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A Roadmap for Implementing Minichannels in Refrigeration and AirConditioning Systems—Current Status and Future Directions
Satish G. Kandlikara
a
Mechanical Engineering Department, Rochester Institute of Technology, Rochester, NY, USA
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DOI: 10.1080/01457630701483497
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A Roadmap for Implementing
Minichannels in Refrigeration and
Air-Conditioning Systems — Current
Status and Future Directions
SATISH G. KANDLIKAR
Mechanical Engineering Department, Rochester Institute of Technology, Rochester, NY, USA
The evolution of compact heat exchangers was driven by the need for reducing their size and weight and enhancing individual
components and overall system performance. These needs are also becoming relevant in stationary refrigeration and air
conditioning applications. The reduction in volume and footprint of the systems is foremost in the minds of urban architects,
while refrigerant charge reduction and performance enhancement remain important safety and operating considerations.
Minichannels provide a viable solution to address these concerns and are expected to be used widely in large as well as small
commercial, industrial, and residential systems. This paper provides a roadmap for their implementation, with a critical
review of the current status of our understanding in this field, recommendations for obtaining fundamental performance data
and correlations, and an emphasis on innovative designs of minichannel heat exchangers.
INTRODUCTION
Challenges Facing Today’s Refrigeration and Air
Conditioning Industry
Refrigeration and air conditioning systems have become an
integral part in the design of spaces occupied by people and
intended to provide controlled thermal, humidity, cleanliness,
and/or other process requirements. As we strive to design these
systems, the constraints imposed on the size and/or specific dimensions of the system components are becoming increasingly
important, while the operating cost remains a critical factor with
high energy prices. The impetus toward improving the energy
efficiency of refrigeration and air-conditioning equipment becomes clear when we recognize that this sector accounts for
15 percent of total energy consumption worldwide [1]. Another
issue of great environmental concern is the release of large quanThe paper was presented as a keynote lecture at the IIR Conference on
Innovative Equipment and Systems for Comfort and Food Preservation held
in Auckland, New Zealand, on February 16–18, 2006, organized by the International Institute of Refrigeration, 177 Boulevard Malesherbes, 75017 Paris,
France; Tel.: 33 1 4227 3235; Fax: 33 1 4763 1798; E-mail: iifiir@iifiir.org;
Web site: http://www.iifiir.org
Address correspondence to Professor Satish G. Kandlikar, Mechanical Engineering Department, Rochester Institute of Technology, Rochester, NY 14623,
USA. E-mail: sgkeme@rit.edu
tities of refrigerants. The risk of accidental leakage, combined
with the increased capital investment (energy and resources required to manufacture the refrigerants), makes it desirable to
have a low refrigerant inventory in the system.
The size and floor space occupied by refrigeration and airconditioning equipment continues to be a major concern in urban environment. With soaring real estate prices, reductions in
volume and floor space of the system components are very desirable. In modern buildings, small equipment size also translates
into smaller passageways for transporting the equipment in and
out, and lower reassembly costs. The size restrictions are becoming important even for the window room air-conditioners
in metropolitan apartments. The major challenges facing the refrigeration and air conditioning industry can be summarized as
follows:
•
•
•
•
improving the system COP (coefficient of performance)
reducing the total refrigerant charge in the system
reducing the footprint and size of the equipment
meeting these challenges in a cost-effective manner
The role of heat transfer in the refrigeration industry was
addressed in a recent editorial [2]. In a conventional refrigeration/air conditioning unit, the heat exchangers offer
significant potential for addressing the above challenges, which
are somewhat similar to those faced in other applications,
including absorption refrigeration, air liquefaction, automotive
973
974
S. G. KANDLIKAR
air conditioning, and high flux chip cooling. Drawing from
the experience gained in these applications, the potential of
minichannel heat exchangers is highlighted and a roadmap for
their implementation is presented.
In refrigeration applications, heat exchangers with minimum
channel dimensions between 200 µm and 3 mm can therefore
be described as minichannel heat exchangers.
CHANNEL SIZE CLASSIFICATION
ADVANTAGES OF MINICHANNEL HEAT
EXCHANGERS
In our efforts to arrive at a common terminology for smaller
diameter channels, several classification schemes based on hydraulic diameters were proposed in literature. A simple channel
classification scheme [3] was based on the hydraulic diameter
as follows:
•
•
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microchannels: 1 µm to 100 µm
meso-channels: 100 µm to 1 mm, and
• conventional passages: >6 mm.
Another approach is to classify the channels according to the
specific processes they are being employed for. In two-phase
applications, Fukano and Kariyasaki [4] and Serizawa
et al. [5]
proposed the use of Laplace length scale, L = σ/(ρ L − ρV ),
for channel size classification based on the relative magnitudes
of the surface tension and gravity forces. Kew and Cornwell
[6] proposed the ratio of the Laplace length scale to channel
hydraulic diameter as the criterion in flow boiling application.
