Proceedings of the 14th International Heat Transfer Conference
IHTC14
August 8-13, 2010, Washington D.C., USA
DRAFT
IHTC14-22647
EXPLORATION OF EXPERIMENTAL TECHNIQUES TO DETERMINE THE
CONDENSATION HEAT FLUX IN MICROCHANNELS AND MINICHANNELS
Melanie M. Derbya, Hee Joon Leea, Rose C. Craftb, Gregory J. Michnac, Yoav Pelesa, Michael K. Jensena
aDepartment
of Mechanical, Aerospace, and Nuclear Engineering, Rensselaer Polytechnic Institute
Troy, NY, USA
bHeat Transfer Research, Inc., College Station, TX, USA
cMechanical Engineering Department, South Dakota State University, Brookings, SD, USA
ABSTRACT
This study seeks to analyze and explore experimental
methods to study condensation heat transfer in micro- and minichannel. Following, an experimental setup was built and initial
results are presented. Several experimental techniques were
reviewed, while two, thermoelectric coolers and a copper-heatflux-sensor were analyzed in detail for condensation heat flux.
It was concluded that thermoelectric coolers were not suitable
as heat flux sensors for single-phase validation, but the copperheat-flux-sensor was appropriate to measure heat transfer
coefficients at the mini-scale. Condensation heat transfer
coefficients were obtained experimentally in seven parallel
square minichannels of diameter 1mm. Existing condensation
correlations were applied to these data; micro- and mini-scale
correlations captured the appropriate trends but underpredicted
the data.
experimental methods and created new experimental techniques
to measure heat flux and wall temperature, and ultimately
condensation heat transfer coefficients. To increase the
measured heat duties, some studies used fluid-to-fluid heat
exchangers comprising of parallel microchannels. FernandezSeara et al. [1] reviewed the Wilson plot— a method frequently
used for macro-scale heat exchangers. The condensation heat
transfer coefficient can be inferred through a series of
resistances without direct measurement of the wall temperature.
However, Garimella and Bandhauer [2] estimated in an
uncertainty analysis that a 1:6 resistance ratio of condensation
to coolant was needed to obtain the heat transfer coefficient
within ± 15%. Webb and Ermis [3] used the Modified Wilson
plot to determine the condensation heat transfer coefficient in
multi-port channels with hydraulic diameters of 0.44 - 1.56
mm. The test section was designed to produce cooling water
heat transfer coefficients higher than the refrigerant. Measured
condensation heat transfer coefficients were reported to be
within ± 10.6%. The Wilson plot method and its derivatives at
the micro and mini-scale required sufficiently high coolant heat
transfer coefficients, which in turn created low log mean
temperature differences. Therefore, temperature measurement
error could increase heat transfer coefficient uncertainty.
Rather than using resistances to find the condensation heat
transfer coefficient as in the Wilson plot method, the heat flux
was measured by conducting an energy balance on a test
section’s coolant. Cavallini et al. [4] utilized a test section with
three fluid-to-fluid heat exchangers. The first fluid-to-fluid heat
exchanger superheated the refrigerant, the second partially
condensed to the desired inlet quality, and the third was the
cooled measurement section. The measurement section had
thirteen square 1.4 mm channels, although there was no flow in
the outer two channels because thermocouples were inserted in
them to measure the wall temperatures. Uncertainties in heat
transfer coefficients of ±7% were reported. But it is important
to note that a large cooling water temperature difference is
INTRODUCTION
Most current electronic cooling systems use air as a working
fluid. However, as transistor and power densities increase, more
aggressive cooling schemes, such as the use of vaporcompression cycles, are needed in high power density systems.
The entire vapor-compression cycle needs to be optimized to be
compact and lightweight. Boiling heat transfer in small
diameter channels has received significant attention in the last
decade. Good boiling heat transfer and pressure drop
correlations are available for design because heat flux in miniand microchannel boiling experiments can be easily obtained
by measuring heater power. However, significantly less
literature is available for condensation heat transfer, with a lack
of fundamental knowledge on governing processes and accurate
correlations. Difficulty in measuring both the heat flux and wall
temperature at the mini- and micro-scales yields high
uncertainties in heat transfer coefficients.
