Heat transfer engineering an international journal, tập 32, số 7 8, 2011

Heat Transfer Engineering, 32(7–8):525–526, 2011
Copyright C Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
DOI: 10.1080/01457632.2010.506388

editorial

Selected Papers from the Seventh
International Conference on
Nanochannels, Microchannels, and
Minichannels
SATISH G. KANDLIKAR
Mechanical Engineering Department, Rochester Institute of Technology, Rochester, New York, USA

It gives us great pleasure to present this special issue highlighting some of the papers presented at the ASME Seventh International Conference on Nanochannels, Microchannels, and
Minichannels, held at the Pohang University of Science and
Technology (POSTECH), in Pohang, South Korea, June 22–24,
2009. The conference was held under the sponsorship of ASME
and was co-hosted by Dr. Moo-Hwan Kim, professor and director of the Two Phase Flow Laboratory at POSTECH. On behalf
of the conference organizing committee and the participants, we
would like to thank him and his team of students and staff for
putting together a world-class event.
Pohang, home of the Pohang Steel Corporation, is a prosperous port city on the eastern side of Korea. As one of Korea’s top
universities dedicated to science and engineering, POSTECH offers 4-year programs in 10 departments and POSTECH’s Graduate School offers programs in 14 departments. The excellence
of the university extends far beyond the campus, as POSTECH
has international cooperative agreements in place with 68 sister universities. We had more than 150 papers presented over 3
days in 24 sessions. The conference theme of interdisciplinary
research was once again showcased with researchers working
in diverse areas such as traditional heat and mass transfer, labon-chips, sensors, biomedical applications, micromixers, fuel
Address correspondence to Professor Satish G. Kandlikar, Mechanical Engineering Department, Rochester Institute of Technology, James E. Gleason

Building, 76 Lomb Memorial Drive, Rochester, NY 14623-5603, USA. E-mail:


cells, and microdevices, to name just a few. Selected papers
in the field of heat transfer and fluid flow are included in this
special volume.
There are 19 papers included in this special volume.
The topics covered include review of cooling technology using microchannels, single-phase flow in microchannels with
porous/fibrous structures, boiling and bubble dynamics, Tjunction micromixers for two-phase flow, capillary filling, wetting in microgrooves with liquid metals, gas flow in rough
nanochannels, effect of ultrasound on subcooled flow boiling,
explosive boiling, flow patterns, and falling film flow on periodic structures. These topics indicate that the microchannels are
now being used in many diverse applications.
The conference organizers are thankful to all authors for
participating enthusiastically in this conference series. Special thanks are due to the authors of the papers in this special issue. The authors have worked diligently in meeting
the review schedule and responding to the reviewers’ comments. The reviewers have played a great role in improving the quality of the papers. The help provided by Enrica
Manos in the Mechanical Engineering Department at Rochester
Institute of Technology with this special issue is gratefully
acknowledged.
We thank Professor Afshin Ghajar for his dedication to this
field and his willingness to publish this special issue highlighting
the current research going on worldwide. He has been a major
supporter of this conference series, and I am indebted to him for
this collaborative effort.

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EDITORIAL

Satish G. Kandlikar is the Gleason Professor of Mechanical Engineering at Rochester Institute of Technology (RIT). He received his Ph.D. degree from the
Indian Institute of Technology in Bombay in 1975
and was a faculty member there before coming to
RIT in 1980. His current work focuses on heat
transfer and fluid flow phenomena in microchannels
and minichannels. He is involved in advanced singlephase and two-phase heat exchangers incorporating

heat transfer engineering

smooth, rough, and enhanced microchannels. He has published more than 180
journal and conference papers. He is a fellow of the ASME, associate editor
of a number of journals including ASME Journal of Heat Transfer, and executive editor of Heat Exchanger Design Handbook published by Begell House
and Heat in History Editor for Heat Transfer Engineering. He has received
RIT’s Eisenhart Outstanding Teaching Award in 1997 and Trustees Outstanding Scholarship Award in 2006. Currently he is working on a Department
of Energy-sponsored project on fuel cell water management under freezing
conditions.

vol. 32 nos. 7–8 2011


Heat Transfer Engineering, 32(7–8):527–541, 2011
Copyright C Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
DOI: 10.1080/01457632.2010.506390

A Review of Cooling in Microchannels
JAMI F. TULLIUS, ROBERT VAJTAI, and YILDIZ BAYAZITOGLU
Department of Mechanical Engineering and Material Science, Rice University, Houston, Texas, USA

Advancements in electronic performance result in a decrease in device size and increase in power density. Because of these

advancements, current cooling mechanisms for electronic devices are beginning to be ineffective. Within the localized hot
spots, the materials of the components are reaching temperature values that can lead to improper functioning of the device.
Many techniques have been successful in the past, such as heat sinks, cavities or grooves, micro pin-fins, etc., but still do not
provide adequate cooling necessary to maintain temperature values low enough for the electronic components to operate.
Microchannels, with their large heat transfer surface to volume ratio, cooled with either gas or liquid coolant, have shown
some potential. By modifying the walls of the microchannel with fins, pins, or grooves, the cooling performance can be
improved. A possible fin material used to increase the surface area of a microchannel is carbon nanotubes, which possess
excellent thermal and mechanical properties. Numerical and computational methods needed to analyze flow at the micro- and
nano-scale are also introduced. The numerical methods such as lattice Boltzmann, molecular dynamics, and computational
fluid dynamics may lessen the cost and time that often accompany experimentation.

INTRODUCTION
Electrical gadgets are continuously advancing in society by
creating larger computing power in more reduced physical dimensions than ever before. The heat produced per unit area
has increased, because of the reduction of the size of these
electronic devices. With the increase in power and heat, overheating of these electrical components has caused concern [1].
The overall well-being of the component, as well as its proper
functioning, is being threatened by the elevated temperatures.
Semiconductor components must maintain a relatively low constant surface temperature. Therefore, the development of electronic technology is limited by the efficient cooling methods
necessary to maintain the operation of the mechanisms. Many
techniques have been studied, such as thermal interface materials, heat spreaders and heat sinks, and microchannels. A method
of appropriate cooling is necessary to allow for more advancement in the years to come while maintaining proper functioning.
This paper provides a brief overview on the thermal cooling of
microchannels.
Microchannels have been proven effective in cooling small
surfaces of electrical components such as microchips. These
This work was partially supported by LANCER directed research funds
from Lockheed Martin POTT0715421 and Alliances for Graduate Education
and the Professoriate (AGEP) program through the NSF grant HRD-0450363.
Address correspondence to Professor Yildiz Bayazitoglu, Department of

Mechanical Engineering and Material Science, Rice University, 6100 Main,
Houston, TX 77005-1827, USA. E-mail:

channels act as heat exchangers or heat sinks, which can efficiently cool the microchip. The high temperatures can be dissipated through the modified surfaces of the microchannel with
natural or forced convection of the fluid flowing within the channel [2–6]. Microchannels contain a much higher heat transfer
surface area to fluid volume ratio, which allows the convection
to be enhanced when compared to the macro-scale systems. As
the hydraulic diameter decreases in a microchannel, the heat
transfer coefficient increases, providing an excellent cooling
mechanism. However, these small channels experience a very
high pressure drop. A basic microchannel with a smooth wall
surface has demonstrated to cool a heat flux of approximately
790 W/cm2 at a temperature of 71◦ C, while the pressure drop
was roughly 214 kPa. In altering the channel surface with small
cavities or fins, the performance of the channel can improve
with a slight increase in pressure drop [7, 8].
The most commonly used fluids in microchannels are air,
water, and refrigerants; however, there are limitations to their
heat transferring capabilities due to their transport properties.
Air has been a preferred fluid used in microchannels to cool
electronic components. However, with heat fluxes going beyond
100 W/cm2, air cooling methods have become inadequate for
most applications. Liquids have a much higher convection heat
transfer coefficient providing a better performance in cooling [7,
9]. Fluids with higher convection heat transfer coefficients and
higher specific heats are more effective in reducing heat from
the surface. The increase of the heat transfer coefficient or the
surface area of the finned structures can help reduce convection

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J. F. TULLIUS ET AL.

