Assessment of the eﬀectiveness of nanoﬂuids for single

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International Journal of Heat and Mass Transfer 50 (2007) 452–463

Assessment of the effectiveness of nanofluids for single-phaseand two-phase heat transfer in micro-channels

Jaeseon Lee, Issam Mudawar *

Boiling and Two-Phase Flow Laboratory, School of Mechanical Engineering, Purdue University, West Lafayette, IN 47907, USA

Received 24 March 2006; received in revised form 2 August 2006Available online 2 October 2006

Abstract

Experiments were performed to explore the micro-channel cooling benefits of water-based nanofluids containing small concentrationsof Al2O3. The high thermal conductivity of nanoparticles is shown to enhance the single-phase heat transfer coefficient, especially forlaminar flow. Higher heat transfer coefficients were achieved mostly in the entrance region of micro-channels. However, the enhancementwas weaker in the fully developed region, proving that nanoparticles have an appreciable effect on thermal boundary layer development.Higher concentrations also produced greater sensitivity to heat flux. Despite this enhancement, the overall cooling effectiveness of nano-particles was quite miniscule because of the large axial temperature rise associated with the decreased specific heat for the nanofluid com-pared to the base fluid. For two-phase cooling, nanoparticles caused catastrophic failure by depositing into large clusters near thechannel exit due to localized evaporation once boiling commenced. These and other practical disadvantages bring into question the over-all merit of using nanofluids in micro-channel heat sinks.� 2006 Elsevier Ltd. All rights reserved.

1. Introduction

Solid particles generally possess far greater thermal con-ductivity than common heat transfer liquids. Mixing solidparticles in a liquid can therefore enhance the coolingpotential of the liquid by increasing the thermal conductiv-ity of the suspended fluid.

In the past, this goal was sought by using relatively large(‘milli’ or ‘micro’ scale) particles. While these fluids doprovide the aforementioned cooling benefit, their imple-mentation is complicated by their tendency to damagemechanical pumps and valves, as well as clog flow passagesdue to particle settling.

Recent developments in nanotechnology and relatedmanufacturing techniques have made possible the produc-tion of far smaller – nanosized – particles. Such particlesgreatly reduce the damage to flow loop components whileproviding the cooling benefits. Choi and co-workers at

0017-9310/$ - see front matter � 2006 Elsevier Ltd. All rights reserved.

doi:10.1016/j.ijheatmasstransfer.2006.08.001

* Corresponding author. Tel.: +1 765 494 5705; fax: +1 765 494 0539.E-mail address: [email protected] (I. Mudawar).

Argonne National Laboratory proposed the use of nano-particles to enhance the thermal conductivity of liquids,and coined the term ‘nanofluids’ for the resulting mixtures[1–3]. Later studies experimentally demonstrated thisenhancement effect. Lee et al. [3] used 50 nm or smallerAl2O3 and CuO particles to enhance the thermal conduc-tivity of their nanofluids. They showed the percentageenhancement in thermal conductivity was not only a func-tion of concentration and conductivities of the particlematerial and liquid, but particle size and shape as well.Xuan and Li [4] reported that copper particles providedgreater thermal conductivity enhancement than metaloxide particles. Xue [5] showed analytically that carbonnanotubes produce the highest conductivity enhancementof all nanofluids.

The aforementioned clogging problem is perhaps mostserious in micro-channel flow devices recently recom-mended for high-heat-flux electronic and aerospace applica-tions. These devices offer a number of cooling advantagessuch as high convective heat transfer coefficients, small sizeand weight, and reduced coolant inventory. Nanofluids are

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Nomenclature

CHF critical heat fluxcp specific heat (kJ/kg K)D diameterDh hydraulic diameterf fanning friction factorh heat transfer coefficient (W/m2 K)Hch micro-channel heightHtc distance from thermocouple to base of micro-

channelk thermal conductivity (W/m K)_m mass flow rate (g/s)Nu Nusselt numbern solid particle shape factorP pressure (bar)Pr Prandtl numberDP pressure drop (Pa, bar)Q heat input to heat sink (W)q00 heat flux through heat sink base area (W/m2 or

W/cm2)Re Reynolds number based on tube diameterReDh Reynolds number based on channel hydraulic

diameterT temperature (�C)tc temperature measurement points inside heat

sink blocku mean fluid velocity [m/s]

Wch micro-channel widthWw half-width of copper sidewallx axial distance (m)

Greek symbols

g fin efficiencyl viscosityq densityu volumetric concentration of nanoparticles

Subscripts

1,2,3,4 temperature measurement locationsbf base fluidch channelD diameterHFE 7100 HFE 7100 base fluidin channel inletm bulk fluidnf nanofluidout channel outletp solid particless solid metal for heat sinksp single-phaseWater water base fluidw channel wallx local condition

Fig. 1. Uniformly heated flow in circular tube.

