miniature heat sinks

7/30/2019 Miniature Heat Sinks

1/13

#P'ZRGTKOGPVCN5VWF[QPVJG2GTHQTOCPEGQH/KPKCVWTG*GCV5KPMUHQT

(QTEGF%QPXGEVKQP#KT%QQNKPI

V. Egan, J. Stafford, P. Walsh, E. Walsh, R. Grimes,

Stokes Research Institute,Department of Mechanical & Aeronautical Engineering

University of LimerickLimerick, Ireland

Phone: (+353) 61 213487

Fax: (+353) 61 212039

Email: [email protected]

ABSTRACTIn recent years the design of portable electronic devices must

incorporate thermal analyses to ensure the device can beadequately cooled to acceptable temperatures. Consumer

demand for smaller, more powerful devices has lead to an

increase in the heat required to be dissipated and a reduction in

the surface area both of which result in an increased heat flux.In this paper, an experimental study is performed on one of the

smallest commercially available miniature fans, suitable forcooling portable electronic devices, used in conjunction with

both finned and finless heat sinks. Previous analysis has

shown that due to fan exit angle, flow does not enter the heat

sinks parallel to the fins or bounding walls. This results in anon uniform flow rate within the channels of the finned and

finless heat sink along with impingement of the flow at the

entrance giving rise to large entrance pressure losses. In thispaper straightening diffusers were attached at the exit of the

fan which resulted in aligning the flow entering the heat sinks

with the fins and channel walls. In designing the finned heatsink current optimization criterion for finned heat exchangers

has been applied to ensure maximum heat transfer rates; the

finless heat sink was designed to the same specifications. The

maximum overall footprint area of the cooling solution is534mm2 with a profile height of 5mm. The thermal

performance of each cooling solution was investigated byquantifying its thermal resistance over a range of fan speeds

and comparing the results to cases without diffusers. In orderto investigate the flow field, detailed velocity measurements

were obtained using Particle Image Velocimetry, which

provided a further insight into the physics of the flow in such

miniature geometries and in designing the straighteningdiffusers. The thermal analysis results indicate that the cooling

power of the solution is increased by up to 20% through the

introduction of a diffuser. Hence, demonstrating the need forintegrated fan and heat sink design of low profile applications.

KEY WORDS: low profile, forced convection cooling,

miniature fan

NOMENCLATURE

Aconv convective surface area, m2

b fin spacing, mmFOV Field of view, m

h convective heat transfer coefficient, W/m2 K

hFC forced convective heat transfer coefficient,W/m2 K

hf total Convective heat transfer coefficient,

W/m2 Kk

heat sinkthermal conductivity of heat sink material

W/m.K

L Length of heat sink, mmN number of pixels in interrogation region

'P pressure drop, PaQ volumetric flow rate, m3/s

QFC power dissipated by forced convection, W

Qinput input power, WQLosses power dissipated by secondary cooling

mechanisms, WRES resolution of camera, Pixels

RthT total thermal resistance,o C/W

RthFC thermal resistance of Forced Convection

Cooling, o C/W

RthLosses thermal resistance of Natural ConvectionCooling, o C/W

RPM revolutions per minute, RPMSP static pressure rise, Pa

tfin fin thickness, mm

Tf ambient air temperature,o C

Ts surface temperature,o C

Umax Maximum velocity, m/sux,y Velocity at local coordinates (x,y), m/s

W Total width of heat sink, mm

978-1-4244-1701-8/08/$25.00 2008 IEEE

497


2/13

X Half of channel width, mmY Half of channel height, mm

Greek Symbols

D thermal diffusivity, m2/sP viscosity , N.s/m2

Subscripts

conv convective

FC forced ConvectionIR interrogation Region

Max maximum

s SurfaceT Total

f Ambient

INTRODUCTION

The design and integration of miniature cooling solutions intotodays portable electronic devices is moving to the forefront

of electronics cooling research. The drive behind this researchis consumer demand for smaller more powerful portable

devices. As the surface area available for power dissipation

reduces, along with technology advancements increasing the

devices power consumption, the resulting heat fluxes areconsiderably elevated. At present most manufacturers rely on

natural convection in conjunction with heat spreadingtechnologies to maintain temperatures at acceptable consumer

levels. However, a major flaw with this method of heat

dissipation is the resulting elevation in device case

temperatures. If the envisaged trend in increasing heat fluxescontinues, then natural convection and conduction

technologies will no longer be able to maintain devices atacceptable temperatures. Hence, there is a need for research

into cooling technologies that can offer higher heat transfer

rates compared to current natural convection and heat

spreading techniques.

Forced convection cooling is the preferred new coolingtechnology among designers of portable electronic devices.

This is due to large number of factors including, cost,

reliability, power consumption, footprint area, and profile

height. Liquid cooling [1] and heat pipe [2] technologiesrequire secondary coolers but for forced convection air is the

working medium which is in abundant supply. Unlike phasechange materials [3], forced convection is capable of

efficiently dissipating heat over extended time periods. An

example of this is the unconventional design proposed by

Walsh et al [4], which gives a thermal resistance value of7.5C/W in a volume of 4.56mm2 with a profile height of 4mm

including the fan. As air cooling is currently in widespread usein industry the implementation costs are relatively low.Finally, with a sufficiently small footprint and profile, heat can

be dissipated directly at the source via a path of low thermal

resistance, resulting in a reduction in the temperature of allother components and surfaces of the device.

