Garg, J., Arik, M., Weaver, S., Wetzel, T., Saddoughi S., Meso Scale Pulsating Jets for Electronic Cooling

8/3/2019 Garg, J., Arik, M., Weaver, S., Wetzel, T., Saddoughi S., Meso Scale Pulsating Jets for Electronic Cooling

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Jivtesh Garg

Mehmet Arik

GE Global Research Center,

Thermal Systems Laboratory,

Niskayuna, NY 12309

Stanton WeaverGE Global Research Center,

Micro and Nano Structures Technology Lab,

Niskayuna, NY 12309

Todd WetzelGE Global Research Center,

Thermal Systems Laboratory,

Niskayuna, NY 12309

Seyed SaddoughiGE Global Research Center,

Propulsion Technologies Laboratory,

Niskayuna, NY 12309

Meso Scale Pulsating Jetsfor Electronics Cooling

Microfluid devices are conventionally used for boundary layer control in many aerospaceapplications. Synthetic jets are intense small-scale turbulent jets formed from periodicentrainment and expulsion of the fluid in which they are embedded. The jets can be madeto impinge upon electronic components thereby providing forced convection impingement

cooling. The small size of these devices accompanied by the high exit air velocity pro-vides an exciting opportunity to significantly reduce the size of thermal management hardware in electronics. A proprietary meso scale synthetic jet designed at GE Global

Research is able to provide a maximum air velocity of 90 m/s from a 0.85 mm hydraulicdiameter rectangular orifice. An experimental study for determining the cooling perfor-mance of synthetic jets was carried out by using a single jet to cool a thin foil heater. Theheat transfer augmentation caused by the jets depends on several parameters, such as,driving frequency, driving voltage, jet axial distance, heater size, and heat flux. Duringthe experiments, the operating frequency for the jets was varied between 3.4 and 5.4 kHz,

while the driving voltage was varied between 50 and 90 V RMS. Two different heater powers, corresponding to approximately 50 and 80 °C, were tested. A square heater with

a surface area of 156 mm2 was used to mimic the hot component and detailed tempera-ture measurements were obtained with a microscopic infrared thermal imaging technique.

A maximum heat transfer enhancement of approximately 10 times over natural convec-tion was measured. The maximum measured coefficient of performance was approxi-

mately 3.25 due to the low power consumption of the synthetic jets.DOI: 10.1115/1.2065727

Introduction

Thermal management of tight space, low power electronics ap-

plications is one of the current challenging engineering problems.

More functionality, faster signal speed and tighter enclosures have

resulted in an increase in volumetric heat generations. Typically,

these devices are built in tightly sealed housings where space is at

a premium. To improve the heat transfer path, a possible replace-

ment for natural convection heat transfer is impingement heat

transfer using micro/meso scale devices like synthetic jets.

Heat transfer paths in wireless communication devices, such ascellular phones, pagers, two-way radios, and PDAs, etc. were pre-

sented by Lee et al. 1. Much of the heat generated in IC pack-

ages is first conducted into printed circuit boards, then transported

to the housing interior walls by conduction, convection through

air and radiative processes. The heat must then be conducted

through the case walls, and finally removed to the ambient. The

low thermal conductivity of the fiberglass epoxy resin-based

printed circuit boards leads to a high thermal resistance between

the heat source and the ambient. With the advent of smaller en-

closures, higher digital clock speeds, greater numbers of power

emitting devices and shorter design cycle times, thermal manage-

ment becomes critical and must be considered early in the design

cycle. Recently, Chriac et al. 2 described further thermal prob-

lems in power amplifiers in mobile devices. CFD simulations for

a power amplifier device were presented, and tight space naturalconvection was analyzed.

A study of synthetic jets, for thermal management of compactelectronic devices such as cell phones was presented by Min-ichielo et al. 3. They used synthetic jet technology to promoteactive cooling in a cellular phone mock-up. The authors observed

heat transfer enhancement of 2–4 times over natural convection. A

similar enhancement over natural convection in air was also ob-

served by Gillespie 4. Parametric variation of enhancement in

natural convection and derivation of optimum parameters was pre-

sented.