J. Lee, I. Mudawar / International Journal of Heat and Mass Transfer 50 (2007) 452–463 453

considered an ideal fit for micro-channel devices becausethey are thought to virtually eliminate the clogging prob-lem. Furthermore, they can enhance the already substantialheat transfer coefficients realized with micro-channel flowby the added benefit of greater thermal conductivity.

Investigating the combined advantages of micro-channelflow and nanofluids is precisely the goal of the presentstudy. Those advantages are examined experimentally forboth single-phase and two-phase situations, and practicalconclusions are drawn concerning the implementation ofnanofluids in small coolant passages.

2. Enhancement for laminar versus turbulent flow

Aside from thermal conductivity, other fluid propertieswill change by the particle inclusion, and it is the combinedeffect of these property variations that dictates the overallenhancement effect for a nanofluid. This point is illustratedwith the aid of a simple example that will also show howthis effect is quite different for laminar flow as comparedto turbulent flow.

Fig. 1 shows a fluid being heated uniformly as it flowsinside a circular tube. Ignoring entrance effects for now(these effects will be discussed latter in this paper), the flowis assumed both hydraulically and thermally fully devel-

oped. The Nusselt number is constant for laminar flowand follows the Dittus–Boelter equation for turbulent flow.

Nu ¼ 4:36 for laminar flow ð1Þ

and

Nu ¼ 0:023Re0:8Pr0:4 for turbulent flow ð2Þ

Lee et al. [3] showed that the thermal conductivity ofAl2O3 nanofluids follows the two-components mixture con-ductivity model of Hamilton–Crosser model fairly well.

knf ¼kp þ n� 1ð Þkbf � n� 1ð Þu kbf � kp

� �kp þ n� 1ð Þkbf þ u kbf � kp

� �" #

kbf ; ð3Þ

Tab

le1

Al 2

O3

nan

op

arti

cle

susp

ensi

on

flu

idp

rop

erti

esat

30�C

and

100

kP

a

Wat

erb

ase

HF

E71

00b

ase

u=

0%1%

2%3%

4%5%

u=

0%1%

2%3%

4%5%

kn

f(W

/mK

)0.

603

0.62

00.

638

0.65

60.

675

0.69

30.

0678

0.06

990.

0720

0.07

410.

0763

0.07

85q

nf

(kg/

m3)

995.

710

21.7

1047

.710

73.8

1099

.811

25.9

1455

.414

76.9

1498

.315

19.8

1541

.215

62.7

ln

f(k

g/m

s)7.

977�

10�

48.

177�

10�

48.

376�

10�

48.

576�

10�

48.

775�

10�

48.

974�

10�

46.

556�

10�

46.

720�

10�

46.

884�

10�

47.

048�

10�

47.

212�

10�

47.

376�

10�

4

c p,n

f(k

J/k

gK

)4.

183

4.14

94.

115

4.08

14.

046

4.01

21.

193

1.18

91.

184

1.18

01.

176

1.17

2

So

lid

(Al 2

O3)

pro

per

ties

:q

s=

3600

kg/

m3,

ks

=36

.0W

/mK

,c p

,s=

0.76

5k

J/k

gK

.

454 J. Lee, I. Mudawar / International Journal of Heat and Mass Transfer 50 (2007) 452–463

where knf, kbf and kp are the thermal conductivities of thenanofluid, base fluid and solid particles, respectively, u isthe volumetric concentration of nanoparticles, and n issolid particle shape factor (n = 3 for spherical particles).

Following Wen and Ding [6], the Einstein equation isused for dynamic viscosity.

lnf

lbf

¼ 1þ 2:5u; ð4Þ

where lnf and lbf are the viscosities of the nanofluid andbase fluid, respectively.

