International Journal of Heat and Mass Transfer 64 (2013) 1205–1215
Contents lists available at SciVerse ScienceDirect
International Journal of Heat and Mass Transfer
journal homepage: www.elsevier.com/locate/ijhmt
Pool boiling heat transfer enhancement over cylindrical tubes
with water at atmospheric pressure, Part I: Experimental results
for circumferential rectangular open microchannels
Jeet S. Mehta, Satish G. Kandlikar ⇑
Department of Mechanical Engineering, Kate Gleason College of Engineering, Rochester Institute of Technology, 76 Lomb Memorial Drive, Rochester, NY 14623, USA
a r t i c l e
i n f o
Article history:
Available online 4 May 2013
Keywords:
Pool boiling
Heat transfer enhancement
Open microchannels
Cylindrical tube
Rectangular groove
Critical heat flux
a b s t r a c t
A two-part experimental study is conducted on pool boiling heat transfer over enhanced cylindrical
microchannel test surfaces with water at atmospheric pressure. The objective of this work is to investigate the heat transfer enhancement and study the effects of geometric parameters on the pool boiling
performance of the open microchannel surfaces over circular tubes. The effects of the horizontal and vertical orientation on heat transfer enhancement are also studied. In this part of the study, the results for
the circumferential rectangular microchannels are presented. A maximum heat transfer coefficient of
129 kW/m2 K was achieved with test surface CRM3 in the horizontal orientation at a heat flux of
1095 kW/m2. The corresponding values for the vertical orientation are 109 kW/m2 K and 1093 kW/m2,
respectively. The critical heat flux limit was also extended by a factor of 1.6 or more over a plain tube.
Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction
Over the past few decades, extensive research towards augmenting the nucleate boiling heat transfer has been conducted. A
number of different techniques have been employed to enhance
the pool boiling heat transfer rates. These techniques can be
broadly classified into the following four main categories – re-entrant cavities, porous surfaces, surface roughness and tube orientation, and microchannels/integral fins. Table 1 shows a
comprehensive summary of some of the representative literature
related to heat transfer enhancement over tubular surfaces. It summarizes the enhancement technique employed to modify the surface, the material, diameter and length of the tested tube,
operational heat flux range and the overall enhancement factors
observed by different researchers.
Heat transfer enhancement using re-entrant cavities has been
widely studied in literature. These re-entrant cavities consist of
trapped vapor which act as nucleation sites and aid in the nucleation process thereby augmenting the heat transfer performance
by approximately 3–4 times. Numerous techniques of generating
re-entrant cavities have been developed and tested by various
researchers in the previous three decades. In 1988, Ayub and Ber-
Abbreviations: CHF, critical heat flux; CRM, circumferential rectangular microchannels; ONB, onset of nucleate boiling.
⇑ Corresponding author. Tel.: +1 585 475 6728.
E-mail address: [email protected] (S.G. Kandlikar).
0017-9310/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved.
http://dx.doi.org/10.1016/j.ijheatmasstransfer.2013.03.087
gles [1] tested GEWA-T re-entrant cavity tubes with water. In their
study they reached a heat flux of 80 kW/m2 with a wall superheat
of 4 K, which yielded a heat transfer coefficient of approximately
20 kW/m2 K. Webb and Pais [2] in 1992 tested four GEWA series
tubes and a TURBO-B tube. The re-entrant cavities on these commercially available tubes are generated by performing further processes on the finned tubes. They tested these surfaces with R11,
R12, R22, R123 and R134a at saturation temperatures of 4 °C
(39 °F) and 27 °C (80 °F) and concluded that higher heat transfer
coefficients were obtained at higher saturation temperatures of
27 °C (80 °F). Memory et al. [3] used re-entrant cavity tubes such
as the commercially available GEWA series, Thermoexcel and Turbo tubes. Their experimental results showed up to 5.5 times heat
transfer coefficient enhancement for the GEWA series tubes and
up to 20 times enhancement for the Thermoexcel and the Turbo
tubes, but only at low heat flux conditions. At higher heat fluxes
the performance deteriorated significantly and was fairly similar
for all tubes. Huebner and Kuenstler [4] tested similar tubes with
n-hexane and propane and observed enhancements in the range
of 2.4–4 times. Tatara and Payvar [5] used R134a with TURBOBII-HP tube and reported 60–90% further increase in performance
compared to the previous generation TURBO-B tube. Rajulu et al.
[6] 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 observed a slight drop in the enhancement factor with an increase in the heat flux condition. They developed a
correlation for the enhancement factor as a function of the heat
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J.S. Mehta, S.G. Kandlikar / International Journal of Heat and Mass Transfer 64 (2013) 1205–1215
Nomenclature
As
h
I
k
L
qa;l
qh
qr
q00r
r1
r2
projected surface area at the outer diameter, m2
heat transfer coefficient, W/m2 K
current supplied, A
thermal conductivity of copper, W/m K
test section length, m
axial heat losses, W
total heat input, W
resultant radial heat output, W
resultant radial heat flux at the outer diameter, W/m2
radius for thermocouple location inside the test
section, m
test section outer surface radius, m
T1–T4
Ts
Tave
Ts
DT
UT S
U 00qr
Uh
V
temperature inside test section at locations 1–4, °C
bulk liquid temperature, °C
average temperature inside the test section at radius r1,
°C
surface temperature at the outer diameter, °C
wall superheat, K
uncertainty in the surface temperature, °C
uncertainty in the resultant radial heat flux, W/m2
uncertainty in the heat transfer coefficient, W/m2 K
voltage applied, V
Table 1
Summary table of the available literature for the various pool boiling heat transfer enhancement techniques over cylindrical surfaces.
Authors/year
Material/diameter
(mm)/length (mm)/
working fluid
Heat flux
range
(kW/m2)
Enhancement technique
Enhancement
factor
Comments and conclusions
Webb and
Pais [1]/
1992
Memory et al.