Because of the low heat duties encountered in mini- and
microscale condensation, researchers have adapted macro-scale
1
Copyright © 2010 by ASME
necessary for the temperature measurement’s uncertainty to be
small compared to the measured temperature difference.
Aside from conducting an energy balance on the test section,
one could conduct an energy balance before and after the test
section to get the heat transfer rate. Wang et al. [5] studied the
effects of stratified and annular flow regimes on condensation
of R134a in ten 1.46-mm diameter channels cooled in a cross
flow air heat exchanger. Two methods of calculating heat duty,
an air-side energy balance and the pre- and post-heater energy
balance, were compared; the difference of ± 10% between the
methods was concluded to be heat losses. This work shows the
effect of heat losses on the measured heat duty, but also the
difference between two methods for obtaining the heat duty,
possibly resulting from measurement uncertainty. Condensation
heat transfer coefficients were determined from seven cross
flow heat exchangers. Reported uncertainties for the heat
transfer coefficient are ± 8.2%.
However, these fluid-to-fluid heat exchangers had some
drawbacks, including low temperature differences across
coolant flows and difficulty measuring wall temperature in
some parallel channel geometries. As a result of measurement
issues associated with fluid-to-fluid heat exchangers, other
studies modified the standard fluid-to-fluid heat exchangers.
Garimella and Bandhauer [2] adopted the Wilson plot method
and modified it to be more suitable at the micro and miniscales. First they studied a square multi-port channel of 0.76mm hydraulic diameter, and then Bandhauer et al. [6] expanded
it to other shapes and diameters. Their test section consisted of
a counterflow fluid-to-fluid heat exchanger. A low quality
change was desirable across the test section, thus low heat
duties needed to be measured, which resulted in a very low
temperature rise across the test section coolant. The researchers
decoupled this problem by using a “thermal amplification”
loop, separating the cooling portion into two loops; the primary
loop cooled the test section and the secondary loop measured
heat duty in the primary loop, but had a larger temperature
change. Using this technique, the total uncertainty of heat duty
was estimated to be ± 10%, and the condensation heat transfer
coefficient was estimated on average within ± 21%.
Still others utilized experimental techniques that measured
the heat transfer coefficient via electrical means. Shin and Kim
[7, 8] compared an air-cooled test section to an electrically
heated test section. This innovative approach measured heat
transfer rates of 10 W or less in circular and rectangular single
channels of diameters ranging from 0.493 to 1.067 mm. Two
identical copper tubes with attached fins were placed in a
crossflow air-cooled heat exchanger. Refrigerant flowed
through one tube, while the other tube contained a resistive
heater, controlled by a DC power supply. The heat transfer rate
was assumed to be the same in both tubes when comparable
surface temperatures were read by thermocouples at identical
fin positions, although the tube boundary conditions differed.
Therefore, this technique was not dependent on an air side
energy balance. Uncertainties in the heat transfer coefficient of
±3.5-19.9% were reported. Two practical barriers to
implementing this method were measuring extremely low mass
flow rates and manufacturing two identical finned copper test
sections. Alternatively, thermoelectric coolers (TECs) provided
an electrical method to determine heat flux. Baird et al. [9]
studied condensation in single channels, 0.92 and 1.95 mm in
diameter. Thermoelectric coolers (TECs) provided an electrical
method to determine heat flux. Ten TECs were arranged along
the channel for cooling, dividing the channel into ten “quasilocal” energy balances. Applied voltages created temperature
differences across each TEC, producing a cold side attached to
a load, and hot side that rejected heat to a heat sink. Each TEC
was tested by placing the TEC between two copper blocks
whose temperatures were controlled by circulating water.
Because the TECs were checked against the manufacturer’s
equation, but not calibrated, the uncertainties in the heat flux
could be greater than their claimed ±5% and, therefore, some
TECs may have been accepted that, in reality, were outside of
the acceptable range. Additionally, the wall temperatures were
measured 2 mm away from the channel in a copper block to
obtain heat transfer coefficients. Baird et al. [9] reported
uncertainties in the heat transfer coefficient of ±20%. The TEC
method showed promise due to the ability to measure quasilocal heat transfer coefficients in several positions along a
single channel, but appropriate calibration is a concern.
Since simultaneous measurement of the heat transfer rate
and wall temperature presented difficulty in condensation,
boiling heat transfer literature was investigated for
experimental methods that could be applied to condensation.