All other properties are linearized relative to nanoparti-cle volumetric concentration (Pak and Cho [7]). This yieldsrelationships for density and specific heat, respectively,

qnf ¼ ð1� uÞqbf þ uqp ð5Þ

and

cp;nf ¼ 1� uð Þcp;bf þ ucp;p: ð6Þ

The system shown in Fig. 1 yields the following varia-tions of bulk mean fluid temperature and inside walltemperature, respectively, with distance x from the tubeentrance.

T m;x ¼ T in þq00pDx_mcp;nf

ð7Þ

and

T w;x ¼ T m;x þq00

h: ð8Þ

Two kinds base fluids are examined in this example,water and HFE-7100. The latter is a 3M dielectric fluidused in electronics cooling and general heat transfer appli-cations. The thermal conductivity for this fluid is signifi-cantly smaller than for water. Table 1 shows propertiesof these base fluids as well as for different concentrationsof Al2O3.

Fig. 2 shows heat transfer results for the laminar case.For each base fluid, increasing nanoparticle concentrationis shown in Fig. 2(a) monotonically increasing the heattransfer coefficient. This is the result of the heat transfercoefficient being proportional to thermal conductivity forlaminar flow. Notice that the heat transfer coefficient val-ues with HFE-7100 are much smaller than for water. Over-all, the enhancement ratio is almost the same for water andHFE 7100 nanofluids.

However, as shown in Fig. 2(b), increasing nanoparticleconcentration increases both fluid and wall temperaturesfor the nanofluids relative to their base fluid counterparts.This temperature increase is the result of the decreased spe-cific heat of nanofluids caused by the solid particles havinglower specific heat than fluids. Notice for HFE-7100 thelarge axial rise in both fluid and wall temperature. Thistrend is the result of the small specific heat for this fluid.

Fig. 3 shows the results for the turbulent case. Accord-ing to Eq. (2), h � k0.6 for turbulent flow compared toh � k for laminar flow. This implies the enhancement effect

Fig. 2. Axial variations of (a) heat transfer coefficient and (b) bulk andwall temperatures for laminar flow and different nanoparticle concentra-tions in water and HFE7100 base fluids.

Fig. 3. Axial variations of (a) heat transfer coefficient and (b) bulk andwall temperatures for turbulent flow and different nanoparticle concen-trations in water and HFE7100 base fluids.


due to the increased thermal conductivity of nanofluids issignificantly weaker for turbulent flow than for laminar.The enhancement in turbulent flow is also dependent onflow rate in addition to viscosity and specific heat. Sinceh � k0:6l�0:4c0:4

p , and because increased nanoparticle con-centration enhances viscosity and degrades specific heat,the enhancement effect of nanoparticles in turbulent flowis further reduced compared to thermal conductivity alone.This effect is quite evident in Fig. 3(a).

This simple analysis shows how the enhancement effectof nanofluids is more effectively realized in laminar flow.Interestingly, most micro-channel flows of practical interestare laminar because of both their small hydraulic diameterand small flow rates. Implementing nanofluids in micro-channel devices therefore appears to be an advantageousstrategy in pursuit of enhanced cooling performance.

3. Experimental methods

3.1. Nanofluid preparation

The solid particles used in this study were Al2O3. Thenanofluid was produced by gas condensation by Nano-

phase Technologies. Insoluble solid Al2O3 particles withan area-averaged size of about 36 nm came dispersed inhigh concentration in water, which formed the base fluid.This high concentration fluid was diluted with de-ionizedwater to the desired specified volume concentrationdemanded for individual experiments.

Two volumetric concentrations of Al2O3 nanofluid, 1%and 2%, were tested in the present study. The particle sizeand shape were identical to those used by Lee et al. [3], whomeasured about 6% enhancement in thermal conductivitywith 2% Al2O3.

3.2. Test loop and operating procedure

The enhancement effects of the Al2O3 nanofluids wereexamined with the aid of a test facility that was used in ear-lier studies of single-phase and two-phase micro-channelheat sink performance by Qu and Mudawar [8–10]. Asshown in Fig. 4, water was pumped from a reservoirthrough a filter followed by one of two parallel flowmeters,and a constant-temperature bath before entering the micro-channel test module. Exiting the test module, the fluid waspassed through a water-cooled condenser before returning

Fig. 4. Schematic of flow loop.


to the reservoir. Since the flow meters were not cali-brated for the 1% and 2% Al2O3 concentrations, abypass was included immediately downstream of the flow-meters to calibrate the flowmeters for these concentra-

Fig. 5. (a) Construction of micro-channel test module and (b)

tions. The calibration was achieved by dividing themass of nanofluid accumulated in a beaker by the fill time.A precision digital bench scale was used for the massmeasurements.

thermocouple locations in copper block inside test module.

