houshmand and peles 2014 impact of flow dynamics on heat transfer of bubbly flow in a...

Farzad Houshmande-mail: [email protected]

Yoav Peles1

e-mail: [email protected]

Department of Mechanical,

Aerospace, and Nuclear Engineering,

Rensselaer Polytechnic Institute,

110 8th Street,

Troy, NY 12180

Impact of Flow Dynamicson the Heat Transfer of BubblyFlow in a MicrochannelDuring nucleate flow boiling, the bubble dynamics affect the liquid flow field and the cor-responding heat transfer process through several distinct mechanisms. At the microscale,this effect is different than at the macro scale partly because the bubble dimensions arecomparable to the characteristic length scale of the channel. Since the process involvesseveral mechanisms, an attempt to isolate and study them independently from one anotheris desired in order to extend knowledge. To remove the evaporation effect from the heattransfer process, noncondensable gas bubbles were introduced upstream of a1 mm� 1 mm heater into a 220 lm deep and a 1.5 mm wide microchannel and the heattransfer coefficient was measured and compared to single-phase liquid flow. High speedimaging and micro particle image velocimetry (l-PIV) measurements were used to eluci-date the bubble dynamics and the liquid velocity field. This, in turn, revealed mechanismscontrolling the heat transfer process. Acceleration and deceleration of the liquid flow dueto the presence of bubbles were found to be the main parameters controlling the heattransfer process. [DOI: 10.1115/1.4025435]

Keywords: bubbles, microchannel, convective heat transfer, two-phase flow, micro-PIV

1 Introduction

In flow boiling, the critical heat flux (CHF) condition declineswith increasing mass quality [1]. Thus, to dissipate very high heatflux while avoiding CHF, low mass quality is required. Since lowmass quality corresponds to nucleate flow boiling, research perti-nent to bubbly flow in micro domains is important to enable highheat flux applications, such as cooling high performance micro-processors, laser diodes, and high power radars.

Flow nucleate boiling heat transfer is complex and involvesseveral distinct mechanisms including quenching, evaporation,and single-phase liquid flow [2,3]. In conventional scale studies,these three processes are often considered as the primary (if notthe only) mechanisms involving flow boiling heat transfer. How-ever, other mechanisms—such as enhanced mixing away from thebubble due to rapid bubble growth/collapse, liquid flow modifica-tion downstream of an attached or detached bubble, and heatingof a thin liquid layer sandwiched between a large bubble and theheat transfer surface—might also be significant. Heat transfermechanisms have been discussed before in connection to flow nu-cleate boiling at the microscale [4–9]. Due to the large number ofpossible mechanisms involved in microchannels (and perhaps inconventional channels) and the interaction between them, often itis challenging to identify the dominant mechanism/s.

Ideally, to the extent possible, this can be resolved by studyingeach mechanism independently. Eliminating the evaporation pro-cess by introducing gas bubbles to the flow at high subcoolingconditions, several mechanisms can be turned off. Betz andAttinger [10] studied gas segments in liquid flow in microchannelsand demonstrated that heat transfer coefficients can be muchenhanced. Poulikakos et al. [11] studied convective heat transferin a segmented immiscible liquid–liquid flow in a microchanneland similar to Betz and Attinger, observed significant enhance-ment in the heat transfer coefficients. In several other studies, air

bubbles [12] and vapor bubbles [13] were used in an attempt tosuppress flow boiling instabilities and enhance heat transfer.

In this study, immiscible gas (air) bubbles were introducedupstream of a 1 mm� 1 mm heater in a 220 lm deep microchan-nel and the effect of bubbles on the average heat transfer coeffi-cient was investigated. Micro particle image velocimetry (l-PIV)measurements were performed to elucidate the effect of bubbleson the carrier liquid flow and to infer the mechanisms affectingthe heat transfer process.

2 Experimental Setup and Method

2.1 Micro Device. A 220 lm deep, 1.5 mm wide, and15.5 mm long microchannel was fabricated by bonding two proc-essed Pyrex substrates between an epoxy-covered vinyl layer (Fig.1). Water entered from an inlet manifold and flowed about 7 mmbefore passing over the 1 mm� 1 mm heater. Air bubbles wereintroduced into the channel through two orifices upstream of theheater: Orifice I (D¼ 350 lm) and Orifice II (D¼ 250 lm) located0.5 mm and 4 mm upstream of the heater, respectively.

