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Proceedings of NHTC'00:
National Heat Transfer Conference
Pittsburgh, Pennsylvania, August 20-22, 2000
3 4 th
NHTC
2000 - 12104
THE EFFECTS OF FLOW RATE AND SODIUM LAUREL SULFATE SURFACTANT
CONCENTRATION ON SINGLE- AND TWO-PHASE HEAT TRANSFER
CHARACTERISTICS IN A MICROCHANNEL
Michael S June
Graduate Student
Kate Gleason College of Engineering,
Rochester Institute of Technology,
Rochester, NY 14623
ABSTRACT
The processing speed of modem microchips and the
density of components at the card level has been increasing at a
steady rate. With these increases, come an increase in heat
generation and a need to dissipate very high heat fluxes. Heat
sinks that employ very narrow channels, on the order of 58 600 m wide have been shown to dissipate heat fluxes as high
as 10,000 kW/m 2 (Tuckerman, 1984) while maintaining chip
temperatures below 130 C. The current study investigates the
heat transfer characteristics of single and two-phase flows in a
200 m wide channel. Water was used as a working fluid at
three flow rates and at three surfactant concentrations. The
results showed a strong dependence on flow rate in the single
phase region, and up to 9120 kW/m 2 was dissipated while
maintaining a surface temperature of 115 C. The data also
suggest a reduction in heat transfer efficiency with the addition
of the surfactant sodium laurel sulfate (0.01%) in the single
phase. There was a delay in the onset of nucleate boiling,
however as expected for the same surfactant concentration. In
the two-phase region, there was no significant change in the
heat transfer characteristics with the addition of surfactants.
INTRODUCTION
Driven by the needs of the electronics industry,
investigations have been underway to find more efficient ways
to transfer heat away from electronic components. This has
been of particular interest in the electronics packaging industry.
Silicon chips must be packaged prior to being installed on a
circuit board. The package provides protection, electrical
connection to the circuit board, and heat dissipation. Both the
reliability of the chip and the reliability of the package depend
on temperature. The processing speed of the chip can also be
Satish Kandlikar*
Professor and Head
Mechanical Engineering Department,
Kate Gleason College of Engineering,
Rochester Institute of Technology
Rochester, NY 14623
*-corresponding author
affected by high temperature. Chips operate best below certain
prescribed temperatures, usually around 130 degrees Celsius.
As the processing speed of a chip increases, so does the
generation of heat. Due to the high heat fluxes generated by
modem chips and the ever increasing density of electronic
components on circuit boards it was estimated in the early
1980s that there would be a need to dissipate as much as 100
W/cm 2 by the end of the decade Prior to 1981, the highest
cooling rates were achieved by specialized cooling units made
by IBM and Honeywell. These devices could dissipate only 20
W/cm 2, which was considered as the practical upper limit to
heat dissipation in electronics packaging.
In their pioneering works, Tuckerman and Pease (1982,
1984) showed that enhanced heat transfer performance could
be achieved using heat exchangers incorporating very small
channels for the working fluid to flow. They performed
optimizations that showed 58 m (0.002 in) to be the ideal
channel width for heat transfer with water as a working fluid.
Channel width was optimized
by minimizing hydraulic
diameter ( D h ) in order to minimize convective thermal
resistance. The lower bound of 58 m was determined based
on the viscosity of the working fluid, water, and a channel
length of lcm. With these dimensions, a system pressure of 50
psi would need to be maintained for laminar flow. These
parameters were chosen to achieve the goal of compactness,
and to keep system pressures comparable to contemporary
cooling systems. Other researchers use the term microchannel
to refer to channels as large as 600 m in width. They have
been fabricated from various materials such as silicon,
aluminum, and indium phosphide, all yielding similar heat
transfer characteristics.
The heat transfer performance of microchannels is
significantly higher than conventional sized channels, but
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Copyright © 2000 by ASME
present drawbacks as well. They are difficult to fabricate and
clog easily. For these reasons, the trend has been toward larger
channels, which are still small enough to maintain the high
performance of microchannels. These are sometimes referred to
as millichannels.