[2]/1995
Copper/17.5 and 19.1/
152.4/R11, R12, R22,
R123, R134a
Copper/15.9/190/R114,
R114 oil mixtures
3–80
1.4–2
1.4–2
The heat transfer coefficient increased with an increase in
saturation temperature at a given heat flux.
2–20
3.6–18
1.7–4
Steady performance drop was observed for all the re-entrant
cavity tubes and porous surface tubes with an increase in the
heat flux.
Huebner and
Kuenstler
[3]/1997
Tatara and
Payvar [4]/
2000
Rajulu et al.
[5]/2004
Copper/14.55–15.81/
200/n-hexane, propane
2–30
2.4–4
1.6
T-shaped and Y-shaped re-entrant cavity tubes showed good
heat transfer performance at intermediate heat flux conditions.
Copper/19.05/211.51/
R134a
8–41
Re-entrant cavities
Microchannels/integral
fins
Re-entrant cavities
Porous surface
Microchannels/integral
fins
Re-entrant cavities
Microchannels/integral
fins
Re-entrant cavities
4.9–7.9
TURBO-BII-HP tube showed 60–90% enhancement in the heat
transfer over standard TURBO-B tube with R134a.
Brass/33/218/
isopropanol, ethanol,
acetone, water
11–42
Re-entrant cavities
1.2–2.65
Jung et al. [6]/
2004
Copper/18.6–18.8/152/
R22, R134a, R125, R32
10–80
1.64–8.77
1.09–1.68
Jung et al. [7]/
2005
Copper/18.6–18.8/152/
R1270, R290, R600,
R600a, RE170
Copper/19.05/554/
R134a
10–80
Re-entrant cavities
Microchannels/integral
fins
Re-entrant cavities
Microchannels/integral
fins
Re-entrant cavities
Porous surface
Acetone and isopropanol performed well on the tubes having
cavity mouth width of 0.3 mm, whereas ethanol and water
performed well on the tubes having cavity mouth width of
0.2 mm.
THERMOEXCEL-E tube showed highest heat transfer
enhancement.
2–9.4
1.2–2.4
40% higher enhancement using flammable refrigerants compared
to halogenated refrigerants.
1.8–7
4.9–21.3
–/18.50–19.09/1088–
1100/R134a
9–90
4
1.6
Chien and
Webb
[10]/1998
Chien and
Webb
[11,12]/
1998
Chien and
Webb
[13]/2001
Kim and Choi
[14]/2001
Copper/18–19.5/140/
Methanol
2–70
Re-entrant cavities
Microchannels/integral
fins
Re-entrant cavities (with
pored foil) tube orientation
At low heat fluxes the porous surfaces showed superior
performance whereas at higher heat fluxes the performance
dropped and was similar to other tubes.
Re-entrant cavity tubes with narrower mouth widths performed
well at lower heat fluxes whereas the tubes with wider mouth
widths performed well at higher heat fluxes.
10–20% lower heat transfer rates were observed in the vertical
orientation compared to horizontal orientation.
Copper/19.1/140/R11,
R123
2–70
Re-entrant cavities (with
pored foil)
–
Copper/18.5–19.1/140/
R134a, R22
2–80
Re-entrant cavities (with
pored foil)
–
Copper/18.8/170/R11,
R123, R134a
1–50
5–6.5
Kulenovic
et al. [15]/
2002
Chen et al.
[16]/2004
Copper/19/115/propane
Up to
100
Re-entrant cavities (with
pores and connecting
gaps)
Re-entrant cavities (with
structured pores)
Carbon steel/19/115/
propane, iso-butane
2–30
Re-entrant cavities (with
elliptical pores and
subsurface tunnels)
2–4
Ribatski and
Thome
[8]/2006
Ji et al. [9]/
2010
0.5–100
20–70
2.5 (approx.)
2–3
Greater tunnel height and smaller tunnel pitch were preferred for
achieving higher heat transfer enhancements. Dry-outs in the
tunnels were observed at certain heat fluxes when the liquid in
the tunnels was depleted.
Sharp rectangular corners at the fin base provided better
performance over circular bases. R22 yielded 100% greater
enhancement with rectangular bases compared to circular bases.
Subsurface connected gaps acted as an additional route for the
liquid supply and delayed the dry-outs in the tunnel.
Enhanced heat transfer performance was observed in the low to
medium heat flux ranges. Deactivation of the re-entrant cavities
was noted at higher heat flux conditions.
They concluded that the enhancements were due to the
evaporation of liquid film inside the tunnels.
J.S. Mehta, S.G. Kandlikar / International Journal of Heat and Mass Transfer 64 (2013) 1205–1215
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Table 1 (continued)
Authors/year
Material/diameter
(mm)/length (mm)/
working fluid
Heat flux
range
(kW/m2)
Enhancement technique
Enhancement
factor
Comments and conclusions
Chien and
Huang
[17]/2009
Kotthoff et al.
[18]/2006
Copper/19/100/R134a
Up to 55
Re-entrant cavities (with
wire mesh)
7–8
Good enhancements with copper mesh wrapped around finned
tubes. Higher fin density and height were advantageous.
Copper/25.4/-/R134a, 2propanol, propane
Up to
100
Re-entrant cavities
(micro cavities)
1.35–1.45
Gorenflo et al.
[19,20]/
2010
Copper/25/-/R125
Up to
100
Re-entrant cavities
(micro cavities)
–
Hsieh and
Yang [21]/
2001
Cieslinski
[22]/2002
Copper/20/210/R134a,
R600a
0.1–30
1.2–2.3
Stainless steel/7.88–
23.57/250/distilled
water
Platinum/0.39/40/FC72
20–1030
Porous surface (with
surface roughness and
wrapped helical wire)
Porous surface (with
surface roughness)
Up to
400
Porous surface
6 (approx.)
Stainless steel/8.15–
23.60/250/distilled
water, R141b
0.1–100
Porous surface
-
Copper alloy
(90Cu:10Ni)/19/56/
pentane
Copper, brass, stainless
steel/19/255/R11, R12,
R22, R123, R134a
Stainless steel/12.7–
19.1/540/water
10–50
Porous surface
5
They observed circumferential temperature distribution from
intermediate up to high heat fluxes. Optimized the re-entrant
cavity surface by narrowing mouth openings.