A 30 nm thick titanium layer along with a 1 lm thick aluminumlayer were deposited on the bottom Pyrex substrate in consecutivesputtering processes. The heater (1 mm� 0.94 mm due to overetching) and the aluminum vias were formed after removing theextra material through chemical wet etch processes. A 600 nmthick SiO2 film was then deposited on the heater and vias for elec-trical insulation. Inlets and outlet ports were formed by drillingholes through the wafer. The microchannel was formed by cuttingthe pattern of the channel in the vinyl layer and attaching the topand bottom Pyrex substrates on the opposite side of the vinyllayer. Before attachment, a couple of holes were drilled on the topPyrex substrate as well as the vinyl layer to provide access to theelectrical pads. Finally, individual devices on the attached sub-strates were separated from each other using a die-saw cuttingmachine.

The micro device was seated in a package—built from Delrinusing a computer numerical control (CNC) machine—which con-nected the fluidic ports in the microchannel to external fittings andto the measurement apparatus. The micro device was held in place

1Corresponding author.Contributed by the Heat Transfer Division of ASME for publication in the

JOURNAL OF HEAT TRANSFER. Manuscript received October 24, 2012; final manuscriptreceived August 13, 2013; published online November 7, 2013. Assoc. Editor: AliEbadian.

Journal of Heat Transfer FEBRUARY 2014, Vol. 136 / 022902-1Copyright VC 2014 by ASME

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by a cover plate bolted to the package, and sealed through a set ofminiature O-rings (Fig. 2).

2.2 Experimental Setup. Figure 3 schematically depicts theopen-loop experimental setup. Distilled water was pressurizedinside the tank and propelled to the micro device through a controlvalve. An Omega FL-3600 series rotameter was used to measurethe flow rate, and a filter was mounted in the flow line to preventdebris from entering the micro device. To introduce bubbles, acontrollable syringe pump delivered the air stream into thechannel.

A direct current (DC) power supply was connected to the heaterthrough gold-plated spring-loaded connectors. Voltage across theheater and the current was measured by two multimeters to calcu-late the power and the electrical resistance of the heater. The heat-er’s electrical resistance was used to infer the average walltemperature. All the measurements were recorded in a personalcomputer (PC) and processed in National Instrument (NI)LabVIEW.

A Zeiss inverted microscope (Observer Z1m) was used to visu-alize the flow. The package and external connections weremounted on a support structure above the microscope and illumi-nated by a halogen lamp and a double-pulse 120 mJ Nd:YaG laser

(Solo 120XT from New Wave). A charge coupled device (CCD)camera (Imager Pro X2M from Lavision) as well as a high speedcomplementary metal-oxide semiconductor (CMOS) camera wereused to capture the images. A timing board mounted on the PCsynchronized the CCD camera and the laser pulses.

2.3 Experimental Procedure. Following the fabrication pro-cess, the heater’s electrical resistance was measured at differenttemperatures in a controllable oven and a calibration curve wasgenerated. With the calibration curve, the heater was also used tomeasure the average wall temperature during experiments.

After setting the water flow rate—controlled by the tank’s pres-sure and a valve—and the air flow rate, the voltage and currentacross the heater were recorded, and the heater power and averagetemperature—based on the calibration curve—were calculated.To prevent gaseous cavitation in the microchannel, the distilledwater was initially degassed for several hours—using a vacuumpump—and then pressurized by helium.

To estimate heat loss through conduction, the device was vac-uumed and the corresponding steady state heater temperatureswere recorded for a range of powers. Heat loss was calculatedbased on the heater temperature and then subtracted from the totalpower to obtain the effective heat dissipated to the flow.