Over the past few years a number of groups have initiatedresearch to address the developing market of miniature forced

convective air cooling. The first series of papers on such

research examined the validity of the conventional scaling

laws [5-9] as the fan scale is reduced to miniature dimensions.The authors found that at low Reynolds numbers boundary

layer phenomena result in reducing the flow coefficient and

elevating the power coefficient. Day et al[10] also reported a

scale effect which resulted in the flow rate delivered beingsignificantly less than expected when operating at low

Reynolds numbers.

With regard to the integration of miniature fans with heat

sinks it is noted that impingement cooling emerges as one of

the best methods to dissipate large heat loads over small areas,with thermal resistances as low as 1.4 oC/W being reported

[11,12] for pin-fin heat sinks. However, to achieve such

performances, footprints areas of order 2000 mm2

and profileheights of order 100 mm were reported. Whilst this footprint is

relatively small, it is noted that such large profile heights

would most likely deem this technology unacceptable for usein most portable electronic devices. Also, when considering

miniature scale fans, the flow rates and pressure rises

attainable are considerably lower than those attained in theaforementioned studies for impingement cooling. Hence in

order to satisfy the constraints of low profile forced

convection cooling a fan and heat sink in parallel is the mostpractical solution. A recent analysis of fan and heat sink

arrangement in term of performance optimisation has been

carried out by Walsh et al [13, 14]. It was shown that bestthermal performance was achieved when the fan exit angle is

matched to the heat sink fin angle, i.e. flow enters the heat

sink parallel to the fins. A recent study by Egan et al[15, 16]investigated the thermal performance of a miniature fan used

in conjunction with a finned and finless heat sink. The overall

footprint area of the solution was 456mm2 with a profile of

5mm. Velocity measurements of the flow in the heat sinksshowed that flow exited the fan at an angle to the heat sink

walls and the fins. In the case of the finned heat sink it was

seen that the flow strongly impinged on the heat sink fins atthe entrance to the channels. The current study is focused on

developing the previous analysis of Egan et al [15, 16] to

reflect the finding of Walsh et al [13, 14] in designing asuitable cooling solution.

The primary objective is to determine the effect of

straightening the flow prior to entering the heat sinks through

the use of a diffuser. A number of different diffusers were

manufactured to obtain the optimum design giving parallelflow at entry to the heat sink. The optimum diffuser design

was determined from flow velocity measurements obtainedusing particle image velocimetry. Upon finding such a design

the fan, diffuser and heat sink solutions were characterised by

repeating the heat transfer measurements carried out in Egan

at al [15, 16]. Such results would conclusively show if thediffuser resulted in an overall enhancement of the heat transfer

rate of the finned and finless solutions. Particle imagevelocimetry measurements of the flow in the heat sink at

speeds ranging from 3000RPM to 8000RPM were also

obtained and provide a further insight into the flow field

within the heat sinks. The results of this study are valuable todesigners of such cooling solutions as enhanced heat transfer

498


3/13

rates may be obtained by simply adding a diffuser between thefan and heat sink.

HEAT SINK AND DIFFUSER DESIGN

The fan tested for the current work is the Micronel U16LM-9having a footprint of 100mm2 and profile of 5mm (See Fig. 4).

This is currently one of the smallest commercially available

fans for forced convection cooling of electronics. The

pressure versus flow rate characteristics supplied by themanufacturer for a nominal speed of 6000RPM are shown in

Fig. 1. In determining this curve it is possible that only a small

number of experimental data points were recorded by thesupplier, hence resulting in a sharp change in flow rate once a

certain static pressure is reached. Also shown on this graph are

the curves for 3000RPM, 5500RPM and 8000RPM. Thesecurves were obtained using the fan scaling laws detailed in

Bleier [17] and defined here in equations 1 and 2.

vQ RPM (1)

2vSP RPM (2)

0

1

2

3

4

5

6

7

8

9

10

11

0 0.00005 0.0001 0.00015 0.0002 0.00025 0.0003

Volume Flow Rate [m3/s]

StaticPressure[Pa]

Supplier Curve Scaled to 8000RPM Supplier Curve Scaled to 5500RPM

Supplier Curve Scaled to 3000RPM Pressure V's Flow curve at 6000RPM

Fig. 1. Pressure versus flow rate characteristics for the

Micronel U16LM-9 low profile fan as supplied by Micronelfor 6000 RPM and scaled for speeds from 3000PRM to

8000RPM using the fan scaling laws.

Experiments were carried out in both finned and finless heat

sinks. In designing the finned heat sink the pressure versusflow rate characteristics at a fan speed of 8000RPM were

used. Fig. 2 (a) shows a schematic of this heat sink and

highlights the variables which were optimized; these include

the fin spacing, (b), and fin thickness, (tfin). The former ofthese was optimized using Eq. (3) which was developed by

Bejan [18] for flow between parallel plates using theintersection of asymptotes between a fully developed flow

limit and a flat boundary layer flow limit.1

4

22.73

PD ' u optimumb

L P

L(3)

Fig. 2. Schematic of finned and finless heat sinks outlining

the overall dimensions and the optimised parameters, b and tfin

In determining the optimum fin spacing, the pressure drop

('P) across the bank of channels, was set equal to 6.5Pa sothat the fin spacing of the heat sink is optimised for a fan

speed of 8000RPM and should correspond to a total

volumetric flow rate of approximately 2.5E-04m3/s passing

through the heat exchanger. In order to achieve this Eq. (3)requires that the fin spacing be approximately 1.1mm. The

second parameter examined is the thickness of the fins, tfin.This was optimized by ensuring that the fin efficiency is

greater than 99% and Ellison [19] shows this to be the case

when Eq. (4) is satisfied.