One such method was a heat flux sensor developed by
Kedzierski and Worthington [10]. Oxygen-free copper was
selected because of high thermal conductivity, thus reducing the
overall uncertainty resulting from the thermal conductivity.
Embedded thermocouples in the copper block were used to
obtain the wall temperature and heat flux from the temperature
gradient in the block. Therefore, the dominant uncertainty in
the heat flux was the contribution from the temperature
gradient. In the pool boiling experiments conducted by
Kedzierski [11] using R123 and hydrocarbon mixtures, the
measured heat flux by the copper-heat-flux sensor was
estimated to be ± 10% at a flux of 10 kW/m2, and ± 3 - 6% for
fluxes above 30 kW/m2. These low uncertainties showed that
this method was promising to consider for condensation heat
transfer experiments.
Existing experimental methods for measuring condensation
heat transfer coefficients were identified in this study. Among
those, TECs and conduction in a copper block were examined
further because they do not rely on measuring a cooling water
temperature differences as in a fluid-to-fluid heat exchanger,
and exhibit potential to accurately measure the condensation
heat transfer coefficient.
NOMENCLATURE
𝐴
Area [m2]
D
Diameter [m2]
G
Mass flux [kg/m2s]
𝐼
Current [A]
𝑘
Theraml conductivity [W/m2K]
2
Copyright © 2010 by ASME
𝑁𝑢
𝑞"
𝑄̇
𝑅𝑒
𝑇
𝑤
𝑦
Figure 1 Loop schematic
Nusselt number
Heat flux [W/m2K]
Heat transfer rate [W]
Reynolds number
Temperature [oC]
Uncertainty
Position [m]
to copper blocks which provided an interface between the TEC
and the tube. To maintain good thermal contact between the
copper and TEC interface as well as the TEC and heat sink
interface, thermal grease was used along with thermal pads to
fill in air gaps. Thermal grease also surrounded the tube to
provide good thermal contact with the copper blocks.
The test section was separated into ten modules insulated
from each other by a sheet of 0.4 mm Teflon. Each module
consisted of a TEC and two copper blocks, which surrounded
the tube. Each of these copper blocks had a 1.5 mm diameter
cutout to allow the tube to pass through. The tube was centered
inside the copper blocks by rubber o-rings, which allowed for
tubes of diameters up to approximately 3 mm to be used. All
ten TECs shared one water-cooled aluminum heat sink to which
the TECs rejected heat. Bolts provided vertical compression to
reduce contact resistances. The modules were held in place by a
combination of a Lexan end block, a Lexan locator tray and a
slider block, which provided horizontal compression, shown in
Fig. 3. The test section was insulated with plastic and balsa
wood.
Subscripts
𝑏
Block
𝑐
Cold
𝐶𝑢
Copper
𝐷
Diameter
𝑒𝑥𝑝
Experimental
𝑓
Fluid
𝑔
Gradient
ℎ
Hot
𝑝𝑟𝑒𝑑
Predicted
𝑤
Predicted
EXPERIMENTAL APPARATUS - TEST FACILITY
The same basic refrigerant loop (Fig. 1) was used for two
experimental test sections: thermoelectric-cooled and copperheat-flux-sensor test sections. The loop was designed to
accommodate a wide range of mass fluxes and qualities. A
pump produced flow rates range from 10 to 180 mL/min,
measured by a four tube rotameter. Pressure in the loop was
controlled by adjusting the pressure in a bladder accumulator. A
preheater set the test section inlet quality and an upstream
throttling valve was added to suppress boiling instabilities.
Temperature and pressure probes at the inlet and outlet of the
test section yielded an energy balance during single-phase
validation. An air-cooled post condenser was added for use with
the copper-heat-flux-sensor test sections.
THERMOELECTRIC COOLERS - DATA ANALYSIS
The test section was instrumented to determine the heat flux,
wall temperature, and, ultimately, the heat transfer coefficient.
The necessary variables for the heat flux were the TEC hot side
temperatures, cold side temperatures, voltages and currents.
