th
Proceedings of the ASME 2012 10 International Conference on Nanochannels, Microchannels, and Minichannels
ICNMM2012
July 8-12, 2012, Rio Grande, Puerto Rico
ICNMM2012-73175
HEAT TRANSFER ENHANCEMENT DURING POOL BOILING OF WATER OVER
HORIZONTAL AND VERTICAL TUBES WITH MICRO STRUCTURED SURFACES
Jeet S. Mehta, Satish G. Kandlikar
Mechanical Engineering Department
Rochester Institute of Technology
Rochester, New York, USA
[email protected], [email protected]
ABSTRACT
Pool boiling is a stable and an efficient method for
transferring large quantities of heat. This mode of heat transfer
is used in a wide range of applications, including steam
generation in boilers, petrochemical, pharmaceutical, cryogenic
and many other industrial processes. It also holds promise for
cooling of microelectronic devices, such as lasers,
microprocessors and others. The objective of this work is to
investigate the heat transfer augmentation due to an array of
micro structured surfaces over a circular tube. The effects of
horizontal and vertical orientation of the tubular test section on
heat transfer enhancement are also studied. The bubble
nucleation, growth and interactions over the micro structured
surfaces are analyzed using high speed cameras to understand
the bubble dynamics.
entrant cavities have been developed and tested by various
researchers in the previous two decades. Webb & Pais [1] in
1992 tested three commercially available GEWA series tubes.
The re-entrant cavities are generated by performing further
modification on finned tubes. They tested these surfaces with
five refrigerants at saturation temperatures of 40 °F (4 °C) and
80 °F (27 °C) and concluded higher heat transfer coefficients
are observed at higher saturation temperatures. Memory et al.
[2] used re-entrant cavity tubes such as commercially available
GEWA series, Thermoexcel and Turbo tubes. Their results
showed heat transfer enhancements of up to 5.5 times with
GEWA series tubes at low heat fluxes. Whereas the
Thermoexcel and Turbo tubes showed up to 20 times the
enhancements at similar heat flux conditions. At higher heat
fluxes the performance was similar for all the different tubes.
Huebner and Kuenstler [3] used similar tubes with n-hexane
and propane and observed enhancements in the range of 2.4-4
times. Tatara and Payvar [4] used R134a with Turbo-BII-HP
tube and reported 60-90 % further increase in performance
compared to the Turbo-B tube. Rajulu et al. [5] fabricated
simple re-entrant cavities by modifying the tips of finned tubes.
Their results showed an enhancement of up to 2.5 times and
they also observed an increase in the enhancement factor with
increasing heat flux. They developed a correlation for the
enhancement factor as a function of the heat flux and the cavity
width of the re-entrant channels. Jung et al. [6, 7] studied the
performance of two re-entrant cavity tubes and observed
significant performance enhancements at low heat fluxes. They
also concluded that the rate of increase of heat transfer
compared to the increase in heat flux was small, possibly
because of the blockage of liquid re-entry into the pores and
tunnels at higher heat fluxes. Their experimental data showed
40 % greater enhancements with flammable refrigerants
compared to halogenated refrigerants. Ribatski & Thome [8]
used GEWA-B, Turbo-CSL and Turbo-BII-HP tubes and
INTORDUCTION
Over the past few decades, extensive research towards
augmenting the nucleate boiling heat transfer has been
conducted. Modifications on boiling surface have proved to be
a viable means of enhancing the heat transfer performance. A
number of different techniques have been employed to enhance
the pool boiling heat transfer performance. These techniques
can be broadly classified into the following five categories:
1. Re-entrant cavities
2. Porous layer coating
3. Surface roughness
4. Tube orientation
5. Microchannels / Integral fins
Heat transfer enhancement using re-entrant cavities have
been widely studied in literature. These re-entrant cavities
continuously trap vapor at the nucleation sites and aid in the
nucleation process thereby augmenting the heat transfer
performance 3-4 times. Numerous techniques of generating re-
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Copyright © 2012 by ASME
obtained enhancement factors of 2.4-5.2, 2.4-2.9 and 1.9-7.0,
respectively, from higher to lower heat fluxes. Ji et al. [9] in
2010 tested four re-entrant cavity tubes with refrigerant and
lubricant mixtures. They concluded that tubes with smaller
cavity mouth widths performed better at low heat fluxes
whereas tubes with larger cavity mouth widths performed better
at higher heat fluxes. They also observed that at higher heat
fluxes the enhancement was relatively lower.