Turbulent Flow in Microchannels
In general, for internal flows through conventional sized
pipes and ducts, flows are considered to be laminar when the
Reynolds number is less than 2300. In microchannels it has
been observed that fully turbulent flow occurs at Reynolds
numbers between 1000 to 1500 (Peng and Wang, 1993, Wang
and Peng, 1994, Sivagnanam, Balakrishnan, and Varma, 1994).
Effect of Surfactant
The addition of surfactants to the working fluid has been
found to slightly increase the heat transfer coefficient and delay
the onset of nucleate boiling. Kandlikar (1999) presents a table
summarizing results of studies on the effects of various
surfactants on heat transfer. While most studies have found a
slight increase in heat transfer with surfactants such as ethylene
glycol, there is one study by Wang and Harnett (1994) that
found a slight decrease in heat transfer with the addition of
sodium laurel sulfate.
It was also stated by Dhir (1999), that as surface wetability
increases, the boiling curve tends to shift to the right.
Increasing wetability is an effect of surfactants.
Objective
The objective of the current work was to construct an
apparatus which could be built from easily obtainable
components, and would allow the study of heat transfer
characteristics of water in a microchannel. In addition, the
apparatus needed to allow for visualization of the flow.
The effect of flow rate, and surfactant concentration were
studied while using a microscope to view the flow and observe
the presence and nature of bubble formation.
NOMENCLATURE
A
Cross sectional area of a resistor
Ac
Channel cross sectional area
B
Bias
CHF
Critical heat flux
Cp
Specific heat
D
Diameter
Hydraulic diameter
Dh
h
Heat transfer coefficient in the microchannel
k
ONB
P
Average heat transfer coefficient in the microchannel
Thermal Conductivity
Onset of nucleate boiling
Perimeter
P
Pr
q"
R
R"
R2
Re
1%
SLS
To
Tlmtd
To
Tsat
Tw
U
O~
/.t/
/.iv
P
Precision
Prandtls number
Heat flux
Resistance
Total thermal resistance
Curve fit
Reynolds number
Initial resistance
Sodium laurel sulfate
Initial temperature
Log mean temperature difference
Outlet temperature
Saturation temperature
Heating element (wall) temperature
Total uncertainty
Temperature coefficient of resistivity
Viscosity of liquid water
Viscosity of water vapor
Density
PROCEDURE
The apparatus was intentionally kept very simple.
Available literature contains descriptions of many microchannel
fabrication techniques such as chemical etching and microsawing, all requiring larger budgets and possibly outside
vendors. The objective was to build a simple inexpensive
apparatus that could be built from easily obtainable
components and that could be built in an average shop with no
exotic tooling. In addition, the apparatus had to be easily
disassembled for reasonably efficient repair in case of heating
element burn out. The apparatus designed for this experiment
meets these needs.
Apparatus
The base and test section were built primarily of glass and
acrylic. An acrylic base was bonded to a piece of ordinary 6.35
mm (1/4 inch) thick glass, which had holes to allow the water
to pass through at the inlet and the outlet of the channel (Figure
1). The glass also had holes for the bolts which applied
pressure to hold the various layers together. Larger holes were
bored into the acrylic, concentric to the water inlet and outlet
holes in the glass. Brass inserts were machined and bonded into
these holes to serve as attachment points for the plumbing
fittings and as electrical conductors to the heating element. The
heating element (Omega cat. no. SPPL-008, 0.008 in diameter
pure Platinum wire) was threaded through the brass inserts so
that it ran along the surface of the glass between the holes, and
was secured inside the brass inserts with set screws.
Microscope slides (VWR Scientific cat. no. 48300-025, 25
X 75 X t mm) were placed tightly against both sides of the
heating element to form the channel. Another ~Ain thick piece
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Copyright © 2000 by ASME
of glass matching the perimeter of the microscope slides was
placed on top forming the top of the channel and providing a
window into the channel for observation.
The channel dimensions were such that any attempts to
place adhesives between the glass layers resulted in excess
adhesive moving into the channel by capillary action. This
immediately clogged the channel rendering it useless. Pressure,
therefore was necessary to hold the layers together and to
insure water would flow in the channel and not between the
layers. This was accomplished by milling a pocket and a
window in another piece of acrylic to hold the topmost piece of
glass, and bolting this structure to the base. Plumber's putty
was placed under the holding fixture and outside the perimeter
of the microscope slides to plug the ends of the channels. To
ensure a good seal, silicon bathtub caulk was applied along the
perimeter of the holding fixture between the holding fixture
and the base.