Synthetic jets are small scale devices and they do not require an

external fluid source eliminating the need for input piping and

complex fluidic packaging. Researchers in several universities and

companies have focused on both theoretical and experimental as-

pects of how to implement these “zero-net mass-flux” jets in real

applications for a decade. Promising thermal findings have been

published for small volume, high flux applications such as cell-

phones and PDAs.

In recent years, plane and round synthetic jets, that by time-

periodic alternate ejection and suction of the working fluid

through an orifice, have been investigated both experimentally

and numerically 5. The studies have emphasized a compact flow

generator in which the orifice forms one of the surfaces of an

otherwise sealed shallow cavity, where the flow is driven by the

motion of a diaphragm that is built into one of the cavity walls.

For a given actuation input, the effectiveness of the flow generator

can be maximized when the diaphragm and cavity are driven at a

coupled resonance that depends on both the cavity flow and the

structural characteristics of the diaphragm. An isolated synthetic

jet in the absence of a cross flow is produced by the interactions of

a train of vortices that are typically formed by alternating momen-

tary ejection and suction of fluid across an orifice such that the net

mass flux is zero. Whereas the nominally axis-symmetric flow

during the suction stroke may be thought of as similar to the flow

induced by a sink that is coincident with the jet orifice, the flow

during the ejection stroke is primarily confined to a finite narrow

domain in the vicinity of the jet centerline. During the momentary

ejection, the flow separates at the sharp edges of the orifice and

forms a vortex sheet that typically rolls into a vortex 6.

Synthetic jet enhancement of natural convection and pool boil-

ing heat transfer in an enclosure filled with a dielectric liquid

FC-72 was presented by Garg et al. 7. The jet actuator used in

Contributed by the Electronic and Photonic Packaging Division of ASME for

publication in the JOURNAL OF ELECTRONIC PACKAGING. Manuscript received October

11, 2004; final manuscript received April 20, 2005. Review conducted by: Stephen

McKeown. Paper presented at the 2004 ASME Heat Transfer/Fluids EngineeringSummer Conference HT-FED2004, July 11, 2004–July 15, 2004, Charlotte, North

Carolina, USA.

Journal of Electronic Packaging DECEMBER 2005, Vol. 127 / 503Copyright © 2005 by ASME

Downloaded 19 Oct 2011 to 128.113.107.4. Redistribution subject to ASME license or copyright; see http://www.asme.org/terms/Terms_Use.cfm



this study, produced planar, submerged liquid jets, which im-pinged upon the flat foil heater and spread laterally along its sur-face. Augmentation of convective thermal transport from the

heated surface was found to peak around 300 Hz and 120 V RMS,yielding nearly fourfold improvement over the natural convection.The highest enhancement of pool boiling was observed near boil-ing incipience, with the synthetic jet producing an earlier transi-tion to nucleate boiling than encountered in a quiescent pool. Thedevelopment of a closed-loop two-phase microchannel coolingsystem using electro osmotic pumping was presented by Jiang etal. 8. The silicon heat exchanger achieved junction-fluid resis-

tance near 0.1 K/W, using 40 channels with hydraulic diameter of 100 m. The electro-osmotic pump with working volume of

1.4 cm3 achieved a maximum backpressure and a flow rate of 160

kPa and 7 ml/ min, respectively, using deionized water. Theclosed-loop system removed 38 W with pump power of 2 W, and

junction-ambient thermal resistance near 2.5 K/W.An alternative technology to the synthetic jets is cantileverlike

piezofans. Although they do not produce high magnitude flows,their effective flow area is higher than concentrated impingingsynthetic jets. Recently, piezoelectric fans for the thermal manage-ment of electronic devices, were presented by Acikalin 9. Sev-eral piezoelectric fans were fabricated and tested. For a piezoelec-tric fan with a steel shim length of 5 cm and 35% PZT patchcoverage, a tip deflection of more than 2 cm was attained under 60VAC and 20 Hz. The piezo fans had a maximum power consump-tion of 10 mW. The fans increased the heat removal rate for a heatsource by more than a factor of 2 relative to natural convection ina test setup comparable to a cell phone. When the orientation wasaltered, the enhancement due to the piezoelectric fan decreasedsignificantly. In a commercially available laptop, piezoelectricfans resulted in a temperature drop of more than 6 °C in nearbypower electronics, and more than an 8 °C temperature drop in thesurrounding air, over and above the effects of the rotational faninstalled in the computer. The use of micro synthetic jets to dis-rupt the laminar flow in forced air convection in microchannelsand improve the heat transfer to the cooling air was numericallystudied by Timchenko et al. 10. They reported that a fully opti-mized synthetic jet has the potential to significantly improve heattransfer from an integrated circuit to the cooling fluid.