Chien & Webb [10-13] in 1998 developed re-entrant
cavities by covering finned tubes with pored foils. They
performed a parametric study of the pore diameter, tunnel pitch,
tunnel width and fin height and concluded that a greater fin
height and a smaller tunnel pitch resulted in better performance.
They also concluded the evaporation of the liquid filled in the
corners of sharp edged tunnels was responsible for subsurface
heat transfer. They recommended finned tubes with rectangular
bases and fin heights of 0.7-1.0 mm. Kim & Choi [14]
fabricated similar re-entrant cavities consisting of pores with
subsurface connections. Observed an enhancement of 5.0-6.5
times over plain tubes performance and concluded that the
subsurface connections continuously supplied liquid to the
heated surfaces and delayed dry-outs. Kulenovic et al. [15] used
structured pores and observed enhancements at low heat fluxes.
Chen et al. [16] conducted a parametric study over the channel
widths and elliptical pore dimensions. They observed 2-4 times
performance enhancement over plain tubes. Chien & Huang
[17] used a brass wire mesh cover over the finned tube and
observed enhancements of up to 7-8 times with R134a at 5 °C
saturation temperature. Gorenflo et al. [18-20] and Kotthoff et
al. [21] developed and extensively studied the macro and micro
re-entrant cavities. These surfaces showed a significant increase
in the number of nucleation sites. They also observed a
circumferential variation in temperature while testing in the
horizontal orientation and observed lower wall superheats near
the bottom of the tube. They observed a 45 % increases in heat
transfer rates and concluded cavity structures with narrower
mouth openings performs better which was similarly observed
by Ji et al. [9] in their study.
Another commonly used technique to enhance the heat
transfer is by application of a porous layer coating over the
boiling surface. The use of porous layer significantly increases
the number of active nucleating sites on the surface. Hsieh &
Yang [22] in 2001 studied the heat transfer mechanism in a
porous layer matrix. Their results confirmed the previous
speculations of nucleation and vaporization taking place in the
porous matrix under steady boiling conditions. They also
observed an effect of porous layer thickness on the heat transfer
performance. Cieslinski [23] conducted an extensive parametric
study of porous layer coatings of different materials such as
copper, aluminum, molybdenum, zinc and brass. He used water
as the working fluid and concluded aluminum coatings offered
the greatest heat transfer enhancement. He also observed that
the boiling commenced at much lower wall superheats of
around 0.1 °C. His data showed a maximum heat transfer
coefficient of approximately 48 kW/m2·K. Kim et al. [24] used
a porous layer coating over a thin platinum wire and observed
an enhancement of 3.75 times compared to the plain wire. They
concluded that the porous layer coatings reduced the bubble
departure diameter and increased the bubbling frequency.
Dominiczak & Cieslinski [25] tested porous aluminum coatings
over stainless steel tubes with distilled water. Lower
circumferential variation in temperature was observed by the
application of porous layer coating over the top surface of the
tube to increase its heat transfer rate. Few researchers [2, 8, 26]
used the commercially available Highflux porous layer coated
tubes and observed high heat transfer enhancements of up to 10
times at very low heat fluxes. These researchers have similarly
observed a decline in the enhancement factor with increasing
heat flux.
The literature of re-entrant cavities and porous layer
coating shows significant heat transfer enhancements at lower
heat fluxes. In most cases the enhancement decreases as the
heat flux increases. Also the heat transfer coefficients observed
in literature using these enhancements are in the range of 10-50
kW/m2·K. It has been observed that none of these enhancement
techniques aid in increasing the critical heat flux limit.
Studies were conducted by Hsieh & Yang [22] and Ribatski
& Jabardo [27] by varying the average roughness on the boiling
surface and observed better performance with higher average
roughness.