Additional evaporation into the bubbles sliding upwards along
the superheated boundary layer caused the circumferential
temperature variation with a minimum superheat developing on
the lower part of the tube at intermediate heat flux conditions.
Nucleation and vaporization took place inside the porous matrix.
Heat transfer enhancement was affected by the porous layer
thickness.
Nucleation commenced at wall superheats as low as 0.1 K.
Aluminum porous coating showed better performance than
porous layers of other materials.
Microporous coatings augmented the performance through
increased latent heat transfer at low heat fluxes and increased
convective heat transfer at high heat fluxes.
They reduced the circumferential temperature distribution
observed in the horizontal orientation by partially coating the top
region of the tube with porous layer coating. Their technique also
reduced the average wall temperatures.
Good heat transfer enhancements with HIGHFLUX tubes over
their operating heat flux range.
0.6–120
Surface roughness
–
15–90
Tube inclination angle
1.25–1.4
(approx.)
1.5–70
Microchannels/integral
fins
1.3–2.4
Kim et al.
[23]/2002
Dominiczak
and
Cieslinski
[24]/2008
McNeil et al.
[25]/2002
Ribatski and
Jabardo
[26]/2003
Kang [27]/
2003
Saidi et al.
[28]/1999
Copper/17.1/554/R123
flux and the cavity width of the re-entrant channels. Jung et al.
[7,8] 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
coefficient compared to the increase in the heat flux was small,
possibly because of the blockage of liquid re-entry through the
pores into the tunnels at higher heat fluxes. Their experimental
data showed 40% greater enhancements for flammable refrigerants
compared to halogenated refrigerants. Ribatski and Thome [9] used
GEWA-B, TURBO-CSL and TURBO-BII-HP tubes and obtained
enhancement factors of 2.4–5.2, 2.4–2.9 and 1.9–7.0, respectively,
in the high to low heat flux range. Ji et al. [10] tested four re-entrant cavity tubes with refrigerant and lubricant mixtures. They
concluded that tubes with narrower cavity mouth widths performed better at low heat fluxes whereas tubes with wider cavity
mouth widths performed better at higher heat fluxes. They also observed that at higher heat fluxes the enhancement factor was relatively poor.
Chien and Webb [11–14] developed re-entrant cavity surfaces
by wrapping pored foils around the outer surface of finned tubes.
They performed a parametric study on 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 that 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 for enhancing the heat
transfer performance. Kim and Choi [15] fabricated similar re-entrant cavity surfaces consisting of pores with subsurface connec-
2.6 (approx.)
More active nucleation sites on the top surface of the tube
compared to the bottom. Natural convection might be the
responsible for such behavior.
Decrease in bubble slug formation at the bottom of the tube and
concluded that the inclination angle was partially responsible for
enhanced performance.
They studied two microchannel geometries and concluded higher
density channeled tubes showed comparatively better and
consistent performance over the entire heat flux range.
tions and observed enhancements of 5–6.5 times, and concluded
that the subsurface connections continuously supplied liquid to
the heated regions and delayed dry-outs. Kulenovic et al. [16] used
structured pores and observed good enhancements in the low to
medium heat flux ranges. But under high heat flux conditions, they
observed deactivation of the re-entrant cavities. Chen et al. [17]
conducted a parametric study of the channel widths and the elliptical pore dimensions. Their results showed 2–4 times performance
enhancement over plain tubes. Chien and Huang [18] used a brass
wire mesh cover over the finned tube and observed enhancements
of up to 7–8 times with R134a at a saturation temperature of 5 °C
(41 °F). Kotthoff et al. [19] and Gorenflo et al. [20,21] developed
and extensively studied macro re-entrant cavities with simple
shapes to understand the mechanisms behind heat transfer
enhancement over the micro cavities within the roughness texture
of the surface. 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.
Their results showed heat transfer enhancements up to 45%. They
concluded re-entrant cavities with narrower mouth openings performed better as was concluded by Ji et al. [10] in their study.
Another commonly used technique to enhance the boiling
heat transfer is by generating a porous surface over the heated
tube. The porous surfaces provide significantly higher number
of active nucleation sites. Hsieh and Yang [22] in 2001 studied
the heat transfer mechanism in a porous layer matrix. Their results confirmed the earlier hypothesis of nucleation and evaporation taking place in the porous matrix under steady boiling
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J.S. Mehta, S.G. Kandlikar / International Journal of Heat and Mass Transfer 64 (2013) 1205–1215
conditions. Cieslinski [23] conducted an extensive parametric
study using porous layer coatings of different materials such as
copper, aluminum, molybdenum, zinc and brass. He used water
as the working fluid and concluded that aluminum coatings offered the greatest heat transfer enhancement. He also observed
that the boiling process commenced at much lower wall superheats of approximately 0.1 K compared to 8 K which was observed with a smooth tube. Kim et al. [24] applied a porous
layer coating over a thin platinum wire and observed an
enhancement of up to 3.75 times. They concluded that the porous layer coatings reduced the bubble departure diameter and
increased the bubbling frequency. Dominiczak and 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 the localized heat transfer
rates. A few researchers [3,9,26] used the commercially available
HIGHFLUX porous surface tube and observed high heat transfer
enhancements of up to 10 times at low heat fluxes. These
researchers have similarly observed a decline in the enhancement factor with an increase in the heat flux.
Surface roughness and tube orientation have shown to have a
small effect on the heat transfer enhancement. Hsieh and Yang
[22] and Ribatski and Jabardo [27] conducted experiments by varying the average roughness on the boiling surface. Their experimental results showed slight enhancements in the heat transfer
performance for higher average roughness surfaces. Kang [28] in
2003 performed a series of experiments by varying the inclination
angle of the tube from horizontal to vertical orientations in steps of
15°. Water was used as the working fluid over a stainless steel
cylindrical surface. Relative enhancement of up to 1.4 times was
observed at an inclination of 15° from the horizontal. He concluded
that the reduced bubble slug formation at the bottom surface of
the tube was due to the slight inclination. Chien and Webb [11]
also achieved a better heat transfer performance in horizontal orientation compared to vertical orientation.