Figure 1 Thermal properties of different fluids of convection flow. Adapted from [11], [12], and [13].

resistances [10]. In Figure 1, a qualitative comparison of different heat transfer coefficients is presented [11–13]. Two-phase
systems have an advantage over one-phase systems, because of
the latent heat during the phase change process [1].
Nanofluids consist of small nanosized particles usually no
bigger than 100 nm in size in a base fluid such as water, ethylene
glycol, engine oil, or refrigerant. In recent studies, nanofluids
have emerged, having unique properties that consist of a very
high thermal conductivity and maintaining stability. Metallic
materials that have been used for these nanoparticles are oxide ceramics (Al2 O3 , CuO), nitride ceramics (AlN, SiN), carbide ceramics (SiC, TiC), metals (Cu, Au, Ag), semiconductors
(SiC, TiC2 ), carbon nanotubes, and some composite materials
(Al70 Cu30 ). The most common materials used are the oxide ceramics. These nanofluids increase the thermal conductivity and
will in turn increase the heat transfer performance. Although
adding nanoparticles to a base fluid can influence the cooling
process positively, there are still challenges. These fluids leave
sedimentation of particles, fouling, high pressure drop, and erosion, and may even clog the channel over time [14–19].
This paper reviews the effects of a microchannel with various fluids as it flows in both one and two phases. Many methods
have been proven to be efficient by thermally enhancing the
channel. Some methods include treating the surface with cavities, fins, micro pin-fins, and increasing the roughness on the
surface. Carbon nanotubes (CNT) are considered to improve the
heat dissipation in a microchannel with their excellent thermal
and mechanical properties. A section describing the mechanical behaviors of CNT will also be discussed. At the nanoscale,
Navier–Stokes equations (NS) are proven to be inaccurate because the assumptions made at the continuum level are no longer

valid. A review of the mathematical methods needed to calculate
micro- and nano-scale problems is discussed.
heat transfer engineering

SINGLE-PHASE FLUID
Single-phase fluids such as air and liquid, before it reaches
the saturation temperature, have been used in microchannels
to effectively enhance performance. Fluids flowing through
the channel can convectively transfer heat from the bottom surface and, as shown in Figure 1, different fluid properties can
influence the amount of heat that can be cooled due to their
heat transfer coefficients. Microchannels can be influenced by
many factors other than the different fluid properties, including
the shape of the channel, the surface roughness of the channel
walls, the cavities machined on the channel surface, etc.
The shapes of the cross-sectional area of the microchannels
themselves can affect cooling. Sadeghi et al. [20] examined a
laminar forced convection channel with an annular cross section
while maintaining a uniform temperature at the inner wall and
an adiabatic outer wall. With lower slip velocity, the exchange
of momentum at the liquid/solid interface is also lowered, providing a decrease of the friction factor and an increase of the
Knudsen number (Kn). Because of the thermal resistance at
the interface, the Nusselt number (Nu) in the smooth channel
decreases [20]. Nonino et al. [21] investigated a developing
laminar flow in a microchannel with different cross sections
and uniform wall heat flux. Channel cross-sectional shapes of
a rectangle, trapezoidal, and hexagonal were examined for this
study. Nonino et al. discovered that viscous dissipation close
to the entrance and the temperature dependent viscosity should
not be ignored. The Nu is greatly affected by the viscosity. The
shape of the channel can influence how well the channels performance can be; however, it does not have much influence on

the effect of pressure drop, which is mainly influenced by the
temperature-dependent viscosity. The channel shape can influence the cooling performance of the microchannel.
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Surface roughness is also a major factor in optimizing the
thermal performance. Microchannels can have smooth surface
walls or can contain small structures meant to disturb the fluid
as it flows. Shokouhmand et al. [22] studied the surface roughness effects of a fully developed, laminar, rough rectangular
microchannel analytically using the Gaussian technique. The
aspect ratio was varied from 0 to 1 and the relative roughness
from 0 to 0.15. For roughness values less than 0.01, there was
little effect on the friction factor; however, for roughness values between 0.01 and 1 with an aspect ratio of 1, there was an
increase of 11.3%, with an aspect ratio 0.5 there was a 5.5%
increase of the friction factor, and for an aspect ratio of 0.1 there
was a 1.7% increase. For the convective heat transfer coefficient,
there is a parabolic profile with the values of the lower aspect
ratio close to 0 and the higher aspect ratio near 1 being high
and the values with the aspect ratio of 0.5 reaching a minimum.
Decreasing the relative roughness of the channel, the heat transfer coefficient also decreases slightly. For a relative roughness
of 0.01, the aspect ratio has little effect. As the relative roughness value increases, so does the friction factor, while Nu is not
dependent on the roughness scale. With an increase in surface
roughness the convection heat transfer coefficient will increase
slightly [22].
A method used to impact the cooling performance is applying small grooves to the surface. Lee and Teo [8], Solovitz
[23], and Baghernezhad and Abouali [9] all adjusted the wall
surface of a microchannel with grooves. These openings can
induce more disturbances in the flow, providing a more effective cooling mechanism. When applying gaps in the surface, the

pressure drop was maintained—that is, it did not increase from a
smooth microchannel—and the heat transfer performance was
increased by roughly 12%. The spacing and the size of the
grooves are still being tested to obtain the maximum efficiency
of the channel [8]. Solovitz [23] modeled a two-dimensional
(2D) simulation with a small dimple-like groove imbedded in
the channel surface. When varying the dimensions of the cavity
and the Reynolds number, there was a 70% increase in the heat
transfer performance with only a 30% increase in pressure drop
when compared to a smooth base model using a depth/diameter
ratio of 0.4 and a Reynolds number of 1000. The depth of the
cavity was proportional to the cooling performance of the channel. Baghernezhad and Abouali [9] compared the shapes of the
grooves used to disrupt the flow. A rectangular groove and an
arc-shaped groove were compared, and it was found that both
shapes can improve the cooling performance but the arc-shaped
one is more effective. This is probably due to the aerodynamics of the flow past the gap. From these studies, grooves can
increase performance while maintaining pressure drop.
Micro pin–fins, micro-studs, pillars, and square pin–fins are
all synthetically engineered structures, usually made of silicon but also of other thermally conducting materials, which
have shown significant improvements in removing heat. These
structures protrude out of the surface, increase the wall surface
area, and interrupt the steady flow of the fluid. They can take
different shapes and sizes and be placed in different patterns
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529

Figure 2 Different shapes used for fins on the surface of the microchannel.

to improve the thermal heat transfer performance. Vanapalli

et al. [24] investigated the pillar “fin” shape, which contains
the lowest friction factor with nitrogen gas flowing through the
microchannel. These pillars are used to increase the contact between the surface and the fluid with minimal thermal resistance.
The geometries tested were circles, squares, rhombus, elliptical, eye-shaped, and sine-shaped cross sections in staggered
arrangements across the surface. Pillars with the sine-shaped
cross sections, when compared to all of the other geometries,
have the lowest friction factor. A three-dimensional (3D) representation of similar shapes that Vanapalli et al. [24] used is
shown in Figure 2. Shapes of the fins can affect the motion of
flow. When the pillars are aerodynamic in shape, there is less
separation of the fluid from the solid body, creating less thermal
resistance at the interface.
Lee et al. [25] implemented oblique fins into a microchannel
to understand the effects of the local and overall heat transfer performance and pressure drop. By introducing the oblique
silicon fins to replace the conventional microchannel heat sink
with continuous fins, the thermal boundary layer development
along the channel surface is disrupted and a secondary flow of
the fluid is created. The opening between the fins disrupts the
momentum and the trailing edge of the thermal boundary layer
of each oblique fin. This causes the leading edge to redevelop,
allowing the flow to remain in the developing state. This in turn
enhances the heat transfer performance. Also the secondary flow
can produce mixing of the flow as the fluid flows through the
fin opening, improving the performance. The heat transfer coefficient of the channel with the oblique fins was enhanced by
80% when compared to the conventional channel. Within this
investigation, Lee et al. [25] studied the pressure drop effects
of the oblique finned channel and a conventional channel. Minimal differences were obtained. With the oblique fins, and the
working conditions already described, there was a significant
heat transfer performance enhancement with little effects on
pressure drop.
A key factor that can influence the performance of heat

transfer is the thermal conductivity of the fins, pins, and micro pin–fins used on the surface of the microchannel. With a
material that has a higher thermal conductivity, the thermal
resistance is decreased and the temperature decreases. Zhong
et al. [2] investigated the effects of varying the properties of an
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array of microstructures placed along the bottom surface of the
channel. When the thermal conductivity of the microstructures
was varied, the temperature decreased and the pressure drop
remained fairly constant. With a material with a higher thermal
conductivity, the resistance at the interface decreases and the
convection heat transfer performance is increased.
Pasupuleti and Kandlikar [1] have applied many of these
factors—fluid properties, material properties, fin shape, etc.—to
their investigation where they studied the effects of refrigerant
R-123 as the working fluid in a single-chip module setup. A
single-chip module is essentially a heat sink on a silicon microchip. The silicon wall is coated with mini ellipse-shaped fins
to increase the surface area. In contrast to water, this refrigerant
is considered a safe working fluid for electronic devices that do
not require corrosion inhibiters or biocide. The results of refrigerant R-123 were compared to those of water and it was found
that they had similar results [1].
Countless modifications to the surfaces of microchannels
have been extensively studied and tested to improve the performance of the cooling devices. Surface roughness, grooves, and
microfins are among the few alterations made to the microchannels in order to remove temperature from the surface using a
single-phase laminar flow. Despite these encouraging effects,