Fig. 6. Variations of (a) pressure drop across test module and (b) frictionfactor across micro-channels with Reynolds number for pure water andfor 1% and 2% Al2O3.


Fig. 5 shows a schematic of the micro-channel test mod-ule. Rectangular 215 lm wide by 821 lm deep grooveswere milled into the top surface of an oxygen-free copperblock. This block was inserted into a G-7 plastic housingand sealed atop with a polycarbonate plastic cover plate.This formed 21 parallel micro-channels with a hydraulicdiameter of 341 lm and a total planform area 1 cm wideby 4.48 cm long. Heating was provided by 12 cartridgeheaters that were embedded in the underside of the copperblock. The G-7 housing featured inlet and exit plenumswhere the fluid temperature and pressure were measuredwith the aid of type-K thermocouples and pressure trans-ducers, respectively.

Fig. 5(b) shows the detailed construction of the copperblock. Notice the three deep cuts in the copper block virtu-ally dividing the block into four cells, which helped confinethe power dissipated by each set of cartridge heaters to themicro-channel flow region immediately above, thus mini-mizing axial heat flow. Four type-K thermocouples wereinserted along the flow direction to determine the channelbottom wall temperatures. A Yokogawa WT210 powermeter measured electric power input to the copper block.All the facility’s sensor signals were input to an HP3852A data acquisition system.

Cooling performances of the nanofluids were bench-marked against that for pure water. The pure water experi-ments were performed first. The test facility was drainedcompletely and flushed very carefully with de-ionized waterbefore it was filled with a new fluid. Operating conditions forthe study were as follows: ReDh = 140–941, Q = 100–300 W,Tin = 30 �C, Pin = 1.17–1.36 bar, and Pout = 1.12 bar.

Measurement uncertainties were ±0.3 �C for tempera-ture, 0.5% for pressure, 0.1% for electrical power, and 2%for flow rate. Heat loss from the copper block was esti-mated at less than 5% of the electrical power input.

4. Single-phase results

4.1. Pressure drop

Suspending solid particles in a fluid generally increasesdynamic viscosity relative to the base fluid. Even withnanoparticles, this phenomenon causes a pressure dropincrease in micro-channels compared to an equal mass flowrate of the base fluid.

Since the present pressure measurements were made inthe inlet and outlet plenums of the test module, theyincluded inlet contraction pressure loss and outlet expan-sion pressure recovery in addition to the pressure dropalong the micro-channels as detailed by Qu and Mudawarfor pure water [8].

Fig. 6(a) shows the variation of the measured pressuredrop between the inlet and outlet plenums with Reynoldsnumber for adiabatic flow. A strong particle concentrationdependency is shown with the total pressure drop increas-ing with increasing nanoparticle concentration. Instead ofusing the viscosity relation given earlier in Eq. (4), ReDh

in Fig. 6(b) was based an empirical relation developedspecifically for water-based Al2O3 nanofluids [11],

lnf

lbf

¼ 123u2 þ 7:3uþ 1: ð9Þ

After subtracting the effects of inlet contraction and out-let expansion from the total measured pressure drop (seeRef. [8] for detailed calculation scheme), a single-phasemicro-channel friction factor was determined from thepressure drop DPch calculated for the micro-channelsalone.

fsp ¼DP chDh

2Lqf u2: ð10Þ

Fig. 6(b) shows particle concentration does not havea clear influence on the friction factor dependence onReynolds number. Error propagation from measurementuncertainties is estimated at 2.9% in the friction factorcalculation.

4.2. Heat transfer coefficient

The single-phase heat transfer coefficient was determinedfrom the measured fluid and heater block temperatures and


the electrical power input using an analytical fin techniquedetailed by Qu and Mudawar [9].