2

fin heat sink 40tt hL k (4)

Forced convection heat transfer coefficients for gases are

typically in the range 25-250 W/m2.K [20]. In the currentstudy the expected maximum velocities, based on the fan

blade top speed, were approximately 0.5m/s, hence a value of60 W/m2K was chosen for hin Eq. (4). Using copper as the

heat sink material, Eq. (4) reveals that the fin thickness should

be at least 0.06mm. However the technique used to

manufacture the heat sink required this dimension to be oforder 0. 3mm, hence ensuring a more than adequate fin

thickness. The result is that a finned heat sink with a

maximum of 6 channels could be manufactured whilstmaintaining the heat sink footprint constraints that were

applied. The finless heat sink was manufactured to have the

identical specifications of profile, footprint area and externalwall thickness and is shown in Fig. 2 (b).

As previously mentioned in order to achieve the optimum

thermal performance of a fan and heat sink in parallel, flow

must enter the heat sink aligned with the heat sink walls and

fins. Fig. 3 is a schematic of a velocity vector diagram for a

backward curved fan with a rotational speed,Z and radial

499


4/13

velocity, Vr. The angle of the flow exiting the fan casing canbe measured using PIV and is dependant on system resistance.

A high system resistance such as that created by a finned heat

sink will result in a lower flow rate through the fan. This in

turn will result in reducing only the radial flow component andhence the angle at which fluid leaves the fan. It can be seen

that high system resistances will reduce J, whilst low systemresistances will result in increasing this angle.

In order to align the flow exiting the fan with the heat sinkwalls and fins a number of diffusers with angles varying

between 25o and 50o with both curved and straight walls weremanufactured from polycarbonate. As discussed above, it was

found that the optimum angle for flow alignment varied

between the finned and finless case. Each diffuser was tested

using PIV to measure the velocity of the flow in each of theheat sinks and it was determined from the measurements that a

25o curved diffuser and a 50o straight diffuser resulted in flowalignment within the finless and finned heat sinks respectively.

Fig. 3. Velocity triangle showing fan exit angle.

EXPERIMENTATIONThis section is structured as follows. Firstly the cooling

solution, encompassing the fan, diffuser and heat sink, is

described followed by a description of the experimentaltechnique employed in determining the thermal resistance.

Particle Image Velocimetry (PIV) was also used to visualizethe change in flow structures encountered by incorporating a

diffuser into the design; hence this technique is described in

the latter part of this section.

Experimental Configuration

The experimental configuration is shown in Fig. 4 consisting

of a finned and finless set-up. The results presented in this

paper are detailed in table 1. The current work presents results

from all four cases however, case 1 and 2 are detailed in a inEgan et al [15,16] and encompass tests performed without a

diffuser.

Fig. 4. Photograph of the experimental configurationemploying the straightening diffusers with the finned and

finless heat sinks.

TABLE 1. Experimental Test Description

Thermal Resistance Measurements

The experimental configuration consisted of a Micronel fan, a

flow diffuser and the optimized heat sink that was designed in

the preceding section. During experimentation the top cover ofthe heat sink, as shown in Fig. 4, was in position so as to

create a fully closed channel. This cooling solution has afootprint of 256mm2 for the fan, either 231mm2 or 278mm2 forthe heat sink and diffuser and a profile of only 5mm. The fan

was powered using a DC power supply and rotational speeds

ranging from 1000RPM to 8500RPM were measured using anoptical tachometer. The heat source used for experimentation

was a 66mm thin-film heater, available from Minco Inc., and

was attached to the base of the heat sinks. Three K-typethermocouples attached using a heat resistant tape to the centre

of the top, base and left wall of the heat sink. The ambient air

temperature was also recorded. The four thermocouples were

connected to a National Instruments 9211 USB High SpeedCarrier and temperatures were recorded and plotted using

LabVIEW 8. During experimentation, the entire heat sink wasfound to be almost isothermal, as a maximum temperature

difference between the hottest (base) and coldest (top) surface

of approximately 5% was measured. Also of significance, was

that the external surfaces of the heat sink were covered usinginsulating tape to reduce the heat loss due to secondary

cooling mechanisms such as natural convection and radiation.The purpose of this is to attempt to simulate the operational

environment for such a system in a typical portable electronic

device where the total set-up is likely to be confined within a

very small cavity.

Test Number Description1 Finless Heat, No diffuser

2 Finned Heat, No diffuser3 Finless Heat, 25o curved diffuser

4 Finned Heat, 50o straight diffuser

Diffusers

Top cover of

heat sink

500


5/13

The thermal performance of any cooling solution can becharacterized by its thermal resistance which is defined by

equation (5).

sT

Input

thT T

RQ

f (5)

QInput is the total power supplied to the thin-film heater, Ts is

the steady state temperature of the heat sink surface and T isthe temperature of the ambient air. However, it should be

noted that this power was not only dissipated by forcedconvection, but also by the secondary cooling mechanisms

already mentioned so that that total power supplied to the heat

sink is dissipated by:

Input FC LossesQ Q Q (6)

Hence, the total thermal resistance of the cooling solution can

also be defined by:

FC Losses

TFC Losses

th thth

th th

R R

R R R (7)

and the forced convection thermal resistance by

T Losses

FC

T Losses

th thth

th th

R RR

R R

(8)

It should be noted, however, that the power dissipated by these

secondary mechanisms is highly dependant on factors that

vary from application to application. Such factors include:

available air for free convection currents; temperature of

surrounding surfaces and any contact of conducting materialswith the heat sink surface. In order to remove this ambiguouspower dissipation from the characteristics reported herein, all

of the experiments were conducted whilst maintaining the heat

sink surface temperature constant at 70oC and ambient at

25oC. An experiment was then conducted with the fanswitched off (i.e. QFC = 0) and resulted in directly measuring

QLosses. Since the same temperatures were maintained

throughout all experiments, the heat lost through secondarymechanisms should remain constant throughout. This allowed

for the calculation of the heat dissipated by just forced

convection giving a more representative result of the forcedconvection cooling.