Type T, 36-gauge thermocouples were installed on the hot and
cold sides of the TECs and attached to the tube wall using
thermal epoxy. The thermocouples were calibrated to ±0.2 oC
THERMOELECTRIC COOLERS - TEST SECTION
The goals of the thermoelectric-cooled test section were to
obtain quasi-local heat transfer coefficients in a wide range of
circular tube diameters. As shown in Figure 2, the complete test
section assembly consisted of an instrumented tube in which
condensation occurred. Thermoelectric coolers were mounted
Figure 2 Cross section view of the test section
Figure 3 TEC test section
3
Copyright © 2010 by ASME
using a National Instruments Data Acquisition system and
LabVIEW and a reference thermometer accurate to ±0.01 oC.
The TECs were connected to a power supply in parallel. Ten
TEC voltages, maximum of 2.3 V, were measured directly via
LabVIEW. Individual currents, maximum of 8.5 A, were
measured by a 15 A current shunt.
The TECs required calibration to determine the heat flux, as
the manufacturer’s curves were only for ideal cases. A
calibration apparatus was constructed, similar to Figure 2, but
an electric heater replaced the tube. Therefore, the heat transfer
rate was measured via the heater voltage and current, and this
could be correlated to cold side temperature, hot side
temperature, TEC voltage and current. Additionally 40 V, 80 A,
and 3.5 mΩ transistors acted as a switch for each TEC.
Feedback control implemented in LabVIEW controlled TEC
cold side temperatures around a set point. The hot side
temperature was controlled by the temperature of the waterfilled heat sink.
The calibration was conducted over a range of 0 – 6 W
heater power and 0-2 V TEC voltage. The cold side temperature
was maintained at 10°C. The data were then correlated in a
form similar to Baird [9],
𝑄̇ = (
𝑇𝑐 − (0.607 𝐼 2 − 8.22 𝐼 + 𝑇ℎ − 1.42)
)
4.96
similar, suggesting that thermal contact is consistent between
TECs. Differences in the individual TECs explain the poor
validation of the heat flux in Eq. (1). In conclusion,
Figure 4 Turbulent Nusselt Numbers using TECs
thermoelectric coolers were suitable for cooling, but not as an
accurate heat flux sensors in condensation experiments.
(1)
COPPER HEAT FLUX SENSOR - TEST SECTION
Because the thermoelectric-cooled test section was not an
accurate heat flux sensor, the copper-heat-flux test section,
shown in Figure 5, was designed. The test section consisted of
seven parallel square minichannels 1 mm in diameter machined
in an oxygen-free copper block of dimensions 20 mm by 38.1
mm by 140 mm. A Lexan plate, clear for future flow
visualization, was sealed to the copper block with an o-ring.
During condensation, a two-phase R134a mixture entered
through a 4.7-mm hole in the Lexan cover plate, flowed into the
inlet header, through the channels, and exited through a
symmetric exit header. To limit heat transfer in the header
region, the total header area accounted for 6.8% of the total
heat transfer area. For heat flux measurements, the copper
block was divided into three measuring segments, with nominal
lengths of 28, 38 and 28 mm and separated by 1.59-mm slots,
for reporting the heat transfer coefficient at average qualities. A
water-filled serpentine cold plate was attached to the bottom of
the copper-heat-flux sensor. The coolant water temperature,
controlled by a thermal bath, created a temperature difference
in the block proportional to heat flux. To measure the heat flux
and wall temperature, 1 mm diameter, 10 mm deep
thermocouple wells were drilled for T-type sheathed
thermocouples, 3, 14, 19, 24, 29 and 34 mm from the bottom of
the channel.
This calibration equation predicted most of the calibration data
within 10%, with better agreement at higher heat transfer rates.
Heat flux was then the heat transfer rate divided by the tube
segment’s surface area. With the heat flux determined from the
TEC calibration in Eq. (1), the wall temperature measured from
the local wall thermocouple, and the fluid temperature for
single-phase flow determined from an energy balance, the heat
transfer coefficient could be obtained through Newton’s law of
cooling.
THERMOELECTRIC COOLERS
- RESULTS AND DISCUSSIONS
The heat flux calibrations were evaluated using laminar and
turbulent single-phase flows in the test section shown in Figure
2. At a Reynolds number of 5420, without feedback control of
the cold side temperature, the measured Nusselt numbers
exhibited wide variation. Although the Gnielinski’s Correlation
[12] predicted a Nusselt Number of 33, the measured values
ranged from 3 to 51 with an average of 13.1, as shown in
Figure 4. Similarly when the cold side temperature was
computer-controlled, Eq. (1) did not adequately determine the
heat flux.