microchannels with single-phase fluids are still not enough to
keep up with the innovations of the electronic industry.
TWO-PHASE FLOW
Phase change of a fluid can cause a substantial amount of
heat to be absorbed. When a liquid turns into a gas or a gas
into a liquid, the temperature reaches a constant temperature
until the percentage composition of the fluid is either solely
liquid or solely vapor. The evaporation of fluid can result in
the absorption of heat during the phase change between liquid
and vapor. More evaporation of the fluid flowing though the
microchannel can result in a higher heat flux [26]. The fluid
surrounding the component reaches a temperature that exceeds
the fluid saturation temperature, and vapor bubbles originate in
the small cavities or pores on the surface.
There are two types of boiling occurrences: pool boiling and
flow boiling. Pool boiling is the thermal cooling of a surface
with a stagnant fluid that can effectively remove heat. Flow
boiling refers to boiling where the liquid has a high-velocity
flow field. Both techniques are limited in the nucleate boiling
regime by the critical heat flux (CHF), which is the maximum
point of operation for engineering system. To influence the flow
boiling and pool boiling process, one should consider reducing
the boiling incipience temperature and increasing the CHF in
efforts to improve the boiling process [12].
Pool Boiling
Pool boiling is the boiling of an inert liquid. At low heat
flux levels, natural convection is dominant, but at high heat flux
heat transfer engineering

levels, nucleate superheating begins to occur. Nucleate boiling is the means by which vapor bubbles form and escape the

heated surface. A larger bubble size on a nucleation site for
a given amount of time results in a more effective thermal
performance. More liquid is being evaporated as the bubble
size increases [27, 28]. When boiling is initiated, the growth
of the bubble from one cavity extends to other nucleation sites,
causing those to initiate. The boiling spreads rapidly over the
surface, increasing the convective heat transfer coefficient and
decreasing the surface temperature [29]. Using liquids with
low boiling points, phase change from a liquid to a gas occurs
more quickly, more heat can be removed, and the pressure drop
decreases.
Nucleate boiling initiates when the temperatures of the fluids
are few degrees higher than the saturation temperature, where
small surface defects appear. As the microchip is heated, the
boiling incipience appears but only with a few vapor bubbles.
Bubbles are generated between the gaps of the fins or surface
cavities. With the evaporation of the surrounding liquid, the bubble grows in between the fins that are confining it. As the vapor
grows, the thin film surrounding the bubble evaporates with a
high heat transfer coefficient value. With its increasing size, the
bubble forces itself to the top of the fin surface. Liquid from
under the rapidly growing bubble attracts the remaining liquid
surrounding the fins. This process enhances the microconvection of the liquid along the walls of the fins. With this suction
reaction, there is an increase in evaporation, which effectively
enhances effective heat transfer. With increasing heat flux, there
is an increase in nucleation sites. When there are more nucleation sites, a domino effect is triggered, creating more vapor
bubbles [27, 28].
When compared to a smooth surface, the optimal effect of
the phase change phenomena ideally produced by microstructures or microcavities in the surface is to have a lower boiling
incipience, decrease the surface temperature, increase the CHF,
and increase the evaporation to obtain more nucleation bubbles

[28, 30]. Many factors must be considered when attempting to
optimize the performance of the cooling method. Surface roughness, orientation of the microchip relative to the flowing fluid,
and geometry or configuration of the fin arrays can all alter the
effect of cooling. A higher thermal conductivity, an enlargement
of the interface surface area, and optimization of fin placement,
geometry, and dimensions are needed to improve the efficiency
and performance of microstructures [2].
The orientation of the channel can influence the thermal
performance. A comparison of a vertically and a horizontally
mounted chip was observed for a smooth surface and a finned
surface using FC-72. The surface orientation of the chip increases the heat transfer performance in the nucleate boiling
regime as the angle increases toward a vertical position for a
smooth surface. It is believed that gravity assists the increase
in performance as the chip is mounted in a relatively vertical
direction. For a treated surface, however, the orientation of the
chip has little or no effect on the heat transfer performance
[27, 29].
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Surface roughness can also influence the thermal performance for a pool boiling process. Testing of surfaces with different degrees of roughness using FC-72 was performed by Honda
and Wei [31]. A high surface roughness lowered the boiling
incipience, increases the CHF, and increases nucleate boiling.
Shokouhmand et al. [22] studied numerically the effects of surface roughness in microchannels on convective heat transfer in
fully developed, laminar flow. With increasing relative roughness, the friction factor increases, the Nu remains unchanged,
and the convective heat transfer coefficient slightly increases.
The use of small cavities is one of the techniques that increase
the heat transfer area. In this process, an array of holes with

precise dimensions is drilled into the silicon surface; in effect,
these holes act as nucleation cavities and enhance nucleate boiling. An investigation of the thermal performance on artificial
micro cavity surfaces was conducted by Yu et al. [12] using a
dielectric fluid, FC-72 as the working fluid. The 16 × 16, 25 ×
25, and 33 × 33 arrays of microcavities were tested varying the
heat flux, diameter, and depth. When increasing the diameter of
the cavity at moderate and high heat fluxes, an earlier decay and
low peak value of the heat transfer coefficient were experienced.
Varying the cavity depth too much can lead to the overall heat
transfer coefficient’s rapid decline. Also, because of the larger
flow resistance created by the deeper cavities, the rewetting of
the surface diminishes. The test section with a 33 × 33 array
of cavities results in an increase of the CHF by a factor of 2.5
when compared to that of the plain silicon surface with a heat
flux value of 30 W/cm2.
Micro pin-fins, micro-studs, pin-fins, and square pin-fins are
manually manufactured structures, usually silicon, proven to
significantly enhance nucleate boiling. Wei et al. [27] submerged two different pin-fin geometries in FC-72 to monitor
the pool boiling performance. These fin types, each in its different topology and shape, improve thermal convection when
they are submerged in the liquid. These cooling mechanisms,
because of the increase of chip surface area, allows for a higher
heat flux to be used. These microchannel modifications have
proven to be effective in decreasing the boiling incipience and
surface temperature and improving the CHF.
The use of nanofluids as the working fluid supplements heat
transfer performance with its higher thermal conductivity than
a pure fluid–air, water, or fluorochemicals [32]. Nanofluids
are colloidal nanoscale metallic or nonmetallic particles in a
base fluid [32–37]. By adding nanoparticles to the volume of
fluid, thermal conductivity can be increased by about 40% [38].

Kedzierski [39] investigated the effect that CuO nanoparticle
concentration had on a roughened horizontal flat surface. A 2%
and a 4% volume concentration of CuO in R-134a were compared in this study. On average the mixtures with a 4% CuO
volume concentration had a 140% larger value for boiling heat
flux than the mixture with only 2% volume [39]. CHF is enhanced up to 45% with a 1% volume concentration of alumina
nanofluid at a mass flux of 2500 kg/m2 [33]. Wright et al. [40]
studied the effects of the percentage of metal particles in the
nanofluid using CNT to increase the thermal conductivity of the
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531

fluids. If the concentration of the metal particles is too low, there
are no significant improvements. At 1% volume concentration of
CNT, there was about a 10–20% increase in the thermal conductivity. The higher concentration of these metal particles inside
the fluids increases the viscosity, making it more difficult for the
fluid to flow through microstructures [40, 41]. An explanation
for the amount of metallic nanoparticles is that the orientation
and alignment of the CNT are random in the fluid and the CNT
need physical contact with each other in order to increase the
thermal conductivity. With a low concentration, there is minimal
contact as there are fewer metallic particles [40]. Gold nanoparticles have also been placed in refrigerant R-141b to increase the
CHF. Boiling coefficients of the fluid increase with an increased
concentration of nanoparticles. For only a 1% volume concentration of nanosized Au particles in refrigerant R-141b, the heat
transfer coefficient doubled when compared to the fluid without
nanoparticles. Nanosized Au particles can significantly increase
pool boiling heat transfer when placed in refrigerant R-141b, but
the surface roughness and the particle size aged after one test,
decreasing the effects [42]. Using nanofluids will improve the
heat transfer due to the nanoparticle interaction with the surface

roughness when compared to those without nanoparticles.
In general, the wettability of the surface can play a critical
role in improving the heat transfer performance. Truong et al.
[30] used nanofluids to investigate how the surface wettability affects the CHF and the heat transfer coefficient. Minimum
wettability contact angle will maximize the CHF. To achieve a
high heat transfer coefficient, the optimal surface is one with low
wettability containing many nucleation sites. Surfaces should be
hydrophilic, having an intrinsic contact angle no higher than 90◦ .
Too much surface roughness can cause a higher contact angle,
leading to a smaller CHF and a hydrophobic surface. Nanoparticles suspended in fluids will not directly affect the heat transfer
coefficient through the contact angle; rather, the nanoparticles
can create many microcavities and therefore nucleation sites.
The heat transfer coefficient strongly depends on the number of
active nucleation sites available for vaporization [30].
Many modifications to microchannel surfaces have been
tested in enhancing the cooling performance. For pool boiling, increasing surface roughness and adding small cavities or
fins to the channel walls can lower the boiling incipience, increase the CHF, and increase the nucleate boiling of the system.
To further enhance the thermal performance, nanofluids with
an optimal amount of volume concentration of particles can
be used. Because of the continuous advancements in electronic
components, further enhancements of the microstructure materials, configurations, geometry, etc. still need to be investigated.
Flow Boiling
Flow boiling has the capability of increasing the thermal performance of the microchannel; it can provide a much higher heat
transfer coefficient than both the single-phase flow and the pool
boiling. Flow boiling is a two-phase process that convectively
removes heat as a fluid that is flowing with some given velocity.
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Figure 3 Flow patterns in flow boiling.

