6% there was

a small but definite reduction of 9% in the heat/mass transfer,

and at 0.86% clearance only a small effect of the wall motion

on the Sherwood number occurred At all higher clearances

no measurable effect of the relative motion on the Sherwood

number was observed

Computational Tip Leakage Flow and Heat Transfer

Multiple numerical studies have been carried out on blade tip

leakage flow associated with heat transfer Numerical prediction

and analysis can provide details of the flow field and thermal

distribution that sometimes are difficult to obtain by

experimen-tal measurements Dorney et al [55] performed a parallelized

unsteady analysis of the effects of tip clearance on the transient

and time-averaged flow fields in a supersonic turbine Results

indicated that improved performance could be traced by a

re-duction in the strength of the shock system in the vane or rotor,

and the reduction in losses was greater than the losses

gener-ated by increasing the tip clearance Green et al [56] conducted

computations and experiments on averaged and time-dependent

aerodynamics of a single-stage high-pressure turbine tip cavity

and stationary shroud The computational results showed good

correlation with the time-resolved data This in turn provided

confidence of the CFD modeling ability to predict turbine

pas-sages, blade tip, and shroud They found the largest amount of

unsteady surface pressure activity at the 15% span location,

es-pecially on the suction surface near the leading edge Past the

leading edge unsteady pressure amplitudes with respect to vane

passing frequency dropped off rapidly, and unsteady pressure

Figure 17 Schematics of blade tips with winglet [57].

amplitudes were much larger for all shroud locations than at theblade tip locations, Also, the results suggested that the bladetip configuration had very little impact on the time-accuratebehavior for the stationary shroud

Saha et al [57] performed calculations to observe the effect

of a winglet on the flow and heat transfer for both a flat tip and

a squealer tip, as shown in Figure 17 All the winglets werelocated on the pressure side only They found that for a flat tipthe winglet resulted in approximately 30% reduction in the localheat transfer coefficient on the tip, and a significant reduction

in the strength of the leakage flow and vorticity, whereas for

B SUNDEN AND G XIE 537

a two-sided squealer tip the winglet produced only marginal

improvements They concluded that the suction-side squealer

with constant winglet width offered better performance with

lower heat transfer coefficient and pressure loss than the others

Lampart et al [58] simulated numerically the effect of

inter-action of the main flow with rotor and tip leakage flows in a

high pressure axial turbine stage The proposed method could

trace and evaluate the process of mixing of the tip leakage and

windage flows with the main stream, and the interaction with

secondary flows and separation

Starodubtsev et al [59] proposed a 3D numerical model for

simulation of the viscous turbulent flow in a one-stage gas

tur-bine and validated the results with experimental measurements

They stated that their method could be applied for a variety of

turbine studies and design task Han et al [60] analyzed

numer-ically 3D flow fields near the tip region in an annular cascade

with tip clearance and rotation and in a linear cascade with the

validation of numerical results by flow visualization Results

showed that rotation could weaken the leakage flow, which

de-creased the size of the separation bubble on the tip surface, and

the tip vortex became larger and moved to the suction side as

the tip leakage flow was increased by an enlarged tip clearance

Intaratep et al [61] studied the interaction between the rotor

blade tip leakage flow and inflow disturbances They found that

the passage flow consisted of shear layers shed from the suction

side tip gap and a high velocity deficit region extending from

the suction side to the pressure side tip gap, and the local

pertur-bations near the blade tip induced the streamwise mean velocity

perturbations in the tip leakage vortex Yang et al [62]

numer-ically simulated the leakage flow and heat transfer on a flat tip,

a double squealer tip, and a single suction side squealer tip of a

scaled up GE-E3 blade The rotational effect was observed

un-der high pressure ratio and high temperature It was found that

the heat transfer coefficient decreased by increasing the squealer

cavity depth, while the shallow squealer cavity was the most

ef-fective in reducing the overall heat load Although the rotation

changed significantly the tip leakage flow pattern and local heat

transfer coefficient distribution on the tip, the area-averaged heat

transfer coefficient was affected only slightly

Mumic et al [63, 64] numerically studied the tip leakage flow

and heat transfer on the first stage of a high pressure turbine

A flat tip and a squealer tip with tip clearance of 1.0%, 1.5%,

and 2.5% blade span were considered Three turbulence models

were used to assess the prediction of the heat transfer It was

found that the three models could provide similar results in

reasonable agreement with the experimental data The low-Re

k-ω model could yield better prediction of blade tip heat transfer

compared to the other two models As shown in Figure 18, the

leakage flow increased and moved toward the trailing edge side

as the tip gap was increased The high heat transfer coefficients

on the rim were increased due to acceleration of the flow going

into the cavity and from the cavity into the rim region, and the

heat transfer coefficient near the leading edge cavity increased

and extended toward the trailing edge The flat tip heat transfer

was higher than the squealer tip heat transfer Mischo et al [65]

Figure 18 Comparison of heat transfer coefficient and simulated flow field [63].

numerically studied the flow field near the blade tip for differentshapes of the recessed cavities An improved design of the bladetip was presented It was found by an appropriate profiling of therecessed shape, the total tip heat transfer Nusselt number wassignificantly reduced by 15% and 7% compared to the flat tipand baseline recessed shape, respectively, as shown in Figure 19.The CFD analysis predicted a 0.38% total efficiency increasefor the rotor equipped with the new recess design compared tothe flat tip

Hamik and Willinger [66] introduced a new concept for sive turbine tip leakage control: A jet was injected roughlyperpendicular to the tip gap flow, as shown in Figure 20 Theyalso presented an analytical model to describe the reduction ofthe tip gap discharged coefficient due to the tip injection Theystated that the blade tip injection could increase the turbine ef-ficiency Prakash et al [67] proposed an improved tip having

pas-a pressure-side inclined squepas-aler shelf pas-and used CFD to studydifferent tip geometries, as shown in Figure 21 It was found thatthe inclined shelf could reduce the leakage flow and improvethe efficiency, indicating that it was superior to a vertical shelf.heat transfer engineering vol 31 no 7 2010

Figure 19 3D CFD flow for different tip shapes [65].

BLADE TIP EXTERNAL COOLING TECHNOLOGY

Needs of Cooling Technology for Blade Tip

Without doubt, cooling is required for the turbine blades,

including all regions being exposed to the high-temperature hot

gas Due to an unavoidable gap clearance between the blade tip

and casing or shroud, the hot gas flowing through the gap results

in a large thermal load on the blade tip The potential damage

due to the large heat load will lead to blade oxidation, as shown

in Figure 22 Hence, the blade tip is a key region that needs

cooling

The turbine blades are cooled by the use of extracted/

bypassed air from the compressor of the gas turbine This

ex-traction results in a reduction of the thermodynamic efficiency

and power output Too little coolant flow results in high blade

Figure 20 Blade tip with internal injection [66].

temperature, while too much coolant flow results in reducedturbine efficiency and power Therefore, it is very important todesign a turbine cooling system considering the balance of theminimum coolant air flow and maximum benefit of a high inlet

Figure 21 Blade tip with squealer shelf [67].

B SUNDEN AND G XIE 539

Figure 22 Material loss due to oxidation [11].

temperature If a proper cooling system is designed, the gain

from high firing temperature is so significant that it can

out-weigh the losses in the efficiency and power output, and offset

the complexity and cost of the cooling technology

The turbine blade tip and near-tip regions are difficult to cool

and are subjected to potential damage because of the high heat

load caused by tip leakage flow A common way to cool the tip

is to extract the cooling air from the internal coolant passages

through some film holes that are located on the blade surface

discretely This cooling is known as film cooling The relatively

cool air passes these holes and forms a thin protective layer/film

to protect the tip surface from the highly hot mainstream

Figure 23 depicts the film cooling concept The performance

of the film cooling depends on the coolant-to-hot mainstream

pressure ratio (blowing ratio), temperature ratio, and the hole

location, configuration (hole size, spacing, shape, angle and

number), distribution (leading-edge, trailing-edge, pressure and

suction side, endwall, tip), and on the representative flow

Figure 23 Schematics of film cooling.

conditions (Reynolds number, Mach number, free-stream lence, and unsteadiness) Obviously, a high and uniform coolingeffectiveness will ensure overall performance of the blade sur-face cooling In general, a higher blowing ratio at a specifictemperature ratio gives a higher film cooling performance, andthereby the heat is transferred to the blade surface and hence theprotection of surface is improved However, too high a blowingratio leads to jet penetration into the mainstream resulting in areduced cooling performance, while too small a blowing ratiodoes not force enough coolant to cover the hot surface Thus, it

turbu-is important to optimize the amount of coolant for film cooling

at the engine operating conditions For a better cooling mance, it is necessary to study the film cooling hole pattern,e.g., shape, angle, location, and distribution, which affect thefilm cooling performance

perfor-Additional review papers related to film cooling of gas bines are exemplified by refs [68]–[71] This paper is limited

tur-to a review of recent publications on blade tip cooling, and thusdoes not include all the results of external film cooling on turbineblades

Blade Tip External Cooling

A summary of Professor D E Metzger’s blade tip coolingstudies on blade tip cooling was published by Kim et al [72].Comparison of various tip cooling configurations and their ef-fects on film effectiveness and heat transfer coefficients werepresented Figure 24 shows the clearance gap and tip film cool-ing configuration and Figure 25 shows the cross sections Fourfilm cooling configurations were tested: (1) discrete slot injec-tion, (2) round hole injection, (3) pressure side flared hole injec-tion, and (4) grooved-tip cavity injection It was found that forcase 4 the overall film cooling performance varied significantlywith injection locations and that among the plane-tip injectionsthe discrete slot injection provided better performance than theothers

Yang et al [73] numerically studied various film hole figurations on plane and squealer tips of a turbine blade Threeconfigurations were tested: (1) the camber arrangement, (2) theupstream arrangement, and (3) the two rows arrangement, asschematically shown in Figure 26 The effects of rotation wereobserved It was found that at high blowing ratios the latter twocases provided better film cooling performance on the planeheat transfer engineering vol 31 no 7 2010

con-Figure 24 Schematics of clearance gap and film cooling configuration.

and squealer tips than the former one Higher blowing ratios

resulted in a higher cooling effectiveness on the shroud for all

cases They also found that rotation decreased the plane-tip film

cooling effectiveness while it slightly affected the squealer-tip

film cooling due to the large cavity depth

Mhetras et al [74] observed the effects of shaped holes on

the tip pressure side, coolant jet impingement on the pressure

side squealer rim from tip holes, and varying blowing ratios

for a squealer tip The film cooling effectiveness distributions

on the blade tip, near-tip pressure side rim, and the inner

pres-sure side rim were meapres-sured using a prespres-sure-sensitive paint

(PSP) technique Numerical simulations were also performed

for prediction of the film cooling It was found that a higher

blowing ratio provided higher effectiveness on the tip rim,

cav-ity flow, and inner rim walls, and the presence of serpentine

passages could supply coolant to the holes so that a significant

impact on film cooling performance was achieved Good

agree-ment between the experiagree-ments and simulation was achieved

Mhetras et al [75] also measured the film cooling

effective-ness of shaped holes near the tip pressure side and cylindrical

1.5W

W

1/3W

d 3d

a

b

Figure 25 Cross section of film cooling configuration.

Figure 26 Various film cooling hole arrangements [73].

holes on the squealer cavity floor using PSP The pressure sidesquealer rim wall was cut near the trailing edge It was foundthat the cutback squealer rim provided high film cooling effec-tiveness in the trailing edge of the blade tip compared to a fullsquealer Due to the combined effect of tip- and pressure-sidecoolant injection, high and uniform effectiveness was found onthe tip rim and inner and outer squealer rim walls

Ameri and Gigby [76] performed computations to predict theheat transfer coefficient distribution on a blade tip with coolingholes The simulation model for prediction of the tip heat transferand cooling effectiveness based on a 3D Reynolds-averaged NSsolver, was assessed by the data of Kim and Metzger [77].Through the numerical flow visualization it was shown that thedistance from the pressure side to the edge of the film coolinghole might be an important parameter Christophel et al [78–80] experimentally investigated the adiabatic effectiveness andheat transfer coefficients along and near the blade tip usingpressure side film cooling holes Results showed that the coolingeffectiveness of the holes was better for a small tip gap than for

a large tip gap With blowing, the tip heat transfer coefficientswere increased above those without blowing, and increased withincreasing blowing ratio The area-averaged net flux reduction

B SUNDEN AND G XIE 541

suggested a small dependence on the coolant flow rate and higher

cooling benefit for a small tip gap

Kwak and Han [81] measured heat transfer coefficient and

film cooling effectiveness on the squealer tip of a gas turbine

blade in a five-bladed linear cascade The blade model was

equipped with a single row of film cooling holes on the pressure

side near the tip region and the tip surface along the

camber-line They found that the overall film cooling effectiveness was

increased but heat transfer coefficients were decreased for the

squealer tip compared to the plane tip at the same tip gap and

blowing ratio High film cooling effectiveness occurred near the

trailing edge cavity because of the coolant accumulation Kwak

and Han [82] also measured the distributions of heat transfer

coefficient and film cooling effectiveness on a turbine blade

tip Three tip gas clearances, i.e., 1.0%, 1.5%, and 2.5%, and

three blowing ratios, i.e., 0.5, 1, and 2, were tested Results

showed that with the increasing blowing ratio, the film cooling

effectiveness increased but the heat transfer coefficient on the tip

slightly decreased The static pressure on the shroud increased,

and with the increase of gap clearance the heat transfer

coeffi-cient and film effectiveness increased By addition of pressure

side injection the film cooling effectiveness could be increased

Ahn et al [83] also observed the effects of the presence of the

squealer tip, the locations of film cooling holes, and the tip gap

clearance on the film cooling effectiveness compared to a plane

tip It was found for the squealer tip with tip and pressure-side

injection that the film cooling effectiveness was higher than that

with only tip injection or with only pressure-side injection For

the plane tip the film cooling effectiveness was significant but

negligible for squealer tip

Gao et al [84, 85] studied the effect of incidence angle on

film cooling effectiveness for a cutback squealer blade tip in

a five-blade linear cascade The film cooling effectiveness was

measured based on mass transfer analogy using PSP techniques

One row of shaped holes was located along the pressure side

just below the tip and two rows of cylindrical holes were

lo-cated on the tip It was found that the film cooling effectiveness

distribution was altered, and the peak of laterally averaged

effec-tiveness was shifted to upstream or downstream depending on

the incidence angle, but the overall area-averaged film cooling

effectiveness was not changed significantly Also, the coolant

jet spread more on the cavity floor at positive incidence angles,

resulting in relatively high and uniform film coverage on the

cavity floor

Other detailed studies related to blade tip heat transfer and

cooling topics have been summarized in many theses These can

be found in refs [86–97]