To capture the sequence of bubble growth and detachment,high speed camera recorded the images through a 5� magnifica-tion lens, while the test section was illuminated by the halogenlamp. The recorded sequence of images was used to track themotion of individual bubbles, and using the time interval betweenthe frames, parameters such as bubble velocity and frequencywere obtained. For frequency calculation, the time interval for 20to 30 bubbles to pass a certain point was measured and averagefrequency was calculated. For micro particle image velocimetry(l-PIV) measurements, the main flow was seeded with 0.71 lmfluorescent particles (peak emission wavelength of 612 nm) anddouble pulses of Nd:Yag laser with wavelength of 532 nm illumi-nated the test section. Double frame images were recorded by theCCD camera through a 10� magnification lens and a 570 nmhigh pass filter to eliminate background light. Finally, doubleframe images were processed in Davis

VR

software through cross-correlation algorithms—after masking out the bubbles—to extractthe velocity field. Initially, ensemble average of a set of—usually500—images (used as background image) was obtained and sub-tracted from individual images to reduce the background noise.For two-phase images, the gas phase sections of the image weremasked out to provide accurate results in the surrounding area.Finally, the images were cross-correlated in 64� 64 and 32� 32

Fig. 1 Micro device’s schematics

Fig. 2 Micro device’s package

022902-2 / Vol. 136, FEBRUARY 2014 Transactions of the ASME


interrogation windows with 50% overlap to extract the velocityfield.

2.4 Data Reduction. For single-phase experiments, Reynoldsnumber was defined as

Re ¼ q �VDh

l(1)

where q (kg�m�3) is the density, l (kg�m�1�s�1) is the dynamicviscosity, �V (m�s�1) is the average velocity, and Dh (m) is the hy-draulic diameter of the channel. Inlet temperature was used forcalculating the thermophysical properties of the liquid flow.

Superficial velocities were used to quantify the flow rates in thetwo-phase regime according to

ji ¼Ci

Ac(2)

where i denotes the fluid (i.e., water or air), Ci (m3�s�1) is the vol-umetric flow rate of fluid i, and Ac (m2) is the cross-sectional areaof the channel.

Based on the volumetric flow rate of the gas and the bubble fre-quency, the volume of an individual bubble was estimated accord-ing to

B ¼ Cg

f(3)

where B is the bubble volume and f is the bubble frequency. Sinceoften the bubble covered the entire height of the channel, planardiameter of the bubbles was estimated as

d ¼ffiffiffiffiffiffiffiffiffiffi4B

p � H

r(4)

where d is the planar diameter of the bubbles and H is the heightof the channel.

To measure the heat transfer coefficient, heat transfer rate andthe corresponding wall temperature were required. The net heattransfer rate to the liquid was obtained by subtracting the heat lossfrom the total applied power as

_Qnet ¼ _Qtot � _Qloss (5)

To compensate for the heat conduction through the thin SiO2

layer, a one-dimensional conduction analysis was performed toinfer the surface temperature based on the heater temperatureaccording to

�Tw ¼ �Theater þ _Qnet

tSiO2

A � kSiO2

(6)

where tSiO2(m) is the SiO2 layer thickness, A (m2) is the heater

area, and kSiO2(W�m�1�K�1) is the thermal conductivity of the

SiO2 layer. This correction for the wall temperature, however,was very small (<1 �C) in the present study. The average heattransfer coefficient was then calculated based on the convectionheat transfer equation

_Qnet ¼ �hAð �Tw � T0Þ (7)

and the corresponding average Nusselt number was

Nu ¼�hDh

kl(8)

where T0 is the inlet temperature and kl is the thermal conductivityof the fluid. Because of the developing thermal boundary layer,the inlet temperature was used in Eq. (7).

The effect of bubbles on the heat transfer coefficient was quan-tified by comparing the heat transfer coefficient at the presence ofthe bubbles to the single-phase liquid flow at the same liquid flowrate

E ¼�hb � �hsp

�hsp(9)

�hb and �hsp denote bubbly flow and single-phase heat transfer coef-ficients, respectively.

2.5 Heat Loss and Uncertainties. As discussed before, heatlosses were measured based on the heater temperature and sub-tracted from total power. The actual heat loss in the presence of

Fig. 3 Schematics of experimental setup

Journal of Heat Transfer FEBRUARY 2014, Vol. 136 / 022902-3


flow might be slightly different from the measured values as dis-cussed in Ref. [14]. However, since the heat loss was between 5%and 7% for the maximum and the minimum flow rates, this dis-crepancy was assumed to be negligible.