To understand the flow boiling process, it is essential to understand how it relates to the macro-scale assumptions. Among
many of the previous works done regarding flow boiling, there
is a controversy about the dominant mechanism driving the heat
transfer at the micro-scale. Is it the conventional nucleate boiling
that is dominant in the macro-scale, or is it the forced convection of the vapor bubble that transfers the most heat? In nucleate
boiling, the heat transfer coefficient is affected by the heat flux
inputted into the system but it is independent of the fluid flow
rate and vapor quality. On the other hand, convective boiling is
driven by the fluid’s flow rate and the vapor quality and is not a
function of the inputted heat flux [43]. Thome et al. [44] believe
flow boiling has the most heat transfer from convection. At the
macro-scale, experimental studies have concluded that nucleate
boiling is the dominant factor in removing heat with little effect
of the convective flow; however, at the micro-scale, the heat
transfer is mainly affected by a thin liquid film that surrounds
the elongated vapor bubbles, not nucleate boiling [45]. Because
of this, using macro-scale assumptions in microchannels is not
realistic when predicting the flow boiling coefficients.
The principal flow regimes in flow boiling are bubbly, elongated bubble (slug), churn, annular, mist, and flows with partial
dry-out. Figure 3 displays a schematic of the main flow patterns that may be experienced with a constant heat flux. Flow
patterns in a small channel can vary slightly, depending on the
orientation of the channel because of the effects of gravity. For
a horizontal channel, when the fluid reaches temperatures just
above the saturation temperature, small bubbles begin to nucleate. This is known as the bubbly regime. As vapor bubbles
increases, a flow pattern develops that entraps the vapor bubbles in the main flowing liquid, known as plug flow. With more

heat, the bubbles grow to be within a few micrometers of the
channel’s hydraulic diameter. The bubble is now confined by
the microchannel and it can no longer grow in diameter, but
elongates, growing in length. This is now the elongated bubble
regime or the slug regime. This pattern is the dominant regime
for flow boiling where the most heat is transferred convectively.

Separating the growing elongated bubbles is a section of fluid
also known as liquid slug. There is a small thin film of liquid surrounding the vapor bubble separating it from the microchannel
wall. If the liquid film enclosing the elongated bubble reaches
a minimum thickness, the region is considered dried out, also
known as the vapor slug. As the length of the vapor continues to
grow, it swallows up the liquid slug until the elongated bubble
emerges to the next cycle or the next patch of vapor. This is
shown in Figure 4. When the dominant flow is vapor and has a
small film of liquid surrounding the bubble, this is known as the
annular regime. This flow pattern may have small droplets of
liquid dispersed throughout the vapor core. The vertical channel
is very similar as it experiences the slug, churn, annular, and
mist pattern. It initiates with the bubbly flow and quickly the
vapor pockets grow to a slug regime. The vapor bubble continues to grow into churn, then annular, and then mist fashion.
Because of the influence of gravity, the horizontal flow patterns are more likely to have intermittent drying and rewetting
of the upper surfaces of the tube for slug and annular patterns
[44, 46, 47].
Figure 4 shows a model used to describe the elongated bubble flow regime in the flow boiling process. At a fixed reference
frame, a liquid slug of some length will pass, followed by an
elongated bubble, and if the thin film dries out before it reaches
the next liquid slug, there will pass a vapor slug. This cycle repeats itself until the vapor is continuous throughout the channel
[44, 46].
To better understand the flow boiling profiles, Thome et al.

[44] developed a three-zone model in an effort to qualitatively
and quantitatively describe the heat transfer effects due to flow
boiling in microchannels. This model describes the evaporation
of the elongated bubble as it flows through a microchannel and
predicts the local heat transfer coefficient in the liquid slug,
evaporating elongated bubble, and a vapor slug or the dry-out region. The three-zone model takes into account the frequency of
the vapor bubbles with respect to time, the minimum liquid film
thickness as dry-out occurs, and the liquid film thickness. When

Figure 4 Model used to describe the flow boiling process.

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J. F. TULLIUS ET AL.

compared to the liquid slug region, the heat transfer coefficient
in the thin film evaporation region (elongated bubble) is several
times higher. The values for the vapor slug region, annular, are
almost negligible as the heat transfer coefficient of air is much
lower.
Like in single-phase flow and pool boiling, adding microstructures to the surfaces of microchannels can enhance the
cooling performance. Chien et al. [48] added square pin-fins of
dimension 400 µm × 400 µm × 400 µm (width × thickness ×
height) to a rectangular 20 mm × 25 mm Cu plate and investigated the heat transfer effects when varying the heat flux and the
flow rate. For both a smooth and pin-finned surface, water and
FC-72 were used as the working fluids in the microchannel. For
water, the performance is influenced by the flow rate for heat flux

lower than 60 W/cm2. At low flow rates, the heat transfer coefficients increase with increasing heat flux. However, at high flow
rates, the heat flux is almost negligible with minimal effect to the
heat transfer coefficient. For lower heat fluxes the performance
is driven by the nucleate boiling, and for higher flow rates, the
forced convection is dominant. However, when using FC-72 as
the working fluid, the most influential parameter is the heat flux.
When varying the flow rate, the heat transfer coefficient curves
were similar, implying that the boiling heat transfer is the more
dominant effect, rather than the forced convection. With the
saturation temperature in FC-72 (56◦ C) being much lower than
that of water (100◦ C), one should expect to see the nucleation
start sooner. More vapor bubbles are observed with Fluorinert.
For low heat fluxes, the heat transfer coefficient increases as the
flow rate decreases, but for lower flow rates, partial dry-out was
observed during the experiments, which drastically degrades the
performance. The pin-finned surface increases the heat transfer
coefficient by about 30% and contains a greater CHF for a fixed
flow rate when compared to the smoothed surface. The convective heat transfer was greater at low flow rates (80–160 mL/min)
and heat fluxes (18–35 W/cm2) for FC-72 than for water. Water,
however, had a superior performance compared with FC-72 for
higher flow rates [48]. Chien et al. [49] in another investigation
compared the effect of FC-72 cooling fluid flow boiling through
two different square pin-fin geometries. A comparison was conducted of square pin-fins with geometries of 400 µm × 400 µm
× 400 µm and 200 µm × 200 µm × 200 µm with various flow
rates (80–960 mL/min) and heat fluxes (18–50 W/cm2). Similar
to the previous investigation, the heat flux is the driving force
influencing the heat transfer coefficient, as opposed to the flow
rate, because of the dominant nucleate boiling effect. With the
square pin-fins, the partial dry-out only occurred at the lowest
flow rate (80 mL/min), unlike with the smoothed surface, because the gaps of the fins prevent the drying out by the liquid

it retains on the surface. Therefore, the structured surface had a
significantly higher heat transfer performance because the surface was kept from drying out for lower flow rates and high
heat fluxes. At high flow rates, both geometries contain similar results, but at the lowest flow rate tested (80 mL/min), the
geometry with the smaller square pin-fins had a 5–10% higher
heat transfer coefficient. For both pin-finned surfaces, there is
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533

a 10–20% increase in the heat transfer coefficient when compared to the smooth surface as the flow rates were between 320
and 960 mL/min [49]. Similar results were found in the study
conducted by Lie et al. [50].
Cetegen et al. [51] used refrigerant R-245fa and passed it
through a force-fed evaporation element and a microgrooved
surface using three mass flow rates. From the results, for all
mass flow rates, there was no variance in the trend below a
heat flux of 320 W/cm2. For lower heat fluxes in this experiment, the heat transfer coefficient was due to the variance of
the heat flux. Therefore, it can be concluded that the dominant
heat transfer mechanism was nucleate boiling with negligible
convective boiling. At a heat flux of 320 W/cm2 there was a
huge temperature jump and a drop in heat transfer coefficient
where it is believed to have reached the CHF. Grooved surfaces,
like finned surfaces, can also influence the thermal performance
of flow boiling process.
For space applications, the effects of gravity on flow boiling
through microchannels can be very useful. Kandlikar and
Balasubramanian [52] varied the orientation of a six-parallelmicrochannel system with flow boiling of water in three
different directions: horizontal, vertical with an upward flow,
and vertical with a downward flow, all while maintaining the
heat and mass flux conditions. For all directions, the flow

regimes encountered were bubbly flow, thin film nucleation,
plug flow, churn flow, and annular flow. The flow patterns seem
to be similar for all orientations of the channel, except the
shapes of the vapor bubbles in the vertical down flow orientation
are more bullet-shaped, while there was an elongated circular
shape for the horizontal orientation. A flow reversal effect was
also encountered for all orientations, but was more noticeable
in the vertical downward flow case. Similar results of the
heat transfer performance for the vertical up flow and the
horizontal flow case were better than that of the vertical down
flow case due to this higher flow reversal encountered. Luciani
et al. [53] compared the effects of microgravity to terrestrial
gravity in a single vertical microchannel using a transparent,
nonflammable and nonexplosive fluid with a low boiling
temperature. A microchannel was placed in an Airbus A300
Zero G, flying in a parabolic fashion, starting from terrestrial
gravity and peaking at microgravity. At microgravity, vapor
patterns lead to larger bubble sizes than in terrestrial gravity.
Churn and slug flow patterns were dominant, while in terrestrial
gravity, for the same conditions, bubbly flow and some slug
flow patterns were visible. The heat transfer coefficient in
microgravity was much higher. Because of the larger vapor
bubbles and the higher heat transfer coefficient, Luciani et al.
[53] concluded his heat transfer performance was driven by the
convective flow and not nucleate boiling. Gravity has little effect
on the heat transfer performance at a terrestrial level, which
seems to be influenced by nucleate boiling. At microgravity
levels, the forced convection of the fluid guides the heat transfer
performance.
As stated earlier, nanofluids can effectively enhance the thermal performance of microchannels. Peng et al. [54] investigated