BLADE TIP INTERNAL COOLING TECHNOLOGY

Apart from external film cooling the blade tip region, a

num-ber of serpentine passages can be used as channels for

inter-nal coolant air to cool the blade These cooling passages wind

Figure 27 A typical serpentine passage inside a blade.

through the blade but are not limited to a simple straight nel A common serpentine passage may consist of a first pass,

chan-a shchan-arp 180◦turn/bend, and a second pass A typical serpentinepassage is schematically shown in Figure 27 The coolant flowsradially outward from the hub and then turns 180◦and travelsradially inward from the tip to the hub Also, rib turbulatorsmight be mounted on the leading or/and trailing walls to en-hance the heat transfer between the blade wall and coolant Theflow field in the turn/bend is very complex, and so is the heattransfer, because the channel configuration, its aspect ratio, theturn geometry, and the rib configuration and location will affectthe flow and heat transfer Because the rotation alters the flowand hence the heat transfer coefficient distribution, the rotationeffect should be considered

This paper does not review all research works on heat transferenhancement in single-pass ribbed channels, but reviews mainlythe findings of flow and heat transfer in two-pass or U-bendchannels with/without rib turbulators, especially in the turn/bendregion

Experimentally Internal Cooling for Blade Tip

Park and Lau [98], Park et al [99–102], Kukreja et al.[103], and Lee et al [104] conducted a series of naphthaleneheat transfer engineering vol 31 no 7 2010

sublimation experiments on local heat/mass transfer

distribu-tions on the leading and trailing walls of rotating smooth and

ribbed two-pass channels The effects of channel orientation,

channel shape, rotation, sharp turn, and angled ribs were

ob-served It was found that rotation did not lower the spanwise

average heat transfer on the leading wall, and the sharp turn

reduced the heat/mass transfer on the leading and trailing walls

Due to the complex flow field with secondary flows, separated

and re-attached flows, and flow recirculation in the turn and

near the ribs on the walls, there existed large variation of local

heat/mass transfer in the turn and immediately downstream the

turn, as shown in Figure 28

Mochizuki et al [105] performed detailed measurements of

the local heat transfer coefficients in turbulent flow through

smooth and rib-roughened serpentine passages with 180◦sharp

bend, and performed flow visualization to reveal the generation

of secondary flows Results showed that for a smooth channel

the heat transfer downstream from the bend was controlled by

secondary flows and the heat transfer coefficients on the wall

surfaces differed from one another For ribbed channel, due to

the interaction of two secondary flows by the ribs and bend, the

ribs could affect strongly heat transfer in the bend and second

pass Chen et al [106] presented 3D detailed mass (heat) transfer

distributions along four active walls of a square duct with a 180◦

bend and ribs in the first pass Results showed that the effect

of the bend was clearly visible in the ribbed duct following the

bend Due to the high velocity resulting from the bend, local

acceleration and turbulence production generated by ribs, the

higher mass transfer rates occurred near the corners of the outer

wall Astarita and co-workers [107, 108] measured the detailed

heat transfer distribution near a 180◦ sharp turn of a square

channel with and without rib turbulators It was observed that

for the smooth channels there were three high heat transfer zones

in the turn, while the only high heat transfer zone left was placed

after the second outer corner and exhibited a smaller extension

The averaged normalized Nusselt number slightly increased for

the both side heating condition compared to that for one-side

heating conditions

Ekkad et al [109] measured the detailed heat transfer

distri-butions inside straight and tapered two-pass channels with and

without rib turbulators It was found that the tapered channel

with ribs provided 1.5–2.0 times higher Nusselt number ratios

over the tapered smooth channel in the first pass, while in the

after-turn region of the second pass the ribbed and smooth

chan-nels provided similar Nusselt number ratios Ekkad et al [110]

and Pamula et al [111] also measured the detailed heat transfer

distribution inside a two-pass square channel connected by two

rows of holes on the divider walls, shown in Figure 29 It was

found that the proposed feed system, from first pass to second

pass using crossflow injection holes, produced higher Nusselt

numbers on the second-pass walls with the enhancement factor

as high as two to three times than that obtained in the second

pass for a channel with a conventional 180◦ turn Son et al

[112] carried out particle image velocimetry (PIV) experiments

to study the correlation between the high Reynolds number

Figure 28 A typical result inside a two-pass channel [98].(Re= 30,000) turbulent flow and wall heat transfer characteris-tics in a two-pass square channel with a smooth wall and a 90◦rib-roughened wall Compared with the heat transfer experimen-tal data of Ekkad and Han [113], the PIV measurement results

B SUNDEN AND G XIE 543

Figure 29 The geometry of a two-pass channel having injection holes.

showed that the flow impingement is the primary factor for the

two-pass square channel heat transfer enhancement rather than

the flow turbulence level itself Besides, the secondary flow

char-acteristics are correlated with the wall heat transfer enhancement

for smooth and ribbed wall two-pass square channels

Chanteloup et al [114] measured flow characteristic effects

on the wall heat transfer distribution of a two-pass internal

coolant ribbed passage of gas turbine airfoils Results showed

that ribs at 45◦ increased the average heat transfer gradients,

and the ratio of high to low Nusselt numbers was up to 6 in

the U-shaped heat transfer distribution downstream the ribs

Chanteloup and Bolcs [115] also measured flow characteristics

in a 180◦bend region and downstream of the bend of two-leg

internal coolant passages of gas turbine airfoils with film

cool-ing hole ejection 45◦ angled ribs were located on the bottom

and top walls of both legs Results showed that adding bleeding

holes having high ratio between the channel inlet mass flow and

the extracted mass flow would affect significantly the flow in the

two-legged cooling channel Due to the high variations in the

streamwise velocity, the large variations in the heat transfer

oc-curred near the upstream part of the bend Iacovides et al [116]

reported flow and heat transfer in a U-bend with 45◦ribs

rotat-ing channel It was found that the Nusselt number in the ribbed

channel was twice that for a smooth channel [117], and the flow

and average Nusselt numbers were relatively unaffected by

ro-tation but led to local hot or cold spots resulting in significant

implications for the level of thermal stresses induced

Hsieh and Liao [118] and Hsieh and co-workers [119–121]

measured the effects of rotation and uneven heating conditions

as well as passage aspect ratio on the local heat transfer and

pres-sure drop in a rotating two-pass ribbed rectangular or smooth

square channel Results showed that higher heat transfer on both

the leading and trailing walls was caused by a complicated 3D

accelerated flow and secondary flow in the U-bend region For

a ribbed channel, steamwise-periodic fully developed flow was

achieved after a sufficient distance The intensity of the shear

layer was greater in the vicinity of the ribs compared to a smooth

surface However, the size of the separation region was smaller

than that of a stationary duct as the rotation number increased

Hsieh et al also found that the rotation makes the turbulent

intensity and shear stress distribution more random in the

trans-Figure 30 Different cross sections of profiled ribs [122].verse direction For a smooth channel, they found no separation

in the first and second channels except for a certain size pocket

of separation on the inner wall in the U-bend region The fluence of the U-bend and rotation on the mean velocity fieldwas apparent, and the rotation may alter the development of themean and fluctuating motion

in-Acharya et al [122] and Nikitopoulos et al [123] investigatedexperimentally the effects of rib with different cross-stream pro-files on the surface mass (heat) transfer distribution along fouractive walls of a square duct having a sharp 180◦bend The crosssections of the profiled ribs are shown in Figure 30 These au-thors found that the profiled ribs enhanced the heat transfer due

to the generation of secondary and longitudinal vorticity thatinteracts with Coriolis-induced secondary flows in the channel

It was suggested that the use of profiled ribs might be a viableand effective solution to local heat transfer enhancement and/orspatial redistribution in actual rotating, ribbed multipass cool-ing channels for gas turbine applications Liou and co-workers[124–134] conducted a series of experiments on flow and heattransfer in rotating two-pass smooth and various angled ribbedchannels using LCT or/and LDV (see Figure 31) The effects

of the divider thickness, rib arrangement, channel cross-sectionshape, channel orientation, and rotation conditions were ob-served in detail

Al-hadhrami and Han [135] tested the effect of various 45◦angled rib turbulators on Nu ratio in a rotating two-pass squarechannel, as shown in Figure 32 It was found that the Nu ratio

in the 180◦ turn region and the differences among differentangled rib orientations were increased with increasing rotationnumber Al-hadhrami et al [136] also studied the heat transfer

in two-pass rotating rectangular channels with five differentorientations of 45◦V-shaped ribs, as shown in Figure 33 Resultsshowed that there was relatively low heat transfer enhancement

in the 180◦turn region due to suppression of the vortices inducedfrom the V-shaped ribs by the turn and no ribs placed at theturn Parallel 45◦rib arrangements provided better heat transfercompared to the other cases

Prabhu and Vedula [137] investigated the local pressure dropcharacteristics in a square-cross-sectioned smooth channel withheat transfer engineering vol 31 no 7 2010

Figure 31 LCT and LDV test facility [124].

a sharp 180◦ bend rotating about an axis normal to the

free-stream direction They found that the local pressure drop

char-acteristics in the bend region are affected by a change in the

rotation number but the influence of the Reynolds number was

weak Another finding is that the friction factor was less

sensi-tive to rotation for a bend with a hydraulic diameter ratio of 0.24

compared to bends with ratios of 0.37 and 0.73, respectively

Ratana-Rao and Prabuhu [138] and Ratana-Rao et al [139]

ex-perimentally studied the effect of several turn treatments on the

pressure drop distribution in smooth and ribbed squared

chan-nels with a sharp 180◦bend Results showed that short and long

guide vanes placed at the center of the bend in a smooth channel

resulted in a reduction of about 28% in overall pressure drop,

and for the ribbed channel a maximum decrease of 15% to 16%

in overall pressure drop was achieved in the case of the long

guide vane located at the center of the bend and multiple 180◦

extended guide vanes

Azad et al [140] measured the heat transfer in a two-pass

rectangular rotating channel with 45◦ angled rib turbulators

Results showed that the heat transfer from the first pass trailing

and second pass leading surfaces was enhanced by rotation

45◦ parallel ribs provided a better heat transfer augmentation

than 45◦ cross ribs Fu et al [141, 142], and Liu et al [143]

reported heat transfer coefficients and friction factors in

two-pass rectangular channels with rib turbulators placed on the

leading and trailing surfaces Five kinds of ribs were considered:

45◦ angled, V-shaped, discrete 45◦ angled, discrete V-shaped,

Figure 32 Two-pass channel with various 45 ◦angled ribs [135].

and crossed V-shaped It was found that due to the turn effectthe rotation effect was greater on heat transfer in the first passthan in the second pass The discrete V-shaped ribs showed thebest overall thermal performance, as shown in Figure 34.Nakayama et al [144] measured flow and heat transfer instationary two-pass channels with a sharp 180◦turn Three turnclearances were considered It was found that flow recirculationappeared in the upstream corner in the turn section as well asalong the divider wall after the turn, and the local maxima ofthe Sherwood number on the short-side walls inside and afterthe turn were mainly caused by the velocity component normal

to each wall

Zhou et al [145] measured the heat transfer and pressure drop

in a rotating smooth two-pass coolant passage It was found thatrotational effects were important in the bend region at lowerReynolds number with significant enhancement along the bend-trailing surface, and a higher density ratio enhanced the heattransfer on both the leading and trailing walls of the inlet, bend,and outlet In the bend region the enhancement was significant

on the leading surface

B SUNDEN AND G XIE 545

Figure 33 Two-pass channel with various 45 ◦V-shaped ribs [136].