Kang [28] in 2003 performed a series of
experiments by varying the inclination angle of the tube from
horizontal to vertical orientations. Water was used as the
working fluid over a stainless steel tubular surface. Relative
enhancement of up to 5-6 times was observed at an inclination
of 15° from the horizontal. He concluded the reduction in the
bubble slug formation at an inclination reduced the wall
superheat and enhanced the heat transfer performance. Chien &
Webb [10] also achieved a higher heat transfer performance in
horizontal orientation compared to vertical orientation.
Microchanneled surfaces or integral finned tubes have
broadly shown two times the heat transfer enhancements over
plain tubes. Many researchers [1-3, 6, 7, 9] have tested
unmodified integral finned tubes and observed good
enhancements in the heat transfer rates. Saidi et al. [29] tested
microchannel grooves with R123 over a copper surface. They
tested two tubes, one with higher channel density and other
with lower channel density but having nearly similar channel
depths. Results showed a stable enhancement factor of up to 2.4
times was achieved with the lower density channeled tube.
Cooke & Kandlikar [30] in 2011 used microchannel
grooves over a flat surface and observed significant
enhancement in the heat transfer coefficient with increasing
heat flux. Testing was performed with water to obtain heat
transfer coefficients of the order of 70 kW/m2·K. They
concluded the re-wetting of the heated surface and the bubble
dynamics are directly responsible for the superior performance.
NOMENCLATURE
𝐼
current supplied, A
𝐿
test section length, m
𝑇𝑆
surface temperature, K
𝑇𝑎𝑣𝑒
average temperature, K
2
Copyright © 2012 by ASME
𝑉
ℎ
𝑘
𝑞𝑎,𝑙
𝑞ℎ
𝑞𝑟
𝑟1
𝑟2
Figure 2. shows a CAD model of the experimental setup
which was modeled using SolidWorks® 3D design software.
The figure shows the location of the test section assembly and
the auxiliary heater in the vertical orientation. The overall
experimental setup dimensions were 180 mm × 180 mm × 100
mm. A 10 mm thick, high temperature resistant borosilicate
glass was used for the windowed regions of the experimental
setup. The experimental setup was laterally compressed using
ten M10 bolts with the aluminum compression plates as shown
in the Fig. 2. Silicone gaskets were used on either side of each
of the glass windows to ensure a leak free setup after
compression.
voltage applied, V
heat transfer coefficient, W/m2·K
thermal conductivity, W/m·K
axial heat losses, W
total heat supplied, W
resultant radial heat, W
thermocouple location radius, m
test section outer surface radius, m
EXPERIMENTAL SETUP
An experimental setup to perform pool boiling over tubular
surfaces was designed. Water was chosen as the working fluid
because of its well-known thermal properties and safety in
handling compared to that of a refrigerant. The experimental
setup was designed so as to have the ability to test in both,
horizontal and vertical orientations of the tube.
The setup consisted of a test section assembly which
housed the circular tube test section. An auxiliary heater was
used in the setup to maintain the temperature of water at
saturation conditions. Large windowed regions were provided
in the experimental setup to allow visual access to the surface
of the test section. Figure 1. shows a schematic of the
experimental setup detailing some of the key components. Two
separately controlled power supplies were used for conducting
the experiments. A 3.3 kW and a 1.5 kW power supplies drove
the test section and the auxiliary heaters respectively. The
heaters used in the setup were FIREROD® cartridge heaters
from Watlow®. The test section heater was rated for 400 W
heat output at 120 V and the auxiliary heater was rated for 200
W heat output at 120V. The test section heater had a diameter of
0.375” (9.525 mm) and a heated length of 0.75” (19.05 mm).
The schematic shows four temperature sensors located inside
the test section and one located in the pool and the details are
discussed in the data acquisition and reduction section.
Figure 2: CAD model of the experimental setup with the test
section in the vertical orientation.
The test section assembly was designed to yield minimal
heat losses. Figure 3. is an exploded view of the test section
assembly that shows the various components in the assembly.
The test sections were made using copper alloy 101, which has
a thermal conductivity of 391 W/m·K at 20°C. The test section
was designed to have a length of 20 mm and an outer diameter
of 15 mm. As shown in the assembly the test section heater was
inserted inside the tubular test section. To ensure a tight fit the
inner surface of the test section was accurately machined to a
diameter of 9.52 mm. Thermally insulating high temperature
ceramic was used on either side of the test section to minimize
the heat losses in the axial direction. Compressible gaskets
were used in the assembly as shown in the Fig. 3. The test
section assembly is axially compressed for sealing.