Fig. 7 shows a unit control volume consisting of a singlemicro-channel and half of the surrounding copper walls.Equating the heat influx to the unit cell through the bottomsolid wall to the heat efflux yields

q00ðW ch þ 2W wÞ ¼ hspðT w � T mÞðW ch þ 2gH chÞ; ð11Þwhere g is the fin efficiency. Since the top wall is adiabatic,the fin efficiency is given by

g ¼ tanh mH ch

mH ch

; ð12Þ

where m is the fin parameter defined as

m ¼

ffiffiffiffiffiffiffiffiffiffiffihsp

ksW w

s: ð13Þ

These three equations were used to determine the single-phase heat transfer coefficient, hsp. The corresponding errorpropagation from measurement uncertainties is estimatedat 5.0%.

Fig. 7. Micro-channel unit cell.

The fin base temperature, Tw, was calculated using theassumption of one-dimensional heat diffusion between theplane of the thermocouple embedded in the copper blockand the plane containing the fin base.

T w ¼ T tc �q00H tc

ks

: ð14Þ

Fig. 8 shows the variation of the single-phase heat trans-fer coefficient along the flow direction for water and thetwo nanofluids for different power inputs. The four mea-surement locations correspond to the thermocouple layoutshown earlier in Fig. 5(b). The two upstream locations arewell within the entrance region, while the third and thefourth indicate the flow is becoming thermally fully devel-oped. The nanofluids exhibit the same spatial trend as thatfor pure water. Fig. 8(a) and (b) show similar trends for 1%and 2% Al2O3, respectively, as compared to pure water. Ineach of these figures, increasing the heat flux has a veryweak effect on the heat transfer coefficient for pure water,but an appreciable effect for the nanofluid. Notice alsothe stronger effect of heat flux for 2% Al2O3, Fig. 8(b),compared to 1% Al2O3, Fig. 8(a).

Keblinski et al. [12] suggested four possible mechanismsfor heat transfer enhancement with nanofluids: Brownian

Fig. 8. Variation of heat transfer coefficient along micro-channel for (a)pure water and 1% Al2O3, and (b) pure water and 2% Al2O3.


motion, liquid layering at liquid/particle interface, phononmovement in nanoparticles, and nanoparticle clustering.While it is difficult to determine which of these four mech-anisms has the greatest impact on enhancement, theincreased dependence of the heat transfer enhancementon heat flux at higher concentrations may be one trendfuture research may rely upon in pursuit of mechanismidentification. In any case, this increased dependenceappears to alter boundary layer development fornanofluids.

Fig. 9 compares the heat transfer coefficient variationwith Reynolds number for water, 1% Al2O3 and 2%Al2O3 for each of the four axial locations along themicro-channel heat sink. For each axial location, 2%Al2O3 shows the greatest heat transfer enhancement. For1% Al2O3, the enhancement is obvious for the first twoupstream locations, but virtually nonexistent for the twodownstream locations where the thermal boundary layeris almost fully developed. In fact, the enhancement effectfor both concentrations appears far more prevalent in theentrance region than the downstream fully developedregion. It can therefore be concluded that nanoparticleshave an appreciable effect on thermal boundary layerdevelopment.

Interestingly, this trend is a general agreement with thefindings of Wen and Ding [6] for Al2O3 nanofluid flowinside a much larger tube (D = 4.5 mm).

Fig. 9. Variation of heat transfer coefficient with Reynolds number for differen(a) tc1, (b) tc2, (c) tc3, and (d) tc4.

4.3. Axial variations of fluid and wall temperatures

As discussed earlier, the thermal performance of a heatsink is dictated not only by the magnitude of the heat trans-fer coefficient, but by the wall temperature of the heat sinkas well. Increasing nanoparticle concentration increases thefluid’s thermal conductivity, thus increasing the heat trans-fer coefficient, especially for laminar flow. On the otherhand, increased nanoparticle concentration also degradesspecific heat, resulting in larger axial rises in both fluidand wall temperatures compared to the base fluid. There-fore, it is not immediately evident if the wall temperatureactually decreases (a sign of better cooling performance)with increasing nanoparticle concentration as a result ofthe higher heat transfer coefficient, given this degradationin the nanofluid’s specific heat.

Fig. 10 shows the measured variation of mean fluid tem-perature between the inlet and outlet of the micro-channelheat sink with mass flow rate. Shown is an increase in tem-perature difference with increased nanoparticle concentra-tion due to the specific heat degradation. This effect ismore pronounced for lower mass flow rates and higher heatfluxes.