Kim et al [146–149] measured the detailed heat/mass

trans-fer and pressure drop in a rotating two-pass duct with transverse

ribs It was found that due to the rotation of the duct, the

Sher-wood number ratios and pressure coefficients were high on the

trailing surface in the first pass and on the leading surface in

the second pass In the turn region of the stationary duct two

Dean vortices were transformed into one large asymmetric

vor-tex cell, which changed the heat/mass transfer and pressure drop

characteristics Cho et al [150] measured the effect of cross ribs

on heat/mass transfer in a two-pass duct under rotating

condi-tions Results showed that for the stationary case the turning

effect dominated the secondary flow at the end of the turn, and

for the rotating case in the first pass the Sherwood numbers

Figure 34 Two-pass channel with various ribs [141].

on the trailing surface were higher than those on the leadingsurface, while in the second pass the Sherwood numbers werehigher on the leading surface Cho et al [151] also measuredthe heat/mass transfer and flow characteristics in a two-pass ro-tating rectangular duct with and without 70◦ angled ribs, andconducted numerical simulations to analyze the flow pattern.Results showed that large overall heat transfer on the leadingand trailing surfaces for the first and second passes depended onthe rotating speed and turn geometry, but the local heat transferwas affected mainly by the rib arrangement

Bunker [152] presented a method to provide substantiallyincreased convective heat flux on the internal cooled tip cap

of a turbine blade, where arrays of discrete shaped pins werefabricated and placed, as shown in Figure 35 The detailed heattransfer distribution over the internal tip cap was obtained basedheat transfer engineering vol 31 no 7 2010

Figure 35 Geometries of tip-cap with pin arrays [152].

on a large-scale model of a sharp 180◦ tip turn Five tip cap

surfaces were tested It was found that the effective heat transfer

coefficient could be increased by up to a factor of 2.5 due to

the combination of impingement and cross-flow convection on

the pins, as shown in Figure 35 The tip turn pressure drop was

negligible compared to that of a smooth surface

Computationally Internal Cooling for Blade Tip

With the fast development of computer resources, the

in-crease of computational power makes it economical to simulate

flow and heat transfer inside turbine blade passages The thermal

and cooling performance then can be optimized and designed

based on numerical analysis Chen et al [153, 154], Jang et al

[155, 156], and Al-Qahtani et al [157, 158] calculated the 3D

flow and heat transfer in rotating two-pass square channel with

smooth walls or 45◦/60◦angled ribs by a second-moment closure

model and a two-layer k-ε isotropic eddy viscosity model Good

agreement with experimental data of Ekkad and Han [113] was

achieved The comparison of the results showed that the

near-wall second-moment closure model provided accurate

predic-tions of the complex 3D flow and heat transfer resulting from

the rotation and strong wall curvatures Also, it was observed

that angled ribs with high blockage ratio and a 180◦sharp turn

produced strong nonisotropic turbulence and heat flux, which

Figure 36 Secondary flow and temperature contours in a rotating smooth channel [157].

affected significantly the flow field and heat transfer coefficient,

as shown in Figure 36

Lin et al [159] performed computations of 3D flow and heattransfer in a U-shaped square duct for rotating and nonrotat-ing conditions The flow streamlines, velocity vector fields, andcontours showed how the fluid flow in a U-duct evolved from aunidirectional one to one with convoluted secondary flows due tothe Coriolis force, centrifugal buoyancy, staggered inclined ribs,and a 180◦ bend, and also how the nature of the fluid flow af-fected the surface heat transfer Suga [160] predicted turbulenceand heat transfer in two types of square sectioned U-bend ductflows with mild and strong curvature by recent second momentclosures Suga and Abe [161] applied a higher order version

of the generalized gradient diffusion hypothesis along with theTCL (two-component-limit) model They found that the secondmoment closure was good enough for predicting flow and heattransfer in the case of mild curvature, but only the TCL modelwas reliable for the strong curvature case

Iacovides [162] carried out computations of turbulent flowsthrough stationary and rotating rib-roughened U-bends to ex-plore both numerical and turbulence modeling Because of usingbody-fitted grids and higher order schemes for the discretization

of the convective transport of all flow variables, grid ments could be reduced Concerning the turbulence modeling,the comparisons suggested that a low-Re second-moment clo-sure becomes necessary, but the second-moment closure cannotaccount for the effects of negative rotation Moreover, the finermesh computations would examine the predicted turbulencefields closely Nikas et al [163] presented computations of heatand fluid flow through a square-ended U-bend that rotates about

require-an axis normal to both the main flow direction require-and also the axis

of curvature The main flow features were well reproduced byall models, but the mean flow within and after the bend wasbetter reproduced by the low-Re models On the other hand, tur-bulence levels within the rotating U-bend were underpredicted,but low-Re DSM models produced a more realistic distribution.Along the leading side, all models overpredicted the heat trans-fer just after the bend, and for the trailing side, the heat transfer

B SUNDEN AND G XIE 547

predictions of the low-Re DSM with a differential length-scale

correction term were close to the measurements Raisee et al

[164] considered the application of low-Re linear and nonlinear

eddy-viscosity models for the numerical prediction of the

veloc-ity and pressure field in flow through two 90◦curved ducts: one

of a square cross section and one of a rectangular cross section

The results indicated that for the bend of square cross section

the curvature induced a strong secondary flow, while for the

rectangular cross section the secondary motion was modified at

the corner regions For both curved ducts, the secondary

mo-tion persisted downstream of the bend and disappeared slowly

Another aspect was that, for the bend of square cross section,

comparisons indicated that both turbulence models could get

reasonable predictions A wider range of data was available for

the bend of rectangular cross section, and it was found the

non-linear k-ε model showed superior predictions of the turbulence

field and the pressure and friction coefficients

Murata and Mochizuki [165] numerically studied the

cen-trifugal buoyancy effect on turbulent heat transfer in a rotating

two-pass square channel with 180◦sharp turns by the large eddy

simulation (LES) It was found that with increasing buoyancy,

the pressure loss coefficient of the sharp turn was decreased and

that of the straight pass was increased in the first pass and

de-creased in the second pass, and due to the aiding and opposing

buoyancy contributions to the main flow the variation caused

by the buoyancy was larger for the heat transfer on the pressure

surface than on the suction surface For the studied buoyancy

range the Colburn j factor was kept almost constant.

Sleiti and Kapat [166] predicted numerically the flow field

and heat transfer of high rotation numbers and density ratio flow

in a square internal cooling channel with U-turn They found that

the four-side-averaged Nusselt number increased linearly with

increasing rotation number but slightly decreased with

increas-ing density ratio At the center of the U-bend the corner vortices

were suppressed with increasing rotation number, while an

in-creased density ratio resulted in a decrease in all surfaces of the

U-turn Sleiti and Kapat [167, 168] also predicted the 3D flow

field and heat transfer in a two-pass rib-roughened square

chan-nel Results showed that in the U-turn high shear stresses were

found near the leading and trailing surface, and were increased

by increasing density ratio

Etemad and Sunden [169, 170], and Etemad et al [171]

used turbulence models with linear and nonlinear expressions

for the Reynolds stresses to investigate turbulent flow and heat

transfer in a square-sectioned U-bend Five turbulence models

were evaluated: Suga’s quadratic and cubic low-Re k-ε, V2F

k-ε, RSM-EVH, and RSM GGDH These models predicted

the stress-induced secondary motion in the straight inlet duct,

and this secondary motion had an impact on the flow in the

bend It was found that Suga’s model performed slightly better

and offered a higher degree of robustness Guleren and Turan

[172] used large-eddy simulation (LES) to carry out

numeri-cal predictions of developing turbulent flow through stationary

and rotating U-ducts with strong curvature Their aim was to

validate LES in a strongly curved U-duct for three different

cases: stationary, positive, and negative rotational cases Theyfound that grid resolution had some effect on the profiles ofthe Reynolds stresses The wall function was responsible forthe excessive turbulent intensities, and LES was superior to thetwo-component-limit turbulence model with the predictions ofmean velocities The primary and secondary flow behavior cangive a better understanding of the origin and development of theflow separation Viswanathan and Tafti [173] predicted turbulentflow field in a two-pass internal cooling duct with normal ribs bydetached eddy simulation (DES) and the unsteady Reynolds av-eraged Navier–Stokes equations (URANS) Results showed thatDES predicted a slower flow development than LES, whereasURANS predicted it much earlier than LES computations andexperiments DES could accurately predict the flow both in thefully developed region as well as in the 180◦bend of the duct.Other related research works about turbulent heat transfer inserpentine passages are available in research theses [174–186].Valuable review articles have been presented [8, 9, 187, 188]

SUMMARY

As the turbine inlet temperature is continuously increasedfor fast development of current gas turbine engines, the heattransferred to the blade is increased To satisfy the even increas-ingly high inlet temperature, turbine blade cooling becomes animportant issue for new designs Such cooling includes bladeend-wall cooling, leading-edge cooling, trailing-edge cooling,and tip cooling The blade tip is one of the critical regions to becooled due to the high thermal load over the tip surface There-fore, highly accurate and highly detailed local heat transfer andflow data related to such regions are needed for analysis, andcooling schemes must be designed to prevent the failure due tothe local hot spots

As reviewed in the preceding sections, more available datafrom experimental measurements and numerical simulations arefor blade tip clearance leakage flow associated with heat transferand for near-tip region flow field and heat transfer Even withsophisticated clearance control methods to employ, the gap isnever eliminated, and thereby the leakage flow occurs due to thepressure difference between the pressure side and suction side.The leakage flow has a pronounced influence on local heat/masscoefficient distribution and hence the heat load Thus, whateverthe tip geometry is and whatever the clearance control strategy

is, to develop novel and optimal techniques will require moreresearch on the detailed and accurate leakage flow and heat/mass transfer characteristics over the blade tip and near-tipregion

The blade tip is the most susceptible region subjected to thelarge thermal load and is difficult to cool sufficiently For exter-nal cooling, a common technique is to add film cooling throughthe tip and near-tip region The cooling performance is affectedsignificantly by most conditions, such as film cooling hole con-figuration, location, and distribution, and the representative flowheat transfer engineering vol 31 no 7 2010

conditions For internal cooling, serpentine cooling passages are

designed inside blades, so that the heat from the pressure side

and suction side is picked up by the turning coolant extracted

from compressors The serpentine channel configuration, aspect

ratio and orientation, rib configuration and location, and rotation

and bend/turn geometry affect significantly the internal cooling

efficiency Several studies have contributed to the cooling issues

However, it is not enough to observe the parametric effects of

film cooling for the blade tip More studies related to combined

film cooling and internal convective cooling are needed Also,

although a large number of research works have concerned

tur-bulent heat transfer and cooling issues inside serpentine

(two-pass, multipass) channels, studies concerning internal blade tip

cooling concept and research are still limited Thus more

stud-ies related to these issues are required Especially, the detailed

flow and heat transfer distribution characteristics and cooling

performance on the tip-cap walls or near-tip region need to be

investigated

The CFD techniques act as an important role in research and

design of gas turbine components and can provide useful data

related to the detailed flow field and heat transfer coefficient

distribution along the gas turbine blades With the fast

devel-opment of CFD techniques as prediction tools, highly accurate

CFD computations are encouraged to provide insight into

com-plex flow and heat/mass transfer as well as the cooling process

on the blade tip, and various turbulence models should be tested

and validated by available experimental data

REFERENCES

[1] Bathie, W M., Fundamentals of Gas Turbines, 2nd ed., John

Wiley & Sons, New York, 1996

[2] Saravanamuttoo, H I H., Cohen, H., and Rogers, G F C., Gas

Turbine Theory, Prentice Hall, Harlow, UK, 2001.

[3] Han, J C., Dutta, S., and Ekkad, S V., Gas Turbine Heat Transfer

and Cooling Technology, Taylor & Francis, New York, 2000.

[4] Goldstein, R J., Heat Transfer in Gas Turbine Systems, Annals

of the New York Academy of Sciences, New York, 2001

[5] Sund´en, B., and Faghri, M., Heat Transfer in Gas Turbines, WIT

Press, Southampton, UK, 2001

[6] Dunn, M G., Convective Heat Transfer and Aerodynamics

in Axial Flow Turbines, ASME Journal of Turbomachinery,

vol 123, no 4, pp 637–686, 2001

[7] Han, J C., and Dutta S., Recent Development in Turbine Blade

Film Cooling, International Journal of Rotating Machinery,

vol 7, no 1, pp 21–40, 2001

[8] Han, J C., Recent Studies in Turbine Blade Cooling,

Interna-tional Journal of Rotating Machinery, vol 10, no 6, pp 1–15,

2004

[9] Han J C., Turbine Blade Cooling Studies at Texas A&M

Uni-versity: 1980-2004, AIAA Journal of Thermophysics and Heat

Transfer, vol 20, no 2, pp 161–187, 2006.

[10] Bunker, R S., A Review of Turbine Blade Tip Heat Transfer,

An-nals of the New York Academy of Sciences, New York, pp 64–

79, 2001

[11] Bunker, R S., Axial Turbine Blade Tips: Function, Design and

Durability, AIAA Journal of Propulsion and Power, vol 22, no 2,

pp 271–285, 2006

[12] Bunker, R S., Gas Turbine Heat Transfer: Ten Remaining

Hot Gas Path Challenges, ASME Journal of Turbomachinery,

vol 129, pp 193–201, 2007

[13] Glezer, B., Harvey, N., Camci, C., Bunker, R., and Ameri,

A A., Turbine Blade Tip Design And Tip Clearance Treatment,

VLI LS 2004-02, Von Karman Institute Lecture Series, Brussels,Belgium, 2004

[14] Garg, V K., Heat Transfer Research on Gas Turbine Airfoils

at NASA GRC, International Journal of Heat and Fluid Flow,

vol 23, pp 109–136, 2002

[15] Ameri, A A., Steinthorsson, E., and Rigby, D L., Effects of

Squealer Tip on Rotor Heat Transfer and Efficiency, ASME nal of Turbomachinery, vol 120, pp 753–759, 1998.

Jour-[16] Ameri, A A., Steinthorsson, E., and Rigby, D L., Effects ofTip Clearance and Casing Recess on Heat Transfer and Stage

Efficiency in Axial Turbines, ASME Journal of Turbomachinery,

[19] Ameri, A A., Heat Transfer and Flow on the Blade Tip

of a Gas Turbine Equipped With a Mean-Camberline Strip,

ASME Journal of Turbomachinery, vol 123, pp 704–708,

2001

[20] Thorpe, S J., Yoshino, S., Ainsworth, R W., and Harvey, N W.,

An Investigation of the Heat Transfer and Static Pressure on theOver-Tip Casing Wall of an Axial Turbine Operating at Engine

Representative Flow Conditions (I) Time-Mean Results, national Journal of Heat and Fluid Flow, vol 25, pp 933–944,

Inter-2004

[21] Thorpe, S J., Yoshino, S., Ainsworth, R W., and Harvey, N W.,

An Investigation of the Heat Transfer and Static Pressure on theOver-Tip Casing Wall of an Axial Turbine Operating at EngineRepresentative Flow Conditions (II) Time-Resolved Results,

International Journal of Heat and Fluid Flow, vol 25, pp 945–

[23] Chana, K S., and Jones, T V., An Investigation on Turbine Tip

and Shroud Heat Transfer, ASME Journal of Turbomachinery,

B SUNDEN AND G XIE 549[26] Azad, G S., Han, J C., Teng, S., and Boyle, R J., Heat Transfer

and Pressure Distributions on a Gas Turbine Blade Tip, ASME

Journal of Turbomachinery, vol 122, pp 717–724, 2000.