Figure 1: Schematic of the experimental setup with the test section
in the horizontal orientation.
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Copyright © 2012 by ASME
Figure 4: (a) Microchanneled test section (b) Details of geometric
parameters for the microchannels.
Table 1: Dimensional details of the rectangular microchannels for
the different test sections.
Test
Section
Channel
Depth
(mm)
Channel
Width
(mm)
Fin
Width
(mm)
Pitch
(mm)
Area
Enhancement
Factor
P0
-
-
-
-
1.0
RRM1
0.300
0.375
0.225
0.600
1.95
RRM2
0.375
0.375
0.225
0.600
2.18
RRM3
0.250
0.375
0.325
0.700
1.67
RRM4
0.400
0.375
0.325
0.700
2.06
Figure 3: Exploded view of the test section assembly.
TEST SURFACES
The test surfaces were manufactured by grooving
structured microchannels into the outer surface of a plain test
section. These microchannels were radially oriented as shown
in Fig. 4(a). Rectangular cross section geometry was used to
create these structured microchannels. The microchannels are
defined by their depth, channel width and fin width as shown in
Fig. 4(b).
Test sections consisting of different microchannel
dimensions were manufactured and tested to study their effects
on the pool boiling heat transfer performance. Table 1. lists the
dimensional details of the four microchanneled test sections
which were labeled as Radial Rectangular Microchannel
(RRM) 1 through 4. Modifications on the surface of the test
section result in an indirect increase in the wetted surface area.
The area enhancement factor is defined as the ratio of the
wetted surface area of the microchanneled test section to that of
the plain test section and is given in Table 1.
The test sections were machined using a Proto TRAK™
CNC lathe to achieve tolerances of less than ±15 µm on the
microchannel dimensions. Each manufactured test section was
dimensionally analyzed using the confocal laser scanning
microscope from Keyence®. The analysis showed a maximum
dimensional error of up to ±10 µm. A 3D surface profile image
generated using the confocal laser scanning microscope is
shown in Fig. 5.
Figure 5: 3D surface profile image of test section RRM1 generated
using the confocal laser scanning microscope.
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Copyright © 2012 by ASME
average temperature, 𝑇𝑎𝑣𝑒 and the resultant radial heat, 𝑞𝑟
through the test section by solving the radial heat conduction
equation. The derived equation is given in Eq. (3).
DATA ACQUISITION AND REDUCTION
Data acquisition systems from National Instruments™
were used to read and record the temperature data generated in
the experimental setup. Thermocouple input module NI 9213
was used with NI cDAQ-9172 USB chassis. A LabVIEW®
Virtual Instrument program was created to read and record the
generated data. The program was also used to determine and
indicate the realization of steady state conditions in the system.
Steady state conditions were attained when the temperature
readings were constant within 0.1 °C over a period of 10
minutes. The program also indicated when the steady state
saturation conditions for the pool of water were reached.
Probe style T-type thermocouple sensors from OMEGA®
were used to measure the temperatures in the experimental
setup. These sensors were calibrated using the ice point and
boiling point calibration technique. These sensors were
positioned in a circumferential plane at a radius of 6 mm and at
a depth of 10 mm into the test section. Thermocouple sensors
𝑻𝟏 , 𝑻𝟐 , 𝑻𝟑 , 𝑻𝟒 were used to measure temperatures inside the test
section and thermocouple sensor 𝑻𝟓 was immersed in the pool
to measure the bulk temperature.
The heat supplied to the test section was varied by
changing the voltage across the test section heater. The current
supplied to the heater at a given voltage (measured at the heater
junction) was recorded to determine the total heat supplied to
the test section and is given by Eq. (1).
𝑞ℎ = 𝑉 × 𝐼
𝑇𝑆 = 𝑇𝑎𝑣𝑒 − (𝑞𝑟 ×
2𝜋𝑘𝐿
)
(3)
Figure 6: (a) Sketch indicating the heat transfer in the test section
(b) Sketch indicating the average and surface temperature radii.