Standard uncertainty analysis methods were used to estimatethe uncertainties of the reduced data [15]. For flow rates—andconsequently superficial velocity and Reynolds number—uncer-tainties were associated with the accuracy of the rotameter mea-surement, which ranged between 63% and 66% for the two-phase experiments. The uncertainty in temperature measurementswas less than 1 �C. Based on Eq. (7), uncertainties in heat transfercoefficient and Nusselt number were estimated to vary between63% and 65% depending on DTs. These values were the totalestimation of the biased and unbiased uncertainties. The standarddeviation of the temperature fluctuations around the time-averaged value is the unbiased part for the uncertainty, and there-fore the true measure of the uncertainty in the enhancement factor,E. This uncertainty was estimated to be 0.1 �C, corresponding to61% in the enhancement factor. Uncertainty in the velocity,through the l-PIV measurements, was estimated to be less than3%, and the uncertainty associated with the bubble frequency wasless than 2%.

3 Results and Discussion

3.1 Bubble Dynamics. Four different liquid flow rates: 13.2,18.8, 24.5, and 30.4 (ml/min), corresponding to Reynolds numbersranging from 310 to 720, were tested; the gas flow rates variedfrom 0.5 to 7 (ml/min). These flow rates provided a range of bub-ble sizes with respect to heater dimension. As discussed in previ-ous studies (e.g., Refs. [16, 17]), bubbles formed from an airstream injected into a liquid flow were affected by the liquid and

gas velocities. A sequence of high speed images is shown inFig. 4 illustrating the shape, size, and frequency of injected bub-bles over the heater area. As expected, the bubbles’ size increasedwith increased gas flow rates and decreased with increased liquidflow rates due to higher drag forces. As a result, the bubbles weresqueezed and stretched. The sequence of bubble growth anddetachment from the orifice (for one case) is presented in Fig. 5.

Frequencies of bubbles were measured for different liquid andgas flow rates and the results are depicted in Figs. 6 and 7 for Ori-fices I and II, respectively. The frequency of the bubbles injectedfrom the smaller orifice (Orifice II) was higher compared to thatof larger orifice (Orifice I) by an average of 37% (for high gasflow rates). This correlates well with the reciprocal ratio of theorifice diameters (350 lm/250 lm¼ 1.4). Considering the sameliquid and gas flow rates, smaller surface tension force (smallerperimeter) withstood smaller drag force exerted by the liquid,

Fig. 4 Bubbles before detachment at different flow rates

Fig. 5 Sequence of high speed images from bubble growth and detachment at 6300 fps;jl 5 0.95, jg 5 0.25 (m/s)

Fig. 6 Bubble frequency (Orifice I)



which led to smaller bubbles at higher frequencies. This force bal-ance also explains the higher bubble frequencies at higher liquidflow rates. At each liquid flow rate, the bubble frequency quicklyreached an asymptotic value with increased gas flow rates. Forvery low gas flow rates, the bubble frequency and the bubble sizefluctuated, and therefore, some data points are not presented in thefigures. This effect can be attributed to slow pressure built-up inthe flow line—due to compressibility of the gas—required toovercome the initial stage of bubble formation. For heat transfermeasurements in these cases, the average values were used.

3.2 Heat Transfer and Flow Modification. The effect of gasbubbles on the heat transfer coefficient was investigated for differ-ent gas and liquid flow rates. Heat transfer coefficients for two-phase bubbly flow were compared to the single-phase results withthe same liquid flow rate. Since the convective heat transfer pro-cess is affected by the fluid flow, l-PIV measurements were per-formed to obtain the velocity field around the bubbles.

Fig. 7 Bubble frequency (Orifice II)

Fig. 8 Average Nusselt number for single-phase water flow

Fig. 9 Temperature change during the bubble injection (OrificeI); jl 5 1.24, jg 5 0.30 (m/s)

Fig. 11 Velocity field around the bubbles injected fromOrifice IFig. 10 Effect of bubbles on the heat transfer (orifice I)



Subsequently, it was used to elucidate the mechanisms controllingthe heat transfer process.

Single-phase heat transfer coefficient measurements were ini-tially obtained for Reynolds numbers ranging from 25 to 2000(Fig. 8), corresponding to 0.05< jl< 3.3 (m/s). To validate themeasurements, the results were compared to a simplified analyti-cal solution for average Nusselt number for laminar flow betweenparallel plates with a fully developed velocity boundary layer anda developing thermal boundary layer and constant heat fluxboundary condition [18]. Since the channel had a high aspect ratioand the heater was located relatively far from the side walls, thisflow condition did not deviate much from the analytical solution’sassumptions. The results compare well with the analyticalsolution.