This paper discusses an experimental study for determining theperformance of a synthetic jet for possible thermal management of

electronic components. The jets were driven by a time harmonicsignal and the operating signal frequency was varied between 3.4and 5.4 kHz. The resonance frequency for a particular jet wasdetermined by measuring the exit air velocity. The effect of vary-ing the signal voltage and the spacing between the synthetic jetand heater on heat transfer augmentation was also studied. A mi-croscopic an infrared thermal imaging technique was used to en-able small area temperature measurements. The use of an infraredthermal imaging technique enabled the temperature distributionfor the entire heater surface to be obtained, instead of just a fewpoint measurements via thermocouples.

Experimental Setup

The test setup was designed in such a way that the temperature

measurements could be made using an IR camera. A thin foilheater was used as the heat source. In order to obtain the tempera-ture profile of the entire surface, it was critical that the entiresurface was visible to the IR camera. Temperature measurementshad to be made while the jets were operating. The only way thiscould be achieved was by allowing the jet to impinge on onesurface of the heater, and making the temperature measurementsfrom the back surface using the IR camera. In order to ensure thata large fraction of the electrical power input to the heater wasdissipated from the surface facing the jets, it was important tominimize the heat losses.

The test set up used to study the heat transfer augmentation dueto synthetic jets is shown in Fig. 1. A thin foil heater was used as

the heat source. The heater was attached to a plexiglass frame toimprove the field of view. The frame was attached to a styrofoaminsulator. A Lexan holder was used to support the structure. TheLexan holder was designed to house a 100 mm diameter, IR-transparent ZnSe window at the center. As can be seen from thefigure, both sides of the window were covered with Styrofoamplates to minimize heat losses. The heater temperature profile was

first obtained under natural convection conditions. The synthetic jet was then turned on and allowed to impinge perpendicularly onthe thin foil heater. Heat transfer augmentation was measured bycomparing the temperature profiles under natural convection con-ditions with the “jet on” conditions. In order to be able to get thetemperature profile for the entire heater with the maximum num-ber of points, it was decided to use state-of-the-art microscopicinfrared imaging for temperature measurement.

Figures 2 and 3 show the pictures of the front and back viewsof the test set up. The jet was mounted on a traverse system,which enabled precise 3D location adjustment of the jet. In orderto ensure that most of the electrical power input to the heater wasdissipated from the surface of the heater facing the jet, it wasessential to minimize the heat losses from the surface facing theIR camera. To avoid heat losses, a zinc selenide ZnSe window

was mounted in between the IR camera and the heater. The win-dow was mounted on the same Lexan holder on which the heaterwas mounted. ZnSe has high transmissivity to infrared radiation in

the wavelength range of 8–10 m. Two properties had to beknown in order to make temperature measurements using IR cam-era; the emissivity of the surface and the transmissivity of the IRwindow. The surface of the heater facing the IR camera waspainted with a black paint, whose emissivity was measured using

Fig. 1 Schematic of the heater arrangement

Fig. 2 Digital image of the jet side

504 / Vol. 127, DECEMBER 2005 Transactions of the ASME




calibration experiments. The transmissivity of the IR transmissivewindow was also measured using separate calibration experi-ments.

The synthetic jet was powered by means of a function genera-tor. A time harmonic signal generated by the function generatorwas fed to a power amplifier. The amplified signal was then fed to

the synthetic jet. An oscilloscope was placed in the power cir-cuitry to precisely measure the frequency and driving voltage.