The water flowed up through the brass fitting in the base
where it made a 90-degree turn into the channel. Here it ran
over the heating element in the channel making another 90degree tum at the end and exiting through another brass fitting
in the base. Electrical current was supplied to the heating
element also through the brass fittings (Figure 1). Drawings of
the apparatus are shown in Figures 1 and 2. A schematic
diagram of the flow path is shown in Figure 2A.
6.35 nn (1/4 in) Glass
Plunbee's
75 x 25 X I mM Microscope
~
\
II
J •
~--Microchan.el
o,eo3e nn (.OOB ~n) Di~ Neoting Etement
I
635 ~
( 1 / 4 i~>
6LQSS
Plunber's
Putty
Holding Fixture
Sibcon ou~w
II
1'
II
25. . . . . ~ . . . . . . . . . .
~oss
i
in
Figure 2 End view of the apparatus with detail of channel
cross section
A t m o s pto
Outlet
heric
Pressure
glides
Hoidimg F i x t u r e
.....
Mr
II
~
3 r 3~
d L
/
Z5.4 nm (l in) AcryLic
Figure 2A Schematic diagram of flow path
Water
Weter
Figure 1 Side view of the experimental apparatus
Test Section
The cross sectional dimensions of the test section were
determined by the diameter of the heating element and the
thickness of the microscope slides. Pure platinum thermocouple
wire was chosen as the heating element because of its ability to
withstand high temperatures and because it's readily available
in a variety of sizes. In preliminary experiments, 0.002 inch
wire was used as the heating element, but it proved too difficult
to work with, and the channels frequently clogged. The wire
used in the final apparatus was .008 inches in diameter, which
resulted in a channel width of 203.2 p.m ( 0.008 inches) which
is roughly 3.5 times wider than Tuckerman's optimized
channels (Tuckerman, 1984).
Since the cross section of the heating element is round and
makes up the bottom of the channel, the channel is not
rectangular. This is not expected to affect the results
significantly, as only one side of the channel is affected. It was
also assumed that the enhancement in heat transfer
characteristics due to the size of the channel would dominate
any due to this departure from rectangularity. The cross section
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Copyright © 2000 by ASME
of the channel is shown in detail in Figure 3. The cross
sectional area of the channel was calculated by considering a
rectangle of the same width as the channel, from the top of the
channel down to the point of contact tangent to the heating
element. One half of the cross sectional area of the heating
element was then subtracted from the area of the
aforementioned rectangle.
C r o s s S e c t i o n a l A r e a : O.1489mm^2
P e r i m e t e r : 2147mm, Dh: 0277mm
/
/
drilled near the bottom for an outlet and another was drilled at
the top for an air inlet. Brass fittings were threaded into the
holes on the outside for the water outlet and the air inlet.
Another fitting was threaded to the same hole as the air inlet,
but on the inside of the vessel. This provided an attachment for
an ordinary water balloon. The air was delivered from a
regulator to the balloon, which forced the water out of the
vessel without introducing air into the water. This allowed the
amount of dissolved air in the water to be minimized. The
pump is shown in Figure 4.
/-
~-
02032nm
/
Air Inlet
/
Figure 3 Cut away cross section of channel with dimensions
Water
It was assumed that the edges of the microscope slides
were in close contact with the heating element. Therefore, the
width of the channel was the same as the diameter of the
heating element. It was assumed that there was water
undemeath the heating element, but it was also assumed that
the water underneath was not flowing with a significant
velocity. Since the water underneath the heating element was
not flowing, it reached an equilibrium condition and did not
contribute significantly to heat transfer from the heating
element. It was further assumed that the lower half of the
heating element was well insulated by the glass base and
channel sides. It was therefore assumed that heat flux to the
working fluid occurred only through the upper half of the
surface area of the heating element.
Heat loss through the apparatus was minimized by the
insulating properties of glass and acrylic. Extra fiberglass
insulation was placed around the apparatus to further minimize
heat loss, especially around the brass fittings. This was verified
by flowing heated water through the test section and
determining the energy loss from the temperature drop across
the apparatus. Energy loss was negligible at less than onepercent throughout the power range.