[27] Azad, G S., Han, J C., and Boyle, R J., Heat Transfer and Flow

on the Squealer Tip of a Gas Turbine Blade, ASME Journal of

Turbomachinery, vol 122, pp 725–732, 2000.

[28] Teng, S., Han, J C., and Azad, G S., Detailed Heat Transfer

Distributions on a Large-Scale Gas Turbine Blade Tip, ASME

Journal of Heat Transfer, vol 123, pp 803–809, 2001.

[29] Azad, G S., Han, J C., Bunker, R S., and Lee, C P., Effect

of Squealer Geometry Arrangement on a Gas Turbine Blade

Tip Heat Transfer, ASME Journal of Heat Transfer, vol 124,

pp 452–459, 2002

[30] Dhadwal, H S., and Kurkov, A P., Dual-Laser Probe

Measure-ments of Blade-Tip Clearance, ASME Journal of

Turbomachin-ery, vol 121, pp 481–485, 1999.

[31] Saxena, V., Nasir, H., and Ekkad, S.V., Effect of Blade Tip

Geom-etry on Tip Flow and Heat Transfer for a Blade in a Low-Speed

Cascade, ASME Journal of Turbomachinery, vol 126, pp 130–

138, 2004

[32] Saxena, V., and Ekkad, S V., Effect of Blade Tip Geometry on

Tip Flow and Heat Transfer for a Blade in a Low-Speed Cascade,

ASME Journal of Turbomachinery, vol 126, pp 546–553, 2004.

[33] Nasir, H., Ekkad, S V., Kontrovitz, D W., Bunker R S., and

Prakash C., Effect of Tip Gap and Squealer Geometry on

De-tailed Heat Transfer Measurements Over a High Pressure Turbine

Rotor Blade Tip, ASME Journal of Turbomachinery, vol 126,

pp 221–228, 2004

[34] Rhee, D H., and Cho, H H., Local Heat/Mass Transfer

Char-acteristics on a Rotating Blade With Flat Tip in Low-Speed

Annular Cascade—Part I: Near-Tip Surface, ASME Journal of

Turbomachinery, vol 128, pp 96–109, 2006.

[35] Rhee, D H., and Cho, H H., Local Heat/Mass Transfer

Char-acteristics on a Rotating Blade With Flat Tip in Low-Speed

Annular Cascade—Part II: Tip and Shroud, ASME Journal of

Turbomachinery, vol 128, pp 110–119, 2006.

[36] Rhee, D H., and Cho, H H., Effect of Vane/Blade Relative

Position on Heat Transfer Characteristics in a Stationary Turbine

Blade: Part I, Tip and Shroud, International Journal of Thermal

Science, vol 47, no 11, pp 1528–1543, 2008.

[37] Rhee, D H., and Cho, H H., Effect of Vane/Blade Relative

Position on Heat Transfer Characteristics in a Stationary Turbine

Blade: Part II, Blade Surface, International Journal of Thermal

Science, vol 47, no 11, pp 1544–1554, 2008.

[38] Matsunuma, T., Effect of Reynolds Number and Freestream

bulence on Turbine Tip Clearance Flow, ASME Journal of

Tur-bomachinery, vol 128, pp 166–177, 2006.

[39] Kwak, J S., and Han, J C., Heat-Transfer Coefficients of a

Tur-bine Blade-Tip and Near-Tip Regions, AIAA Journal of

Thermo-physics and Heat Transfer, vol 17, no 3, pp 297–303, 2003.

[40] Kwak, J S., Ahn, J., and Han, J C., Effects of Rim Location,

Rim Height and Tip Clearance on the Tip and Near-Tip Region

Heat of a Turbine Blade, International Journal of Heat and Mass

Transfer, vol 47, pp 5651–5663, 2004.

[41] Kwak, J S., and Han, J C., Heat Transfer Coefficients on the

Squealer Tip and Near Squealer Tip Regions of a Gas Turbine

Blade, ASME Journal of Heat Transfer, vol 125, pp 669–677,

2003

[42] Kwak, J S., Ahn, J., Han, J C., Lee, P C., Bunker R S., Boyle

R., and Gaugler R., Heat Transfer Coefficients on the Squealer

Tip and Near Squealer Tip Regions of A Gas Turbine Blade With

Single or Double Squealer, ASME Journal of Turbomachinery,

[44] Rehder, H J., and Dannhauer, A., Experimental Investigation

of Turbine Leakage Flow on the Three-Dimensional Flow Field

and Endwall Heat Transfer, ASME Journal of Turbomachinery,

vol 129, pp 608–618, 2007

[45] Govardhan, M., Sasrri, S K., and Vishnubhotla, V S., perimental Investigation of the Three-Dimensional Flow in a

Ex-Large Deflection Turbine Cascade With Tip Clearance, Journal

of Thermal Science, vol 7, no 3, pp 149–164, 1998.

[46] Govardhan, M., Gowda, B H L., and Wankhade, R M., perimental Study of Endwall and Tip Clearance Flows in Two-Dimensional Turbine Rotor Blade Cascade, Effect of Incidence

Ex-Angle, Journal of Thermal Science, vol 9, no 1, pp 63–76,

cade, ASME Journal of Turbomachinery, vol 128, pp 300–309,

[50] Jin, P., and Goldstein, R J., Local Mass and Heat Transfer on a

Turbine Blade Tip, International Journal of Rotating Machinery,

vol 9, no 2, pp 81–95, 2003

[51] Jin, P., and Goldstein, R J., Local Mass/Heat Transfer on

Tur-bine Blade Near-Tip Surface, ASME Journal of Turbomachinery,

vol 125, pp 521–528, 2003

[52] Stephens, J., Corke, T., and Morris, S., Turbine Blade Tip age Control: Thick/Thin Blade Effects, 45th AIAA AerospaceScience Meeting and Exhibit, Paper 2007-0646, 2007.[53] Douville, T., Stephen, S J., Corke, T., and Morris, S., TurbineBlade Tip Leakage Control: By Partial Squealer Tip and PlasmaActuators, 44th AIAA Aerospace Science Meeting and Exhibit,9–12 January, Reno, NV, Paper 2006-20, 2006

Leak-[54] Srinivasan, V., and Goldstein, R J., Effect of Endwall Motion

on Blade Tip Heat Transfer, ASME Journal of Turbomachinery,

vol 125, pp 267–273, 2003

[55] Dorney, D J., Griffin, L W., and Huber, F W., A Study of the

Effects of Tip Clearance in a Supersonic Turbine, ASME Journal

of Turbomachinery, vol 122, pp 674–683, 2000.

[56] Green, B R., Barter, J W., Haldeman, C W., and Dunn, M G.,Averaged and Time-Dependent Aerodynamics of a High Pres-sure Turbine Blade Tip Cavity and Stationary Shroud: Compari-

son of Computational and Experimental Results, ASME Journal

International Journal of Rotating Machinery, vol 12, pp 1–15,

2006

[58] Lampart, P., Gardzilewicz, A., Yershov, S., and Rusanov, A.,

Investigation of Interaction of the Main Flow With Root and

Tip Leakage Flows in an Axial Turbine Stage by Means of a

Source/Sink Approach for a 3D Navier–Stokes Solver, Journal

of Thermal Science, vol 10, no 3, pp 198–204, 2001.

[59] Starodubtsev, Y V., Gogolev, I G., and Solodov, V G.,

Numer-ical 3D Model of Viscous Turbulent Flow in One Stage Gas

Turbine and Its Experimental Validation, Journal of Thermal

Science, vol 14, no 2, pp 136–141, 2005.

[60] Han, S., Han, B., Jin, P., and Goldstein, R J., Numerical

Predic-tion of the Flow Field Near the Tip of a Rotating Turbine Blade,

Journal of Engineering Physics and Thermophysics, vol 74,

no 4, pp 859–869, 2001

[61] Intaratep, N., Devenport, W J., and Staubs, J., The Tip

Leak-age Vortex Shed From an Unsteady Tip Clearance Flow, 34th

AIAA Fluid Dynamics Conference and Exhibit, June 28–July 1,

Portland, OR, AIAA-2004-2430, 2004

[62] Yang, H T., Chen, H C., and Han, J C., Turbine Rotor With

Various Tip Configurations Flow and Heat Transfer Prediction,

AIAA Journal of Thermophysics and Heat Transfer, vol 22,

no 2, pp 201–209, 2006

[63] Mumic, F., Eriksson, D., and Sunden, B., On Prediction of Tip

Leakage Flow and Heat Transfer in Gas Turbines, Proceedings

of ASME Turbo Expo 2004 Power for Land, Sea, and Air, June

14-17, Vienna, Austria GT-2004-53448, 2004

[64] Mumic, F., Eriksson, D., and Sunden, B., A Numerical

Investi-gation of Tip Leakage Heat Transfer and Fluid Flow for a Gas

Turbine Rotor Blade, 4th European Thermal Sciences

Confer-ence, Birmingham, UK, paper no 238, 2004.

[65] Mischo, B., Behr, T., and Abhari, R S., Flow Physics and

Profil-ing of Recessed Blade Tips—Impact on Performance and Heat

Load, ASME Journal of Turbomachinery, vol 130,

GT2006-91074, 2008

[66] Hamik, M., and Willinger, R., An Innovative Passive

Tip-Leakage Control Method for Axial Turbines: Basic Concept

and Performance Potential, Journal of Thermal Science, vol 16,

no 3, pp 215–222, 2007

[67] Prakash, C., Lee, C P., Cherry, D G., Doughty, R., and Wadia,

A R., Analysis of Some Improved Blade Tip Concepts, ASME

Journal of Turbomachinery, vol 128, pp 639–642, 2006.

[68] Han, J C., and Ekkad, S., Recent Development in Turbine Blade

Film Cooling, International Journal of Rotating Machinery,

vol 7, no 1, pp 21–39, 2001

[69] Kercher, D M., A Film-Cooling CFD Bibliography: 1976–

1996, International Journal of Rotating Machinery, vol 4, no 1,

pp 61–72, 1998

[70] Bunker, R S., A Review of Shaped Hole Turbine Film-Cooling

Technology, ASME Journal of Heat Transfer, vol 127, pp 441–

453, 2005

[71] Bogard, D G., and Thole, K A., Gas Turbine Film Cooling,

AIAA Journal of Propulsion and Power, vol 22, no 2, pp 249–

270, 2006

[72] Kim, Y W., Downs, J P., Soechting, F O., Abdel-Messeh, W.,

Steuber, G D., and Tanrikut, S., A Summary of the Cooled

Turbine Blade Tip Heat Transfer and Film Effectiveness

Inves-tigations Performed by Dr D E Metzger, ASME Journal of

Turbomachinery, vol 117, pp 1–11, 1995.

[73] Yang, H T., Chen, C H., and Han, J C., Film-Cooling Prediction

on Turbine Blade Tip with Various Film Hole Configurations,

AIAA Journal of Thermophysics and Heat Transfer, vol 22,

Breath-[77] Kim, Y M., and Metzger, D E., Heat Transfer and Effectiveness

on Film Cooled Turbine Blade Tip Model, ASME Journal of Turbomachinery, vol 117, pp 12–21, 1995.

[78] Christophel, J R., Thole, K A., and Cunha, F J., Cooling the Tip

of a Turbine Blade Using Pressure Side Holes, Part I—Adiabatic

Effectiveness Measurements, ASME Journal of Turbomachinery,

[82] Kwak, J S., and Han, J C., Heat Transfer Coefficients and

Film Cooling Effectiveness on a Gas Turbine Blade Tip, ASME Journal of Heat Transfer, vol 125, pp 494–502, 2003.

[83] Ahn, J Y., Mhetras, S., and Han, J C., Film Cooling ness on a Gas Turbine Blade Tip Using Pressure-Sensitive Paint,

Effective-ASME Journal of Heat Transfer, vol 127, pp 521–530, 2005.

[84] Gao, Z H., Narzary, D., Mhetras, S., and Han, J C., Full erage Film Cooling for a Turbine Blade With Axial Shaped

Cov-Holes, 39th AIAA Thermophysics Conference, 25–28 June,

Mi-ami, Florida, USA AIAA 2007-4031, 2007

[85] Gao, Z H., Narzary, D., Mhetras, S., and Han, J C., Effect

of Inlet Flow Angle on Gas Turbine Blade Tip Film Cooling,

Proceedings of GT2007, ASME Turbo 2007: Power for Land, Sea, Air, May 14–17, Montreal, Canada GT2007-27066, 2007.