The total and forced convective heat transfer coefficients werecalculated using Eq. (9) and Eq. (10) respectively.

1

T

T

th conv

hR A

(9)

1

FC

FC

th conv

hR A

(10)

Where, h, is the heat transfer coefficient in W/m2 K and Aconvis the convective surface area of the finned heat sink in m 2.

The convective surface area of the finned and finless heat

sinks was calculated to be 0.000984m2 and 0.00044m2

respectively with an accuracy of 1E-06m2. The heat transfercoefficient is inversely proportional to the convective surface

area of the heat sink as shown in Eq. (9) and (10), this meansthat the finless heat sink with a lower Aconv will result in a

higher heat transfer coefficient.

Uncertainty analysisAn uncertainty analysis was performed, [21], giving the

thermal resistance an error of less than 5%. The uncertainty

was greatest for natural convection and decreased withincreasing fan speed.

The optical tachometer used to measure the rotational speed ofthe fan has a resolution of 1RPM. However, during

experimentation it was observed that the speed varied up to

50RPM.

Velocity Measurements

The second set of experiments involved detailed velocity

measurements of the flow within the channels of the heat sink.

Particle Image Velocimetry (PIV) was employed to obtainthese velocity measurements. However, in order to obtain the

optical access required for this technique, the copper top cover

of the heat sink was replaced with a glass slide. Hence all PIVresults were obtained without heat transfer. The following

paragraphs detail the principle of the technique and the set-up

used during experimentation.

PIV works on the principle of introducing seeding or tracer

particles to the flow and tracking their motion over known

time intervals. Ideally the particles should be small and theirdensity matching the working fluid density. Gravitational

forces will produce an error depending on the density

mismatch of the particles and if the particles are large theywill be unable to follow the exact motion of the fluid. A water-

based oil consisting of monopropalene glycols or CO2 fog wasused for the seeding which was produced using a JEM Stage

Hazer 2000. The size of seeding particles produced were of

approximately 10Pm diameter with relative density of1.05kg/m3 at 293K, [22]. In order to contain the seeding

particles the experimental apparatus was placed in a glasswalled enclosure of dimensions 400mm(L) 250mm(H)

250mm(W). Once the particles were released into the

enclosure a laser illuminated a plane in the heat sink and thelight subsequently emitted by the particles was recorded on a

camera positioned perpendicular to the illuminated plane.

Velocity measurements were taken for a plane at mid-depth inthe heat sink channels. A schematic of the experimental set-up

is shown in Fig. 5

501


6/13

Fig. 5. PIV Experimental Set-up showing camera and laser

positions with respect to fan and heat sink. The diffuser is

inserted between the fan and heat sink.

Raffel et al. [23] discuss the science and technical aspects of a

PIV system in full. Almost all PIV systems use a laser systemas lasers emit intense and monochromatic light, which after

being passed through a series of lenses is easily shaped into athin light sheet for illuminating the seeding particles. A

Nd:YAG, 532nm, semiconductor laser was used giving a

1.5mm thick light sheet. To measure the velocity of the

particles the laser was pulsed twice with a time delay chosenso that the distance travelled by the seeding particles between

images is discernable and that particles do not leave theilluminated plane within the period. The exposure time of the

camera was long enough to capture the position of the

particles and short enough to avoid blurring of the image. To

ensure this, TSI Inc [24] suggest that the time between pulses

be calculated from:

s in IR

max

*

4U

pixel

FOVN

RES(11)

where FOV is the field of view, RES is the resolution of thecamera in the direction of flow, Npixels in IR is the number of

pixels in an interrogation region and Vmax is the expected

maximum in-plane velocity. Prior to experimentation, in order

to calibrate the subsequent measurements, a scaling devicemust be inserted in the test section to obtain the correct scaling

factor for the velocity measurements. This was achieved byplacing a calibrated ruler in the measurement plane.

A PowerViewTM Plus 2MP CCD camera of resolution 1600

1200 with a 60mm focal length lens was used to record thescattered light from the illuminated particles and their images

appear as bright spots on a dark background. The images arerecorded on separate frames from the double pulsing of the

laser. This technique is called multi-frame/single exposure

PIV. A statistical PIV evaluation using Insight 3G software

supplied by TSI Incorporated was applied to the recorded data

in order to extract the required displacement information. Therecordings were divided into regions known as interrogation

areas and a correlation technique cross-correlation, was

applied to determine the displacement of the particles within

the interrogation area. The size of the interrogation area wasset according to the time between pulses so that the seeding

particles would not travel more than 25% of the interrogationarea length. If seeding particles travel much more than this

distance then too many seeding particles will have left the

interrogation area when the second image is taken. These

particles will not be available for velocity measurement andwill be a source of noise. This problem can be overcome by

overlapping interrogation areas but again care must be taken to

ensure that seeding particle displacements between pulses arenot too small. Interrogation area overlap was set to 50% of the

interrogation area height and width for the current results and

the interrogation area size was set at 32 32 pixels. Thevelocities were calculated from the displacements and the

duration of the illumination pulse. A series of post-processing

steps were applied to the PIV data.

RESULTS AND DISCUSSION

The objectives of this work are to provide optimized thermalmanagement solutions based on a previous analyses, [13-16],

using one of the smallest commercially available fans

combined with finned and finless heat sink geometries.Optimization of such solutions has been considered using

velocity field and thermal resistance measurements presented

in this section. Velocity measurements are considered first inorder to develop a suitable diffuser for each case based on

design criterions previously discussed. Thermal resistance

measurements, heat transfer and flow rate results proceed fromthe flow field analysis, quantifying the influence of the

introduction of diffusers on both heat sinks over a range of fan

rotational speeds.