Pump
A pressure vessel was fabricated from a piece of ¾
inch thick walled tubing, 6 inches in diameter and 8 inches
long. Steel end caps were fabricated and bolted to each end
with rubber gaskets between the mating surfaces. A hole was
Outlet
Figure 4 Air pressure-driven pump
The heating element temperature could not be measured
with a thermocouple as it would become an efficient path for
heat loss. Instead, the resistance of the element was measured
and from this, the average temperature over the length of the
element was determined
Water was pumped using a peristaltic pump from a
constant temperature bath through the apparatus. To achieve
better accuracy during the calibration, current was supplied
using a Keithley model 224 Current Supply. The Hewlett
Packard power supplies used in the experiment were unable to
supply a steady current in the milliamp range needed for
calibration. Calibration was performed using a 75 milliamp
current with the corresponding voltage drop across the heating
element measured with the a Keithley model 199 System
DMM/Scanner with 6 '/2 digit readout. From this data,
resistance was determined at the temperature of the water
pumped through the apparatus. Temperatures were chosen to
correspond to the expected temperature range of the heating
element during the experiment. The calibration plot is shown in
Figure 5.
4
Copyright © 2000 by ASME
The equation the best fit line through the calibration data was
determined to be
R = 0.00025(T - To) + 0.06867
(RZ=0.995)
(1)
where To = 27.2 o C.
From Equation 1, a can be determined to be
0.003640 which is a 7.3% difference from the value given in
the CRC Handbook of Chemistry and Physics (1986) of
0.003927 (over the range of temperatures from 0-100 C.).
0.08
..... y: ......
0.078
"
0.076
E
~e 0.074 •
~ 0.072 •
0.07
0.068
0
5
10
15
20
25
30
35
40
45
T-To (C)
Figure 5 Calibration curve
The working fluid in these experiments was either deionized water, or de-ionized water mixed with the surfactant,
sodium laurel sulfate (SLS, CAS number 151-21-3). The
working fluids were degassed using a vacuum pump and were
left under negative pressure until out-gassing ceased. For
solutions with 0.05% SLS and 0.10% SLS, degassing was not
effective as emulsions were formed which made it impossible
to tell when out-gassing had ended. Consequently, an
experiment was performed using working fluids that were not
degassed, with de-ionized water and de-ionized water with
0.01% SLS. This ensured that the level of dissolved gasses in
the two working fluids were comparable and therefore did not
affect the comparison.
The solutions of 0.05% and 0.10% SLS were made by
dissolving 0.500g and 1.000g respectively of solid sodium
laurel sulfate in room temperature de-ionized water. The final
volume of each mixture was 1.0L therefore the solution
concentrations are given as percent by weight. The 0.01%
solution was prepared by dissolving 10.0g of SLS into 100ml
of de-ionized water, and diluting the subsequent mixture to a
final concentration of 0.01% by weight.
All working solutions were filtered through a 5/Jz., Gelman
Scientific in-line filter prior to reaching the test section.
Prior to recording data, the system was allowed to reach a
steady. Steady state was considered to have been reached when
there was not more than a 0.2 C change in temperature over a
1-minute interval. In most cases steady state was reached
within seconds and lasted until an air pressure variation in the
pump was encountered. Since the results were sensitive to flow
rate, any fluctuation in flow rate caused departure from steady
state.
During certain phases of boiling, temperature variations
were not due to changes in pump air pressure fluctuations, and
could not be kept within 0.2 C. Steady state in these cases was
considered to be the point at which the output temperature
stopped showing an upward trend and fluctuated about a steady
mean.
Once steady state was established, readings of voltage and
temperatures were taken approximately every 15 seconds, for a
total of 10 readings for every power setting.
The quantities read were,
•
Voltage across the heating element.
•
Voltage across the shunt resistor (for determining current
through the heating element).
•
Temperature of the working fluid at the inlet of the
channel.
•
Mixed mean temperature of the working fluid at the exit of
the channel.
•
Volume of water leaving the apparatus over a 1-minute
time interval.