The literature on re-entrant cavities and porous layer coating
shows significant enhancements in the heat transfer performance
but only at low heat flux conditions. In most cases the
enhancement factor 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 for cylindrical
surfaces in either the horizontal or vertical orientation. It has
also been observed that none of these enhancement techniques
aid in extending the critical heat flux limit using the modified
surfaces.
Microchannel surfaces or integral finned tubes have broadly
shown heat transfer enhancements of up to 2–3 times over plain
tubes. Many researchers [2–4,7,8,10] have tested unmodified integral finned tubes and observed good enhancements in the heat
transfer rates. Saidi et al. [29] tested microchannel grooves over
a copper tube with R123. They tested two tubes, one with higher
channel density and other with lower channel density but having
nearly similar channel depths. They studied the microchanneled
tubes over a heat flux range of 1.5–70 kW/m2 and their results
showed a stable enhancement factor of up to 2.4 times with the
higher channel density tube.
Nakayama et al. [30,31], in 1980 conducted pool boiling experiments with R11, water, and nitrogen over flat surfaces. They employed the porous re-entrant cavity surfaces to enhance the heat
transfer performance. They were able to reduce the wall superheat
for a given heat flux by 80–90% with the use of their surfaces compared to a plain surface. They obtained a maximum heat flux of
1.2 MW/m2 with water, yielding a heat transfer coefficient of
approximately 300 kW/m2 K. Cooke and Kandlikar [32,33] used
open microchannel geometries over flat surfaces with water as
the working fluid. They obtained heat transfer coefficients as high
as 269 kW/m2 K at a significantly higher heat flux of 2.44 MW/m2.
Their parametric study of the geometric microchannel dimensions
showed that wider and deeper channels with thinner fins performed comparatively better. They concluded that the re-wetting
of these heated surfaces through the open microchannel network
and the bubble dynamics over them were directly responsible for
the superior performance. Keeping in mind the performance
enhancements that could be obtained under high heat flux conditions using microchannel surfaces, the presented research work
was conducted for microchannel surfaces over cylindrical tubes
with water as the working fluid at atmospheric pressure.
2. Approach
2.1. Experimental setup
An experimental setup to study pool boiling of water over cylindrical tubular surfaces was designed and fabricated. The setup was
designed to test the enhanced tubes in both horizontal and vertical
orientations at atmospheric pressure. Large windows were incorporated to allow visualization of the test section to observe the
bubble dynamics over the microchannel surface. Bubble dynamics
was studied and analyzed and is presented in part II paper [34] of
this work.
A CAD model of the experimental setup using the SolidWorksÒ
3D design software is shown in Fig. 1. The overall dimensions of
the experimental setup were 200 mm 200 mm 125 mm. The
glass windows were held in place by laterally compressing the
assembly using the aluminum compression plates and a set of
ten M10 fasteners. Silicone gaskets were used on either side of
the glass windows to ensure a leak free setup and reduce localized
stresses on the glass. A 9.5 mm thick, high temperature resistant
borosilicate glass was used to ensure safe operation while conducting the experiment and withstand the numerous thermal load
cycles.
Fig. 1. CAD model of the experimental setup with the test section vertically
oriented.
J.S. Mehta, S.G. Kandlikar / International Journal of Heat and Mass Transfer 64 (2013) 1205–1215
Fig. 2. Exploded view of the test section assembly.
Fig. 1 also shows the test section assembly and the auxiliary
heater for the setup in the vertical orientation. The test section
assembly houses the main heater which was used to supply the required heat directly to the test section. The auxiliary heater was
used to increase and maintain the temperature of water at its saturation condition at atmospheric pressure. Two power supplies of
3.3 kW and 1.5 kW from TDK-Lambda having independent controls
were used to drive the main and the auxiliary heater, respectively.
The heaters used in the setup were FIRERODÒ cartridge heaters
from WatlowÒ. The main heater was rated for 400 W heat output
at 120 V and the auxiliary heater was rated for 200 W heat output
at 120 V. The heater with the highest heat output per unit area
available was selected as the main heater, and had a diameter of
9.53 mm and a heated length of 19.05 mm.
An exploded view of the test section assembly is shown in Fig. 2.
The assembly mainly consists of the test section, main heater, thermal insulation, gaskets and holding fixtures. The test sections were
made using copper alloy 101, which has a thermal conductivity of
391 W/m K at 20 °C. The plain test section was designed to be
20 mm long, so as to closely match the heated length of the main
heater. The inner and the outer diameters for the test section were
accurately machined to 9.53 mm and 15 mm, respectively. To ensure good thermal contact between the heater and the test section,
high density polysynthetic silver thermal compound from Artic SilverÒ was used between the two. To minimize heat losses from the
test section in the axial direction and maximize radial heat transfer, thermally-insulating high temperature ceramic was inserted
on either side of the test section. The assembly was axially compressed between the top and bottom aluminum plates using M4
fasteners and gaskets as required, to ensure a leak free setup. The
bottom plate was also used to fasten the test section assembly to
the experimental setup.
1209
test sections were designed and fabricated by varying the orientation of the microchannel grooves and dimensions of the geometric
parameters. The microchannel grooves were oriented either circumferentially or axially over the test section surface. Rectangular
and V-groove cross-section geometries were employed for generating the microchannels. In this part of the study, test sections having circumferentially oriented rectangular microchannel grooves
are designed, fabricated, and tested. Their results are analyzed
and a detailed discussion is given in the following section. In part
II paper [34] of this study, experimental results for circumferential
V-groove microchannel test sections and axial rectangular microchannel test sections are described and discussed in detail.