The significant difference in TEC performance could be a
combination of variation in thermal contact resistance and TEC
performance. Nagy and Buist [13] developed a technique to
determine the variation in thermal contact. Once the TECs were
turned off, the decay of the Seebeck voltage indicated quality of
the thermal contact. If the thermal contact was consistent
between TECs, the slope of the voltage versus time curve
would be the same. Using this technique, the slopes were
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Copyright © 2010 by ASME
along the channel length, as in single-phase validation.
FLUENT simulations utilized single phase flow, and showed
the insertion of air-filled slots between the measuring segments
reduced the axial conduction in a segment. The temperature
distribution in the copper block on two lines parallel and 17.1
mm and 27.1 mm from the bottom of the channels is shown in
Figure 7. Due to two-dimensional conduction, the axial
temperature changed greatly in segments without separating
slots, but the two-dimensional conduction effects decreased
with the insertion of slots between the channels. Thus, onedimensional conduction was an appropriate assumption when
the measuring segments were separated by air-filled slots.
In addition to the heat flux in each segment, a measured wall
temperature was needed to determine the heat transfer
coefficient, and the header heat transfer coefficient was also
needed to determine fluid temperature in single-phase
Figure 5 Copper test section
COPPER HEAT FLUX SENSOR - DATA ANALYSIS
The average heat flux and wall temperature were measured
in three measuring segments along the channel. For validation,
the data reduction method was applied to FLUENT data and the
results were compared. First, the heat transfer rate in each
segment was expressed assuming Fourier’s Law in one
dimension,
𝑄̇ = −𝑘𝑐𝑢 𝐴𝑏
𝑑𝑇
𝑑𝑦
(2)
Subsequently, the channel heat flux was determined by dividing
the segment’s heat transfer rate by its channel area. The
assumption of one-dimensional Fourier’s Law required a linear
temperature gradient. Therefore, FLUENT simulations were
conducted that predicted the linear temperature gradient, shown
in Figure 6, and, in each segment, the five thermocouples
located 14, 19, 24, 29, and 34 mm from the bottom of the
channels measured the temperature gradient.
The uncertainty in the temperature gradient, therefore, was
the slope of the temperature versus position line. Kedzierski
and Worthington [10] developed an expression for the
uncertainty in the temperature gradient, wg,
𝑞"𝐷 2
1
𝑤𝑔 = √𝑤𝑇2 + (
) √ 𝑁
6𝑘𝑐𝑢
∑𝑖=1(𝑦𝑖 − 𝑦̅)2
Figure 6 Temperature Gradient in FLUENT Simulations
(3)
The first term in Eq. (3) represented the uncertainty in the
temperature sensors and the uncertainty in the position of the
thermocouple in the drilled thermocouple well, although the
thermocouple well diameter was selected to minimize this
contribution to the gradient uncertainty. The second term
represented the distribution of the thermocouple wells in the
block, as increasing the distance between thermocouple wells
will reduce the uncertainty in the temperature gradient.
Also, axial conduction along the channel and between
segments contributed to uncertainty in the temperature gradient.
This was more important when the fluid temperature changed
Figure 7 Effect of Slots on Axial Conduction
5
Copyright © 2010 by ASME
experiments and quality in condensation experiments. The wall
temperature was extrapolated with a second order polynomial
curve fit, using the five thermocouples for measuring the
temperature gradient and sixth “near wall” thermocouple,
located 3 mm from the bottom of the channel. The y-intercept
of the polynomial fit was then the wall temperature.
Additionally, the inlet header heat transfer coefficient was
needed to determine the single-phase fluid temperature or twophase inlet quality. FLUENT simulations showed the inlet
header heat transfer coefficient was approximately twice that of
the middle measuring segment, and the outlet heat transfer
coefficient the same as the middle measuring segment.
All these assumptions were validated by inputing FLUENT
simulation data into the developed data reduction program, and
comparing the results of FLUENT and the data reduction
program. The FLUENT simulations incorporated both fluid
motion and conduction in the test section. In the fluid, a very
fine structured hexagonal grid solved the steady incompressible
Navier-Stokes equations using k-ε turbulent modeling, while
unstructured tetrahedron grids solved the conduction equation.