The radial heat flux, 𝑞𝑟′′ over the surface area, 𝐴𝑠 of the test
section was evaluated using Eq. (4). For a microchanneled test
section the surface area was evaluated using the outer diameter
of the test section.
(1)
𝑞𝑟′′ =
A computational heat loss study was conducted using
COMSOL Multiphysics® to determine the total heat loss in the
axial direction. The losses in the test section assembly occur
only at the axial top and bottom interfaces. The heat losses were
minimal in relation to the pool boiling heat flux, because of the
high heat resistance of the ceramic insulation material. The
study evaluated an average percentage heat loss at minimum
and maximum total heat input conditions in the horizontal
orientation as 1.91 % and 0.15 % respectively, and similarly in
the vertical orientation as 1.76 % and 0.17 %, respectively. The
resultant heat supplied in the radial direction through the test
section is given by Eq. (2). Figure 6(a). shows the total heat
supplied, the axial heat loss and the resultant radial heat
transferred through the test section.
𝑞𝑟 = 𝑞ℎ − 𝑞𝑎,𝑙
ln(𝑟 ⁄𝑟 )
(4)
The temperature reading from thermocouple sensor 𝑇5 was
used to check for saturation conditions while performing the
experiment and was used to evaluate the wall superheat. The
heat transfer coefficient, ℎ at different heat flux conditions was
determined using the wall superheat in Eq. (5).
ℎ=
𝑞′′𝑟
(𝑇𝑠 − 𝑇5 )
(5)
The reduced data was used to plot the performance curves
for the different test sections under various testing conditions.
These curves are discussed in detail in the results section of this
paper. The following section details the uncertainty in the
experimental results due to the propagation of errors from the
various parameters.
(2)
In order to calculate the heat transfer performance, the
temperature at the outer surface of the test sections was
evaluated using the steady state one dimensional radial heat
conduction equation. The average temperature was determined
at the radius, 𝑟1 by taking a simple average of the temperature
readings from thermocouples, 𝑇1 , 𝑇2 , 𝑇3 , 𝑇4 . The surface
temperature was evaluated at the outer radius, 𝑟2 of the test
section as shown in Fig. 6(b). In case of the microchanneled
test sections the radius 𝑟2 was taken at the outer surface. The
surface temperature, 𝑇𝑆 was determined as a function of the
UNCERTAINTY ANALYSIS
Uncertainty analysis is an important part of any
experimental work performed. There are a few experimental
parameters where an error can originate and propagate through
the calculations into the results. To understand the uncertainty
in the results an uncertainty analysis has carefully been
performed. As indicated in the previous sections there are five
thermocouples in the experimental setup. Each of these sensors
has an error attributed to them in sensing the temperature at any
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Copyright © 2012 by ASME
given point. The experimental uncertainty value for a T-type
thermocouple can be estimated to ±0.25 °C after carefully
calibration at steady state conditions. Some of the uncertainties
originate from the dimensional measurements of the test section
and hence the uncertainty in each of these parameters was
calculated. Uncertainties in geometric parameters account for a
very small amount since very tight tolerances were used in
while manufacturing these test sections. There was a small
amount of uncertainty in the total heat supplied from the power
supply and was estimated ±0.1 % using the power supply
calibration data. The uncertainty in the thermal conductivity of
copper alloy 101 for the given operating range was estimated to
be ±1 %.
There are several standard techniques to determine the
uncertainties in an experimental result. In this study, the method
of partial sums was used and Eq. (6). gives the general equation
used to determine the uncertainty in any derived parameter, p.
Up is the uncertainty in the calculated parameter p and ua
represents the uncertainties in all the measured parameters, a i .
= √∑
1( 𝑎
𝑎
)
smaller. The uncertainties in the heat transfer coefficients for
results of all the test sections in the horizontal and vertical
orientation are shown in Figs. 12. and 13. respectively.