[86] Ranson, W W., Adiabatic Effectiveness Measurements of age Flows Near the Hub Region of Gas Turbine Engines,M.S thesis, Virginia Polytechnic Institute and State University,Blacksburg, VA, 2004

Leak-[87] Saxena, V., Effect of Unsteady Wake, Free Stream Turbulence,Tip Geometry on Blade Tip Flow and Heat Transfer, M.S thesis,Louisiana State University, Baton Rouge, LA, 2003

[88] Kontrovitz, D M., Effect of Tip Geometry on Blade Tip, Flowand Heat Transfer, M.S thesis, Louisiana State University, BatonRouge, LA, 2002

B SUNDEN AND G XIE 551[89] Nasir H., Turbine Blade Tip Cooling and Heat Transfer, Ph.D.

thesis, Louisiana State University, Baton Rouge, LA, 2004

[90] Staubs, J K., Correlation Between Unsteady Loading and Tip

Gap Flow Occurring in a Linear Cascade With Simulated Stator–

Rotor Interaction, M.S thesis, Virginia Polytechnic Institute and

State University, Blacksburg, VA, 2005

[91] Azad, G M., Measurement and Analysis of Gas Turbine Blade

Tip Heat Transfer, Ph.D thesis, Texas A&M University, College

Station, TX, 2000

[92] Kwak, J S., Measurement and Analysis of Gas Turbine Blade

Tip Heat Transfer and Film Cooling, Ph.D thesis, Texas A&M

University, College Station, TX, 2002

[93] Phutthavong, P., Numerical Investigation of the Unsteady

Aero-dynamics of Blade Tip Leakage Flow Inside Gas Turbine

En-gines, M.S thesis, Concordia University, Concordia, Canada,

2006

[94] Xiao, Y., Numerical Simulations on Flow and Heat Transfer

in Turbine Cascade and Tip Clearance Over Shrouded Blades,

Ph.D thesis, University of Wisconsin–Milwaukee, Milwaukee,

WI, 2001

[95] Teng, S Y., Gas Turbine Blade Film Cooling and Blade Tip Heat

Transfer, Ph.D thesis, Texas A&M University, College Station,

TX, 2000

[96] Yoon, J H., and Martinez-Botas R F., Measurements of

Lo-cal Heat Transfer Coefficient and Film Cooling Effectiveness

in Turbine Blade Tip Geometries, Imperial College of Science,

Technology and Medicine, London, UK, 2001

[97] Mumic, F., Numerical Simulation of Some Heat Transfer and

Fluid Flow Phenomena for Gas Turbine Blades and a Transonic

Turbine Stage, Ph.D thesis, Division of Heat Transfer, Lund

University, Lund, Sweden, 2006

[98] Park, C W., and Lau, S C., Effect of Channel Orientation of

Local Heat (Mass) Transfer Distributions in a Rotating

Two-Pass Square Channel With Smooth Walls, ASME Journal of

Heat Transfer, vol 120, pp 624–632, 1998.

[99] Park, C W., Kandis, M., and Lau, S C., Heat/Mass Transfer

Distribution in a Rotating Two-Pass Square Channel, Part I:

Regional Heat Transfer, Smooth Channel, International Journal

of Rotating Machinery, vol 4, no 3, pp 175–188, 1998.

[100] Park, C W., Kukreja, R T., and Lau, S C., Heat/Mass Transfer

Distribution in a Rotating Two-Pass Channel With Transverse

Ribs, AIAA Journal of Thermophysics and Heat Transfer, vol 12,

no 1, pp 80–86, 1998

[101] Park, C W., Kukreja, R T., and Lau, S C., Heat/Mass Transfer

Distribution in a Rotating Two-Pass Channel With Angled Ribs,

International Journal of Rotating Machinery, vol 5, no 1, pp 1–

16, 1999

[102] Park, C W., Yoon, C., and Lau, S C., Heat (Mass) Transfer in

a Diagonally Oriented Rotating Two-Pass Channel With

Rib-Roughened Walls, ASME Journal of Heat Transfer, vol 122,

pp 208–211, 2000

[103] Kukreja, R T., Park, C W., and Lau, S C., Heat/Mass Transfer

Distribution in a Rotating Two-Pass Square Channel, Part II:

Lo-cal Transfer Coefficient, Smooth Channel, International Journal

of Rotating Machinery, vol 4, no 1, pp 1–15, 1998.

[104] Lee, S W, Ahn, H S., and Lau, S C., Heat (Mass) Transfer

Distribution in a Two-Pass Trapezoidal Channel With a 180-Deg

Turn, ASME Journal of Heat Transfer, vol 129, pp 1529–1537,

2007

[105] Mochizuki, S., Murata, A., Shibata, R., and Yang, J W., DetailedMeasurements of Local Heat Transfer Coefficients in TurbulentFlow Through Smooth and Rib-Roughened Serpentine PassagesWith a 180◦Sharp Bend, International Journal of Heat and Mass Transfer, vol 42, pp 1925–1934, 1999.

[106] Chen, Y., Nikitopoulos, D E., Hibbs, R., Acharya, S., andMyrum, T A., Detailed mass transfer distribution in a ribbedcoolant passage with a 180◦bend, International Journal of Heat and Mass Transfer, vol 43, pp 1479–1492, 2000.

[107] Astarita, T., and Cardone, G., Thermofluidynamic Analysis ofthe Flow in a Sharp 180◦Turn Channel, Experimental Thermal and Fluid Science, vol 20, pp 188–200, 2000.

[108] Astarita T., Cardone, G., and Carlomagno G M., ConvectiveHeat Transfer in Ribbed Channels With a 180◦Turn, Experiments

[113] Ekkad, S V., and Han, J C., Detailed Heat Transfer

Distribu-tions in Two-Pass Square Channels With Rib Turbulators, national Journal of Heat and Mass Transfer, vol 40, pp 2525–

[120] Hsieh, S S., and Chin, H J., Turbulent Flow in a Rotating Two

Pass Ribbed Rectangular Channel, ASME Journal of

Turboma-chinery, vol 125, pp 609–622, 2003.

[121] Hsieh, S S., Chen, P J., and Chin, H J., Turbulent Flow in a

Rotating Two Pass Smooth Channel, ASME Journal of Fluids

Engineering, vol 121, pp 725–734, 1999.

[122] Acharya, S., Eliades, V., and Nikitopoulos D E., Heat

Trans-fer Enhancements in Rotating Two-Pass Coolant Channels With

Profiled Ribs, Part I—Average Results, ASME Journal of

Tur-bomachinery, vol 123, pp 97–106, 2001.

[123] Nikitopoulos, D E., Acharya, S., and Eliades, V., Heat Transfer

Enhancements in Rotating Two-Pass Coolant Channels With

Profiled Ribs, Part II—Detailed Measurements, ASME Journal

of Turbomachinery, vol 123, pp 107–114, 2001.

[124] Liou, T M., and Chen, C C., Heat Transfer in a Rotating

Two-Pass Smooth Two-Passage With a 180◦ Rectangular Turn,

Interna-tional Journal of Heat and Mass Transfer, vol 42, pp 231–247,

1999

[125] Liou, T M., Tzeng, Y Y., and Chen, C C., Fluid Flow in a 180

Deg Sharp Turning Duct With Different Divider Thicknesses,

ASME Journal of Turbomachinery, vol 121, pp 569–576, 1999.

[126] Liou, T M., Chen, C C., and Chen, M Y., TLCT and LDV

Measurements of Heat Transfer and Fluid Flow in a Rotating

Sharp Turning Duct, International Journal of Heat and Mass

Transfer, vol 44, pp 1777–1787, 2001.

[127] Liou, T M., Chen, M Y., and Tsai, M H., Fluid Flow and

Heat Transfer in a Rotating Two-Pass Square Duct With

In-Line 90-Deg Ribs, ASME Journal of Turbomachinery, vol 124,

pp 260–268, 2002

[128] Liou, T M., Chen, M Y., and Chang, K H., Spectrum Analysis

of Fluid Flow in a Rotating Two-Pass Duct With Detached 90◦

Ribs, Experimental Thermal Fluid Science, vol 27, pp 313–321,

2003

[129] Liou, T M., Chen, M Y., and Wang, Y M., Heat Transfer, Fluid

Flow and Pressure Measurements Inside a Rotating Two-Pass

Duct With Detached 90 Deg Ribs, ASME Journal of

Turboma-chinery, vol 125, pp 565–574, 2003.

[130] Liou, T M., Chen, C C., and Chen, M Y., Rotating Effect on

Fluid Flow in Two Smooth Ducts Connected by a 180-Degree

Bend, ASME Journal of Fluids Engineering, vol 125, pp 327–

335, 2003

[131] Liou, T M., and Dai, G Y., Pressure and Flow

Characteris-tics in a Rotating Two-Pass Square Duct With 45-Deg Angled

Ribs, ASME Journal of Turbomachinery, vol 126, pp 212–219,

2004

[132] Liou, T M., Hwang, Y S., and Li, Y C., Flowfield and

Pres-sure MeaPres-surements in a Rotating Two-Pass Duct With Staggered

Rounded Ribs Skewed 45-Deg to the Flow, ASME Journal of

Turbomachinery, vol 128, pp 340–348, 2006.

[133] Liou, T M., Hwang, Y S., and Chen, M Y., Heat Transfer

Improvement by Arranging Detached Ribs on Suction Surfaces

of Rotating Internal Coolant Passages, International Journal of

Heat and Mass Transfer, vol 50, pp 2414–2424, 2007.

[134] Liou, T M., Chang, S W., Hung, J H., and Chiou, S F., High

Ro-tation Number Heat Transfer of a 45◦Rib-Roughened

Rectangu-lar Duct With Two Channel Orientations, International Journal

of Heat and Mass Transfer, vol 50, pp 4063–4076, 2007.

[135] Al-Hadhrami, L., and Han, J C., Effect of Rotation on Heat

Transfer in Two-Pass Square Channels With Five Different

Ori-entations of 45◦Angled Rib Turbulators, International Journal

of Heat and Mass Transfer, vol 46, pp 653–669, 2003.

[136] Al-Hadhrami, L., Griffith, T., and Han, J C., Heat Transfer

in Two-Pass Rotating Rectangular Channels (AR = 2) WithFive Different Orientations of 45 Deg V-Shaped Rib Turbula-

tors, ASME Journal of Heat Transfer, vol 125, pp 232–242,

2003

[137] Prabhu, S V., and Vedula, R P., Pressure Drop Characteristics

in a Rotating Smooth Square Channel With a Sharp 180◦Bend,

Experimental Thermal and Fluid Science, vol 21, pp 198–205,

2000

[138] Ratna-Rao, D V., and Prabuhu, S V., Pressure Drop Distribution

in Smooth and Rib Roughened Square Channel With Sharp 180◦

Bend in the Presence of Guide Vanes, International Journal of Rotating Machinery, vol 10, no 1, pp 99–114, 2004.

[139] Ratna-Rao, D V., Babu, C S., and Prabuhu, S V., Effect ofTurn Region Treatments on the Pressure Loss Distribution in aSmooth Square Channel With Sharp 180◦ Bend, International Journal of Rotating Machinery, vol 10, no 6, pp 459–468,

[145] Zhou, F G., Lagrone, J., and Acharya, S., Internal Cooling in

4:1 AR Passages at High Rotation Numbers, ASME Journal of Heat Transfer, vol 129, pp 1666–1675, 2007.

[146] Kim, K M., Lee, D H., and Cho, H H., Detailed Measurement ofHeat–Mass Transfer and Pressure Drop in a Rotating Two-Pass

Duct With Transverse Ribs, Heat and Mass Transfer, vol 43,

pp 801–815, 2007,[147] Kim, K M., Lee, D H., Rhee, D H., and Cho H H., LocalHeat/Mass Transfer Phenomena in Rotating Passage, Part 1:

Smooth Passage, AIAA Journal of Thermophysics and Heat Transfer, vol 20, no 2, pp 188–198, 2006.

[148] Kim, K M., Lee, D H., Rhee, D H., and Cho, H H., LocalHeat/Mass Transfer Phenomena in Rotating Passage, Part 2:

Angled Ribbed Passage, AIAA Journal of Thermophysics and Heat Transfer, vol 20, no 2, pp 199–210, 2006.

[149] Kim, K M., Lee, D H., and Cho, H H., Rotational Effects on

Pressure Drop in Smooth and Ribbed Two-Pass Ducts, AIAA Journal of Thermophysics and Heat Transfer, vol 21, no 3,

pp 644–647, 2007

B SUNDEN AND G XIE 553[150] Cho, H H Lee, S Y., and Rhee, D H., Effects of Cross Ribs

on Heat/Mass Transfer in a Two-Pass Rotating Duct, Heat and

Mass Transfer, vol 40, pp 743–755, 2004.

[151] Cho, H H Lee, S Y., and Rhee, D H., Heat/Mass Transfer in a

Two-Pass Rotating Duct With and Without 70-Deg Angled Ribs,

Heat and Mass Transfer, vol 40, pp 467–475, 2004.

[152] Bunker, R S., The Augmentation of Internal Blade Tip-Cap

Cooling by Arrays of Shaped Pins, Proceeding of GT2007,

ASME Turbo 2007: Power for Land, Sea, Air, May 14-17,

Mon-treal, Canada, GT2007-27009, 2007

[153] Chen, H C., Jang, Y J., and Han, J C., Near Wall

Second-Moment Closure for Rotating Multi-Pass Cooling Channels,

AIAA Journal of Thermophysics and Heat Transfer, vol 14,

no 2, pp 201–209, 2000

[154] Chen, H C., Jang, Y J., and Han, J C., Computation of Heat

Transfer in Rotating Two-Pass Square Channels by a

Second-Moment Closure Model, International Journal of Heat and Mass

Transfer, vol 43, pp 1603–1616, 2000.

[155] Jang, Y J., Chen, H C., and Han, J C., Computation of Flow and

Heat Transfer in Two-Pass Channels With 60 Deg Ribs, ASME

Journal of Heat Transfer, vol 123, pp 563–575, 2001.

[156] Jang, Y J., Chen, H C., and Han, J C., Numerical Prediction

of Flow and Heat Transfer in a Two-Pass Channel With 90 Deg

Ribs, International Journal of Rotating Machinery, vol 7, no 3,

pp 195–208, 2001

[157] Al-Qahtani, M., Jang, Y J., Chen, H C., and Han, J C.,

Predic-tion of Flow and Heat Transfer in Rotating Two-Pass

Rectan-gular Channels With 45-Deg Rib Turbulators, ASME Journal of

Turbomachinery, vol 124, pp 242–250, 2002.