Calibration of the Emissivity and the Transmissivity. Twomajor components involved in the IR temperature measurementswere the emissivity of the painted surface and the transmissivityof the IR window. If those parameters are not calibrated accu-rately, significant temperature measurement errors could persist.Therefore, a careful calibration procedure was developed and car-ried out. In order to measure the emissivity of the black paint, acalibration plate was fabricated out of highly conductive alumi-num. The plate was 50.8 mm in diameter with a thickness of 6mm. A thin foil heater was attached at one face of the calibrationplate. On the other face of the plate, 0.5 mm deep grooves weremade for mounting T-type thermocouples. Thermocouples weretightly located in the grooves and the remaining clearances were

filled with thermally conductive epoxy. The face of the plate withthe thermocouples was painted with black paint.

The foil heater was heated to several power levels to reach thedesired plate temperatures between 30 °C and 90 °C. At eachpower level, the temperature was measured using thermocouplesas well as the IR camera. By adjusting the emissivity parameter inthe IR camera, temperatures were matched for both measurementtechniques. The adjusted emissivity was taken to be the emissivityof the black paint.

The thermocouple reading at steady state for the first calibrationpoint was 48.4 °C. The IR point measurement for the thermo-couple location, with an emissivity setting of 0.98 was 48.2 °C.Two areas were selected near the thermocouple location. Theiraverage temperatures at steady state were 49.2 °C. Thus, an emis-sivity of 0.98 gave excellent agreement with thermocouple mea-

surements. A higher temperature setting of close to 80 °C wasanother point for paint calibration. A sample IR image is shown inFig. 4. The thermocouple measurement for this setting was80.4 °C, while the IR point measurement was 79.9 °C for theemissivity of 0.98. Once again two areas were selected in thevicinity of the thermocouple location. The area averages of thetwo areas were 80.1 °C and 80.2 °C. Good agreement was ob-tained between the thermocouple measurements and IR measure-ments by using an emissivity of 0.98 for the black paint.

The next step was to obtain the transmissivity of the IR win-dow. In order to obtain the transmissivity of the window, the samecalibration plate was used. The heat flux to the heater was ad-

justed such that the temperature of the calibration plate was in the

vicinity of 80 °C. The steady state temperature of the calibrationplate was first measured using IR camera alone without the win-dow. Then the window was placed in between the IR camera andthe calibration plate. At this point the transmissivity of the exter-nal optics was set to 1. Since the insertion of the ZnSe window

caused the amount of radiation received by the camera to drop, thetemperature recorded by the camera also dropped. The transmis-sivity value was then adjusted to make the recorded temperatureequal to the steady state temperatures recorded without the win-dow. Two different transmissivity values were used, 0.96 and0.97. The results of this study are shown in Fig. 5. The steadystate temperature recorded without the window was 82.0 °C. Thetemperature recorded by the camera, upon insertion of the IRwindow, but without making any adjustment to the transmissivitywas 80.2 °C. The temperature recorded with a transmissivityvalue of 0.96 was 82.1 °C and that with a transmissivity of 0.97was 81.7 °C. Thus for an emissivity of 0.96 there was a goodagreement between the temperatures recorded by the IR camerawithout the window, and the temperatures recorded using the win-dow. Therefore 0.96 was taken to be the value of the transmissiv-

ity of the IR window. The measured transmissivity agreed verywell with the manufacturer’s data 11.

An array of 400400 pointwise temperatures on the heaterwere recorded using a microscopic IR system. This would allowpointwise Nu numbers and heat transfer coefficients in the theo-retical study of the synthetic jets later.

Velocity Measurements. The jet velocity was measured bymeans of a TSI IFA 100 hot wire anemometry system. The jetoperates by periodic ejection and suction of the fluid. In order tocapture both the suction and the ejection velocities the probe wasplaced close to the jet orifice at a distance of approximately 1 mm.A computer code was used to separately average the ejection

Fig. 3 Digital image of the camera side

Fig. 4 IR image at high temperature

Fig. 5 Transmissivity calibration

Journal of Electronic Packaging DECEMBER 2005, Vol. 127 / 505




peaks and suction peaks. A typical velocity profile for the syn-thetic jet is shown in Fig. 6. The average exhaust velocity andaverage suction velocity was calculated. In order to study thevariation of velocity with the driving frequency, the frequencywas varied and average exhaust velocity was measured at eachfrequency.