In addition, the following dimensions were measured using
micrometers and scales where appropriate. Ten measurements
of each dimension were averaged to obtain the final results.
•
Length of the heating element.
•
Length of the channel.
•
Diameter of the heating element. (width of the channel)
•
Thickness of the microscope slides (height of the channel)
RESULTS
Effect of Flow Rate on Single Phase Data
For single phase heat transfer, heat fluxes as high as
5000kW/m 2 were achieved while keeping the heating element
temperature under 100 C. Figure 6 shows the heating element
temperature as a function of heat flux for three flow rates. The
three flows are all turbulent and in the single-phase region.
Clearly, as flow rate is increased, higher heat fluxes can be
dissipated.
Figure 7 shows the average heat transfer coefficient (~)
for the same three turbulent flows in Figure 7. All flows show a
nearly constant value for ~, as expected, and ~-increases with
increasing flow rate.
5
Copyright © 2000 by ASME
200
9000
0~180
8000
~ 160
• 0.46 mils (Re = 1200)
• 0.83 mils (Re = 2000)
7000
~14o
• 1.0 mils (Re = 2600)
• 0.46 mils (Re = 1200)
110.83 mils (Re = 2000)
• 1.0 mils (Re = 2600)
~---6000-
12o
!
=
~ 50
~ 90
~ 40
<
eee•ee
•
•
•
2000
3000
4000
Heat Flux (kWlm z)
5000
00
6000
40
60
80
100
120
Figure 8Boiling curve including single phase data for three
flow rates
N~'140
E
16o.
~120
¢
• 0.46 ml/s (Re = 1200)
• 0.46 mils (Re = 1200)
• 0.83 ml/s (Re = 2000)
• 1.O mils (Re = 2600)
J
100
• 0.83 mils (Re = 2000)
.y120
8 90
• 1.0 m l / s (Re = 2600)
o
50
m
o
<J 80
40
&
I
20
o:
Average Heating ElementTemperature(°C)
Figure 6 Average heating element temperature at various
heat fluxes with flow rate as a parameter
~ 20
.:.~.'.."
I000-
i"
#*
2000-
1000
"~,m.
g
Io
3000-
20
z
• &&~•
5000x=
E 4000-
~ t00
•
•!
tI000!
• e l • elle
• •emcee• 1 • •
20
<
•
0
00
1000
2000
3000
4000
5000
6000
7000
8000
9000
Heat Flux (kWIm 2)
01000
2000
3000
4000
5000
6000
m=Fh~0et~r~
Figure 9 Average heat transfer coefficient at various heat
fluxes with flow rate as a parameter for single phase data
Figure 7 Average heat transfer coefficient at various heat
fluxes with flow rate as a parameter for single phase data
Effect of Flow Rate on Two-Phase Data
For two-phase heat transfer, heat fluxes as high as
8000kW / m 2 were achieved while keeping the heating
element temperature under 106 C. Figure 8 shows the heat flux
as a function of heating element temperature for three flow
rates in the single and two-phase regimes. Figure 9 shows
for the three flow rates in single and two-phase heat transfer.
Single-phase heat transfer shows an increase with increasing
flow rate. Values of h become coincident in the fully developed
boiling region.
Figure 10 shows ~ as a function of cooling water inlet
temperature subtracted from heating element temperature. Once
again, h is directly dependent on flow rate. Hysterisis can be
seen in the 0.83 ml/s (8.3E -4 k g / s ) flow rate data in the twophase region.
120
A
100
==
60
~
60
I-
40
• 0.46 ml/s (Re = 1200)
• 033 ml/s (Re = 2000)
• 1.0 mils (Re = 2600)
•
g
E
0
eO
A&
• • ~,A=
0
•
• ¢~
•
•
•
•
20
0
20
40
60
80
T.,. - Ti.l.* (°C)
Figure 10 Average heat transfer coefficient for single-and
two-phase data with flow rate as a parameter
6
Copyright © 2000 by ASME
Effect of Surfactant on Single Phase Data
Figure 11 shows heat flux as a function of heating element
temperature. There is no apparent enhancement of heat transfer
characteristics shown from the addition of surfactants in this
data.