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J. F. TULLIUS ET AL.

the effects of a refrigerant-based nanofluid, CuO/R-113, as the
working fluid in a flow boiling process for a rectangular, smooth
tube. This nanofluid was analyzed by varying the mass flux (100
kg/m2-s, 150 kg/m2-s, 200kg/m2-s), the heat flux (3.08 kW/m2,
4.62 kW/m2, 6.16 kW/m2), and the volume concentration of the
nanoparticles in the base fluid (0%, 0.1%, 0.2%, 0.5%). The results conclude that the heat transfer coefficient increases as the
vapor quality of the fluid increases. When comparing the heat
transfer coefficients with respect to mass flux, for a 0.5% volume
concentration, there is a 29.7%, 22.7%, and 25.6% enhancement
of heat transfer coefficient with mass fluxes of 100 kg/m2-s, 150
kg/m2-s, and 200 kg/m2-s, respectively, when compared to the
base fluid with 0% concentration. The thermal performance of
the channel will increase due to the enhancement of the heat
transfer coefficient provided by nanofluids.
Flow boiling in microchannels can improve the thermal performance. Based on these papers, flow boiling is made up of
nucleate boiling and forced convection. The dominant form of
heat transfer in flow boiling varies depending on the heat flux
for nucleate boiling and the mass flow rate for forced convection
dominance. Flow boiling is influenced by microstructures, fluid
flow rates (in some cases), heat flux variance (in some cases),
gravity, etc.

CARBON NANOTUBE STRUCTURES ACTING AS

MICROFINS
Carbon nanotubes (CNT) may become a novel material that
can considerably improve microchannel cooling performance.
The unique molecular structure of CNT results in excellent
physical and mechanical properties such as great mechanical
strength, great flexibility, and low weight. Their mechanical
and chemical stability provides resistance against damage from
external physical and chemical factors applied by their environment [4, 55–57]. The main sp2 hybridized bonds of CNT,
similar to the in-plane ones in graphite, place them among the
strongest materials known today. CNT have a Young’s modulus
as high as 1000 GPa, which is approximately 5 times higher than
steel, and a tensile strength of about 63 GPa, which is almost
50 times higher than steel [4, 58–61]. These cylindrical tubes
remain stable up to very high temperatures, similar to graphite,
with values approximately 4000 K [62].
There are two types of CNT: single-walled carbon nanotubes (SWNT) and multiwalled carbon nanotubes (MWNT).
An SWNT is a cylindrical graphene shell with diameters ranging
from 0.45 to 2.5 nm; it can be considered as a giant molecule.
An MWNT consists of several concentric cylindrical shells with
the outer diameters ranging from 2.5 to 60 nm and inner diameters between 1.5 to 40 nm. MWNT can be considered materials
similar to graphite. The distance between the concentric shells
of a MWNT is approximately 3.4 Å [58, 60].
CNT are single sheets of graphite, named graphene, made up
of a honeycomb-shaped lattice representing an atomic layer of
the crystalline material, rolled up to make tubes with diameters
heat transfer engineering

of 0.45 to about 100 nm. The circumference of a CNT is equal
to the length of the chiral vector describing its symmetry and
unit cell; the chiral vector is created from the unit vector as

Ch = na1 + ma2 . The integers n and m specify the chiral vector
and therefore the chirality or helicity and the atomic structure of
the tube. Specific chiralities are zigzag and armchair structures.
The classification of nanotubes can also be given by the chiral
angle (θ) at which the graphite sheet is rolled to create that
cylindrical shape. When the chiral angle and m in the chiral
vector are zero it is known as zigzag. Armchair is a nanotube
with the chiral angle equaling 30◦ and n = m in the chiral vector
[60, 63].
For an ideal individual SWNT, the thermal conductivity has
been reported to be higher than diamond, roughly 6000 W/m-K.
As a comparison, graphite has a thermal conductivity of about
2000 W/m-K and diamond between 2000 and 2500 W/m-K [60].
However, measurements on larger number of nanotubes resulted
in thermal conductivity values as low as about 250 W/m-K for
SWNT samples and 20 W/m-K for MWNT samples [40, 57, 64].
Thermal conductivity is determined by many factors: e.g., for
an ideal SWNT, its (m,n) chiral vector and its length. Molecular
dynamics (MD) studies, explained later in this paper, have been
conducted investigating the thermal conductivity of CNT as
they vary in length and chirality of SWNT with a finite length.
Higher thermal conductivity has been obtained for SWNT with
lower diameter (chirality (5,5)). The length and the thermal
conductivity of a SWNT are proportional to each other, while the
thermal conductivity does not depend linearly on the diameter
[65–68].
CNT can be grown directly onto silicon substrate by Chemical Vapor Deposition (CVD). CVD is a low cost process and can
produce CNT in many fashions, in either bulk quantities or predefined micropatterns. In most of the CVD processes, substrates
are heated in a furnace while a hydrocarbon gas is flowing
through the reactor. A catalyst—iron, nickel, or cobalt—is deposited on the substrate [69] or fed into the reactor together with

the hydrocarbon source [70]. In the first, more widely used, case,
CNT grow in the location of the catalyst and form 3D structures
erected above the 2D mold defined from the catalyst materials.
Similar patterning is possible in the floating catalyst method as
the CVD growth is template dependent [71–76]. In this CVD
method, ferrocene (Fe(C5 H5 )2 ) is dissolved in xylene (C8 H10 ),
at concentrations of ∼0.01 g/mL, preheated at about 150◦ C,
co-evaporated, and fed into the reaction zone. After passing the
desired reaction time, estimated by the desired nanotube length,
the carbon source and catalyst valves are closed and the tube is
cooled down in an argon atmosphere. Growth of several tens of
micrometers thick, uniform, vertically aligned multiwalled nanotube films (with a narrow diameter distribution of 20–30 nm)
can be produced on silica substrates with a growth rate of ∼10
µm/min. Using silicon/silica patterned substrates, the deposition
of nanotubes is extremely selective as nanotube structures grow
aligned with respect to each other and normal to the substrate, on
the oxide only (but not on Si nor on the native oxide layer of Si).
This selectivity is retained down to the small lithographically
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J. F. TULLIUS ET AL.

535

Figure 5 Carbon nanotube pillar (fin) structures defined by (a and b) selective CVD (reprinted with permission [74], copyright [2003] IEEE), (c) laser ablation
technique (reprinted with permission from [5], copyright 2007, AIP), and (d) solvent treatment [78].

patterned micro-size dimensions of SiO2 on the SiO2 /Si substrate, and below that limit the cooperative growth phenomenon
is weakened and the order of the structure is diminished. This

simple characteristic of the substrate provides an opportunity
to build ordered nanotube fin structures by designing the SiO2
patterns in terms of dimension, thickness, shape (cross section),
and uniformity. High aspect structures also show deviation
from the exact perpendicular growth.
Another way of pattern generation is processing the uniformly grown CNT layers by laser ablation [5, 77] or solvent
treatment [78, 79]. During laser treatment, the pattern of the
structure is given by a computer controlled laser beam and a
wide variety of patterns can be generated. The solvent treatment
method can be used to create predefined structures or random
cell structures of CNT. In both structures the nanotubes are
pulled together by capillary forces and kept together due to the
van der Waals interactions [78, 79]. The solvent-treated structures have higher CNT population in the pillar and wall. This
higher density of the nanotubes is advantageous for reaching
better mechanical properties and higher thermal conductivity
and heat capacity values. Figure 5 shows images of CNT fin
structures grown using CVD, laser ablation processes and solvent treatment.
Mo et al. [64] compared a smooth microchannel to a channel
with carbon nanofins. The nanotube fins lowered the temperature by 6◦ C. The flow rate of the coolant in the CNT coated
channel was decreased by 12% and the heat input increased to
23%, yet the nanotubes still maintained a lower temperature.
This more efficient method decreased the temperature without
a significant drop in pressure. Zhong et al. [2], using computational fluid dynamics (CFD), also explained further in this
paper, examined the effects of CNT microstructures as they
were coated on the surface of the microchannel walls. Figure 6 shows a schematic of the structural composition of CNT

Figure 6 Microstructure makeup when microchannel is coated with CNT.

heat transfer engineering

microstructures consisting of fins organized into arrays. The
microstructures are roughly about 1 mm × 1 mm × 100 µm
(length × width × height) placed on the surface of the microchannel. If you magnify the scale further, the microstructure
is made up of hundreds of microfins across that surface area.
These are the clusters of CNT forced together mainly by the van
der Waals interaction. Each microfin is made up of hundreds of
CNT.
The effects of CNT clustered together to form microstructures are still under investigation. Microstructure can be modeled as a macro-scale continuum; however, as the scales decrease, alternative methods must be introduced into the problem.
CNT are an excellent conducting material that has been tested
in the use of increasing the performance of the microchannel
heat sink. They are simple to fabricate with the CVD process
and they contain excellent physical and mechanical properties
among the top known today.