The exhaust velocity increased with driving frequency until itreached a peak and then started dropping. So there existed anoptimum frequency at which the exhaust velocity was maximum.The exhaust velocity also increased with driving voltage. The

variation of maximum exhaust velocity with driving voltage isshown in Fig. 7. It can be seen from the figure that the meanexhaust velocity increased sharply from a driving voltage of

50 VRMS to 90 VRMS, but increased much less from 90 VRMS to

110 VRMS. The mean exhaust velocity at 50, 90, and 110 VRMS

was 59, 86, and 92 m/s, respectively. It was observed that the

velocity behavior started saturating after 90 VRMS. Further in-crease in the driving power would not increase the velocities asmuch as at low voltages.

Uncertainty Analysis. Standard techniques were followed toobtain the experimental uncertainty 12. The basic uncertaintysources were temperature and power measurements, heat losses,and calibration errors. The IR measurements were calibratedagainst standard T-type thermocouples. The uncertainty associatedwith temperature measurement using the thermocouples was

±1°C. The error in temperature measurements using the IR cam-era is introduced at two stages, first due to the emissivity of thepaint, and second due to the IR transmissive window. The error in

temperature measurements due to emissivity is ±0.5°C. The totalerror in the measurement of temperature using IR camera is 1.1 °Cin the temperature range where the experiments have been per-formed. It is interesting to note that most of this uncertainty wascontributed by the thermocouple.

A standard 0.1 precision resistor was placed in the power

circuit to measure the current. Total uncertainty in the power mea-surements was estimated to be approximately 0.05%.

Experimental Procedure and Results

The thin foil heater dimensions were selected to be 12.5

12.5 mm2. The heat transfer augmentation experiments wereconducted at two different heater power levels. The heat flux lev-els from the heater surface were chosen such that the maximumheater temperatures in natural convection were in the vicinity of 50 °C and 80 °C. At each heat flux level the variation of heattransfer augmentation with axial distance between the jet andheater as well as with driving voltage was studied. The axial dis-tance between the heater and the jet was varied between 2 mm and24 mm at seven distinct locations. In addition, the driving voltage

was varied between 50 and 90 VRMS.The driving frequency during the experiments was kept con-

stant and equal to the resonance frequency. The resonance fre-quency was previously determined from two different experi-ments. First, the resonance frequency of the jet was determinedfrom the variation of the velocity with frequency experiments.

Secondly, separate heat transfer experiments were conductedwhere heat transfer augmentation was measured as a function of driving frequency. From these experiments the resonance fre-quency was measured to be 4400 Hz.

In the first set of heat transfer augmentation experiments con-ducted, the power to the heater was set to such a level that theheater temperature was in the vicinity of 50 °C. The total powerinput to the heater for this setting was 0.3 W. The IR image of theheater under natural convection conditions is shown in Fig. 8.Temperature at the lexan frame was found to be around 33 °C,while the main styrofoam structure was at 25 °C or near ambient.

After the system reached steady state under natural convectionconditions, the IR temperature distributions were recorded. There-after the axial distance between the jet and the heater was set to 2mm. The synthetic jet was turned on, with the driving jet fre-quency set to 4400 Hz. The driving voltage for the first set of

experiments was 50 VRMS. After the system reached steady state,the IR image was recorded. The axial distance was then variedfrom 2 mm to 24 mm, in several steps, and at each step the steadystate temperatures were recorded. The procedure was repeated for

driving voltages of 70 VRMS and 90 VRMS. Figure 9 presents thevariation of the temperature rise above ambient with the axialdistance for various driving voltages. The heater temperature wasfound to decrease as the axial spacing between the jet and theheater was increased, until it reached a minimum. Thereafter in-creasing the distance led to an increase in the heater temperature.Optimum cooling was obtained at a distance of approximately 8mm. The temperature rise above ambient under natural convection

Fig. 6 Typical velocity response of a synthetic jet

Fig. 7 Variation of the exhaust velocity with driving voltage

Fig. 8 Temperature distribution under natural convection con-dition for lower heat flux





conditions was 27.8 °C, while the minimum value under “jet on”conditions was 7.2 °C. Thus the temperature rise decreased by afactor of 3.9 with the jet on.