[158] Al-Qahtani, M., Jang, Y J., Chen, H C., and Han, J C., Flow

and Heat Transfer in Rotating Two-Pass Rectangular Channels

(AR= 2) by Reynolds Stress Turbulence Model, International

Journal of Heat and Mass Transfer, vol 45, pp 1823–1838,

2002

[159] Lin, Y L., Shih, T I P., Stephens, M A., and Chyu, M K., A

Numerical Study of Flow and Heat Transfer in a Smooth and

Ribbed U-Duct With and Without Rotation, ASME Journal of

Heat Transfer, vol 123, pp 219–232, 2001.

[160] Suga, K., Predicting Turbulence and Heat Transfer in 3-D Curved

Ducts by Near-Wall Second Moment Closures, International

Journal of Heat and Mass Transfer, vol 46, pp 161–173, 2003.

[161] Suga, K., and Abe, K., Nonlinear Eddy Viscosity Modeling for

Turbulence and Heat Transfer Near Wall and Shear-Free

Bound-aries, International Journal of Heat Fluid Flow, vol 21, pp 37–

48, 2000

[162] Iacovides, H., The Computation of Turbulent Flow Through

Sta-tionary and Rotating U-Bends With Rib-Roughened Surfaces,

International Journal for Numerical Methods in Fluids, vol 29,

pp 865–876, 1999

[163] Nikas, K P., and Iacovides, H., The Computation of Flow and

Heat Transfer Through an Orthogonally Rotating Square-Ended

U-Bend Using Low-Reynolds-Number Models, International

Journal of Rotating Machinery, vol 3, pp 232–243, 2005.

[164] Raisee, M., Alemi, H., and Iacovides, H., Prediction of

Devel-oping Turbulent Flow in 90◦-Curved Ducts Using Linear and

Non-Linear Low-Re k–ε Models, International Journal for

Nu-merical Methods in Fluids, vol 51, pp 1379–1405, 2006.

[165] Murata, A., and Mochizuki, S., Centrifugal Buoyancy Effect on

Turbulent Heat Transfer in a Rotating Two-Pass Smooth Square

Channel With Sharp 180-Deg Turns, International Journal of Heat and Mass Transfer, vol 47, pp 3215–3231, 2004.

[166] Sleiti, A K., and Kapat, J S., Fluid Flow and Heat Transfer

in Rotating Curved Duct at High Rotation and Density

Ra-tios, ASME Journal of Turbomachinery, vol 127, pp 659–667,

2005

[167] Sleiti, A K., and Kapat, J S., Effect of Coriolis and CentrifugalForces on Turbulence and Transport at High Rotation and Den-

sity Ratios in a Rib-Roughened Channel, International Journal

of Thermal Science, vol 47, pp 609–619, 2008.

[168] Sleiti, A K., and Kapat, J S., Effect of Coriolis and Centrifugal

Forces at High Rotation and Density Ratios, AIAA Journal of Thermophysics and Heat Transfer, vol 20, vol 1, pp 67–79,

2006

[169] Etemad, S., and Sund´en, B., Heat Transfer Analysis of

Turbu-lent Flow in a Square-Sectioned U-Bend, ASME International Mechanical Engineering Congress and Exposition, November

13–19, Anaheim, CA, Paper no IMECE2004-60977, 2004.[170] Etemad, S., and Sund´en, B., Analysis of Developing and FullyDeveloped Turbulent Flow and Heat Transfer in a Square-

Sectioned U-Bend, ASME Summer Heat Transfer Conference,

July 17–22, San Franciso, CA, Paper no HT2005-72233, 2005.[171] Etemad, S., Sund´en, B., and Daunius, O., Turbulent Flow and

Heat Transfer in a Square-Sectioned U-Bend, Progress in putational Fluid Dynamics, vol 6, pp 89–100, 2006.

Com-[172] Guleren, K M., and Turan, A., Validation of Large-Eddy tion of Strongly Curved Stationary and Rotating U-Duct Flows,

Simula-International Journal of Heat and Fluid Flow, vol 28, pp 909–

Gas-[175] Peeyush, A., Heat–Mass Transfer in Smooth and Ribbed angular Serpentine Passages of Different Aspect Ratios and Ori-entation, Ph.D., Louisiana State University, Baton Rouge, LA,2004

Rect-[176] Stephens, M A., Modeling and Simulation of Turbulent Flowand Heat Transfer in Turbine-Blade Coolant Passages, Ph.D.thesis, Carnegie Mellon University, Pittsburgh, PA, 1996.[177] Park, J S., Turbulent Heat Transfer Enhancement in Rectangu-lar Channels With Two Opposite Rib-Roughened Walls, Ph.D.thesis, Texas A&M University, College Station, TX, 1986.[178] Chandra, P R., A Study of Local Heat–Mass Transfer Distribu-tions in Multipass Channels for Turbine Blade Cooling, Ph.D.thesis, Texas A&M University, College Station, TX, 1989.[179] Jang, Y J., Numerical Prediction of Flow and Heat Transfer inthe Coolant Passages of Gas Turbine Rotor Blade, Ph.D thesis,Texas A&M University, College Station, TX, 2000

[180] Al-Qahtani, M S., Computation of Flow and Heat Transfer inRotating Rectangular Channels With Angled Rib Turbulatorsfor Gas Turbine Blade, Ph.D thesis, Texas A&M University,College Station, TX, 2001

[181] Al-Hadhrami, L M., Rotating Heat Transfer in Turbine RotorBlade Cooling Channels With Turbulence Promoters, Ph.D the-sis, Texas A&M University, College Station, TX, 2002

heat transfer engineering vol 31 no 7 2010

[182] Liu, Y H., Effect of Rib Spacing on Heat Transfer and Friction

in a Rotating Two-Pass Rectangular (AR=1:2) Channel, M.S

Thesis, Texas A&M University, College Station, TX, 2005

[183] Su, G G., Numerical Simulation of Flow and Heat Transfer of

Internal Cooling Passage in Gas Turbine Blade, Ph.D thesis,

Texas A&M University, College Station, TX, 2005

[184] Fu, W L., Aspect Ratio Effect on Heat Transfer in Rotating

Two-Pass Rectangular Channels With Smooth Walls and Ribbed

Walls, Ph.D thesis, Texas A&M University, College Station, TX,

2005

[185] Sethuraman, E., Heat/Mass Transfer in Rotating, Smooth, High

Aspect-Ratio (4:1) Coolant Channels With Curved Walls, Ph.D

thesis, Louisiana State University, Baton Rouge, LA, 2006

[186] Sewall, E A., Large Eddy Simulations of Flow and Heat

Trans-fer in the Developing and 180◦ Bend Regions of Ribbed Gas

Turbine Blade Internal Cooling Ducts With Rotation—Effect of

Coriolis and Centrifugal Buoyancy Forces, Ph.D thesis, Virginia

Polytechnic Institute and State University, 2005

[187] Han, J C., and Chen, H C., Turbine Blade Internal Cooling

Passages With Rib Turbulators, AIAA Journal of Propulsion and

Power, vol 22, no 2, pp 226–248, 2006.

[188] Han, J C., and Wright, L M., Enhanced Internal Cooling of

Tur-bine Blades and Vanes, in The Gas TurTur-bine Handbook, chapter

4.2.2.2, USA, NETL (National Energy Technology Laboratory),

http://204.154.137.14/technologies/coalpower/turbines/refshelf

html, 2006

Bengt Sund´en received his M.S and Ph.D from

Chalmers University, G¨oteborg, Sweden He is rently professor of heat transfer and department head

cur-at Lund University, Sweden His main research terests are computational heat transfer, heat exchang- ers, transport phenomena in fuel cells, gas turbine heat transfer, combustion-related heat transfer, and enhanced heat transfer He has published more than

in-450 articles in well-recognized journals, books, and proceedings He has edited 25 books He is the editor-

in-chief of the International Journal of Heat Exchangers, and editor-in-chief for a book series, Developments in Heat Transfer In addition, he is on the edi-

torial boards for another four journals He is a fellow of the ASME and served

as associate editor of journal of Heat Transfer, 2005–2008 He is a honorary

professor of Xi’an Jiatong University, Xi’an, China.

Gongnan Xie is currently a postdoc at Lund

University, Sweden He received his Ph.D from the School of Energy and Power Engineering in Xi’an Jiaotong University in 2007, Xi’an, China He re- ceived his B.S degree in thermal and power engi- neering in 2002 from Guangdong Ocean University, Zhanjiang, China His research interests include com- putational fluid dynamics, numerical heat transfer, heat exchangers, gas turbine heat transfer, and appli- cation of computational intelligence in thermal engi- neering He is the author or co-author of more than 30 papers in international journals or conferences.

Copyright
C Taylor and Francis Group, LLC

ISSN: 0145-7632 print / 1521-0537 online

DOI: 10.1080/01457630903425429

Analysis of Infiltration, Solidification, and Remelting of a Pure Metal in

Subcooled Porous Preform

1Department of Mechanical and Aerospace Engineering, University of Missouri, Columbia, Missouri, USA

2College of Energy and Power Engineering, University of Shanghai for Science and Technology, Shanghai, China

The parts fabricated by selective laser sintering of metal powders are usually not fully densified and have porous structure.

Fully densified parts can be obtained by infiltrating liquid metal into the porous structure and solidifying the liquid

metal When the liquid metal is infiltrated into the subcooled porous structure, the liquid metal can be partially solidified.

Remelting of the partially solidified metal can also take place and a second moving interface can be present Infiltration,

solidification, and remelting of metal in a subcooled porous preform obtained by laser sintering of metal powders are

analytically investigated in this article The governing equations are nondimensionalized and the problem is described using

six dimensionless parameters The temperature distributions in the remelting and uninfiltrated regions were obtained by

an exact solution and an integral approximate solution, respectively The effects of porosity, Stefan number, subcooling

parameter, and dimensionless infiltration pressure are investigated.

INTRODUCTION

Selective laser sintering (SLS) [1] is a rapid

prototyp-ing/manufacturing technology that can fabricate functional parts

from powdered materials by a directed laser beam Fabrication

of a three-dimensional part by selective laser sintering is a

layer-by-layer additive process in which each layer is formed by

se-lectively sintering the powders with a focused laser beam The

laser beam scans a selective area and consolidates thin tracks of

powder Upon completion of sintering of one layer, the whole

powder bed is lowered and a fresh powder layer is spread to the

build zone The sintering process is repeated until the entire part

is fabricated Depending on the powder materials, the powder

particles can be bond together by glass transition of a polymer

melt (for thermoplastic powder, such as polycarbonate or nylon),

solid-state sintering (for ceramic or metal powders) or

liquid-phase sintering (for metal powders) The current state of SLS

in terms of material, laser, and process control of laser-material

interaction has been reviewed [2–4]

Support for this work by the Office of Naval Research (ONR) under grant

N00014-04-1-0303 and Chinese National Natural Science Foundation under

grant 50828601 is gratefully acknowledged.

Address correspondence to Professor Yuwen Zhang, Department of

Me-chanical and Aerospace Engineering, University of Missouri, Columbia, MO

65211, USA E-mail: zhangyu@missouri.edu

The parts produced by SLS with single- or component metal powders are usually not fully densified andhave porous structure In order to produce fully densified parts,

multiple-a postprocessing step is necessmultiple-ary The existing ing techniques include sintering, hot isostatic pressing (HIP)[5–7], and infiltration [8–10] Compared to sintering and HIPprocesses, the advantage of infiltration is that the full densitycan be achieved with hardly any shrinkage in postprocessing.The additional advantages of postprocessing with infiltration in-clude that it is relatively inexpensive and the tooling is similar

postprocess-to casting process Infiltration is a process where a liquid metal

is drawn into the pores of a porous solid (part produced by SLS

of metal powder) by capillary forces The liquid, as it advancesthrough the solid, displaces gas(es) from the pores and leavesbehind a relatively dense structure The rate of infiltration isrelated to the viscosity and surface tension of the liquid and tothe pore size of the SLS parts The infiltration process requiresthat the liquid is able to wet the solid and that the surface tension

of the liquid is high enough to induce capillary motion of theliquid metal into the pores of the porous solid In addition tocapillary force, the infiltration process can also be influenced bythe gravitational forces

In order to allow the liquid metal infiltrate into the pores inthe SLS parts, the pore structure generated during SLS needs

to be interconnected A large pore size is not desirable because

555

it cannot produce sufficient capillary force to drive the liquid

metal flow into the pore structure On the other hand, small pore

size will provide a small path for the liquid metal flow and higher

friction, which is not favorable to the infiltration Fluid flow in

porous media can find applications in many areas, and it is well

documented in the literature [11–14] While postprocessing of

laser sintered metal part by liquid metal infiltration and

solidifi-cation SLS is fairly new, infiltration and solidifisolidifi-cation have been

used in fabrication of metal-matrix composite (MMC) for a long

time [15] Mortensen et al [16] derived a general expression for

fluid flow and heat transfer during infiltration of pure metal into

a fibrous preform and obtained a similarity solution for the case

of unidirectional infiltration The same group also carried out

experimental study on infiltration of pure aluminum into a

fi-brous alumina perform, and good agreement with their model

was obtained [17] Infiltration and solidification/remelting of

the pure aluminum into a two-dimensional metal matrix were

studied numerically by Tong and Khan [18] The body-fitted

coordinates were used to deal with the transient and irregular

domains formed during infiltration The progress on

numeri-cal simulation of metal matrix composite and polymer matrix

composites processing by infiltration was reviewed by Lacoste

et al [19] During fabrication of MMC, external pressure is

usually applied to drive the liquid metal into the preform D¨uke

et al [20] developed a model describing infiltration of liquid

Sn60PbAg into laser-sintered bronze-nickel parts and validated

their model with experiments Yu and Schaffer [21] examined

the microstructural evolution during pressureless infiltration of

aluminum alloy parts fabricated by selective laser sintering

When infiltration is employed in the postprocessing of the

subcooled SLS parts, solidification accompanies the infiltration

process If the initial temperature of the preform is too low,

the solidification of the liquid metal may completely block the

path of liquid flow and prevent the liquid metal from thoroughly

infiltrating into the pores in the SLS parts Therefore, the SLS

part must be preheated to a temperature near the melting point

of the liquid metal On the other hand, the temperature of the

liquid metal that infiltrates the SLS parts must not be too high

because melting of the preform may occur and the part may be

distorted Therefore, the temperatures of both liquid metal and

preform are important processing parameters of postprocessing

of laser-sintered parts by infiltration In this paper, infiltration,

solidification, and remelting of metal in subcooled laser sintered

porous structure are analytically investigated The capability of

the developed model is demonstrated by studying infiltration

of pure liquid aluminum into a copper part fabricated by

se-lective laser sintering The developed model can be applied to

both pressured and pressureless infiltration because the pressure

difference for liquid metal is an independent parameter in this

paper The temperature distributions in the remelting and

unin-filtrated regions are obtained by an exact solution and an integral

approximate solution, respectively A processing map that can

aid the selection of various processing parameters is presented

The effects of porosity, Stefan number, subcooling parameter,

and dimensionless infiltration pressure are also investigated

Figure 1 Physical model.