NUMERICAL AND COMPUTATIONAL ANALYSIS
OF MICROCHANNELS
Experimentation with microchannels can be very costly in
both time and money. The experimental setup has to be nearly
perfect to get the desired effect and satisfy boundary conditions. In testing more than one experiment with varying parameters, the setup alone can take weeks. Numerical and computational analysis can cut costs and time while investigating the
thermal performance. The analysis can help predict reasonable
testing parameters for the experimentation of microchannels.
The lattice Boltzmann method (LBM) can numerically calculate the interaction of small particles and fluid flow. To simulate particles at the atomic level, computer software such as
MD is used. CFD is computer software that simulates the fluid
as a continuum using NS. These numerical and computational
methods can shorten the design cycle and lessen experimental
costs.
In microchannels, the heat and fluid flow are very different from those in macrochannels. Because of the high
surface–volume ratio the surface defects affect the domain in
small devices. The macroscopic no-slip boundary conditions
are not valid at this micro- and nano-scale under some circumstances [80]. The classification of the fluid regime is measured

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J. F. TULLIUS ET AL.

Figure 7 Flow characteristics based on the Knudsen number.

by the Kn, which is defined by Kn = λ/H, where λ is the mean
free path and H is the characteristic length of the channel in
which the fluid flows. The Kn characterizes the different flow
regimes for which certain numerical equations can be used when
calculating the flow patterns. This is shown in Figure 7.
If Kn ≤ 10−3 the flow is assumed to be a continuum and the
NS with the no-slip boundary conditions can be utilized. For
the flow regime where Kn > 10, the flow is known as a free
molecular flow. Free molecular flow is where the molecules are
larger than the size of the chamber or the object being tested,
known as a vacuum. A rarefied gas is neither a continuum nor a
free molecular flow but its Kn is between the ranges. Between
the continuum flow and the free molecular flow regimes, there
are slip flow and transition flow characteristics. When the flow
regime is 10−3 < Kn ≤ 10−1, the NS with the slip boundary
conditions can be used to assure accuracy. When the rarefaction
factor becomes greater than 10−1 the macroscopic method based
on the NS will no longer suffice and a more accurate method
must be used. When the rarefaction factor is for 10−1 < Kn ≤ 10,
particle-based methods should be used. LBM, MD, and direct
simulation Monte Carlo (DSMC) are such methods that can sufficiently calculate the fluid flow regimes as the scales are reduced

[20, 81– 83]. Therefore, with the flow regime characteristics, microchannels with characteristic dimensions between 1 µm and 1
mm can be modeled using the NS equation as the fluid will follow macroscopic behavior. Channels with dimensions less than
1 µm will follow microscopic flow and NS can no longer be
used [24, 83–87]. Harley et al. [88], Araki et al. [89], and Arkilic
et al. [90, 91] all investigated the flow continuum in microchannels, concluding that the conventional equations were no longer
adequate in predicting the flow patterns. Harley et al. [88] investigated gas flow through a channel with varying depths of
0.5–20 µm and a width of 100 µm using nitrogen, helium, and
argon gas. The results were compared to the macroscopic calculations and they did not correlate. Araki et al. [89] studied
the frictional characteristics of nitrogen and helium and found
that the frictional resistance of the gaseous flow is smaller in
microchannels than that of a traditional channel. Arkilic et al.
[90, 91] investigated the deviation of the transport of gas in
microchannels versus the continuum.
heat transfer engineering

LBM can accurately simulate fluid flowing through a microchannel [84, 92–96]. This simplified method of the kinetic
equations is derived from the Boltzmann equations. Unlike other
mathematical methods such as MD and DSMC, LBM does not
depend on distribution of the number of molecules; rather, it
concentrates on the distribution function dependence of the velocity coordinates. This method focuses on the local velocity
averages at distinct locations [81, 84, 86, 94, 97]. As shown in
Figure 7, these equations are suitable for all flow regimes. LBM
is considered to be computationally stable, accurate, efficient,
and easy to use. This method is capable of solving complex
geometries and of effortlessly implementing the boundary conditions, and is less difficult to compute, while avoiding the need
to follow every particle in the system like MD, and DSMC [84,
94, 97]. Darbandi and Setayeshgar [81] use the LBM to investigate a fluid flowing past a confined cylinder in a microchannel
with slip flow regime. A result justification through a comparison and an examination of the continuum and noncontinuum
flow past a confined cylinder in a microchannel was conducted.
The results suggested that the Kn increases when the cylinder

was placed in the flow path, decreasing the hydraulic diameter.
MD is a particle-based computer simulation program that
can compute the thermal transport of properties in nanostructures at the atomic scale using classical and quantum physics.
This method makes it possible to attain the thermal resistances
for a solid/solid, solid/liquid, and liquid/liquid interface. MD is
made up of two processes: equilibrium and nonequilibrium MD.
The equilibrium MD is calculated based on the Green–Kubo relations, and nonequilibrium MD is calculated using the Fourier
laws [24]. Shibahara et al. [98] examined the thermal resistance effects of a liquid/solid interface using MD simulations.
The resistance was calculated by the heat flux and the temperature jump at the interface, and it was found that by increasing the
density of the fluid, the thermal resistance decreases between
the thermal transport of the solid and liquid. MD can evaluate
the data necessary to predict the interaction of the fluid flowing
through CNT as well as the thermal conductivity through each
graphite tube as it varies with length and chirality. This method
can isolate and solve one CNT to understand the performance
and the heat effects as it stands by itself. Recognizing the heat
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J. F. TULLIUS ET AL.

effects through the lone nanotube can be useful when calculating the effects of many nanotubes close together. Hu et al. [99]
implemented distinct boundary conditions to simulate one CNT
in the midst of a square array of CNT. It was discovered that
the CNT should contain a small gap big enough for a fluid to
flow through the tube separation. This will maximize the heat
transfer due the increase of surface area. Kharazmi and Kamali
[100] investigated the effects of surface roughness or deviations
along the surface of a fluid in a nanochannel using MD simulations. The wall–fluid interaction and the surface irregularity are
factors important to the disturbance of the flow. The cavitations

and wall microstructures can vary the local density pattern yet
still maintain an overall average [100]. MD can simulate the interaction at the solid–fluid interface accurately in the molecular
scale and also save time and money.
The CFD method, calculated through a commercial software
package, is used to approximate the fluid mechanics and heat
transfer characteristics of microchannels. CFD can be used in
parallel to experimental setups in an effort to predict the flow and
heat effects of the given surface modifications and the specified
parameters and boundary conditions of a microchannel. CFD is
based on the NS equations derived from Newton’s second law
of motion. NS is a set of equations that describe the fluid flow
behavior in a continuum. CFD simulations have been used to
study the many different aspects of microchannels. Zhong et al.
[2] modeled a 2D simulation of a microchannel with arrays of
microstructures on the bottom surface. A study of the effects
of varying the geometry of the microstructures was completed,
varying the fluid speed and changing the thermal conductivity
of the fin material as they influence the thermal performance.
When the results were compared to the mathematical calculations of the conservation of energy equations, they were almost
identical to the CFD simulation results [2]. Srivastava et al.
[101] used the CFD software package Fluent to understand the
effect of roughness on fluid flow characteristics. Three rectangular channels were simulated: two with different geometries of
roughness, compared to a smooth microchannel [101].

CONCLUSIONS
This paper has reviewed different techniques used in efforts
to modify microchannels in both single- and two-phase laminar flow. Some surface modifications discussed include adding
micro-fins, adding grooves, increasing surface roughness, etc.
Many of these modifications have been successful in increasing the thermal performance; however, with advancements in
technology, heat transfer in microchannels still needs to be further enhanced. CNT built onto the surface of microchannels can

improve thermal performance, because of impressive material
and mechanical properties. Different configurations and geometries of CNT microstructures in mini-/microchannels need to be
further numerically and experimentally evaluated to maximize
cooling of the electronic components. Mathematical analysis of
microchannels can be done through LBM at both the microheat transfer engineering

537

and nano-scale, whereas at the atomic scale one can use MD
and CFD for the continuum regime. Each of the mathematical
processes can accurately approximate the thermal performance
of the channel for their specified Kn. A cooling mechanism
needs to be designed or modified to efficiently cool electronic
components and keep them from malfunctioning in order for
electronic innovation to continue [102].
NOMENCLATURE
CFD
CHF
CNT
CVD
DSMC
H
Kn
LBM
MD
MWNT
NS
q
SWNT
u

computational fluid dynamics
critical heat flux
carbon nanotubes
chemical vapor deposition
direct simulation monte carlo
characteristic length of the channel
Knudsen number
lattice Boltzmann method
molecular dynamics
multiwalled carbon nanotube
Navier–Stokes equation
heat flux (W/m2)
single-walled carbon nanotube
velocity (m/s)