Fig. 11 shows the net effect of increasing thermalconductivity and decreasing specific heat on the wall tem-perature, which is perhaps the most realistic measure ofthe effectiveness of a given cooling scheme. These wall

t heat inputs and different Al2O3 concentrations at measurement locations

Fig. 10. Variation of fluid temperature rise along micro-channel for (a)pure water flow and 1% Al2O3, and (b) pure water and 2% Al2O3.

Fig. 11. Variation of wall temperature along micro-channel for differentflow rates and Al2O3 concentrations and heat inputs of (a) 100 W and (b)300 W.


temperatures were calculated using Eq. (14) with measuredthermocouple data. The only noticeable decrease in walltemperature is realized near the inlet for the lower massflow rate at the lower heat input. All other conditionsand axial locations show virtually no measurable enhance-ment. Furthermore, any differences in the wall temperatureare within the range of measurement uncertainty of thestudy, let alone the difficulty of achieving an exact flow ratewhen comparing different cases. These findings provenanoparticles are not an effective means for enhancing sin-gle-phase cooling performance of a micro-channel heatsink.

4.4. Fluid stability

Particle settlement is an obvious concern with solid par-ticle suspensions that are intended for cooling applications.As mentioned earlier, nanoparticle suspensions are farmore stable than suspensions of larger particles.

This issue was examined in the present study by repeat-ing heat transfer coefficient measurements in tests that wereseparated by many hours. Fig. 12(a) and (b) shows theseresults for 2% Al2O3 and Reynolds numbers of 202 and

530, respectively. These results show fairly repeatableresults over the indicated time span.

Another method of assessing nanofluid stability was tovisually inspect fluid samples over an extended period oftime. The samples were sealed in glass beaker for the entireperiod. Fig. 13 compares picture of 1% and 2% Al2O3

nanofluids taken shortly after initial mixing and 30 dayslater. The latter pictures show some settlement and a con-centration gradient in both nanofluids. Such effects cancause long-term degradation in thermal performance dueto settling inside the cooling system’s reservoir. Further-more, most cooling systems contain feedback control sys-tems that attenuate coolant flow rate or stop it altogetherin response to the fluctuating heat load of the device beingcooled. Slow or zero flow rate accentuate nanoparticle set-tling and may even lead to clogging of individual compo-nents of the cooling loop. This issue points to the needfor long-term testing of a nanofluid cooling loop beforeactual deployment.

Another long-term concern with nanofluids is surfaceabrasion. Like most metal surfaces, copper exhibits someoxidation when exposed to pure water. This effect wasmore pronounced with increasing surface temperature.

Fig. 12. Heat transfer coefficient versus time for 2% Al2O3 and Reynoldsnumbers of (a) 202 and (b) 530.


The oxidize layer acts as an additional thermal resistanceand tends to degrade the heat transfer performance.Interestingly, the micro-channel surface appeared cleanerand shinier following each single-phase experiment withthe nanofluids. This phenomenon is believed to be theresult of erosion of the copper surface by the nanoparticles.While this may seem an advantage for nanofluids, special

Fig. 13. Nanofluid settlin

attention must be paid to the possibility of the removedoxide getting entrained with the flow, altering the enhance-ment effect of the nanoparticles, and contributing longterm to particle settling or even clogging of flow passages.

5. Flow boiling results

Flow boiling can greatly enhance the cooling perfor-mance of a micro-channel heat sink by increasing the heattransfer coefficient. Furthermore, since flow boiling reliesto a great degree on latent heat transfer, better temperatureaxial uniformity is realized both in the coolant and the wallcompared to a single-phase heat sink. The question hereis whether nanoparticles could further enhance an alreadysuperior performance.

Recently, several research articles examined the effectsof nanoparticles on pool boiling [13–15]. Degradationin nucleate boiling performance was related to surfacesmoothing, resulting from bubble-generated cavity fillingwith particles and particle deposition on the heating sur-face. However, nanofluids netted a measurable increase incritical heat flux (CHF).