Eight enhanced Circumferential Rectangular Microchannel
(CRM) test sections were generated by varying the dimensions of
the microchannel geometric parameters. A representative CAD
model of the CRM test section is shown in Fig. 3(a). These microchannels were defined and differentiated by their depth, channel
width, and fin width as shown in Fig. 3(b). The dimensional details
of test sections CRM1–CRM8 are given in Table 2. Surface modifications result in an increase in the total wetted surface area. The
area enhancement factor is a ratio of the total wetted surface area
on a microchannel test section to that of the plain test section at
their outer diameters. The area enhancement factors for all the test
sections were evaluated using SolidWorksÒ CAD models and are
listed in the table.
The test sections and the microchannel grooves were machined
using a ProtoTRAK™ CNC lathe to achieve high accuracy in the
dimensions and tolerances within ±15 lm. Every microchannel
test section was dimensionally analyzed using a confocal laser
scanning microscope and the investigation showed an average
dimensional variation of less than 10 lm. A 3D surface profile generated by the microscope for a CRM test section is show in Fig. 4.
2.3. Data acquisition and reduction
The experimental setup consisted of five probe style T-type
thermocouples from OMEGAÒ for temperature measurement.
Thermocouples T1T4 were positioned circumferentially at a radial
distance of 6 mm and a depth of 10 mm inside the test sections. An
average temperature inside the test section at a radius of 6 mm
was evaluated using the measurements from these thermocouples.
Fifth thermocouple T5 was immersed into the pool of water to
measure the bulk fluid temperature while testing. This thermocouple was located in the central region between the test section
assembly and the auxiliary heater. A data acquisition system from
National Instruments™ was used to read and record the generated
2.2. Test surfaces
As discussed earlier in the introduction, modifications on the
surface were seen to be very effective in passively enhancing the
heat transfer performance. In this work, microchannel groove
enhancement technique was employed for this reason. The plain
test section geometry described earlier, was further machined to
incorporate the desired microchannel geometry. Twenty dissimilar
Fig. 3. (a) CRM test section, (b) key geometric parameters required to define a CRM
test section.
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J.S. Mehta, S.G. Kandlikar / International Journal of Heat and Mass Transfer 64 (2013) 1205–1215
Table 2
Dimensional details of geometric parameters for CRM test sections.
Test section
Depth (mm)
Channel width (mm)
Fin width (mm)
Pitch (mm)
Area enhancement factor
CRM1
CRM2
CRM3
CRM4
CRM5
CRM6
CRM7
CRM8
0.29
0.39
0.25
0.41
0.30
0.37
0.26
0.36
0.38
0.36
0.38
0.40
0.29
0.28
0.30
0.30
0.21
0.23
0.32
0.30
0.21
0.22
0.30
0.29
0.59
0.59
0.70
0.70
0.50
0.50
0.60
0.59
1.95
2.18
1.67
2.06
2.13
2.40
1.80
2.11
Fig. 5. (a) Sketch showing heat input and output surfaces, (b) sketch indicating the
radii for average and surface temperatures with the thermocouple locations.
Fig. 4. 3D surface profile of CRM4, generated using the confocal laser scanning
microscope.
data. A thermocouple input module NI 9213 was used with NI
cDAQ 9172 USB chassis to interface with the controller. A LabVIEWÒ Virtual Instrument program was created to read and record
the generated data. The program was designed to indicate the realization of steady state condition for each test. Testing was performed at a saturation temperature corresponding to the local
atmospheric pressure with water as the working fluid. Steady state
conditions were assumed when the fluctuations in temperature
readings from all the thermocouples were within ±0.1 °C for a period of 10 min. The program graphically displayed the temperature
data from all the five sensors in real-time. The data was recorded at
a sampling rate of 5 Hz. The voltage applied across the main heater
and the current supplied were inputted into the program to evaluate the heat flux, wall superheat, and the heat transfer coefficient
using the various equations detailed below in this section.
The heat supplied to the test section was varied by changing the
voltage v across the main heater. The current supplied to the main
heater was measured at the heater junction and recorded to determine the total heat supplied qh to the test section and is given by
Eq. (1).
qh ¼ V I
ð1Þ
A computational heat loss study was conducted to determine
the total heat losses in the axial direction using COMSOL MultiphysicsÒ. The heat loss occurred at the top and bottom faces of
the test section. The study indicated an average percentage heat
loss at minimum and maximum total heat input conditions in
the horizontal orientation of 1.9% and 0.2%, respectively, and in
the vertical orientation of 1.8% and 0.2%, respectively. The resultant
heat supplied in the radial direction through the test section is given by Eq. (2). The total heat supplied to the test section, the axial
heat lost from the top and the bottom interfaces and the resultant
radial heat outputted through the test section tubular surface are
illustrated in Fig. 5(a).
qr ¼ qh qa;l
ð2Þ
To evaluate the heat transfer performance, the temperature at
the outer surface of the test section was evaluated using the steady
state one dimensional radial heat conduction equation. The average temperature was determined at the radius r1 by averaging
the temperature readings from the four thermocouples inside the
test section. The surface temperature was evaluated at the outer
radius, r2 of the test section as shown in Fig. 5(b). For the microchannel test sections the radius r2 was taken at the outermost surface. The surface temperature Ts was determined using the average
temperature Tave and the resultant radial heat, qr by solving the one
dimensional radial heat conduction equation and is given by Eq.
(3).
lnðr2 =r 1 Þ
T s ¼ T av e qr 2pkL
ð3Þ
The radial heat flux, q00r over the projected surface area, As at the outer diameter of the test section was evaluated using Eq. (4).
q00r ¼
qr
As
ð4Þ
The bulk water temperature recorded using thermocouple T5 was
used to evaluate the wall superheat on the test section surface.
The heat transfer coefficient h at different heat flux conditions
was evaluated using Eq. (5).
h¼
q00r
Ts T5
ð5Þ
The reduced data was used to plot the performance curves for the
different test sections under various testing conditions.