A constant temperature boundary condition was prescribed on
the bottom surface of the copper block. A wide range of singlephase Reynolds numbers was tested, as higher Reynolds
numbers exhibited high heat transfer coefficients and low fluid
temperature changes along the channel, akin to condensation.
As shown in Table 1, there was good agreement between the
wall to fluid temperature difference, especially at lower
Reynolds numbers, and outside of the laminar range, the
difference in heat flux was less than 4%.
Condensation data were graphed in Figure 9 for a mass flux
of 250 kg/m2s and a saturation temperature of 45 oC. The trend
of increasing heat transfer coefficient with increasing average
quality was expected. In addition, the uncertainties in the
condensation heat transfer coefficient are lower than in single
phase because a larger temperature gradient occurred at the
higher condensation heat fluxes. In Figure 10, the data were
compared with various existing condensation correlations. The
Shah [16] and Akers et al. [17] macro-scale correlations, both
did not capture the trend of the data. However, the Bandhauer
[6] and Agarwal [18] correlations, developed for micro and
mini-scale circular and non-circular channels, respectively,
underpredicted the data. The behavior of these correlations may
be explained by differences in boundary conditions, as the
correlations were developed for four sided cooling, and the
experimental apparatus utilizes three sided cooling.
Additionally, the Soliman [19] correlation, developed for the
conventional scale condensation mist flow regime, also predicts
the appropriate trend of the data. Therefore, the trends of the
data were best fit by micro and mini-scale correlations, but
difference could be caused by the different boundary
conditions.
Table 1 Mean Average Errors Between FLUENT and Data
Reduction Program
Re
MAE
Tf - Tw
MAE
Heat Flux
COPPER HEAT FLUX SENSOR
- RESULTS AND DISCUSSIONS
Single-phase experiments validated the determination of
wall temperature and heat flux. Nusselt and Reynolds numbers
were calculated at local fluid properties and plotted in Figure 8.
The first segment Nusselt numbers were corrected for turbulent
developing flow using the Al-Arabi [14] correlation. To obtain
the turbulent entrance length for a rectangular channel, Hartnett
et. al. [15] proposed the developing length over diameter ratios
were 40 for a Reynolds number of 3000 and less than 20 for
Reynolds numbers greater than 4000. Therefore, a turbulent
entrance length to diameter ratio of thirty was assumed because
the flow was in the transitional regime. The experimental data
changed with Reynolds number to the power of 1.08, close to
the power of one used in the Gnielinski’s correlation [12]. As
these Reynolds numbers were at the lower end of the
correlation’s range, and the Gnielinski’s correlation was not
developed for the micro/mini-scales and three-side cooling, a
higher value of Nusselt number was reasonable. Additionally, a
single-phase energy balance and the heat transfer rate
calculated from the temperature gradient and headers agreed
within 3.5-8%. Good agreements in Nusselt number and energy
balance showed this approach is valid for determining the
condensation heat transfer coefficient.
930
3,200
6,475
32,000
63,000
3.0%
2.4%
4.0%
8.6%
9.5%
8.3%
2.7%
3.9%
3.5%
3.2%
Figure 8 Nu Versus Re Using Copper Test Section
6
Copyright © 2010 by ASME
uncertainty analysis showed that this approach can yield high
accuracy heat transfer coefficient measurements.
3. The copper block heat flux sensor was used to determine the
condensation heat flux in mini-channels. The measured heat
transfer coefficient indicated that the trend is physically correct.
The existing mini-scale correlations were the best predictor for
current measured condensation heat transfer coefficient.
Therefore, this method is promising for accurately measuring
the heat transfer coefficient.
ACKNOWLEDGEMENTS
This work was supported in part by the Office of Naval
Research (ONR) under the Multidisciplinary University
Research Initiative (MURI) Award GG10919 entitled "SystemLevel Approach for Multi-Phase, Nanotechnology-Enhanced
Cooling of High-Power Microelectronic Systems."