Greek Symbol
λ mean free path of fluid
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Jami F. Tullius is a graduate student pursuing a doctorate degree in mechanical engineering at Rice University, Houston, TX. She received her bachelor’s degree in mechanical engineering from the University
of Texas at El Paso, El Paso, TX, in 2008. She was
one of the recipients of a Graduate Education and the
Professoriate (AGEP) program scholarship through
an NSF grant. Currently she is an HENACC Scholar,
GEM Fellow, and recipient of a NASA–Harriett G.
Jenkins Pre-Doctoral Fellowship.
Robert Vajtai is a faculty fellow of the Department
of Mechanical Engineering and Materials Science at
Rice University, Houston, TX. He received his master’s, doctoral, and Ph.D. degrees at University of
Szeged, Szeged, Hungary. He has more than 100
publications in international scientific journals and
these papers received more than 2500 citations. He
is editor of Nanopages and Fluctuations and Noise

Letters and an editor for the Springer Handbook of
Nanomaterials. He leads and takes part in projects
for applications of nanomaterials in thermal managements.
Yildiz Bayazitoglu joined Rice University, Houston,
TX, in 1977, and since 1996 has been H. S. Cameron
Chair Professor of Mechanical Engineering in the
Department of Mechanical Engineering and Materials Science. She received her M.S. and Ph.D. degrees
from University of Michigan, Ann Arbor. Her current
research interests include containerless processing of
materials, solution to electromagnetic radiation equation, molecular dynamics studies for nano heat transfer, microchannel fluid and heat transfer, and bio heat
transfer. She is the editor in chief (Americas) of the International Journal of
Thermal Sciences (IJTS), which is published by Elsevier. She received several
awards due to her achievements in teaching, mentoring, and research. She is a
fellow of ASME and AAAS.

vol. 32 nos. 7–8 2011


Heat Transfer Engineering, 32(7–8):542–553, 2011
Copyright C Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
DOI: 10.1080/01457632.2010.506397

Experiment Investigation of R134a
Flow Boiling Process in Microchannel
With Cavitation Structure
HONG ZHANG CAO, HONG BO XU, NAN LIANG, and CHANG QING TIAN
Institute of Engineering Thermal-Physics, Chinese Academy of Sciences, Beijing, China

One cavitation structure in which the channel cross section expanded suddenly was introduced in single straight microchannel.

The experiment was carried out with R-134a as the fluid medium, which was driven by a gear pump. The flow pattern was
observed by a charged coupled device (CCD) camera and microscope. The average boiling heat transfer coefficient was
estimated with the calculation method proposed in this paper. The experimental results show that the boiling began at the
cavitation structure, and stable flow boiling was maintained. When heat power rose, the boiling became strong; then the
pressure drop in the micro-channel increased and heat transfer was enhanced. The liquid percentage in two-phase flow
increased and the length of boiling area became small when the liquid subcooling degree rose with the fixed heat power.

INTRODUCTION

EXPERIMENTAL FACILITY

Many experimental results for flow boiling in unchanged
cross section microchannel or channels have presented unsteady
different patterns of alternating flow [1–8]. Especially in slug
flow, the fast growth of a bubble could induce local dryout in a
channel. In several investigations [9–13], a restrictive inlet (e.g.,
inlet orifice) where cavitation was caused by static pressure
decrease and velocity increase with sudden reduction in the
flow area was used in microchannel or channels to suppress
flow boiling oscillations. Following the results of macro-scale
cavitation showing that cavitation in a submersed jet could reach
onset in the low-pressure kernel of a vortex and the cavitation
bubble collapsing could induce a microjet [14, 15], we supposed
that this phenomenon could occur in a microchannel and the
microjet could break the growing bubbles surrounding it. So for
this paper, one cavitation structure in which the channel cross
section expanded suddenly was introduced in a single straight
microchannel and the experiment was carried out with R-134a.

The experimental apparatus is shown in Figure 1. A gear

pump was used to force the liquid through the test section
with a constant flow rate measured by a flow meter (K-F mass
flow meter, accuracy ±0.1%). The work fluid through the condenser was cooled by cooling water whose temperature could be
adjusted.
The test section is shown in Figure 2. The microchannel was
fluted with an electric line cutting technique in a 2-mm-thick,
25-mm-wide, 76-mm-long copper plate, which connected to an
inlet tank and outlet tank made from Teflon. The channel depth
was 0.5 mm and the width changed from 3 mm (46 mm long)
to 0.5 mm (10 mm long), then to 1 mm (20 mm long). A glass
plate was put above the copper plate for visual observation. For
estimating the pressure drop of fluid, the whole channel was
divided into seven parts; part 1 is the channel entrance, part 2
is the 3-mm-wide and 46-mm-long channel, part 3 is the sudden reduction between part 2 and 4, part 4 is the 0.5-mm-wide
and 10-mm-long channel, part 5 is the sudden expansion between part 4 and 6, part 6 is the 1-mm-wide and 20-mm-long
channel, and part 7 is the channel exit. The cavitation structure is where the channel cross section expands suddenly, i.e.,
part 5. A film heater (15 mm wide, 20 mm long, 27 ) shown
by the dashed shape in Figure 2 was set on the copper plate back
side and covered by an insulated layer. Heat power was supplied by direct current, and voltage was measured by voltmeter

This study was supported by Natural Science Foundation of China (grant
50676099).
Address correspondence to Dr. Hong zhang Cao, Institute of Engineering
Thermal-Physics, Chinese Academy of Sciences, 100190 bei si huan xi lu 11,
Beijing, China. E-mail:

542


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543

Figure 1 Schematic of experimental apparatus. 1—pump, 2—flow-meter,
3—inlet-tank, 4—outlet-tank, 5—text-section, 6—CCD & microscope,
7—condenser

(accuracy: ±1%). The channel entrance and exit pressures (Pin ,
Pout ) were measured by pressure sensors (pressure resistance,
accuracy: ±0.5%) in the inlet tank and outlet tank, respectively.
Thermocouples (T type, accuracy: ±0.1◦ C) were used to measure fluid temperatures in the inlet tank and outlet tank (T in ,
T out ), as well as the wall temperatures (T 1 , T 2 , T 3 , T 4 , T 5 , T6 )
at different locations along the microchannel, shown in Figure
2. A CCD camera (Panasonic WV-CP240EXCH) and a microscope located above the test section were used to observe flow
patterns. The fluid was driven by a gear pump, and the flow rate
was changed according to the test requirement.

CALCULATION OF AVERAGE BOILING HEAT
TRANSFER COEFFICIENT
The average boiling heat transfer coefficient is defined by
h boil = Q hboil /Aboil (Tavewall − Tavesat )

(1)

where T avewall is the average wall temperature of channel, T avesat
is the average saturation temperature of work fluid, Aboil is the
heat transfer area of boiling part in channel, and Qhboil is the
heat transfer in the boiling part. The summation of Qhboil and
Qhliquid , the heat transfer in the single-phase liquid flow part, is
the heat power supplied by the film heater to the test section.

Cavitation structure, 1st OLNB

7

Film heater
6
5

T avewall , T avesat , and Qhliquid are defined by:
Tavewall ≈ ( Ti )/n,

n = 6, i = 1, . . . , 6

(2)

Tavesat ≈ (TOLNBsat + Tout )/2

(3)

Q hliquid = C p × G × (TOLNBsat − Tin )

(4)

2nd OLNB

G

4
3

2

1

G

T 6 T5 T4 T3

Figure 3 Single-phase liquid flow: (a) flow pattern; (b) pressure drop vs. fluid
mass flux; (c) friction factor vs. fluid mass flux. The bottom of channel was
not smooth. Wales [bright lines showed in part (a)] at bottom of channel were
caused by manufacturing effect.

T2 T1
Figure 2 Test section.

heat transfer engineering

where TOLNBsat is the fluid temperature, which is the saturation temperature corresponding to the pressure (POLNB ) at the
originating location of nuclear boiling (OLNB) observed by the
CCD camera and microscope, Cp is liquid specific heat, and G
is fluid mass flux. The pressure (POLNB ) would be calculated approximately from the experimental results of single-phase liquid
flow.
In the experiments for single-phase liquid flow, the pressure
difference between the channel entrance and exit is equal to the
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544

H. Z. CAO ET AL.

Figure 4 Flow pattern while pump power frequency is 30 Hz: (a) heat power
8 W and mass flux 4.48 kg/h; (b) heat power 15 W and mass flux 4.06 kg/h; (c)
heat power 23W and mass flux 3.82 kg/h.