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J.S. Mehta, S.G. Kandlikar / International Journal of Heat and Mass Transfer 64 (2013) 1205–1215
2.4. Uncertainty analysis
An uncertainty analysis was performed by considering the
propagation of errors. The experimental uncertainty value for measuring the temperature at a point using a T-type thermocouple was
estimated to be ±0.1 °C after carefully calibrating these thermocouples at different temperatures under steady state conditions. An
uncertainty in the total heat supplied from the power supply was
estimated to ±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%.
The method of partial sums was used to calculate the uncertainties. Eq. (6) gives the general equation used to determine the
uncertainty in any derived parameter, p. Up is the uncertainty in
the derived parameter p and U ai represents the uncertainties in
all the measured parameters, ai.
vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
u n 2
uX @p
Up ¼ t
U ai
@ai
i¼1
ð6Þ
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 in the surface temperature for all the
test sections at different heat flux conditions was in close proximity
of ±0.61%. The uncertainty in estimating the radial heat flux was
±0.72% as determined from Eq. (8).
UTs
¼
Ts
"
( 2
2 2 2 )#1=2
U ðr2 =r1 Þ
U qr 2
U T av e
1
Uk
UL
þ
þ
þ
þ
2p
T av e
qr
ðr 2 =r 1 Þ lnðr 2 =r 1 Þ
k
L
ð7Þ
U q00r
¼
q00r
"
U qr
qr
2
2 2 #1=2
U r2
UL
þ
þ
r2
L
ð8Þ
The equation for heat transfer coefficient was used to derive its
uncertainty equation and is given by Eq. (9). The uncertainty calculations for the heat transfer coefficients showed a reduction in the
percentage uncertainty in the range from low to high heat fluxes.
The uncertainties in the heat transfer coefficients for the results in
the horizontal orientation are detailed in Table 3. The performance
was mainly evaluated at high heat flux conditions where the error
was relatively small.
"
2 2
U qr 2
U r2
UL
ðT s Þ2
þ
þ
þ
qr
r2
L
ðT s T 5 Þ2
(
( 2
2 2 2 ) )
2
U qr
U ðr2 =r1 Þ
U T av e
1
Uk
UL
þ
þ
þ
þ
2p
T av e
qr
ðr 2 =r 1 Þ lnðr 2 =r 1 Þ
k
L
2 #1=2
UT5
ð9Þ
þ
Ts T5
Uh
¼
h
2.5. Experimental procedure
The entire setup was assembled and thoroughly inspected for
any leaks. Distilled water was used for conducting each test to
avoid any effects of impurities on the experimental results. The
main heater and auxiliary heater were initially operated at high
heat flux conditions to raise the temperature of water to its saturation conditions. After reaching the saturation temperature, the
water was allowed to boil for at least 15–30 min to degas the dissolved gases present in the water. Also it was assumed that the
temperature in the pool of water was uniform within an uncertainty of ±0.1 °C, since the auxiliary heater was used to assist the
main heater in maintaining the pool temperature at saturation
conditions. For all test sections, the experiment was first carried
Table 3
Uncertainties in the heat transfer coefficients for all the test sections in the horizontal
orientation.
Test section
Uncertainty in the heat transfer coefficient, Uh
Lowest q00r
2
P0
CRM1
CRM2
CRM3
CRM4
CRM5
CRM6
CRM7
CRM8
Highest q00r
W/m K
%
W/m2 K
%
1133
1456
925
1644
1720
1350
1281
1111
1069
16.5
18.7
14.9
19.5
20.1
17.8
17.5
16.3
15.9
1551
6205
6327
10218
7293
7937
7876
6710
5659
4.1
6.3
6.3
7.9
6.8
7.1
7.1
6.5
6.0
out in the horizontal orientation and then in the vertical
orientation.
The power supplied to the test section heater was adjusted to
deliver the desired radial heat flux through the test section. The
setup was maintained at those conditions until the LabVIEWÒ program indicted that the test section had attained steady state conditions. Data was logged over an interval of 20 s at a sampling rate of
5 Hz. The voltage across the test section heater was incremented
and the same procedure was followed for recording the data at
the next radial heat flux condition. These steps were repeated to
generate temperature data for a given test section over a wide
range of heat fluxes. For safety reasons the experiments were conducted only up to a radial heat flux of approximately 1100 kW/m2.
At heat fluxes above 900 kW/m2, care was taken while increasing the heat supplied to the test section so as not to reach the CHF
condition. Necessary precautions were taken to check for the CHF
condition, by closely monitoring the temperature inside the test
section and also visually monitoring the bubble growth and departure from the test section surface through the large window regions while increasing the voltage across the main heater. If CHF
condition was reached the power supplied to the main heater
was immediately shut off and sub-cooled water was injected over
the test section so as to break the vapor film encapsulating the surface. Under CHF conditions the temperature of the test section
spikes to around 400–500 °C in a short period of time, since the
heat removal capacity of the surface is greatly diminished. It was
important to avoid the CHF condition since it may have led to
the failure and destruction of the test section and the experimental
setup.
3. Results and discussion
In this part of the study, the circumferentially grooved rectangular cross-sectional microchannels over cylindrical tubes were
tested. Testing was performed with water as the working fluid at
its saturation temperature at atmospheric pressure. The objective
of these experiments was to evaluate the effect of the various
microchannel geometric parameters on the heat transfer performance. Table 2 shows the channel depth, channel width, and fin
width for all the CRM test sections tested. As discussed earlier, each
of these test sections were tested in both horizontal and vertical
orientation under increasing as well as decreasing heat flux
conditions.
3.1. Effect of tube orientation
The data reduction equations were used to evaluate the resultant radial heat flux, the wall superheat, and the heat transfer coefficient under different testing conditions. This reduced data was
1212
J.S. Mehta, S.G. Kandlikar / International Journal of Heat and Mass Transfer 64 (2013) 1205–1215
Fig. 6. Boiling curves for the CRM test sections in the horizontal orientation.
Fig. 7. Boiling curves for the CRM test sections in the vertical orientation.
used to plot the performance curves for the plain surface test section P0 and the CRM test sections. The boiling curves generated for
the test sections in the horizontal orientation are shown in Fig. 6.