In this research, flow boiling experiments were con-ducted with the micro-channel heat sink using pure waterand 1% Al2O3. Fig. 14(a) shows the data for pure water fol-low a typical boiling curve. This curve is initiated at lowheat fluxes with single-phase forced convection, followedby an appreciable slope change following the onset of boil-ing, and culminating in a sudden unsteady temperature riseat CHF. While single-phase forced convection proceeded inthe same manner with increasing heat flux for 1% Al2O3,flow boiling was never achieved with this nanofluid. Itappears fluid evaporation associated with the initial incip-ience of boiling near the channel exit caused rapid deposi-tion of nanoparticles in the same region. Despite the smallsize of these particles, the deposition rate was so severe thatrelatively large particle clusters could be seen near the exit.This clustering effect quickly propagated upstream andfilled the entire channel, preventing the coolant from enter-ing the heat sink. The heat sink responded with a sudden

g over 30-day period.

Fig. 14. (a) Flow boiling curve at measurement location tc4 for pure waterand 1% Al2O3 and (b) photo of particles after being removed from micro-channels.

Table 2Advantages and disadvantages of nanofluids

Advantages Disadvantages

Higher single-phase heat transfercoefficient, especially for laminarflow, due to increased thermalconductivity

Increased axial rise in walltemperature due to degradedspecific heat

Increase pumping power due togreater pressure dropLong-term fluid settling andpotential clogging of flow passagesPossible damage to flow loop partsby erosionInability to sustain flow boilingHigh cost of nanoparticlesuspensions


unsteady temperature rise. This phenomenon constitutes acatastrophic failure for a cooling system. Fig. 14(b) showsphotos of the deposited particles after they were removedby a fine needle.

6. Conclusions

This study is an assessment of the effectiveness of Al2O3

nanoparticles at enhancing single-phase and two-phaseheat transfer in micro-channel heat sinks. Key conclusionscan be summarized as follows:

(1) The high thermal conductivity of nanoparticles rela-tive to common pure fluids enhances the single-phaseheat transfer coefficient for fully-developed laminarflow. The enhancement is far weaker for turbulentflow because of a weaker dependence of the heattransfer coefficient on thermal conductivity as wellas decreased specific heat and increased viscosity withincreased nanoparticle concentration.

(2) Increasing nanoparticle concentration increases sin-gle-phase pressure drop compared to pure fluids atthe same Reynolds number.

(3) Higher single-phase heat transfer coefficients areachieved in the entrance region of micro-channelswith increased nanoparticle concentration. However,the enhancement is weaker in the fully developedregion, proving that nanoparticles have an apprecia-ble effect on thermal boundary layer development.Higher concentrations also produce greater sensitiv-ity to heat flux.

(4) Despite their ability to enhance to the single-phaseheat transfer coefficient due to the increased thermalconductivity, the overall cooling effectiveness ofnanopartcles is quite miniscule. Simultaneous degra-dation of the specific heat leads to larger fluid temper-ature rise along the micro-channel. The net effect ofincreasing thermal conductivity and decreasingspecific heat on wall temperature is virtually non-existent.

(5) Single-phase cooling performance with nanofluids isquite repeatable over extended durations. However,very long-term use may lead to particle settling andpotential clogging of flow passages. This issue mustme carefully examined by long term testing of theentire cooling system, including variations of, orcomplete stoppage of the coolant flow rate, beforedeployment. Nanoparticles also cause erosion tometal surfaces and, while this may help maintain sur-face texture, the removed oxide may ultimately alterthe enhancement effect of nanoparticles, accentuatesettling, or even clog flow passages.

(6) Nanoparticles should not be used in two-phasemicro-channel heat sinks. Once boiling commences,particles begin to deposit into relatively large clustersnear the channel exit due to localized evaporation.This clustering phenomenon quickly propagatesupstream to fill the entire channel, preventing coolantfrom entering the heat sink and causing catastrophicfailure of the cooling system.

Table 2 summarizes the practical implications from thepresent study. While this study does demonstrate the effec-tiveness of nanoparticles at enhancing the single-phase heattransfer coefficient by increasing the nanofluids thermalconductivity, it also shows only a miniscule enhancementin cooling effectiveness compared to the pure fluid, andreveals several important disadvantages. These disadvan-tages bring into question the overall merit of using nano-fluids in micro-channel heat sinks. A far more effectivestrategy may be to simply optimize micro-channel geo-


metry in pursuit of a higher heat transfer coefficient andlower pressure drop using pure fluids [16,17].

Acknowledgements

The authors greatly appreciate the support of the Officeof Naval Research (ONR) for this study.

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Documents

Assessment of the eﬀectiveness of nanoﬂuids for single