REFERENCES
[1] Fernandez-Seara, J., Uhia, F. J., Sieres, J., Campo, A., 2007,
“A General Review of the Wilson Plot Method and its
Modifications to Determine Convection Coefficients in Heat
Exchanger Devices,” Applied Thermal Engineering, 27 (1718) , pp. 2745-2757.
[2] Garimella, S., Bandhauer, T. M., 2001, “Measurement of
Condensation Heat Transfer Coefficients in Microchannel
Tubes,” Proc. ASME International Mechanical Engineering
Congress and Exposition, New York, NY, pp. 1-7.
[3] Webb, R. W., Ermis, K., 2001, “Effect of Hydraulic
Diameter on Condensation of R-134a in Flat, Extruded
Aluminum Tubes,” Enhanced Heat Transfer, 8 (2), pp. 77-90.
[4] Cavallini, A., Censi, G., Del Col, D., Doretti, L., Longo, G.
A., Rossetto, L., 2003, “Experimental Investigation on
Condensation Heat Transfer Coefficient Inside Multi-port
Minichannels,”
First
International
Conference
on
Microchannels and Minichannels , Rochester, NY, pp. 691-698.
[5] Wang, W. W., Radcliff, T. D., Christensen, R. N., 2002, “A
Condensation Heat Transfer Correlation for Millimeter-scale
Tubing with Flow Regime Transition,” Experimental Thermal
and Fluid Science, 26 (5), pp. 473-485.
[6] Bandhauer, T. M., Akhil, A., Garimella, S., 2006,
“Measurement and Modeling of Condensation Heat Transfer
Coefficients in Circular Microchannels,” Journal of Heat
Transfer 128 (10), 1050-1059.
[7] Shin, J. S., Kim, M. H., 2004, “An Experimental Study of
Flow Condensation Heat Transfer Inside Circular and
Rectangular Mini-channels,” Second International Conference
on Microchannels and Minichannels, Rochester, NY, pp. 633640.
[8] Shin, J. S., Kim, M. H., 2004, “An Experimental Study of
Condensation Heat Transfer Inside a Mini-channel with a New
Measurement Technique,” International Journal of Multiphase
Flow 30 (3), pp. 311-325.
[9] Baird, J. R., Fletcher, D. F., Haynes, B. S., 2003, “Local
Condensation Heat Transfer Rates in Fine Passages,”
International Journal of Heat and Mass Transfer 46 (23), pp.
4453-4466.
Figure 9 Condensation Data, G = 257 kg/m2s
Figure 10 Comparison of Data with Correlations
CONCLUSIONS
Existing
experimental
techniques
for
measuring
condensation heat transfer have been reviewed. In this study,
thermoelectric coolers and a copper-heat-flux sensor were built
and analyzed as potential approaches for studying flow
condensation in micro and minichannels. Single-phase
validation experiment results and condensation data were
presented for these two methods. The main conclusions of this
study are as follows:
1. Thermoelectric coolers were suitable for cooling but are not
appropriate as a heat flux sensors.
2. The single-phase heat transfer coefficient results from the
copper block heat flux sensor experiments agreed well with the
nGnielinski’s correlation [12] for transitional flows. An
7
Copyright © 2010 by ASME
[10] Kedzierski, M. A., Worthington III, J. L., 1993, “Design
and Machining of Copper Specimens with Micro Holes for
Accurate Heat Transfer Measurements,” Experimental Heat
Transfer 6 (4), pp. 329-344.
[11] Kedzierski, M., 2000, “Enhancement of R123 Pool Boiling
by the Addition of Hydrocarbons,” International Journal of
Refrigeration 23 (2), pp. 89-100.
[12] Gnielinski, V., 1995, “New method to calculate heat
transfer in the transition region between laminar and turbulent
tube flow,” Forschung im Ingenieurwesen 61 (9), pp. 240.
[13] Nagy, M. J., Buist, R. J., 1996, “Transient Analysis of
Thermal Junctions Within a Thermoelectric Cooling Assembly.
15th International Conference on Thermoelectrics,” pp. 288292.
[14] Al-Arabi, M., 1982, “Turbulent Heat Transfer in the
Entrance Region of a Tube,” Heat Transfer Engineering 3 (3-4),
pp. 76-83.
[15] Hartnett, J. P., Koh, J. Y., McComas, S. T., 1962, “A
Comparison of Predicted and Measured Friction Factors for
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