ℎ
{(
2
)
2
1
{(
2𝜋
)
2
(
2
)
)×𝑙 (
)
)
( )
𝑟
2
2
1⁄
2
( ) ]
𝐿
2
( )
2
( )
( −
𝐿
)
)
)
)
2
1
2𝜋
1⁄
2
2
)
{(
( )
2
2
( ) }}
𝑘
𝐿
(
2
−
) ]
Lowest Heat Flux
Highest Heat Flux
W/m ·K
%
W/m2·K
%
P0
1219.6
17.7
1641.2
4.4
RRM1
1568.5
20.2
6611.4
6.7
RRM2
995.7
16.1
6745.4
6.8
RRM3
1771.7
21.0
10923.8
8.5
RRM4
1853.8
21.6
7781.4
7.2
EXPERIMETAL PROCEDURE
One plain and four microchanneled surfaces were tested
for the scope of this paper. Distilled water was used for each
test to avoid any effects of dissolved gases on the experimental
results. Initially the experimental setup is assembled and tested
for any leaks in the system. The test section and the auxiliary
heaters were used to raise the temperature of water to the
saturation conditions. The water in the setup is allowed to boil
for some time before commencing testing so as to degas any
dissolved gasses and attain steady state conditions. A picture of
the experimental setup is shown in Figure 7.
The power supplied to the test section heater was adjusted
to deliver the desired heat flux to the test section. Also the
power supplied to the auxiliary heater was set so as to maintain
the water temperature at saturation condition. Data was logged
over an interval of 20 seconds at a sampling rate of 5
samples/sec at a constant heat flux. The power to the test
section heater is incremented or decremented depending upon
the test being performed. After attaining steady state at the new
heat flux condition the next reading is recorded. These steps are
repeated to generate temperature data in the setup over a range
of heat fluxes.
At high heat fluxes of about 900 kW/m2 care was taken
while increasing the power supplied to the heater. Precaution
was taken to check the critical heat flux condition by closely
inspecting the bubble dynamics over the test section surface
2
𝑘
Similarly the percentage uncertainty of ±0.72 % in the
radial heat flux at the outer diameter was evaluated using its
equation of uncertainty given by Eq. (8).
)
𝑟
)×𝑙 (
Test Section
(7)
(
)
Uncertainty in Heat Transfer Coefficients at
(6)
𝐿
2
(
(
2
( ) }]
)
(
(
Table 2: Uncertainties in the heat transfer coefficients for the
different test sections in the horizontal orientation.
1⁄
2
= [(
2
2
(
(
)
2
)
(9)
The equation for uncertainty in the surface temperature at
the outer diameter was derived using Eq. (3) and is given by
Eq. (7). The evaluated percentage uncertainty for the generated
data was in close proximity of ±0.62 % of the surface
temperature.
= [(
= [(
(8)
The equation for heat transfer coefficient was used to
derive its uncertainty equation and is given by Eq. (9). The
calculations showed an increase in the uncertainty in the range
from low to high heat fluxes. But importantly the evaluation of
the percentage uncertainty showed a reduction in the range
from low to high heat fluxes. The uncertainty in the heat
transfer coefficients in the horizontal orientation for the
different test sections are detailed in Table 2. The uncertainties
observed in the various parameters and the heat transfer
coefficients are in an acceptable range. Also the performance is
evaluated at very high heat fluxes where the error is relatively
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Copyright © 2012 by ASME
through the large windowed regions while performing any test.
The test section temperatures are also closely monitored for any
indications of instantaneous temperature spikes. At critical heat
flux condition the boiling heat transfer mechanism changes
from nucleate boiling to film boiling. Once the critical heat flux
condition is reached the temperature in the test section
increases to around 400-500 °C due to high heat flux input and
low heat dissipation from the outer surface of the test section.
Power to the test section heater is immediately cutoff so as to
avoid any damage to the experimental setup.
Figure 8: Performance curves for all the test sections in the
horizontal orientation.
Figure 7: Picture of the experimental setup for testing in the
horizontal orientation.
RESULTS
The results for the tested tubular test sections with micro
structured surfaces are presented in this section. Each test
section was tested in the horizontal and vertical orientations
with increasing and decreasing heat flux conditions. The
reduced data for the different test sections were graphically
compared using their respective boiling curves. Figures 8. and
9. plots the heat flux, ′′ (W/m2) as a function of the wall
superheat, ∆𝑻 (K) in the horizontal and the vertical orientations,
respectively. The results significant performance enhancements
for micro structured surfaces when compared to plain surfaces.