Fig. 6 and Fig. 7 depict the flow at the entrance regions to the

heat sink for finned and finless geometries without the use of a

diffuser [15,16]. While Fig. 8 and Fig. 9 present the velocitymagnitude data for flow in the finless and finned heat sinks

with the inclusion of diffusers at the entrance to the heat sink.A range of diffusers were examined for both heat sink

geometries producing various heat sink fan orientations from

25o to 60o. Therefore due to the quantity of PIV results

analyzed, the data presented for this work is only consideredfor the diffusers which would optimize flow alignment and

distribution in both heat sink cases. It was found that a 25ocurved diffuser for the finless heat sink, and a 50o straight

diffuser for the finned heat sink were beneficial.

The velocity results presented in all figures are velocitymagnitudes with vector length and color varying accordingly;

longer vectors represent high velocity regions. Therelationship between color and velocity magnitude is indicated

in the legend provided with each figure. The data presented for

the current work was obtained by averaging 70 instantaneous

results, each taken at a frequency of 14.5Hz.

502


7/13

The PIV results for the finless heat sink and 25o curveddiffuser are shown in Fig. 8 for fan rotational speeds between

3000RPM and 8000RPM. Each velocity magnitude plot

shown in Fig. 8 presents the fluid flow through the diffuser

and heat sink, with the dashed line indicating the entrance tothe heat sink from the diffuser. Considering the data presented

in Fig. 8 and Fig. 6 for a fan speed of 3000RPM, it is possibleto distinguish noticeable differences in the flow distribution

due to the introduction of a diffuser. In Fig. 8 the flow is

distributed relatively evenly in the channel of the finless heat

sink for all three speeds tested. This implies that the effects offorced convection should also be relatively even across the

channel section, reducing the possibility of raised

temperatures in localized areas of the heat sink. However, inFig. 6, a large region of low velocity exists near the right wall

at the entrance. Hence forced convective cooling in this region

will have little effect on the right wall temperature. Byintroducing the diffuser at the entrance to the heat sink, flow

now enters the finless channel parallel to the heat sink walls as

seen from the velocity magnitude plot of Fig. 8.

The results from PIV analysis on the finned heat sink and 50o

straight diffuser are presented in Fig. 9. The diffuser and heatsink are labeled in Fig. 9(a) and also color contrasted to

distinguish the fins (black) from their shadows in the diffuser

(grey). These were unavoidably created by the laser beamused to illuminate the seeding particles within the heat sink.

Consequently particle tracking was not possible in these

regions of the diffuser. However, this does not affect thevalidity of the current work as flow structures are still visible

in the diffuser, and concentration on flow characteristics

within the heat sink is of primary interest for optimizing this

thermal management solution.

Fig. 6. PIV velocity magnitude plot, in m/s, from Egan et al

[16] of the flow in the entrance region of the finless heat sinkwith no diffuser attached for 3000RPM. All velocities are

normalised with respect to the maximum velocity of 4m/s.

For each case considered in Fig. 9, the flow entering the heat

sink shows minimal impingement on the fins as opposed to

Fig. 7, where no diffuser is used at 8000RPM. Under fullydeveloped conditions the maximum and minimum velocity

and hence flow rate is recorded in channel 3 and channel 6

respectively for all three fan speeds (see equations 12 and 13).The percentage difference in flow rates between channel 3 and

6 for fan speeds of 3000RPM, 5500RPM and 8000RPM is

42%, 48% and 43% respectively. However the standard

deviation of the flow rates and velocities in for all six channelsis higher for a fan speed of 8000RPM suggesting that the flow

distribution between channels is not as uniform at higherspeeds. It can be seen from Fig. 9(c) that the angle of the flow

entering the diffuser changes at 8000RPM and that a longer

diffuser but of the same angle may be required.

Fig. 7. PIV velocity magnitude plot, in m/s, from Egan et al.

[15] showing the flow field in the entrance region to the finned

heat sink at 8000RPM with no diffuser. All velocities are

normalised with respect to the maximum velocity of 2.7m/s.

Following the selection of a diffuser for the finned and finlessheat sinks using the flow field analyses discussed, the

experimentation was arranged to obtain a measure of the total

thermal resistance for a range of fan speeds from 0RPM to

8000RPM. The data recorded are plotted in Fig. 10. Thisfigure, along with all subsequent figures discussed, also

compares data recorded for the finned and finless cases whereno diffuser was used [15,16]. From Fig. 10 it can be seen that

a reduction in thermal resistance can be obtained through the

use of a diffuser. For the finned heat sink at low velocities, the

measured thermal resistances are of similar magnitude forboth set-ups. However, as rotational speed of the fan

increases, the heat sink and diffuser combination clearly outperforms the solitary heat sink case. For a finless heat sink

there is a constant reduction in total thermal resistance

achieved when using a diffuser between approximately

4000RPM and 8000RPM. Below this range of speeds, large

reductions in thermal resistance are seen by using a diffuser toalign the flow exiting the fan. In order to achieve a thermalresistance of lower than 40oC/W using a finned heat sink anddiffuser combination, the fan must operate at a rotational

speed greater than 5000RPM. Alternatively, the use of a

finless heat sink and diffuser combination produces similarthermal resistances from a fan speed of just 3000RPM. A

benefit of this is that lower operating speeds are sufficient for

similar cooling effects when using the optimized finlessarrangement. In terms of fan reliability, noise and pumping

power this proves highly advantageous.