During the second set of experiments the heater power was setto a level such that the heater temperature was in the vicinity of

80 °C. The heater temperature profiles under natural convectionand synthetic jet operating conditions are shown in Figs. 10 and11, respectively. Figure 12 presents the variation of the tempera-ture rise with axial distance for various driving voltages at thissecond heat flux level. It can again be seen that there is a maxi-mum in heat transfer enhancement at around 8 mm “jet to heater”axial distance. The heater temperature rise above ambient under“jet on” conditions decreased by a factor of 4.1 compared to natu-

ral convection, for the high heat flux case. Going from 50 V RMS

excitation to 70 VRMS only decreases the temperature by 1.5 K,

while further increasing the jet driving power to 90 V RMS did notfurther improve cooling. In order to study the enhancement pro-vided by the jet, an enhancement factor was defined. The defini-tion of enhancement factor is given in Eq. 1,

Enhancement =

h jet

hnc =

q jet

qnc

T s − T airnc

T s − T air jet 1

In Eq. 1, T s is the average heater temperature and q jet is theheat flux being dissipated from the surface of the heater facing the

synthetic jet. Clearly, q jet was not equal to the total heat input tothe heater, as there was a finite heat loss from the backside of theheater as well as by conduction from the heater to the support

fixture and from the support fixture to ambient air. Similarly, qnc is

the heat flux being dissipated from the surface of the heater facingthe ambient air and not facing the IR window. Again this heat fluxwas not equal to the total heat input to the heater as there was heatloss by conduction to the support fixture as well as from the sur-face of the heater facing the IR window. In order to estimate theenhancement factor correctly, it was important to estimate these

losses and use the correct values of q jet and qnc. An experimentalapproach was adopted to estimate these losses and is describedbelow.

The approach consisted of three steps. In step 1, a thin foilheater was held in ambient air, using the electrical wires for sup-port. In this configuration, both surfaces of the heater, were ex-posed to the same ambient conditions. The heat flux to the heaterwas turned on and steady state temperatures were recorded. It canbe expected that in this configuration, the heat dissipation fromboth sides of the heater was the same and equivalent to the freeconvection heat transfer from the heater. Therefore heat dissipa-tion from one side of the heater was half of the total input power.Heat input to the heater was varied from 0.028 W to 0.39 W in sixsteps. At each heat flux, steady state heater temperatures wererecorded. A function relating the natural convection heat dissi-

pated to the average temperature difference between the surface of the heater and the ambient was obtained. Regression analysis wasdone to fit a second order polynomial to the experimental data.The function relating the heat flux to temperature difference isgiven in Eq. 2,

Qnc = 0.00003395T 2 + 0.0053T − 0.003838 2

where

T = T s − T air

Fig. 9 Variation of the temperature rise with jet axial distancefor low heat flux case

Fig. 10 Temperature profile under natural convection for highheat flux case

Fig. 11 Heater temperature profile with the jet on „V =50 V,Distance=8 mm…

Fig. 12 Variation of the temperature rise with jet axial distancefor high heat flux





The data from this first step provides an estimate for free con-

vection losses from the heater to ambient. The data for the naturalconvection experiments with the heater in the fixture was used, inconjunction with the data from first step, to estimate the conduc-tion and convection heat loss from the back of the heater the sidefacing the IR camera for a given heater temperature. There weretwo natural convection runs, with the heater inside the fixture: onewith the heater at approximately 50 °C and the other with theheater at approximately 80 °C. The function from step 1 was usedto estimate how much heat was being dissipated from the front

surface Qfront of the heater for these two cases. Then by sub-

stracting Qfront from Qtotal, heat dissipated from the back surface

was obtained Qloss for these two temperatures see Table 1. The

natural convection heat transfer coefficients are presented in thelast column of Table 2, and simply defined as

h =

Qfront

AT 3

It is to be noted that Qfront was a sum of heat dissipated by both

convection and by radiation. Thus h, obtained by using the aboveequation was an effective heat transfer coefficient. The reason to

choose effective h instead of a pure natural convection h, was that

in any real application the heat dissipated from the surface would

be a sum of both convection and radiation components. The h

used henceforth in this paper, is based on heat lost by both con-

vection and radiation.