After conducting single-phase experiments, air bubbles wereintroduced into the channel and heat transfer coefficients weremeasured for different flow rates. Both liquid and gas streamswere maintained at the room temperature (�23 �C) to isolate theeffect of temperature difference of the secondary flow on the heattransfer coefficient.

Figure 9 illustrates the average wall temperature changes duringthe bubble injection period. After triggering the syringe pump, ittakes a short period for bubbles to reach a steady state condition atwhich the stable pressure is built in the flow line. Smaller bubbleswere observed before the steady state condition, which indicateslower gas flow rate condition. Similar behavior—but moregradual—was observed when the syringe pump was switched off,namely, the bubbles gradually diminished until the single-phasecondition restored. It is worth nothing that Fig. 9 was plotted for aspecific flow combination at orifice I, however, because the injec-tion mechanism was the same for all bubbles, a similar behaviorwas observed for Orifice II.

The heat transfer coefficients of bubbly flow were compared tosingle-phase heat transfer coefficients for the same liquid flowrate, and the enhancement factor, E, was calculated. Figure 10shows the results for bubbles introduced close to the heater fromOrifice I. At low liquid flow rates, the heat transfer coefficient ini-tially increased and then began to decline. The downward trendled to heat transfer coefficients even lower than those for single-phase liquid flow. At higher liquid flow rates, a similar trend wasobserved, although the heat transfer coefficient was higher thanfor liquid single-phase case for all the gas flow rates. A peak atrelatively low gas flow rate was observed indicating an optimumgas flow rate for heat transfer enhancement. At gas flow ratesabove the optimum condition, lower temperatures were observed

before and after the steady state condition indicating a higher heattransfer coefficient at a lower gas flow rates.

l-PIV velocity measurements around the injected bubbles wereobtained to reveal the mechanism of the observed trends. Theresults for six different cases taken at midplane of the channel arepresented in Fig. 11. Flow fields at different stages of the bubbledevelopment were also captured and are presented in Fig. 12. (Itshould be noted that in Fig. 12, because of the low frequency ofthe CCD camera—maximum 15 Hz—the images were not takenfrom the same bubble at different stages; instead images from var-ious bubbles at different stages were obtained and the sequencewas reconstructed.) The results identified several mechanismscontrolling the enhanced heat transfer process during bubble for-mation: (a) acceleration of the carrier liquid flow corresponding toreduced effective cross-sectional area because of the presence ofbubbles and (b) acceleration around the bubbles. On the otherhand, low velocity regions downstream of the bubbles hinder heattransfer. The mixing in the shear layer between the high and lowvelocity regions can enhance the heat transfer. Oscillations of thebubbles displace the fluid around them and create high velocityregions especially when the bubbles detach and the tail recoilsinto the bubble (Fig. 12(b)). Another factor that affects the heattransfer, especially at high gas flow rates, is related to the regiontrapped beneath the bubbles. In these regions, a thin liquid filmforms between the wall and the bubble interface. The heat transfer

Fig. 12 Velocity field around developing bubbles injected from Orifice I; jl 5 0.95, jg 5 0.25 (m/s)

Fig. 13 effect of bubbles on heat transfer (Orifice II)



process is mainly dependent on the thickness and velocity of theliquid film. When the bubbles are attached to the wall, a stagna-tion region is formed in the film beneath the bubbles, and the liq-uid velocity is smaller compared to the single-phase case. Thisthin liquid layer can heat up, and if not quickly removed, can ele-vate the surface temperature and impede the heat transfer process.

It should be noted that the total effect of the bubbles on the heattransfer was reflected through the area average heat transfer coef-ficient over the entire heated area.

At low liquid flow rates, the effect of bubbles on flow accelera-tion was comparably small and the bubbles were bigger. Conse-quently, a thin liquid layer over a large segment of the heaterformed and as the liquid layer stagnated, lower heat transfer coef-ficients were observed. At higher liquid flow rates, the flow accel-eration became more appreciable and the bubbles were smaller.As a result, a larger heater area was influenced by the favorableeffects of the bubbles on the heat transfer process, while a smallerarea was exposed to their adverse effects. In addition, higher bub-ble frequency helped replenish the liquid film near the heatermore frequently, thus, preventing it from heating up.