PHYSICAL MODEL

Figure 1 shows the physical model of the infiltration problemunder consideration It is assumed that the liquid velocity is uni-form, i.e., a one-dimensional slug flow model is adopted Theporosity and initial temperature of the preform are uniformly

equal to ϕ and T i , respectively At time t = 0, the liquid metal

with a temperature T0 infiltrates the porous preform due to apressure difference induced by capillary force or external pres-

sure Since the initial temperature of the preform, T i, is wellbelow the melting point of the liquid metal, heat transfer fromliquid metal to the preform takes place as liquid metal infiltratesthe porous preform As a result, the liquid metal is cooled as itcontacts the porous preform Once the temperature of the liq-uid metal decreases to the melting point, partial solidification

of liquid metal occurs and the available flow area for the liquidmetal is reduced As the hot liquid metal flows into the preform,the partially solidified metal may be remelted Therefore, thereexist three regions: (1) remelting region, (2) solidification re-gion, and (3) uninfiltrated region, separated by two interfaces:

(1) remelting front (x = s), and (2) infiltration front (x = l)

(see Figure 1) Since infiltration is a slow process, it is assumedthat the liquid metal and the preform are in local thermal equi-librium so that the local instantaneous temperatures of liquidmetal and the preform can be represented by a single value [22,23] The temperature in the remelting region is higher than themelting point of the liquid metal, while the temperature in theuninfiltrated region is lower than the melting point of the liquidmetal In the solidification region, the temperature is uniformlyequal to the melting point of the liquid metal

In the remelting and solidification regions, the superficialvelocity of the liquid metal can be described by Darcy’s law:

u= −Kµ

∂p

where K is permeability In arriving at Eq (1), it is assumed

that viscosity of the liquid metal is constant In the remelting

front, and infiltration front, respectively In the remelting region,

the permeability can be expressed as [13]

K1= d

2

pϕ3

where d p is the diameter of the particle in the laser sintered

porous preform The permeability obtained from Eq (4) is valid

for packed spherical particles It is commonly used to obtain the

permeability of the sintered metal (especially copper) particle

wick in the heat pipes In the solidification region, the particle

size is increased to d psdue to solidification at the surface of the

particle If the fraction of liquid metal solidified in the

solidi-fication region is f , the particle diameter in the solidisolidi-fication

region can be obtained by a simple volume balance:

K2= d

2

ps[ϕ(1− f )]3

where ϕ(1− f ) is the porosity in the solidification region.

Combining Eqs (2) and (3) and considering Eqs (4)–(6), the

superficial liquid metal velocity becomes

where
p = p0−p lis the pressure difference between the inlet

and the infiltration front For postprocessing of the laser sintered

metal part, this pressure difference is created by capillary force

at the infiltration front or higher pressure of the liquid metal at

inlet The superficial velocity is related to the location of the

infiltration front by

u= ϕdl

dt (8)

where dl/dt is the velocity of the infiltration front Combining

Eqs (7) and (8) yields

where the subscripts p and l represent preform and liquid metal,

respectively The boundary conditions of Eq (10) are

The energy equation in the uninfiltrated region is(1− ϕ)ρp c p

∂T

∂t = k p,eff

∂2T

∂x2, x > l, t >0 (16)where the effective thermal conductivity in the uninfiltrated re-gion (sintered metal powder) can be expressed as

heat transfer engineering vol 31 no 7 2010

where L is a characteristic length, the Eqs (8)–(10), (13)–(15),

and (18)–(22) can be nondimensionalized to

The infiltration, solidification, and remelting problem is now

described by Eqs (24)–(34) and the dimensionless temperature

distribution is shown in Figure 2

SEMI-EXACT SOLUTION

At the point that is sufficiently far away from the infiltration

front in the uninfiltrated region, the dimensionless temperature

is equal to−Sc as indicated by Eq (34) One can define the

Figure 2 Dimensionless temperature distribution.

dimensionless thermal penetration depth,
, beyond which the

temperature of the uninfiltrated region is not affected by theliquid metal infiltration (see Figure 2), i.e., the dimensionlesstemperature satisfies the following two conditions at the thermalpenetration depth:

∂θ

∂X

Heat transfer in the uninfiltrated region can be solved by using

an integral approximate solution [14] Integrating Eq (30) in the

interval of (,
), one obtains

X =
− ∂θ

∂X

X =

(37)

where the first term on the right-hand side is zero according to

Eq (36) Equation (37) can be rewritten into the following form

by using Leibnitz’s rule:

d

dτ ( + Sc
) = − ∂θ

∂X

X =

(38)where

=



Assuming the temperature distribution in the uninfiltratedregion is a second-degree polynomial function and determiningthe constants using Eqs (32), (35), and (36), one obtains:

= 2λ√τ (43)

= 2δ√τ (44)where λ and δ are two constants that need to be determined.Substituting Eqs (43) and (44) into Eqs (41) and (42), the

Y ZHANG ET AL 559

following equations about λ and δ can be obtained:

λ= δ+

9δ2− 24

δ= λ +2(1− ϕ)SteSc

It is seen that the solid fraction, f , in the solidification region

appears in Eq (46), and it must be determined before λ and δ

can be solved for from Eqs (45) and (46)

The temperature distribution in the liquid region can be

ob-tained by a similarity solution Inspired by Eqs (43) and (44),

one can introduce the following similarity variable:

is the dimensionless thermal diffusivity in the remelting region

The energy Eq (26) of the remelting region can be transformed

into the following ordinary differential equation:

where

σ = γϕ[(1− ϕ) + γϕ]√α¯c

(50)

is the heat capacity of liquid in the remelting region The

bound-ary conditions specified by Eqs (27) and (28) become

is a constant that describes the location of remelting front The

general solution of Eq (49) is

e−η 2

dη (55)

which is an odd function that satisfies erf(−η) = −erf(η)

Af-ter deAf-termining the two constants C1 and C2 in Eq (54) from

Eqs (51) and (52), the temperature distribution in the remelting

where the solid fraction, f , in the solidification region is still

unknown at this point

Substituting Eqs (43), (44), and (53) into Eq (25), an tion for the solid fraction is obtained as follows:

which serves as a bridge between the solutions in the filtrated region and the remelting regions The infiltration, so-lidification, and remelting problem is now described by four

unin-unknowns, δ, λ, β and f , which can be solved iteratively from

four algebraic equations, Eqs (45), (46), (58), and (59)

RESULTS AND DISCUSSION

The infiltration, solidification, and remelting problem is scribed by six dimensionless parameters: the heat capacity ra-tio, γ, thermal conductivity ratio, κ, porosity, ϕ, Stefan number,

de-Ste, subcooling parameter, Sc, and dimensionless pressure ference, P While there is no apparent relationship among these

dif-six parameters, improper combination of these parameters mayresult in complete solidification of the liquid metal near theinlet and further infiltration will not be possible On the otherhand, if the liquid metal inlet temperature and/or the initialtemperature are sufficiently high, the solidification region maynot appear

A simple processing map can be obtained by analyzing the

energy balance when a preform with a small volume,
V , and

porosity ϕ is infiltrated by the liquid metal Although the sult of such analysis is accurate only if the heat conduction inboth remelting and uninfiltrated regions is negligible, this willprovide a first-order estimation on the appropriateness of theprocessing parameters The amount of sensible heat required tobring the temperature of the preform to the melting point of the

re-liquid metal is q p,s = (1 − ϕ)
V ρ p c p (T m − T i) The amount

of sensible heat that can be released by the liquid metal whenits temperature decreases from its initial value to its melting

point is q l,s = ϕ
V ρ l c l (T0− T m) If all of the liquid

infil-trated into
V is solidified, the amount of latent heat released is

q l,l = ϕ
V ρ l h sl Therefore, the condition under which liquidheat transfer engineering vol 31 no 7 2010

is not completely solidified is q l,s + q l,l > q p,s, i.e.,

ϕ
V ρ l c l (T0− T m)+ ϕ
V ρ l h sl

>(1− ϕ)
V ρ p c p (T m − T i) (60)Substituting Eq (23) into Eq (60) yields

Sc < Sc max = ϕγ

1− ϕ



1+ 1Ste



(61)

where Sc max is the maximum allowable subcooling parameter,

above which infiltration becomes impossible

On the other hand, solidification will occur only if q l,s < q p,s,

where Sc minis the minimum subcooling parameter below which

there will be no solidification Figure 3 shows a processing map

for infiltration of liquid aluminum into the preform fabricated

by laser sintering copper powder particles The heat capacity

ratio and thermal conductivity ratio for this combination are

γ = 0.664 and κ = 0.257, respectively As the Stefan

num-ber increases, Sc max significantly decreases, while Sc min is not

affected by change of Stefan number Although Sc max is the

upper limit of subcooling parameter above which infiltration is

impossible, Sc minis only the limit below which no solidification

will occur When Sc is below Sc min, infiltration is still possible

except there will be no solidification region and the solution will

be much simpler A subcooling parameter below Sc min means

that the preform must be preheated to an initial temperature very

close to the melting temperature of the liquid metal In this

pa-per, only the cases when the subcooling parameters are between

Sc min and Sc maxare investigated

Figure 3 Processing map.

Figure 4 Temperature distributions at different porosity (Ste= 0.1, Sc = 2, and P= 20).

Figure 4 shows the temperature distributions at differentporosity while all other parameters are kept at Ste= 0.1, Sc =

2, and P = 20 For the aluminum–copper system, the Stefannumber of 0.1 and subcooling parameter of 2 correspond to a

liquid metal inlet temperature of T0= 967◦C and an initial

tem-perature of T i = 865◦C For the porous preform obtained by

sintering 50-µm copper particles, the dimensionless pressure

of P = 20 corresponds to a pressure difference of 1.39 MPa,which is within the range used by Masur et al [17] in theirexperiments The values of β that represent the locations of theremelting front for the three porosities are 0.59, 0.54, and 0.43,respectively Thus, the velocity of the remelting front decreaseswith increasing porosity The values of λ that represent the loca-tions of infiltration fronts are 2.67, 2.19, and 1.72, respectively.Therefore, the velocity of the infiltration front significantly de-creases with increasing porosity The thermal penetration depthalso decreases with increasing porosity The values of δ thatrepresent the thermal penetration depth are 3.89, 3.10, and 2.44,respectively The solid fractions for these three cases are 0.430,0.451, and 0.490, respectively

Figure 5 shows the temperature distributions at different fan number while all other parameters are kept at ϕ= 0.4, Sc =

Ste-2, and P = 20 It can be seen that the velocity of the remeltingfront increases with increasing Stefan number The values of βfor the three Stefan numbers are 0.29, 0.54, and 0.79, respec-tively The effect of Stefan number on the velocity of infiltrationfront is more significant than its effect on the remelting front:The values of λ for the three Stefan numbers are 1.09, 2.19, and2.86, respectively The values of δ for the three Stefan numbersare 1.72, 3.10, and 4.23, respectively The solid fractions forthese three cases are 0.660, 0.451, and 0.346, respectively.Figure 6 shows the temperature distributions at differentsubcooling parameters while the other parameters are kept atSte= 0.1, ϕ = 0.4, and P = 20 The velocity of the remelting

front increases with increasing subcooling parameter The ues of β for the three subcooling parameters are 0.48, 0.54, and

val-Y ZHANG ET AL 561

Figure 5 Temperature distributions at different Stefan number (ϕ= 0.4,

Sc = 2, and P = 20).