A linear equation relating Qloss to the temperature difference

between heater and ambient was obtained:

Qloss = 0.007T + 0.00162 4

In step 3, the heat transfer coefficients with the jet on were

determined. Heat losses from the back face of the heater were

determined by using Eq. 4. By subtracting the heat losses from

Table 1 Heat loss from the heater back face during natural convection

Table 2 Heat transfer coefficients for low heat flux condition





total heat input, the heat flux dissipated from the front face of theheater was determined. This value was then used to determine theheat transfer coefficient with the jet on.

The data for the heat transfer coefficient for natural convectionand “jet operating” conditions for low heat flux conditions is pre-sented in Table 3. Figure 13 presents the heat transfer enhance-ment for the low heat flux condition for three jet excitation powerlevels. When the jet was located at the closest location i.e., 2mm of 2.3 jet diameters from the heater surface, the overall en-hancement was measured to be 5.5. The enhancement was found

to be maximum at an axial distance of approximately 8 mm, or 9.3

jet diameters. The maximum enhancement was found to be ap-

proximately 10 times that of natural convection.

Although the reason for an optimum spacing is being further

studied, there are several possibilities that can account for it. First,

at low spacing, the area of the heater being cooled by impinge-

ment is much smaller compared to larger spacing. This implies

that the region of high heat transfer coefficient is smaller at asmaller spacing, leading to a lower average heat transfer coeffi-

cient. This in turn leads to higher average heater temperature.Another important aspect is the coolant temperature. The air

exiting the jet during the exhaust cycle is hotter than the ambient,mainly because of the heat generated within the jet itself. When

the jet is placed very close to the heat source, the temperature of the impinging air is close to the temperature of the air exitingfrom the jet. When the distance between the jet and the heater isincreased, ambient air gets entrained, causing the temperature of

the air impinging upon the heater to decrease substantially. Inorder to study this effect more carefully, air temperature measure-ments were made at different distances from jet orifice in the

absence of a heater. At 90 V RMS driving voltage, the air tempera-ture increased by 11 °C above ambient at a distance of 1 mm from

the jet orifice. The temperature increase at a distance of 8 mm wasclose to 4 °C. When the jet was located further, the performancedegraded. This is likely due to reduced velocities due to expansionof the jet.

The variation in jet thermal performance for three excitation

Table 3 Heat transfer coefficients for high heat flux condition

Fig. 13 Enhancement factor for lower heat flux case





levels was found to be within 10%. The heat transfer coefficientsfor the higher flux condition are presented in Table 3. Figure 14presents the results for the higher power boundary condition.Similar enhancements were obtained for the high power condi-tion. The 70 V and 90 V cases showed very similar behaviors,while the 50 V condition had about a 12% lower performance. Forthe same heater it is interesting to see the independence of thermalperformance from the heat flux of the heater. This will provide awider design space for the engineers. This is a strong indicationthat the synthetic jet is producing forced convection that is over-whelming heat-flux-driven free convection effects.

A coefficient of performance was defined to study the heattransfer augmentation relative to the power consumed by the de-vice. The COP was defined as the ratio of heat removed by the jetto the power consumed by it for a temperature rise of 15 °C abovethe ambient.

COP = Q jet

PT =15°C

5

The values of COP for 3 different jet power levels are presented inTable 4.

It can be seen from Table 4 that the COP is maximum at adriving voltage of 50 V, but decreases rapidly after that. In fact at90 V driving voltage, the power consumed by the device is higherthan the amount of heat removed from the heater. This clearlyindicates a benefit in operating the jet at lower driving voltages.

The tests were not conducted at voltages below 50 VRMS, and thevariation of COP with driving voltage at lower voltages remainsto be studied.

Predicting Pulsating Flow Heat Transfer. Impingement flowsallow for short flow paths on the surface and then relatively highheat transfer rates are achieved. Over the last five decades, a largenumber of experimental and numerical studies have been per-formed to understand the flow behavior and heat transfer charac-teristics. Many correlations have also been proposed by research-ers, to compare the synthetic jet behavior with a conventionalsteady impinging jet. The empirical correlation from a single slotnozzle impingement jet is given as the following 13:

Nu= Pr0.42G r

Dh

, H

Dh

F 1Re 6

where

F 1 = 2 Re1/21 + 0.005 Re0.551/2

and

G = Dh

r

1 −1.1 Dh

r

1 + 0.1 H

Dh

− 6 Dh

r

h =Nu k

Dh

This expression is valid for a range of Reynolds number be-

tween 2,000 and 400,000, while H / Dh ratio should be between 2

and 12, as well as r / Dh ratio should be between 2.5 and 7.5. Inorder to use a steady jet correlation to predict the heat transfercoefficient for a pulsating jet, a suitable jet exit velocity had to beused. For the current study, a time-averaged velocity was chosenas the velocity to be used in the correlation. So for a driving

voltage of 50 VRMS, even though the peak exit velocity was

59 m / s, the time averaged velocity was only 29.3 m / s. Similarly

for 90 VRMS driving voltage, the peak exit velocity and time-

averaged velocities were 86 m/s and 44.9 m/s, respectively. Forthe current experiments the Re number was found to be between

1500 and 2200 for 29.3 and 44.9 m/s, respectively.

The comparison between predicted and experimental heat trans-fer coefficients is shown in Fig. 15. While the steady jet heattransfer coefficient continued to increase as the distance was de-creased, the pulsating jet heat transfer coefficient was found to beoptimum at a spacing of 8 mm. Decreasing the distance beyond 8mm was found to decrease the pulsating jet heat transfer coeffi-

cient. At a distance of 8 mm, the predicted h was 14% higher than

the experimental heat transfer coefficient for 50 VRMS drivingvoltage. As the distance was increased beyond 8 mm, the experi-mental heat transfer coefficient was found to decrease moresharply than the steady jet. At a distance of 24 mm, the predictednumber was 30% higher than the experimental findings. For

90 VRMS driving voltage, the discrepancies were found to behigher, being 31% at 8 mm spacing and 44% at 24 mm spacing.

Since, the primary goal of the current study was determining

the effectiveness of the synthetic jets for thermal management of confined high flux applications, no correlation development was

Fig. 14 Enhancement factor for higher heat flux case

Table 4 Coefficient of performance

Fig. 15 Predicted and experimental heat transfer coefficients





sought. A further study based on an extensive DOE i.e., design of experiments with jet parameters such as size, driving power, fre-quency, orifice size, hydraulic diameter, and heater geometric con-siderations such as axial distance, heated surface size, and surfacecharacteristics, etc. would be useful to find a suitable correlationto predict the Nu numbers and heat transfer coefficients in un-steady jets.

Summary and Conclusions

An experimental investigation of synthetic jet devices to ob-serve the unsteady impingement heat transfer over a small heatersurface area was implemented to determine the full-field thermalmaps and heat transfer characteristics. The point-wise tempera-

tures for pixel sizes as small as 30 m were determined withmicroscopic IR measurements. The micro fluidic jet was excitedbetween 0.15 mW and 0.67 mW. The resonance frequency wasfound to be 4400 Hz. It was found that jet effectiveness stronglydepends on the axial distance and driving frequency. The maxi-mum heat transfer enhancement over the natural convection wasfound to be approximately 10 at a jet distance of 8 mm or 9.3 jetorifice diameters. It was found that the enhancement was slightlylower at higher heat fluxes. This was mainly because of an in-crease in natural convection heat transfer coefficient at higherheater temperatures. The maximum experimentally measured heattransfer coefficient, based on heater surface area and average

heater temperature, was found to be approximately 240 W/m2 K.

Nomenclature A surface area of the heater m2

COP coefficient of performance heat removed/ consumed power

D hydraulic diameter mh heat transfer coefficient W / m2 K

H jet axial distance mIR infrared

k thermal conductivity of air W/m Km constant in equation

Nu Nusselt numberP power consumed by the jet W

Pr Prandtl number

Q heat Wq heat flux W / m2

Re Reynolds numberr radius mT temperature °CU velocity m / s x half of heater length mV voltage V

W jet slot width m

Greek symbols

density kg/m3

T temperature difference K dynamic viscosity kg m/ s2

Subscriptsair ambient

e experimentfront surface of the heater facing the synthetic jet

jet synthetic jeth hydraulic

loss

heat dissipated from the surface of the heaterfacing the IR cameranc natural convection

off jet off on jet on p predicted

RMS root mean squares heater surface

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Documents

Garg, J., Arik, M., Weaver, S., Wetzel, T., Saddoughi S., Meso Scale Pulsating Jets for Electronic Cooling