6000
0.46 mils (Re = 1200) Water
5000
• 0,46 mils (Re = 1200) 0.05% SLS
~P'4000
41
•
•
1,)
".:,
• 0.46 mils (Re = 1200) 0,10% SLS
o•
• •
~3000
,"C
lw*
Z 2000
Higher Heat Flux Data
Figure 13 and 14 show the results of the repeat experiment
comparing de-ionized water to a lower SLS concentration.
Figure 13 shows heat fluxes on the order of 6 0 0 0 k W / m 2 were
dissipated while keeping the heating element temperature
below 100 C. Figure 14 shows g as a function of of heat flux,
and like the previous surfactant data, show decreasing ~ with
surfactant.
Figure 13 shows heat flux as a function of heating element
temperature. Here the delay in the onset of nucleate boiling can
be observed, but hysterisis is not observed.
Bubbles were observed in the test section, but appeared to
be forming only in areas where the seal was not tight. The
bubbles appeared to be going off to one side and may have
been escaping between the layers of glass.
1000
10000
9000
20
40
60
80
A v e r a g e H e at in g E l e m e n t T e m p e r a t u r e (*C)
100
120
• 0.46 mils (Re = 1200) 0.01% SLS
8000
• 0.46 ml/s (Re = 1200) Water
7000
Figure 11 Heat flux as a function of average heating
element t e m p e r a t u r e with surfactant concentration as a
parameter.
I
~ 6000
5000
4000
|*
Z
3000
2000
1000
20
Effect of Surfactant on T w o - P h a s e Data
Figure 12 shows h as a function of heat flux. The effect of
the surfactant is unclear from this data, but significant
hysterisis can be seen in the two-phase region. The data shows
a possible downward trend in g as surfactant concentration is
increased.
40
60
80
100
A v e r a g e Heating E l e m e n t Te mpe r a tur e (~C)
120
140
Figure 13 H e a t flux as a function of average heating
e l e m e n t t e m p e r a t u r e with surfactant concentration as a
parameter
140 • 0.46 mils (Re = 1200) 0.01% Surfactant
~E
~
120
•
• 0,46 mils (Re = 1200) NO Surfactant
90 •
,F
• 0.46 ml/s (Re = 1200) Water
~= 8 0 .
Q
'0
• 0.46 ml/s (Re = t 200) 0,05% Suffactant
100
80"
¢
• 0.46 ml/s (Re = 1200) 0.10% Suffactant
70
60-
° 60.
~ 40 "
•
•
•
•
g
~ 20-
50
#
g 40
0
<
3c
0
1000
2000
3000
4000
5000
6000
2000
4000
6000
H e a t F l u x ( k W l m 2)
8000
10000
Figure 14 H e a t transfer coefficient as a function of heat flux
with surfactant concentration as a p a r a m e t e r
H e a t Flux ( k W l m 2)
Figure 12 Average heat transfer coefficient as a function of
heat flux with surfactant concentration as a parameter.
7
Copyright © 2000 by ASME
DISCUSSION
Comparison to available literature must be made with some
qualification. Available literature deals with the cooling of a
heat source (microchip) via a heat sink, whereas this study
deals with the direct cooling of a heating element. In the
available literature, therefore, the heat source is separated from
the working fluid. The heat sink conducts the heat to the
working fluid on three sides of the channel. In the current
study, the heat source is directly immersed in the working fluid,
which is heated by only one wall of the channel.
0.12
-
0.097
0,1
0,085
~'~'0.08
0.072
~.~
0.04
0.02
-
-
Phillips
Dissipated Heat Flux
Despite the aforementioned differences, the data are similar to
published results. Figure 15 shows heat fluxes dissipated at
given surface temperatures for the current study as well as
previous studies by Tuckerman (1984) and Phillips (1987). As
seen in Figure 15, the heat fluxes dissipated in the current study
are within 14% of the heat fluxes dissipated in current studies
at similar surface temperatures.
Figure 16 shows a comparison of total modified thermal
resistance between the current study and previous studies by
Tuckerman (1984) and Phillips (1987). The results of the
current study differ from the results of Phillips (1987) by as
much as 35%, however differences in flow rate may account
for the difference in results. Results reported by Phillips (1987)
were for Re = 1500 while the results of the current study are for
Re = 1200.