Results for the test section RRM3, showed the greatest
performance when compared with other test sections in both the
orientations. A heat transfer coefficient of 129.1 kW/m2·K was
obtained at a heat flux of 1095.2 kW/m2. It was also noted that
even though the area enhancement factor for test sections
RRM1 and RRM2 are higher, their relative performance was
poorer when compared to test sections RRM3 and RRM4 as
shown in Fig. 9. From the results it can be concluded that the
microchannel geometry play an important role in the overall
performance enhancement. The growth of the bubbles over the
microchannel groove when pinned to the fin tips as observed by
Cooke & Kandlikar [30] affirms this conclusion.
Figure 9: Performance curves for all the test sections in the
vertical orientation.
Figure 10. shows the effects of the channel depth on the
heat transfer performance, for test sections RRM3 (channel
depth-250 µm) and RRM4 (channel depth-400 µm) in the
horizontal and vertical orientations. The results clearly indicate
that shallower channeled test sections show superior
performance when all other dimensions are the same. This
inference can also be validated by comparing the performance
of test sections RRM1 and RRM2 in Fig. 9. From the Fig. 10 it
can also be noted that the performance of test sections is
superior in the horizontal orientation compared to the vertical
orientation. Table 3. details the heat transfer coefficients for
different test sections at their respective maximum heat flux
conditions in the horizontal and vertical orientations. Hence it
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Copyright © 2012 by ASME
can be concluded that the heat transfer performance in the
horizontal orientation is comparatively superior.
bars in these figures. From the results of the microchanneled
test sections it was observed that the critical heat flux limit was
greatly extended. The plain test section was successfully tested
only up to a heat flux of 670 kW/m2, whereas the
microchanneled test sections were tested up to 1,100 kW/m2.
Figure 10: Plot showing the effects of channel depth on the
performance in the horizontal and vertical orientation.
Table 3: Results for test sections at maximum heat flux condition.
Horizontal Orientation
Test
Section
𝑞𝑟′′
∆𝑇
2
𝑞𝑟′′
ℎ
2
Figure 11: Plot showing the effects of fin width on the performance
and the observed hysteresis in the horizontal orientation.
Vertical Orientation
2
∆𝑇
ℎ
kW/m
K
kW/m ·K
kW/m
K
kW/m2·K
P0
667.3
17.8
37.5
667.2
18.8
35.5
RRM1
1092.9
11.0
99.1
1070.7
11.2
95.6
RRM2
1082.9
10.9
99.8
1082.7
12.0
90.0
RRM3
1095.2
8.5
129.1
1093.0
10.0
109.1
RRM4
1095.2
10.1
108.1
1090.8
11.9
91.4
Figure 11. shows the effects of the fin width for the test
sections RRM2 (fin width-225 µm) and RRM4 (fin width-325
µm) in the horizontal orientation. The results showed that the
test sections with wider fin widths performed comparatively
better. Similar trend was observed for the results of test sections
RRM1 and RRM3 in the horizontal and vertical orientations.
Figure 11. also shows the hysteresis observed in the generated
results. Hysteresis was observed for all test sections in the
horizontal orientation. However the hysteresis was negligible in
the vertical orientation.
In literature it has been observed that the heat transfer
coefficient for most of the test surfaces gradually decreases as
the heat flux increases. In contrast, the results obtained for the
microchanneled surfaces show a steady increase in the heat
transfer coefficient with increasing heat flux. Figures 12. and
13. plot the heat transfer coefficient as a function of heat flux in
the horizontal and vertical orientations, respectively. The
uncertainty in the heat transfer coefficients are shown as error
Figure 12: Results of heat transfer coefficient as a function of the
heat flux in the horizontal orientation.
CONCLUSIONS
An experimental investigation of pool boiling of water
over circular tubes with micro structured surfaces was
conducted. Testing was performed in horizontal and vertical
orientations and their effects were studied. A parametric study
of the channel depth and fin width was also performed.
8
Copyright © 2012 by ASME
Tube Geometries," International Journal of Heat and Mass
Transfer, 35(8), pp. 1893-1904.