503


8/13

(a) Finless Heat Sink at 3000RPM

(b) Finless Heat Sink at 5500RPM

(c) Finless Heat Sink at 8000RPM

Fig. 8. PIV velocity magnitude plots depicting the flow within the diffuser and finless heat sink for fan speeds of (a)3000RPM (b) 5500RPM and (c) 8000RPM respectively. The respective velocity scale bars in m/s are shown above each image.

The dashes lines represent the entrance to the heat sink from the diffuser.

504


9/13

(a) Finned Heat Sink 3000RPM

1

2

3

45

6

Diffuser Heat Sink

(b) Finned Heat Sink 5500RPM

(c) Finned Heat Sink 8000RPM

Fig. 9. PIV velocity magnitude plots of flow in finned heat sink and 50o straight diffuser at 3000RPM, 5500RPM and8000RPM. Respective scale bars for velocities in m/s are given above each result. Flow is from left to right. Regions in gray are

shadows from the heat sink fins which result from the laser light entering from the right.

505


10/13

A plot of the forced convection thermal resistance comparingall cases with and without diffusers is presented in Fig. 11.

This plots only the forced convection cooling achieved from

the fan and not the heat dissipated by secondary cooling

mechanisms. For the current work these secondary coolingmechanisms were found to give a thermal resistance, Rth-Losses,

of 63oC/W. Hence, by using Eq. (9) and the results for Rth-Tpresented in Fig. 10, a corresponding forced convection plot of

the forced convection thermal resistance, Rth-FC was obtained

and is presented in Fig. 11 . For a fan speed of 8000RPM,

using the diffuser designed for the finned heat sink, a decreaseof approximately 23% in forced convection thermal resistance

is obtained. At the same operating speed, the finless heat sink

and combined diffuser provide a decrease of 15.3% over afinless heat sink without diffuser. It is therefore apparent that

in the case of a fan and heat sink in parallel aligning the flow

with the heat sink fins produces lower thermal resistances thanimpinging flows [13].

At 8000RPM the finned diffuser case results in a lower

thermal resistance compared to the finless diffuser solution.

This is an expected result as the design of the finned heat sinkwas optimized for 8000RPM, based on boundary layeranalyses discussed in the preceding section. As previously

mentioned, a difference in velocity magnitudes for both

geometries can be seen in Fig. 8 and Fig. 9. This can be

accredited to a higher system resistance due to the finned heatsink, which implies that higher mass flow rates are evident in

the finless heat sink for each rotational speed recorded. Highermass flow rates commonly result in increased heat transfer

rates; however Fig. 10 and Fig. 11 enforce the influence of

boundary layer analysis used in the finned heat sink design,

i.e. the optimum spacing of fins in a heat sink for a givenpressure drop is determined based on the boundary layers

merging just at the exit of the heat sink. Thermal resistance

measurements for the finned case drop below that of thefinless case at approximately 7500RPM. It is also apparent

that the boundary layer optimization criterion is proven for

both situations with and without the use of a diffuser for lowprofile solutions. However, the thermal resistance values for

the finless-diffuser case are lower compared to finned case.

(test 1). Even at 8000RPM where the forced convectivethermal resistance appears to be reaching a constant value with

increased rotor speed, there is a reduction of 11.3% over the

forced convective thermal resistance of the finned heat sinkwithout a diffuser. This is significant considering the 25o

diffuser used for the finless heat sink only increases the

overall footprint area of the thermal solution by approximately

6%. It was also noted that the 50o

diffuser used to optimize thefinned heat sink increases the overall footprint area of the

thermal solution by approximately 16%. This resultdemonstrates the need to design integrated and optimized fan

and heat sink solutions for the low profile market.

0

10

20

30

40

50

60

70

0 1000 2000 3000 4000 5000 6000 7000 8000 9000

Rotor Speed (RPM)

Rth(oC/W)

Finned + diffuser

Finless + diffuser

Finned

Finless

Fig. 10. Plot showing total thermal resistance for fan speedranges between 0RPM and 8300RPM.

0

25

50

75

100

125

150

175

200

225

250

275

300

3000 3500 4000 4500 5000 5500 6000 6500 7000 7500 8000 8500 9000Rotor Speed (RPM)

RthFC(oC/W)

Finned + Diffuser

Finless + Diffuser

Finned

Finless

Losses

Fig. 11. Plot of forced convection thermal resistance withvarying RPM. Also indicated are the effects due to secondary

cooling mechanisms (Losses).

Similarly to a combined fan and heat sink cooling solution,sizing issues may also be associated with the power supply

necessary for operation. Considering the power requirements

of a cooling solution, it is reasonable to measure shaft powersupplied to the rotor when characterizing fan performance.

This provides designers with power requirements for the fan

alone which, may deviate considerably from the motor powerrequired especially when considering small scale fans. The

current work is not examined in this manner, as there is an

enclosed motor and fan assembly designed for combined use.Power requirements are therefore presented for the motor and

fan collectively. The manufacturer provides a nominal

operating speed of 6000RPM. To achieve this rotationalspeed, a power input of 0.163W is required. This value

506


11/13

increases to 0.320W to reach an operating speed of 8000RPM.The scale law for power consumption predicts a 1.77 factor

[(8000/6000)3] increase however, the power is seen to almost

double. This is due to the efficiency of the driving motor

dropping off at higher speeds as bearing losses become larger.Power is therefore conserved to a greater extent when

operating the cooling solution below the nominal operatingspeed of 6000RPM as opposed to above this speed where

power requirements rise sharply for small increases in

rotational speed. Hence, depending on the level of restrictions

evident with power supply choice and the thermalperformance requirements, it may be beneficial to integrate a

finless heat sink and diffuser design as it outperforms the

finned heat sink cases below 6000RPM, shown in Fig. 10 andFig. 11. It is also of interest to note that this level of power

consumption could be supplied by a high end mobile phone

battery (Nokia BP-4L Battery) for a period of 34hrs ofcontinuous use. In reality however, the fan would only be in

use while the device is in high processing mode so the actual

lifetime is likely to be limited by phone functionality ratherthan by cooling solution.