0.60, respectively The effect of the subcooling parameter on the

velocity of the infiltration front is more significant than its

ef-fect on the remelting front: The values of λ for three subcooling

parameters are 1.76, 2.19, and 2.55, respectively The thermal

penetration depth significantly increases with increasing

sub-cooling: The values of δ for the three subcooling parameters are

2.48, 3.10, and 3.69, respectively The solid fraction decreases

with increasing subcooling parameters: The solid fractions for

these three cases are 0.531, 0.451, and 0.388, respectively This

is because the width of the solidification region significantly

increases with increasing subcooling parameter

Figure 7 shows the temperature distributions at different

di-mensionless pressure differences while the other parameters are

kept at Ste = 0.1, Sc = 2, and ϕ = 0.3 It can be seen that

the effect of P on β is not significant: The values of β for the

three pressure differences are 0.62, 0.54, and 0.50, respectively

Figure 6 Temperature distributions at different subcooling parameters (Ste =

infil-CONCLUSIONS

Infiltration, solidification, and remelting of metal in cooled laser sintered porous structure are analyzed in this paper.The governing equations are nondimensionalized and the prob-lem is described by six dimensionless parameters: the heat ca-pacity ratio, γ, thermal conductivity ratio, κ, porosity, ϕ, Ste-

sub-fan number, Ste, subcooling parameter, Sc, and dimensionless pressure difference, P A processing map that identifies the

conditions of complete solidification and no solidification wasobtained by analyzing the overall energy balance of the con-trol volume As the Stefan number increases, the maximumallowable subcooling parameter significantly decreases, whilethe minimum subcooling parameter, below which no meltingwill occur, is not affected by change of Stefan number

The effects of porosity, Stefan number, subcooling eter, and dimensionless pressure difference on the infiltrationare investigated The velocity of the remelting front decreasesheat transfer engineering vol 31 no 7 2010

param-with increasing porosity and dimensionless pressure difference,

but increases with increasing Stefan number and subcooling

parameter The velocity of infiltration front increases with

de-creasing porosity and pressure difference, and with inde-creasing

Stefan number and subcooling parameter The solid fraction in

the solidification region increases with increasing porosity and

dimensionless pressure difference, and with decreasing Stefan

number and subcooling parameter

NOMENCLATURE

c specific heat (J/kg-K)

d p diameter of the particle in the laser-sintered

preform (m)

d ps particle diameter after partial solidification (m)

f mass fraction of solid in the solidification region

h sl latent heat of melting or solidification (J/kg)

P dimensionless pressure difference

s location of remelting front (m)

S dimensionless location of remelting front, s/L

Sc subcooling parameter, (T m − T i )/(T0− T m)

Ste Stefan number, c l (T0− T m )/ h sl

T0 inlet temperature of liquid metal (K)

T i initial temperature of preform (K)

[1] Beaman, J J., Barlow, J W., Bourell, D L., Crawford, R H.,

Marcus, H L., and McAlea, K P., Solid Freeform Fabrication: A New Direction in Manufacturing, Kluwer Academic Publishers,

Dordrecht, The Netherlands, 1997

[2] Kruth, J P., Wang X., Laoui, T., Froyen, L., Lasers and Materials

in Selective Laser sintering, Assembly Automation, vol 23, no 4,

pp 357–371, 2003

[3] Kumar, S., Selective Laser Sintering: A Qualitative and Objective

Approach, JOM—Journal of Minerals, Metals, Material Society,

Rapid Prototyping Journal, vol 4, pp 112–117, 1998.

[6] Das, S., Wohlert, M., Beaman, J J., and Bourell, D L., ProducingMetal Parts With Selective Laser Sintering/Hot Isostatic Press-

ing, JOM—Journal of the Minerals, Metals and Material Society,

Y ZHANG ET AL 563

[11] Nield, D A., and Bejan, A., Convection in Porous Media,

Springer-Verlag, New York, 1992

[12] Kaviany, M., Principles of Heat Transfer in Porous Media, 2nd

ed., Springer Verlag, New York, 1995

[13] Faghri, A., Heat Pipe Science and Technology, Taylor & Francis,

Bristol, PA, 1995

[14] Faghri, A., and Zhang, Y., Transport Phenomena in Multiphase

Systems, Elsevier, Burlington, MA, 2006.

[15] Mortensen, A., Cornie, J A., and Flemings M C.,

Solidifica-tion Processing of Metal-Matrix Composites, Journal of Metals,

vol 40, no 2, pp 12–19, 1988

[16] Mortensen, A., Masur L J, Cornie J A., and Flemings, M C.,

Infiltration of Fibrous Preform by a Pure Metal: Part I Theory,

Metallic Trans A vol 20A, pp 2535–2547, 1989.

[17] Masur, L J., Mortensen, A., Comie, J A, and Flemings, M

C., Infiltration of Fibrous Preforms by a Pure Metal: Part II

Experiment, Metallic Trans A vol 20A, pp 2549–2557,

1989

[18] Tong, X., and Khan, J A., Infiltration and

Solidifica-tion/Remelting of a Pure Metal in a Two-Dimensional Porous

Preform, ASME Journal of Heat Transfer, vol 118, pp 173–180,

1996

[19] Lacoste, E., Mantaux, O., and Danis, M., Numerical Simulation

of Metal Matrix Composites and Polymer Matrix Composites

Processing by Infiltration: A Review, Composites, Part A, vol 33,

pp 1605–1614, 2002

[20] D¨uke, J., Mienling, F., Neeße, T., and Ptto, A., Infiltration as

Post-Processing of Laser Sintered Metal Parts, Powder Technology,

vol 145, pp 62–68, 2004

[21] Yu, P., and Schaffer, G B., Microstructural Evolution During

Pressureless Infiltration of Aluminium Alloy Parts Fabricated by

Selective Laser Sintering, Acta Materialia, vol 57, pp 163–170,

2007

[22] Khan, M A., and Rohatgi, P K., Numerical Solution to a

Mov-ing Boundary Problem in a Composite Medium, Numerical Heat

Transfer, Part A, vol 25, pp 209–221, 1994.

[23] Cantarel, A., Lacoste, E., Danis, M., and Arquis, E., Metal trix Composites Processing: Numerical Study of Heat Transfer

Ma-Between Fibers and Metal, International Journal of Numerical Methods for Heat and Fluid Flow, vol 15, no 8, pp 808–826,

2005

Yuwen Zhang is a professor of mechanical and

aerospace engineering at the University of Missouri, Columbia, Missouri His research interests include phase-change heat transfer, heat pipes, ultrafast, ultra- intense laser materials processing, and transport phe- nomena in materials processing and manufacturing.

He is author of more than 100 journal papers and more than 70 conference papers, as well as co-author of two

textbooks: Transport Phenomena in Multiphase tems and Advanced Heat and Mass Transfer He is

Sys-a recipient of the 2002 Office of NSys-avSys-al ReseSys-arch (ONR) Young InvestigSys-ator Award He is a Fellow of the ASME and Associate Fellow of AIAA.

Piyasak Damronglerd is a Ph.D student in the

De-partment of Mechanical and Aerospace ing at the University of Missouri, Columbia, Mis- souri He received a B.S in mechanical engineering

Engineer-in 1999 from Chulalongkorn University, Thailand, and an M.S in mechanical engineering in 2003 from Southern Illinois University, Edwardsville His re- search interests include enhancement of convective heat transfer in duct flows, and transport phenomena

in porous media.

Mo Yang is a professor and director of the Thermal

Engineering Institute at the University of Shanghai for Science and Technology, Shanghai, China He received his Ph.D in engineering thermophysics from Xi’an Jiaotong University in 1991 His main research interests are numerical heat transfer in fluid flow and multiphase fluid flow.

heat transfer engineering vol 31 no 7 2010

Copyright
Taylor and Francis Group, LLC

ISSN: 0145-7632 print / 1521-0537 online

DOI: 10.1080/01457630903425510

Ceramic Miniature Heat Pipes and

Liquid Charging Methods

MINGCON GAO, YIDING CAO, and MARC A ZAMPINO

Department of Mechanical and Materials Engineering, Florida International University, Miami, Florida, USA

Three working-liquid charging methods for miniature heat pipes are introduced, and their advantages and disadvantages

are described The methods are referred to as the micro-syringe method, thermodynamic equilibrium method, and

capillary-tubing method Using these methods, two types of ceramic heat pipes were charged and tested The ceramic heat pipes were

made of alumina and have overall dimensions of 89 mm × 12 mm × 2.9 mm and a designed vapor space of 82.5 mm ×

4.1 mm × 1.27 mm Axial micro-capillary grooves were provided on the top and bottom or sidewalls inside the heat pipes

as wick structures Water was used as the working liquid More than 20 W of heat input was achieved on a 5 mm × 5 mm

heating surface The corresponding heat flux was 80 W/cm 2

INTRODUCTION

Electronic system designers are facing a great challenge to

continuously reduce system volume and increase electronics

complexity and power density Consequently, this trend places

an ever-increasing importance on the thermal management at

the system and packaging levels The thermal management at

system level is usually not a major problem, as adequate cooling

schemes are available However, the packaging level is in

dra-matic need of special cooling techniques due to the local high

heat flux It is not an overstatement that thermal barrier emerges

again as a constraint in the advances in microelectronics and

optoelectronics technology, especially in military applications

Ceramics are widely used in microelectronics technology for

substrates, module covers, sealing material for modules,

com-ponents of thin film conductor, resistors, dielectrics, and so on,

because of their unique combination of mechanical, dielectric,

physical, and chemical properties Among various ceramic

ma-terials, high- and low-temperature alumina substrates are

popu-lar in the electronic packaging industry However, their thermal

conductivity is not adequate for high-power application

By directly integrating miniature heat pipes into substrates

and combining the functions of the electrical interconnection

and heat sink, a uniform temperature field could be obtained

The overall thermal resistance from the heat source to the heat

Address correspondence to Professor Yiding Cao, Department of

Mechan-ical and Materials Engineering, Florida International University, Miami, FL

33174, USA E-mail: Yiding.Cao@fiu.edu

sink could be greatly reduced and the local high heat powercould be efficiently spread onto a large area, which results in

a moderate heat flux that can be handled by the conventionalair cooling This technique would be helpful for the multichip,multilayer substrate module

Liquid charging is critical to any micro/miniature heat pipes

It is extremely difficult to precisely measure and control a smallamount of working liquid on the order of magnitude of 0.01 g

in a vacuum environment In this study, the vapor space is verysmall and the proper charging is less than 0.1 g of water Poorcharging methods may fully fill the heat pipes or leave themwith no liquid at all

The effects associated with less liquid include prematuredry-out at the evaporation section due to an unsaturated wickstructure The dry-out causes a rapid increase in the local tem-perature; hence the electronic components being cooled by theheat pipe can suffer thermal failure A large temperature gra-dient at the evaporator section experiencing dry-out can alsocause a failure in the ceramic material due to thermal stressand/or thermal shock Overcharging the heat pipe can lead to aflooded and blocked condensation section by the excess liquid.This challenge is further complicated by the necessity of an ini-tial high vacuum and its maintenance during the liquid chargingprocess

An initial high vacuum itself is difficult to be obtained inmicro/miniature heat pipes via a small-diameter filling tube Asimple test was conducted to simulate this environment Thesetup, as shown in Figure 1, consists of a turbo molecular vac-uum pump system and two vacuum meters located at both ends

M GAO ET AL 565

Figure 1 Vacuum test setup.

of a Tygon manifold The sensor of vacuum meter 1 is mounted

directly at the pump inlet port and that of vacuum meter 2 is

connected via a small tube at the other end The two sensors are

1 m apart The small tube represents the typical filling tube of

a miniature heat pipe The small tube has a length of 40 mm

and an inside diameter of 0.88 mm For comparison, the test

was carried out with or without the small tube or just with a

1.6 mm of hole The results are shown in Figure 2 It can be

seen from the figure that an obvious vacuum difference exists

between the two meters With a small filling tube, it is difficult

to develop high vacuum inside the heat pipe After evacuating

for 30 min, the pressure reading of meter 2 is still above 10−2

torr The knowledge of this situation is important to choosing a

liquid charging method

In the present paper, emphasis is placed on the liquid

charg-ing techniques and the verification of the techniques through

performance of the heat pipes The design concerns related to

miniature heat pipes and cofire fabricating issues of the

minia-ture heat pipes were addressed in previous publications [1–4]

Plesch et al [5] charged their miniature copper heat pipes

with dimensions of 7 mm× 2 mm × 120 mm by introducing

condensing water vapor They indicated that it was not possible

Figure 2 Vacuum test results.

to control and determine the amounts of water present during theprocess of filling After pinching off the fill tubes, the amount ofcharge was determined by weighing the charged heat pipe Theirrequired liquid was more than 0.15 g, which is much more thanthat in the current study They mentioned that a filling stationwas under construction, which would allow very small amounts

of degassed clean water to be metered in, but the follow-upreport has not been published yet

Badran et al [6] tested a micro heat pipe array anisotropicallyetched on a single crystalline semiconductor silicon wafer Twosets were fabricated, each of which had 73 and 127 micro heatpipes, respectively, in the silicon wafer The condenser section

of the heat pipe array was connected to a common reservoir forfilling procedures In order to measure the amount of workingliquid visually, a Pyrex glass wafer was used to seal the pipearray Similarly, Duncan and Peterson [7] charged their microheat pipe array visually one by one The heat pipes were sealedalso by 200 µm thick Pyrex glass and were evacuated to apressure of 10−3 torr A working liquid was introduced withthe aid of a syringe and a trough The vapor became trapped

in the evaporator end and the fluid filled the remaining volume ofthe heat pipes The wafer was removed from the vacuum systemwhile being heated from the evaporator end The extra liquid wasthen removed by vapor expansion until the desired fluid quantitywas measured via a caliper visually through Pyrex glass Benson

et al [8] tried to measure the methanol quantity during the fillingprocess by pumping the fully filled heat pipe through a seriescombination of a needle valve and heated capillary tubing Thetubing was heated to 125◦C, ensuring that the fluid passingthrough was in the vapor state to allow metering The heat pipearray was immersed in a temperature-controlled bath maintained

at 26◦C during the vacuum pumping process By monitoring thepressure change, the full evacuating time needed for the heatpipe array to reach the completely dry state was determinedpreviously Then by pumping out the fully filled heat pipe arrayfor a fraction of the full time, for example, 80%, the remainingliquid would be 20% of the initial volume This is reasonableunder the assumption that a linear relationship exists betweenmass removal rate and the pumping time

To sum up, there is no existing charging method available forour ceramic heat pipes It is necessary to develop a reliable andpractically feasible charging method

LIQUID CHARGING METHODS

Three methods have been developed by the authors for ing micro/miniature heat pipes, and are discussed in the follow-ing subsections

charg-Micro-Syringe Method

VICI Series A Pressure-Lok syringe, designed primarily forprecision measurement and displacement of liquid or gaseousheat transfer engineering vol 31 no 7 2010