To examine the effect of bubbles introduced farther upstream—and presumably past the acceleration stage—bubbles with thesame gas flow rates were injected from Orifice II, and their effecton the heat transfer coefficient was investigated (Fig. 13). At lowliquid flow rates, after an initial increase, the heat transferenhancement decreased as the gas flow rate increased, but aninflection point was observed. At high liquid flow rates, theenhancement factors increased with increased gas flow rates. Ob-servation of the velocity field (Fig. 14) revealed a low velocityregion in the center of the channel—in the vicinity of the bub-bles—and a high velocity region on the sides. The average heattransfer coefficient may increase or decrease depending on themagnitude and the location of the high velocity region withrespect to the heater. Furthermore, as reported in several studies,e.g., Refs. [19, 20], the liquid film thickness varies with capillarynumber. Assuming constant properties, it can be inferred that thefilm thickness increases with bubble velocity. This concurs withthe higher heat transfer enhancements at higher liquid flow rates.

4 Conclusion

Effect of air bubbles on heat transfer characteristics in microdomains was studied experimentally. Heat transfer coefficient fordifferent water and air flow rates were measured and compared tosingle-phase results. Enhancements up to 16% in the heat transfercoefficient were observed. Moreover, bubble dynamics and veloc-ity field around the bubbles were studied to elucidate the heattransfer mechanisms. Results for the injected bubbles from twodifferent orifices upstream of the heater suggested two simultane-ous and competing processes controlling the effect of bubbles onthe heat transfer. Depending on the dominance of each mecha-nism, the average heat transfer coefficient can be enhanced or hin-dered. The velocity field modification caused by the bubblesplayed a key role in the heat transfer process. Although the impactof adiabatically formed bubbles on the flow field and single-phaseheat transfer was discussed, and it helps elucidating one of themany heat transfer mechanisms in the process, further investiga-tion on subcooled boiling with similar approach can furtherdivulge the processes controlling heat transfer mechanisms. Mix-ing of the liquid flow caused by the bubbles was also believed tobe important in the heat transfer process, but the extent is depend-ent on the thickness of the thermal boundary layer.

Acknowledgment

This work was supported by the Office of Naval Research (Pro-gram Manager Dr. Mark Spector). The authors would like toacknowledge the staff of the Micro and Nano Fabrication CleanRoom (MNCR) at Rensselaer Polytechnic Institute for their assis-tance in fabrication of the micro devices.

Nomenclature

A ¼ heater area (m2)Ac ¼ cross-sectional area (m2)cp ¼ specific heat (J�kg�1�K�1)

Dh ¼ hydraulic diameter (m)

Fig. 14 Velocity field around the bubbles injected fromOrifice II



E ¼ heat transfer enhancementf ¼ frequency (Hz)�h ¼ average heat transfer coefficient (W�m�1�K�1)

H ¼ channel height (m)ji ¼ superficial velocity of fluid i (m�s�1)kl ¼ liquid conductivity (W�m�1�K�1)

kSiO2¼ SiO2 conductivity (W�m�1�K�1)

Nu ¼ Nusselt numberP ¼ pressure (kPa)

_Qloss ¼ power loss (W)_Qnet ¼ net power (W)_Qtot ¼ total power (W)Re ¼ Reynolds number

t ¼ time (s)T0 ¼ inlet temperature (�C)

�Theater ¼ average heater temperature (�C)�Tw ¼ average wall temperature (�C)�V ¼ average velocity (m�s�1)

Vb ¼ bubble velocity (m�s�1)

Greek Symbols

a ¼ thermal diffusivity (m2�s�1)Ci ¼ volumetric flow rate of fluid i (m3�s�1)li ¼ viscosity of fluid i (kg�m�1�s�1)q ¼ density (kg�m�3)t ¼ kinematic viscosity (m2�s�1)r ¼ surface tension (N�m�1)

Subscripts

b ¼ bubbleg ¼ gasl ¼ liquid

sp ¼ single-phase

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houshmand and peles 2014 impact of flow dynamics on heat transfer of bubbly flow in a...

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