In Fig. 12 and Fig. 13, the heat transfer coefficient is presentedfor combined cooling and forced convection cooling. Similar

to the thermal resistance measurements, this data is based on

the average heat transfer coefficient of the heat sink. The data

presented in Fig. 12 is that of the total heat transfer coefficientand it can be seen that the finless cases achieve more than

twice the average heat transfer coefficient compared to thefinned cases over the rotor speeds tested. This is highly

dependant on the convective surface area of both heat sinks as

given in Eq. (9). With reference to Fig. 10 and Fig. 11, which

show thermal resistance plotted against fan speed, it ispossible to conclude that in general over the range of speeds

examined, the finless design will dissipate a similar, and in

some cases favorable magnitude of heat per unit oftemperature compared to the finned design. The heat transfer

coefficient quantifies this relative to the convective surface

area of the heat sink. The finless heat sink considered here hasa convective surface area 44.7% less than the convective

surface area of the finned heat sink. It is therefore expected

that the heat transfer coefficient values will be greater for thefinless design as it dissipates a similar magnitude of heat using

almost half the convective surface area of the finned design.

This also applies to Fig. 13 for the forced convection heattransfer coefficient data. Increasing rotational speed results in

an increasing difference in heat transfer coefficient between

heat sinks using diffusers and heat sinks without diffusers. At

8000RPM, the finned heat sink achieves an increase ofapproximately 27% through the introduction of the selected

diffuser. For the same rotor speed, an increase ofapproximately 15% is achieved for the finless heat sink using

the diffuser to align the flow. The effect of convective surface

area results in a heat transfer coefficient for the finned case

with diffuser to be 51% of the heat transfer coefficient for thefinless case with diffuser, considered for a rotor speed of

8000RPM.

0

10

20

30

40

50

60

70

80

90

0 1000 2000 3000 4000 5000 6000 7000 8000 9000

Rotor Speed (RPM)

h(W/m

2oC)

Finned + diffuser

Finless + diffuser

Finned

Finless

Fig. 12. The total heat transfer coefficient for both finned andfinless heat sinks.

0

5

10

15

20

25

30

35

40

45

50

0 1000 2000 3000 4000 5000 6000 7000 8000 9000

Rotor Speed (RPM)

hFC(W/m

2oC)

Finned + diffuser

Finless + diffuser

Finned

Finless

Fig. 13. The forced convection heat transfer coefficient for

both finned and finless heat sinks.

As seen in Fig. 9 the flow within the finned heat sink channels

is seen to fully develop which makes it possible to describe the

velocity profile across the entire channel cross section usingthe solution provided by Purday [25]. This solution describes

the velocity distribution in rectangular channels of varyingaspect ratios and is adapted in Eq. (12) to describe the flow in

channels with an aspect ratio of , such as those in the heat

sink analysed herein. In this equation X and Y correspond

to half the channel width and half the profile heightrespectively. While x and y correspond to the local

positions measured from the mid-width and mid-depth of thechannel respectively.

507


12/13

2 3

( , ) max 1 1x yx y

u UX Y

.7

(12)

Hence, the volumetric flow rate in each channel can be

calculated from Eq. (13).

( , ) max2 4.743 3.7

Y X

x y

Y X

Q u dydx XY U

(13)

The maximum velocity measured using PIV within each of theheat sink channels was substituted into Eq. (13) which yielded

the corresponding flow rates in each channel. These were

summed for all 6 channels to obtain the total volumetric flowrate passing through the system. The results are presented in

Fig. 14 and compared to volumetric flow rates for the tests

carried out without a diffuser [15]. It is seen that the flow ratein the finned heat sink with the diffuser attached at the

entrance is on average 15% higher that just the fan and heat

sink arrangement. However the increase in volumetric flow

rate at 8000RPM is lower compared to 3000RPM and5500RPM. A possible explanation for this is a higher variation

in flow rate between the heat sink channels at 8000RPM dueto higher speeds affecting the flow distribution. The accuracy

of the velocity measurements are due to a number of

experimental factors and are estimated at 5%.

0.E+00

1.E-05

2.E-05

3.E-05

4.E-05

5.E-05

6.E-05

7.E-05

8.E-05

9.E-05

0 1500 3000 4500 6000 7500 9000

Rotational Speed (RPM)

TotalVolumetricFlowR

ate(m3/s)

Heat Sink +Diffuser

Heat Sink

Fig. 14. Graph depicting the volumetric flow rates through

heat sink measured using Eq. (12) along with the PIV data

presented in Fig. 9 and [13] over the three fan speed tested

CONCLUSIONS

A miniature low profile cooling solution suitable for

integration into a range of portable electronics devices has

been characterized through thermal resistance and velocityfield measurements.

This paper presents an analysis on fan exit angle with resultsshowing that a reduction in thermal resistances are obtained

for flow entering the heat sink aligned with the fins. Flow

alignment has been achieved through the introduction of a

diffuser at the entrance to the heat sink. By accounting for fan exit angles, the thermal

resistance has been reduced by up to 23% for afinned heat sink and 15% for a finless heat sink. In

order to achieve these reductions the footprint area

has been increased by 16% and 6% respectively.

A finless heat sink and diffuser outperforms theconventional finned heat sink solution over a range

of fan speed tested.