The results show significant performance enhancements with
microchannel surfaces compared to a plain surface. The plain test
section (P0) was tested only up to a maximum heat flux of approximately 670 kW/m2, where it recorded a wall superheat of almost
17.8 K, yielding a boiling heat transfer coefficient of 38 kW/m2 K.
At an approximate heat flux of 700 kW/m2, the plain surface
reached its critical heat flux limit and the boiling mechanism transitioned into the film boiling regime from the nucleate boiling regime. The CRM test sections were successfully tested up to a heat
flux of 1095 kW/m2. At these high heat flux conditions a minimum
wall superheat of 8.5 K was observed with test section CRM3,
which yielded a heat transfer coefficient of over 129 kW/m2 K. This
heat transfer coefficient is noted to be the highest recorded for pool
boiling with water over a cylindrical surface in the horizontal orientation. The wall superheats observed for other CRM surfaces
were in the range of 9.6 K to 11.5 K at a heat flux of approximately
1100 kW/m2.
The results for all the test sections in the horizontal and vertical
orientation are given in Table 4. The table details the maximum
heat fluxes, the corresponding wall superheats and heat transfer
coefficients for all the test sections. The results for the test sections
in the vertical orientation are shown in Fig. 7 using the log–log
scale to provide the scatter of data points with the same relative
accuracy throughout the heat flux range. The heat transfer performance with microchannel test sections showed similar enhancements over the plain test section even in the vertical orientation.
Experimental results for the test section CRM3 again showed better performance compared to other test sections. A heat transfer
coefficient of slightly over 109 kW/m2 K was achieved with a wall
superheat of 10 K at a heat flux of 1093 kW/m2. Comparing the results detailed in the table and shown in the two figures it was clear
that the performance observed in the horizontal orientation was
approximately 10–15% greater than the performance observed in
the vertical orientation.
3.2. Enhancement in heat transfer and CHF limit
In literature it has been observed that the enhancement in the
heat transfer coefficient for some of the re-entrant cavity and porous layer coated surfaces gradually decreases as the heat flux is increased. The results obtained for microchannel surfaces in this
study, shows a steady increase in the heat transfer coefficient for
increasing heat flux as shown in Fig. 8. At a heat flux of around
670 kW/m2 enhancement factors of 1.9–2.6 in the heat transfer
coefficients were observed for the microchannel test sections in
the horizontal orientation. Similarly in the vertical orientation,
enhancement factors of 1.6–2.1 were observed at a heat flux of
around 670 kW/m2. The enhancement factors presented here do
not take into account the higher heat flux conditions up to which
the microchannel test sections were successfully tested. The overall heat transfer enhancement factors observed for each CRM test
section over the plain test section (P0) at their respective highest
heat flux conditions are given in Table 4. Overall enhancement factors of 2.5–3.4 were achieved in the horizontal orientation whereas
in the vertical orientation the enhancement ranged from 2.2 to 3.1.
Table 4
Results for all the test sections in the horizontal and vertical orientation at their highest tested heat flux conditions.
Test section
P0
CRM1
CRM2
CRM3
CRM4
CRM5
CRM6
CRM7
CRM8
Horizontal orientation
Vertical orientation
q00r
kW/m2
DT
K
h
kW/m2 K
hCRM =hP0
–
q00r
kW/m2
DT
K
h
kW/m2 K
hCRM =hP0
–
667
1093
1083
1095
1095
1069
1079
1083
1083
17.8
11.0
10.9
8.5
10.1
9.6
9.7
10.5
11.5
38
99
100
129
108
112
112
103
94
–
2.6
2.7
3.4
2.9
3.0
3.0
2.8
2.5
667
1071
1083
1093
1091
1067
1077
1083
1082
18.8
11.2
12.0
10.0
11.9
10.3
11.6
12.5
14.0
36
96
90
109
91
103
93
87
77
–
2.7
2.5
3.1
2.6
2.9
2.6
2.4
2.2
J.S. Mehta, S.G. Kandlikar / International Journal of Heat and Mass Transfer 64 (2013) 1205–1215
Fig. 8. Heat transfer coefficient comparison between the CRM test sections in the
horizontal orientation.
Fig. 8 also shows the uncertainty bars in the heat transfer coefficients for the plain test section and the best performing CRM3
test section. The uncertainty in the heat transfer coefficient for
the test section CRM3 was around ±19.5% at the lowest heat flux
condition and around ±7.9% at the highest heat flux condition. Also
noticeable in this figure is the extension of the critical heat flux
limit. As mentioned earlier, the test section P0 reached its CHF conditions at an approximate heat flux of 700 kW/m2, whereas the
open microchannel test sections were successfully tested up to a
heat flux of 1100 kW/m2. Even at these high heat flux conditions
the CRM test sections did not reach their CHF limits. Hence the
CHF limit was successfully extended to at least 1.6 times over the
limit observed for the plain surface.
1213
Fig. 9. Boiling curve comparison to analyze the effects of the channel depth.
heat flux conditions. The effect of the fin width on the performance
is explained using the results shown in Fig. 11. The boiling curves
for CRM2 and CRM4 having fin widths of 0.23 mm and 0.30 mm,
respectively, are shown in the figure. It is seen that CRM4 with
the wider fin width performed better than CRM2. Similar trend
was observed between the results for CRM1 and CRM3 in the horizontal and vertical orientations as seen from Fig. 6 and Fig. 7,
respectively. The figure also shows the hysteresis observed between the increasing heat flux results and the decreasing heat flux
results. In the horizontal orientation, some hysteresis in the results
was observed for all the CRM test sections. However in the vertical
orientation the hysteresis observed was negligible.