Figure 5 Flow pattern while pump power frequency is 40 Hz: (a) heat power
8 W and mass flux 6.2 kg/h; (b) heat power 15 W and mass flux 5.85 kg/h; (c)
heat power 23 W and mass flux 5.33 kg/h.

sum of pressure drops at seven parts (Figure 2), written as

D2 = 0.86 mm, D4 = 0.5 mm, and D6 = 0.67 mm, f i is the
single-phase friction factor, and ξi is the local loss coefficient.
For the channel in this paper, the cross section of part 2 is
much more than for part 4 or part 6. Therefore, the pressure loss
of part 2 (P2 ) could be neglected compared to the total pressure
drop. The test data for single-phase liquid flow presented in
Table 1 showed the values of Reynolds numbers (Re) in part
4 and part 6 are larger than the critical Re in macro-scale
(2000–2300). That meant the test single-phase liquid flow was
turbulence. So here take f 2 = f 4 = f 6 = f , approximately.
Similarly, the local loss coefficient ξi was valued according to
macro-scale investigating results. Here ξ1 = 0.5, ξ3 = 0.43, ξ5 =
0.25, and ξ7 = 1. Through measuring of mass flux and pressure
drop for the single liquid phase flow, the average single-phase
friction factor (f ) could be calculated. This result could be used
to approximately calculate the pressure (POLNB ) in the flow boiling experiment with visual observation.
Although there is error for the calculation method of the
average boiling heat transfer coefficient used in this paper, it is

applicable to find out the variation of pressure loss and boiling
heat transfer in a microchannel with the cavitation structure.

Pin − Pout =

Pi ,

i = 1, . . . , 7

(5)

where
P1 = ξ1 × ρu 22 /2 for part 1
P2 = f 2 × L 2 /D2 × ρu 22 /2
P3 = ξ3 × ρu 24 /2

for part 7

(7)
(8)

for part 4

for part 5

P6 = f 6 × L 6 /D6 × ρu 26 /2
P7 = ξ7 × ρu 26 /2

for part 2

for part 3

P4 = f 4 × L 4 /D4 × ρu 24 /2
P5 = ξ5 × ρu 24 /2

(6)

(9)
(10)

for part 6

(11)
(12)

and where ρ is liquid density of refrigerant, ui is refrigerant
velocity, Li is channel length, Di is hydraulic diameter, here
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H. Z. CAO ET AL.
Table 1

545

Test data of single phase liquid flow (dynamic viscosity, µ = 0.202 × 10−3 Pa-s, 25◦ C)

Pin (kPa)

688
694
700
718
736
760

Pout (kPa)

G (kg/h)

T in (◦ C)

ρ (kg/m3)

u2 /u4 /u6 (m/s)

615
604
591
579
568
561

4.284
5.07
5.965
6.85
7.692
8.58

24.4
23.7
23.4
23.35
21.9
21.45

1208
1211
1212
1212
1217
1219

0.6567/3.94/1.97
0.775/4.652/2.326
0.911/5.468/2.734
1.047/6.28/3.14
1.171/7.024/3.512
1.303/7.82/3.91

RESULTS AND DISCUSSION
The test data for single-phase liquid flow are presented in
Table 1. The pressure difference between the channel entrance
and exit increases when mass flux rises. On the contrary, the
average single-phase friction factor, f , decreases, as shown in
Figure 3. The single-phase liquid flow pattern is shown in the
same figure.
During the experiment of flow boiling in a microchannel,

the power frequency of the gear pump was set at 30 Hz, 40
Hz, and 50 Hz. At each rotation speed, the heat power was
increased at three levels, 8 W, 15 W, and 23 W. The region
surrounding the cavitation structure (part 5 in Figure 2) was
observed and flow pattern was recorded by the CCD camera and
the microscope. A comparison of flow patterns with different
heat power at each pump rotation speed is shown in Figures
4–6.
The experimental results show that the boiling originated
in the cavitation structure where the vortex cavity was in the
reentrant region, and boiling became strong with the increase
of heat power. Meanwhile the OLNB changed against the flow
direction from part 5 to part 3 in course of the heat power
increasing. The OLNB was in part 3 for two reasons. First,
in part 3 the cross section became small, which induced the
liquid velocity to increase and the pressure to decrease. Second,
with the heat power increasing, the heat flow along the copper
plate induced the wall and liquid temperature to increase in

Re2 /Re4 /Re6
3394/11,780/7893
3995/13,944/9342
4700/16,400/10,990
5402/18,840/12,622
6067/21,158/14,176
6762/23,595/15,808

Pin Pout (kPa)

f

73
90
109
139
168
199

0.24
0.205
0.175
0.168
0.161
0.152

part 3. This means that part 3 is the second cavitation structure.
In this experiment, the OLNB was appointed according to the
flow pattern in Figures 4–6. Corresponding to Figure 4c and
Figure 5c, the OLNB was in part 3, and in the others it was
in part 5. Otherwise, the liquid proportion in two-phase flow
became large when the refrigerant mass flux increased. In the
channel center, liquid flow could be observed clearly in Figures
4–6.
Figure 7 shows the influence of refrigerant subcooling degree
on the refrigerant quality and boiling length. The liquid percentage in two-phase flow increased and the length of boiling area
became small when mass flux rose with the fixed heat power
in the experiment. The mass flux rising induced the pressure
increase in the entrance, corresponding to the liquid subcooling
degree of increase. Thus, boiling was weakened by liquid subcooling degree. For cavitation, it depended on liquid velocity
and temperature.

Figure 8 shows the variation of refrigerant pressures and
temperatures with time, and wall temperature distribution along
the channel at differing heat power and mass flux. The transient
changes of pressures and wall temperatures demonstrate that
flow boiling in the microchannel is stable because there are no
relatively observable oscillations. The irregular oscillation of
temperature in the outlet tank was caused by the alternate touch
of the liquid droplet and vapor to the thermocouple, due to the
spray effect of the microchannel exit in the outlet tank. The
wall temperatures increased and then decreased along the flow

Table 2 Calculation of average boiling heat transfer coefficient (Cp = 1.42 kJ/kg-K; ρ = 1210 kg/m3)
Heat power(W)
Power frequency (Hz)
G (kg/h)
f
Pin (MPa)
T in (◦ C)
Pout (MPa)/T out (◦ C)
OLNB
POLNB (MPa)/T OLNBsat
(◦ C)
Qhliquid (W)
T avewall (◦ C)
T avesat (◦ C)
Aboil (mm2)
hboil (kW/m2-K)

8
30

15
40

50

30

23
40

50

30

40

50

4.48
6.2
7.79
4.06
5.85
7.55
3.82
5.33
7.3
0.23
0.17

0.16
0.25
0.178
0.16
0.265
0.196
0.16
0.683
0.726
0.772
0.674
0.722
0.774
0.654
0.708
0.782
23.8
23.85
23.5
24.2
23.9
23
24.12
24.2
23.5
0.588/21
0.584/20.8 0.577/20.1 0.564/19.6
0.551/18.8
0.541/18.4
0.531/17.9

0.529/17.6 0.526/17.4
Part 3
Part 5
Part 5
Part 3
Part 5
Part 5
Part 3
Part 3
Part 5
0.674/25.45 0.646/24
0.651/24.3 0.667/25.06 0.648/24.13 0.661/24.78 0.648/24.13 0.698/26.6 0.676/25.55
2.9
27.4
23.23
55
22

0.37
27.15
22.4
40
40

2.46
26.66
22.2
40
31

1.38
27.1
22.33
55
51

heat transfer engineering

0.51
27.12
21.46
40
64

5.31
27
21.6
40
44

vol. 32 nos. 7–8 2011

0
27.52
21.02
55
64

5.08
28

22.1
55
55

5.9
28.76
21.48
40
58


546

H. Z. CAO ET AL.

Figure 6 Flow pattern while pump power frequency is 50 Hz: (a) heat power
is 8 W and mass flux 7.79 kg/h; (b) heat power 15 W and mass flux 7.55 kg/h;
(c) heat power 23 W and mass flux 7.3 kg/h.

Figure 7 Flow pattern: boiling area became short when mass flux rose from
1.562 kg/h (a) to1.983 kg/h (b),then to 2.043 kg/h (c), with fixed heat power
23 W.

direction because the subcooling liquid refrigerant absorbed
heat into the saturation liquid first, and then the saturation
temperature went down with the decrease of refrigerant pressure
in the microchannel.
Aboil , T avesat , and Qhliquid could be approximately determined
with the OLNB, as defined, and the average boiling heat transfer coefficient (hboil ) could be calculated with the equations
given earlier. Here the liquid specific heat and the liquid density were maintained constant because the inlet temperatures

were approximately equal. The calculation results are showed in
Table 2.
Figure 9 presents the changes of pressure drop, average boiling heat transfer coefficient, and average wall superheating with
the heat power, respectively. Figure 9a shows that the pressure

drop increased when heat power or mass flux rose. In Figure
9, b and c, the average boiling heat transfer coefficient and
average wall superheat rise with the increase of heat power
when the pump power frequency is 30 Hz or 50 Hz. But the
curves are different when rotation rate is 40 Hz. It could be
explained from Table 2 that the OLNB was in the same part
of the channel when the pump power frequency was 30 Hz or
50 Hz but the OLNB changed with the heat power when the
pump power frequency was 40 Hz. Thus, the experimental results shown in Figure 9 present the average boiling heat transfer
coefficient and average wall superheat decrease with the mass
flux increasing at the same heat power, which means boiling
heat transfer in the microchannel was abated by the mass flux
increasing.

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vol. 32 nos. 7–8 2011


H. Z. CAO ET AL.

547

Figure 8 (a) pressure in inlet-tank and outlet-tank; (b) temperature in inlet-tank and outlet-tank; (c) wall temperature; (d) wall temperature distribution along the
channel.

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548

H. Z. CAO ET AL.

Figure 8 (Continued)

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vol. 32 nos. 7–8 2011