The finned heat sink was designed using anoptimization criterion based on the fin spacing which

results in thermal boundary layers merging at the exit of

the heat sink. This was applied for a fan speed of8000RPM.

Results presented show that this optimization

criterion was confirmed for miniature scalesolutions as the finned heat sink outperforms the

finless case at 8000RPM

Finally the results also provide a basis for designers of heatsto analyze fan exit angles and a move towards integrated fan

and heat sink solutions.

REFERENCES

[1] Vogel, M.R., Liquid Cooling Performance for 3-D

Multichip Module and Miniature Heat Sink, IEEE

Transactions on components, packaging and

manufacturing technology, Part A, Vol. 18 (1) (1995),68-73.

[2] Launay, A., 2004, Experimental Study on Silicon

Micro-Heat Pipe Arrays, J. Applied ThermalEngineering, Vol. 24 (2-3), 233-243.

[3] Tan, F.L., 2004, Cooling of mobile Electronic Devices

using Phase Change Materials, J. Applied ThermalEngineering, Vol. 24 (2-3), 159-169.

[4] Walsh, E.J, Grimes, R. and Punch, J. 2007, "A coolingdevice", WO2007057871

[5] Grimes, R., Walsh, E.J., Quin, D., and Davies, M.R.D.,

2005, Effect of Geometric Scaling on Aerodynamic

Performance, Journal of AIAA, Vol. 43 (11), 2293-2298.

[6] R. Grimes, Walsh, P., Walsh, E., and Egan, V., 2007,The Effects of Diameter and Rotational Speed on the

Aerodynamic Performance of Low Profile Miniature

Radial Flow Fans, Puebla, Mexico, ICNMM2007-

30185.[7] K. Hanly, Grimes, R., Walsh, E., Rodgers, B., and

Punch, J., 2003, The Effect of Reynolds Number on theAerodynamic Performance of Micro Radial Flow Fans,

Proceedings of ASME Heat Transfer Conference, San

Francisco, California, USA, HT2005-72514.

[8] P. Walsh, V. Egan, R. Grimes, E.J. Walsh, 2007,Scaling of Flow Characteristics and Power

Consumption with Profile Height for Miniature

508


13/13

Centrifugal Fans, Proceedings of the 5th InternationalConference on Nanochannels, Microchannels and

Minichannels, ICNMM2007, Puebla, Mexico, 18th

20th June, ICNMM2007-30136.

[9] Grimes, R., Walsh, E.J., Walsh, P. and V. Egan, 2007,The effects of diameter and rotational speed on the

aerodynamic performance of low profile radialminiature flow fans, Proceedings of the 5th

International Conference on Nanochannels,

Microchannels and Minichannels, ICNMM2007,

Puebla, Mexico, 18th 20th June.[10] S.W. Day, Lemire, P.P., Flack, R.D., and McDaniel,

J.C., 2003, Effect of Reynolds Number on Performance

of a small Centrifugal Pump, Proceedings of the 4thASME_JSME Joint Fluids Engineering Conference,

Honolulu, Hawaii, USA, FEDSM2003-45686.

[11] Chien, H., 2001, The Study of Micro-Fin Heat Sinksfor Electronic Cooling Applications, 17th IEEE SEMI-

THERM Symposium.

[12] J.G. Maveety and Jung, H.H., 2002 Heat Transferfrom Square Pin-Fin Heat Sinks using Air

Impingement Cooling, IEEE Transactions on

components, packaging and manufacturing technologyVol. 25 (3), pp. 459-469.

[13] Walsh, E., Walsh, P., Grimes, R., Egan, V., 2007,

Thermal Management of Low Profile Applications,Proceedings of the 5th International Conference on

Nanochannels, Microchannels and Minichannels, 18th

20th June, Peubla, Mexico.

[14] Walsh, E.J., 2006, Low profile thermal management,

EPEC Seminar on High Temperature Electronics and

Thermal Management, 9th 10th November, sud-west

park conference center, Nuremburg, Germany.

[15] Egan, V., Walsh, P., Walsh, E., Grimes, R., 2007, Onthe Characterisation of Finned and Finless Heat Sinks

for Portable Electronics, 5th International Conferenceon Nanochannels, Microchannels and Minichannels,

18th 20th June, Peubla, Mexico.

[16] Egan V., Walsh P., Walsh E.J., Grimes R., 2007,

Characterisation of a Miniature Low Profile CoolingSolution for Portable Electronics, Sixth International

Conference on Enhanced, Compact and Ultra-Compact

Heat Exchangers: Science, Engineering andTechnology, Potsdam, Germany, 16th 21st

September.

[17] F.P. Bleier, 1997, Fan Handbook: Selection,Application and Design, McGraw-Hill.

[18] Bejan, A., Convective Heat Transfer, John Wiley &Sons, Inc. (2004).

[19] G.N. Ellison, 1989, Thermal Computations forElectronic Equipment, Krieger Publishing Company

[20] Incropera F.P., DeWitt D.P., Fundamentals of Heat and

Mass Transfer, John Wiley & Sond, Inc (1996)

[21] Moffat J.R., 1999. Uncertainty analysis. ElectronicsCooling. 5(2).

[22] JEM, Jem "Safety Data Sheet" Pro Haze Fluid. SDS

No. 112., (2005).

[23] M. Raffel, 1998, Particle Image Velocimetry: Apractical guide, Springer.

[24] TSI Inc., TSI Insight 3G Software Manual, (2006).

[25] Purday, H. F. P. 1949, An Introduction to the

Mechanics of Viscous Flow (Streamline Flow) DoverPublications Inc., New York.

miniature heat sinks

Documents