Tuckerman
June
Figure 16 Comparison of total thermal resistance to
previous studies
Comparison to Correlations
Figure 17 compares the results of the current study, to
correlations given by Dittus-Bolter (1930), and Peng and Wang
(1993). Both correlations are for single phase, turbulent flow,
however Peng and Wang (1993) formulated their correlation
specifically for microchannel flows, citing that Dittus-Bolter
under predicted the results of microchannel phenomena. The
Dittus-Bolter correlation is given by,
Nu = 0.023 Re 4/5 Pr °'4
(2)
11000
(110,10590) Phillips
•
10500
(130,10000) T u c k e r m a n ~
10000
A
~lE 9500
!
(115, 9120) June
and the Peng and Wang correlation is given by,
•
TM
e
(3)
9ooo
8500
8000
7500
7000
20
40
60
80
100
120
140
Heating Element Temperature (C)
As seen in Figure 17, data match Peng and Wang for
Reynolds numbers of 1100 and 1921, but results of the current
study differ from both correlations for Reynolds number of
2556. This is due to the scatter in the data at that Reynolds
number. The data average to a higher value than was expected.
Figure 15 Comparison of heat flux results to previous
studies.
8
Copyright © 2000 by ASME
O Peng and Wang
A
.=
• June
x Dietus BoRer
~) lo
=,
,<
Average RE
Figure 17 Comparison of average Nusselt numbers to
correlations
Power data were compared to values predicted by a simple
energy balance for given heating element temperatures.
As expected, the experimental data showed slightly higher
than predicted power was needed to obtain a given heating
element temperature, due to losses. Figure 18 shows the energy
balance results and experimental results for power at given
heating element temperatures.
Effect of Surfactant
Figure 12 shows inconclusive results for the effects of the
surfactant, sodium laurel sulfate on single phase heat transfer.
This is most likely due to the differences in dissolved gasses in
each working fluid. A study by Adams, Ghiaasiaan, and AbdelKhalik (1999) showed that dissolved air in water could enhance
the heat transfer as much as 17%. Due to the foaming of the
surfactant when placed under negative pressure, the working
fluids could not be held under negative pressure for equal
amounts of time and were therefore degassed to different
extents. Also it was thought that the surfactant concentrations
were too high at 0.05% and 0.10%. For this reason the
experiment was repeated for a non-degassed working fluid at a
concentration of 0.01% sodium laurel sulfate. This was
compared to the same de-ionized water that was used to
prepare the surfactant solution to ensure that the levels of
dissolved gasses were as similar as possible between the deionized water and the surfactant solution. Results of this
experiment are given in Figures 13 and 14 where it can be seen
that the heat transfer coefficient was lower for the working
fluid with the surfactant in single-phase heat transfer.
For two-phase heat transfer, Figure 13 shows a delay in the
onset of nucleate boiling as predicted by Dhir (1999).
Hysterisis is also observed in Figure 13 for the de-ionized
water, but not for the surfactant data. This is most likely due to
having too few data for higher heat fluxes in the surfactant
data. Hysterisis in the de-ionized water data is from the sudden
release of dissolved air in the working fluid at the onset of
nucleate boiling.
0.05000
0.04500
0.040Oo
o
0.03500
g
0.03OO0
i
0.02500
UNCERTAINTY
Estimates of uncertainty in this experiment were
calculated using a 95% confidence interval. Uncertainty was
determined from estimates of bias (B) and precision (P) by the
following,
U = [B2 + p2 ]1/2
(4)
o
O ExperimentalData
• Energy Balance
o"
o
0.02000
_o
0.01500
o
0.01000
i
8
0.00500
0,00000
20
40
60
80
100
120
Heating ElementTemperature (C)
Figure 18 Experimental results compared to results
predicted with energy balance
In general, bias was considered to be the least count of the
measuring instrument, calibration data as supplied by the
manufacturer or, in the case of the thermocouples, calibration
done in-house. Precision was determined by multiplying the
standard deviation of 10 measurements by 2.26 (from a
standard Students T-Table). In general, 30 would be the optimal
number of measurements, however difficulties in maintaining
steady state for a long period of time prevented that number of
measurements.
Propagation of error was determined by solving the
propagation equation of Kline and McClintock (1955).