[2] Memory, S. B., Sugiyama, D. C., and Marto, P. J., 1995,
"Nucleate Pool Boiling of R-114 and R-114-Oil Mixtures from
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Shape of the Fins," International Journal of Refrigeration,
20(8), pp. 575-582.
[4] Tatara, R. A., and Payvar, P., 2000, "Pool Boiling of Pure
R134a from a Single Turbo-Bii-Hp Tube," International Journal
of Heat and Mass Transfer, 43(12), pp. 2233-6.
[5] Rajulu, K. G., Kumar, R., Mohanty, B., and Varma, H. K.,
2004, "Enhancement of Nucleate Pool Boiling Heat Transfer
Coefficient by Reentrant Cavity Surfaces," Heat and Mass
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[6] Jung, D., Kwangyong, A., and Jinseok, P., 2004, "Nucleate
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Boiling Heat Transfer Coefficients of Flammable Refrigerants
on Various Enhanced Tubes," International Journal of
Refrigeration, 28(3), pp. 451-455.
[8] Thome, J. R., and Ribatski, G., 2006, "Nucleate Boiling
Heat Transfer of R134a on Enhanced Tubes," Applied Thermal
Engineering, 26(10), pp. 1018-31.
[9] Wen-Tao, J., Ding-Cai, Z., Nan, F., Jian-Fei, G., Numata,
M., Guannan, X., and Wen-Quan, T., 2010, "Nucleate Pool
Boiling Heat Transfer of R134a and R134a-Pve Lubricant
Mixtures on Smooth and Five Enhanced Tubes," Journal of
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[10] Chien, L.-H., and Webb, R. L., 1998, "Visualization of
Pool Boiling on Enhanced Surfaces," Experimental Thermal
and Fluid Science, 16(4), pp. 332-341.
[11] Liang-Han, C., and Webb, R. L., 1998, "A Parametric
Study of Nucleate Boiling on Structured Surfaces. I. Effect of
Tunnel Dimensions," Transactions of the ASME. Journal of
Heat Transfer, 120(4), pp. 1042-8.
[12] Chien, L.-H., and Webb, R. L., 1998, "A Parametric Study
of Nucleate Boiling on Structured Surfaces, Part Ii: Effect of
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120(4), pp. 1049-1054.
[13] Chien, L. H., and Webb, R. L., 2001, "Effect of Geometry
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Pore Enhanced Boiling Surfaces," Journal of Enhanced Heat
Transfer, 8(5), pp. 329-339.
[14] Nae-Hyun, K., and Kuk-Kwang, C., 2001, "Nucleate Pool
Boiling on Structured Enhanced Tubes Having Pores with
Connecting Gaps," International Journal of Heat and Mass
Transfer, 44(1), pp. 17-28.
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Speed Flow Visualization of Pool Boiling from Structured
Tubular Heat Transfer Surfaces," Experimental Thermal and
Fluid Science, 25(7), pp. 547-555.
Figure 13: Results of heat transfer coefficient as a function of the
heat flux in the vertical orientation.
The following conclusions were drawn after analyzing the
experimental results.
 The micro structured test surfaces showed a significant
enhancement over the performance of a plain surface.
 It was also concluded that the heat transfer coefficient
enhancement factor increased with increasing heat flux
with microchanneled surfaces even at very high heat
fluxes.
 The microchanneled test sections showed better
performance in the horizontal orientation compared to the
vertical orientation. In the horizontal orientation, the heat
transfer coefficient enhancement factors for different test
sections were found to be within the range of 2.0-2.6,
whereas in the vertical orientation they ranged between
1.8-2.1 at an approximate heat flux of 670
 kW/m2.
 The best performing test section RRM3, reached a
maximum heat transfer coefficient of 129.1 kW/m2·K at a
heat flux of 1095.2 kW/m2.
 The parametric study for the channel depth showed that
microchannel grooves with smaller depths aided in
enhancing the performance of the surface. It was also
concluded that the fin with wider forms also showed better
performance.
 It was observed that the bubble dynamics over the
microchanneled surfaces played an important role in the
overall performance enhancement.
 The critical heat flux limit was successfully extended from
700 kW/m2 for plain surfaces to over 1,100 kW/m2 using
micro structured surfaces.
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9
Copyright © 2012 by ASME
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Copyright © 2012 by ASME