3.4. Area normalized results
3.3. Effects of microchannel dimensions
The CRM test sections were designed to conduct a parametric
study to analyze the channel depth, channel width and fin width
effects on the boiling heat transfer performance. The effect of the
mcirochannel depth is shown in Fig. 9. This figure shows the boiling curves for CRM3 and CRM4 in the horizontal and vertical orientations having channel depths of 0.25 mm and 0.41 mm,
respectively. Also shown in the figure are the boiling curves for
CRM7 and CRM8 having channel depths of 0.26 mm and
0.36 mm, respectively. All the other dimensions on these two pairs
of test sections were fairly similar and hence a direct comparison
between their results showed the effect of channel depth. The heat
transfer performance for CRM3 was relatively better than that obtained for CRM4 in both the orientations. Similarly CRM7 showed
comparatively better performance than CRM8 which had deeper
microchannels. Hence from these results, it was concluded that
test sections with smaller microchannel depths yielded better heat
transfer performance compared to deeper microchannels. This
inference was also validated by comparing the results for CRM1
and CRM2 from Fig. 7.
The effect of the channel width on the heat transfer performance is discussed using the results shown in Fig. 10. The results
for CRM1 and CRM5 having channel widths 0.38 mm and
0.29 mm, respectively were compared using their respective boiling curves. Similar comparison between the boiling curves of
CRM2 and CRM6 having channel widths of 0.36 mm and
0.28 mm, respectively, is shown in the figure. From these results
it was concluded that narrower microchannels showed some
enhancement in the heat transfer performance. The effect of channel width on the heat transfer rate was more prominent at higher
The heat fluxes described in all the results above were evaluated using the projected surface area at the outer diameter of the
test sections. Any modification on the surface has a direct impact
on the total wetted surface area. As described earlier, the increase
in the wetted surface area is defined by the area enhancement factor. In a number of earlier studies, the enhancements in the heat
transfer rates were achieved by increasing the wetted surface area
of the tube. To explore the effects of other enhancement factors the
results obtained were area normalized. The heat fluxes were recal-
Fig. 10. Boiling curve comparison to analyze the effects of the channel width.
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J.S. Mehta, S.G. Kandlikar / International Journal of Heat and Mass Transfer 64 (2013) 1205–1215
4. Conclusions
An experimental investigation of pool boiling of water over
cylindrical tubes with rectangular microchannel surfaces was conducted in this study. Testing was performed on eight CRM test sections and a plain test section in the horizontal and vertical
orientations. A parametric study of the channel depth, the fin
width and the channel width was performed and their effects on
the heat transfer performance were studied. The following conclusions were drawn after analyzing the experimental results.
Fig. 11. Boiling curve comparison to analyze the effects of the fin width.
culated using the total wetted surface area for each of the CRM test
sections. The boiling curves obtained using the normalized results
for the horizontal orientation data are given in Fig. 12. From these
results it can be noted that the overall heat transfer enhancements
are comparatively lower than the results derived using the projected surface area heat fluxes. But still the normalized results obtained for the microchannel test sections showed better
performance than that obtained with the plain test section.
From these results it was determined that factors other than the
surface area augmentation were responsible for the improved performance with microchannel test sections. It was concluded that
the microchannel geometry played an important role in the overall
performance enhancement for the modified surfaces. The microchannels also aided in the rewetting of the heated surface thereby
extending the critical heat flux limit for the surface. Cooke and
Kandlikar [32,33] analyzed the bubble nucleation and growth over
flat microchannel surfaces and concluded that the bubble dynamics and the rewetting phenomenon were mainly responsible for the
heat transfer enhancement. In part II paper [34] of this study, a discussion on the bubble nucleation, growth and departure from the
microchannel surfaces are observed and analyzed using videos
from high speed cameras is presented.
Significant enhancements in the heat transfer performance
were observed with microchannel test surfaces.
The best performing test section, CRM3 reached a maximum
heat transfer coefficient of 129 kW/m2 K in the horizontal orientation at a heat flux of 1095 kW/m2 while maintaining the wall
superheat of 8.5 K. In the vertical orientation the test section
was able to attain a maximum heat transfer coefficient of
109 kW/m2 K at a heat flux of 1093 kW/m2.
The CRM test sections showed 10–15% better performance in
the horizontal orientation compared to that in the vertical
orientation.
At around 670 kW/m2 representing the mid-range heat flux, the
heat transfer coefficient enhancement factors for different test
sections were in the range of 1.9–2.6 in the horizontal orientation, and 1.6–2.1 in the vertical orientation.
At the highest respective heat flux conditions for the CRM test
sections, enhancement factors in the range of 2.5–3.4 were
achieved in the horizontal orientation. In the vertical orientation the enhancement factors were in the range of 2.2–3.1.
The heat transfer coefficient enhancement factor increased with
increasing heat flux conditions for microchannel surfaces.
The critical heat flux condition for the plain test section was
reached at approximately 700 kW/m2, whereas all the microchannel test sections were successfully tested up to 1100 kW/
m2 without reaching their CHF limits. Hence an extension of
at least 1.6 times in the critical heat flux limit was observed
with the use of microchannel surfaces.
The parametric study for the rectangular microchannel grooves
showed that channels with smaller depths around 0.25 mm
aided in enhancing the performance of the surface. The effect
of the channel width was only prominent at higher heat flux
conditions, where the narrower channel widths in the range
of 0.3–0.35 mm performed slightly better. It was also concluded
that the wider fins of approximately 0.3 mm on the microchannel surfaces yielded better performance. Further study in this
area is recommended.
The normalized results showed that the performance enhancements observed with microchannel test sections were not solely
dependent on the area enhancements. The bubble dynamics
and the rewetting phenomenon also played an important role
in the overall performance augmentation.
Acknowledgements
Fig. 12. Area normalized boiling curves for the CRM test sections.
The experimental work was conducted in the Thermal Analysis,
Microfluidics, and Fuel Cell Laboratory at the Rochester Institute of
Technology in Rochester, NY. Part of this work was supported by
the National Science Foundation EPDT grant #0802100. Special
acknowledgement and thanks to Dave Hathaway, Robert Kraynik
and Jan Maneti at the RIT’s mechanical engineering machine shop
for their support.
J.S. Mehta, S.G. Kandlikar / International Journal of Heat and Mass Transfer 64 (2013) 1205–1215
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