9
Copyright © 2000 by ASME
A summary of bias, precision and total uncertainty values
determined for this experiment is given in Table 1.
Table 1 Summary of bias, precision and total uncertainty
for a representative subset of the data.
q"
Data Set
0.01% SLS
0.46 ml/s
Re = 1200
Water
0.46 ml/s
Re= 1200
(kResult
W/m2)
test section. The number of minor losses between the pressure
reading and the test section needs to be minimized, and more
accurate measurements need to be taken on either side of the
test section. The challenge is eliminating or accurately
accounting for the losses, since the working fluid must flow
through at least one large reduction and make two 90-degree
turns into and out of the test section.
T~
B
P U Result
~c)
I*/,)
(P)'/.(Uo) Result B [ P U
B
(°/o} ('/.) I*/.)
(kW/m2k) (%) (%) ~%)
212
255
297
343
395
499
622
0.019
0.019
0.019
0.019
0.018
0.018
0.018
3.19
3.14
3.14
3.14
3.14
3.14
3.13
3.19
3.14
3.14
3.14
3.14
3,14
3.13
95.3
90.5
95.3
97.7
100.3
103.0
105.8
3.1
3.3
3.1
3.1
3.0
2.9
2.8
2.18
0.62
0.97
0.82
0.50
0.55
0.27
3.83
3.37
3.29
3.18
3.03
2.96
2.85
33
42
47
52
59
72
87
4.93
5.33
5.10
5.01
4.95
4.76
4.73
4.81
3.40
3,66
3,49
3.61
3.40
3.20
6.88
6.32
6.28
6.10
6A3
5.85
5,71
136
169
200
244
290
333
445
561
624
760
912
0.020
0.020
0.019
0.019
0.019
0.019
0.018
0.018
0.018
0.018
0.018
3.19
3.14
3.13
3.14
3,14
3.16
3.14
3.14
3.13
3.13
3.14
3.19
3.14
3.13
3.14
3.14
3.16
3.14
3.14
3.13
3.13
3.14
60.7
65.1
70.8
82.0
90.3
93.7
99.3
102.9
104.4
108.5
115.0
4.9
4.6
4.2
3.7
3.3
3.2
3.0
2.9
2.9
2.8
2.6
2.93
0,60
0.37
0.57
0.62
1.45
0.67
0.55
0.33
0.26
0,56
5.74
4.65
4.25
3.70
3.38
3.51
3.10
2.97
2.89
2.78
2.67
42
46
48
47
48
53
66
82
90
107
127
9,94
8,80
7.64
6.16
5.35
5.20
4.86
4.95
4.95
4.92
5.21
6.73
3.36
3.22
3.28
3.33
3.95
3.34
3.30
3.21
3.31
4.59
12.01
9.42
8,29
6.98
6.31
6,53
5.90
5.95
5,90
5.93
6.95
CONCLUSION
The goal of constructing a low cost apparatus for
investigation of flow boiling phenomena in microchannels, was
achieved, however further improvements are possible. The
small cross section heating element, being only 1.8 cm long,
had a very low resistance that was difficult to measure
accurately. In addition, the low thermal coefficient of resistance
of Platinum, combined with the low resistance, made detecting
the changes in resistance even more difficult. It is suggested
that a different material be found, with a change to a much
larger area. This would approximate the chip/heat sink model,
followed by the majority of the researchers in the current
literature. This would necessitate the construction of a
microchannel heat sink rather than just a single channel.
Further investigation is needed into the effect of the
surfactant concentrations on the heat transfer and boiling
phenomena. In this study, and one other study (Wang and
Harnett, 1994) heat transfer was actually reduced by the
addition of SLS, while in other studies, the heat transfer was
increased. It is possible that above some limiting concentration,
the heat transfer is adversely affected by the surfactants, while
below that concentration, heat transfer is enhanced.
All investigations to date have been unsuccessful in
visualizing bubbles in the channel during boiling. The nature of
the boiling curve, confirms two-phase heat transfer, but bubbles
have been elusive. This apparatus would be ideal used in
conjunction with a high-speed camera, under magnification, or
with strobe light photographic techniques.
Finally, saturation temperature could not be determined
due to inaccuracies in calculating the pressure drop across the
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11
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