The effect of channel size is seen to be different on heat transfer
and fluid flow phenomena. For example, in the case of singlephase gas flow, the flow passage dimensions are compared with
the mean free path λ of the gas molecules (Knudsen number,
Kn = λ/Dh ). In adiabatic two-phase flows, the relationship between void fraction and flow Reynolds number provides guidance. Kawahara et al. [7] studied the variation of void fraction
with volumetric quality in adiabatic two-phase flow for 50, 100,
and 251 µm diameter tubes and noted a marked change in the
relationship for the two smaller diameter tubes. The authors suggested that the transition boundary between microchannels and
minichannels lies between Dh = 100 µm and Dh = 251 µm.
From a practical standpoint, it is useful to consistently employ
the same classification scheme while considering the overall
behavior of commonly encountered processes, such as singlephase gas and liquid flows, two-phase adiabatic flow, flow boiling, and flow condensation processes. Kandlikar and Grande
[8] considered various processes described above and recommended the channel classification shown in the inset to be used
for all processes. Kandlikar et al. [9] modified the above classification to replace the channel hydraulic diameter with the minimum channel dimension D in the original classification scheme.
Channel size classification [8]
Conventional channels
Minichannels
Microchannels
Transitional channels
Transitional microchannels
Transitional nanochannels
Molecular nanochannels
where D is the minimum channel dimension.
D > 3 mm
3 mm ≥ D > 200 µm
200 µm ≥ D > 10 µm
10 µm ≥ D > 0.1 µm
10 µm ≥ D > 1 µm
1 µm ≥ D > 0.1 µm
0.1 µm > D
heat transfer engineering
Minichannel Heat Exchangers: Overview
A typical minichannel heat exchanger (MCHX) employs a
flat tube with triangular or rectangular passages of 1–2.5 mm
hydraulic diameter. Still smaller passages are being applied in
condenser applications, 25-mm or more wide, and length dependent on the refrigerant and operating conditions. These heat
exchangers are generally employed in refrigerant-to-air heat exchangers, with folded louvered fins employed on the air-side.
Figure 1 shows a schematic of an MCHX.
Minichannel heat exchangers were studied experimentally by
Wang et al. [10]. They used copper plates in which channels of
300 µm to 4.5 mm were etched in brass and stainless steel. The
overall dimensions were 296 × 122 × 30 mm, and the highest
surface area density achieved was 1500 m2 /m3 . The resulting
heat transfer coefficients on the volumetric basis were as high
as 7 MW/m3◦ C. Although these heat exchangers were intended
for biological applications, they demonstrated the heat transfer
potential of narrow channels.
Early work of Webb and Jung [11] showed the advantages
of folded louver fins and flat tubes over continuous fins and the
standard round tubes. The heat transfer coefficient was 90 percent higher for the new design while the pressure drop increased
by only 25 percent. The main shortcomings of the round tube
design were the inefficient airflow performance in the wake of
fins and high thermal contact resistance between the round tubes
and the fins. Brazed aluminum condensers employing minichannels for residential application were studied by Webb and Lee
[12]. Westphalen et al. [13] identified some additional benefits
with these heat exchangers, including all-aluminum construction with improved corrosion resistance, ease of leak repair with
Figure 1 Schematic of a multi-louvered fin and flat tube with minichannel
passages [22].
vol. 28 no. 12 2007
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S. G. KANDLIKAR
epoxy, reduced air-side resistance due to flat tubes incorporating
minichannels, and reduced fan power for condenser and evaporator units.
Utilizing MCHX leads to size reductions and performance
improvements in air-conditioning systems. Kim and Bullard [14]
compared the performance of two different types of MCHX with
the baseline case of finned round-tube condenser and found that
for the same Energy Efficiency Ratio (EER), the MCHX designs yielded a 35 percent reduction in refrigerant charge and
35–55 percent reduction in core volume and weight. Kim and
Groll [15] compared the performance of MCHX in a unitary
split system in an outdoor air conditioning/heat pumping coil.
The MCHX had a 23 percent smaller face area and a 32 percent lower refrigerant volume in the core. It resulted in a 1–6
percent improvement in EER, but the defrost cycle and system
performance were dependent on the air-side fin density and coil
orientation. Further need for research on optimum MCHX configuration for a specific application is clearly brought out from
their work.
Rugh et al. [16] presented their experimental results on a
minichannel heat exchanger used as an air heater with a very
high air-side fin density for heating the passenger compartment
of a car. Rectangular minichannels with internal dimensions of
1.52 mm × 14.4 mm were used. The tubes had 0.25-mm ribs
that effectively reduced the height to 1 mm. On the air-side,
high density louvered fins at 20 fins/cm were provided. The
airflow passages were essentially only 0.5 mm wide. Making the
fins louvered at an angle of 25◦ provided further enhancement.
The water side heat transfer coefficients were as high as 35,600
W/m2◦ C. Such high values were made possible by the small
passage dimensions combined with flow interruptions, which
introduce periodic developing flows as new boundary layers are
formed following each interruption in the flow stream.
Heat Transfer Enhancement with Minichannels and
Microchannels
The heat transfer rate in a given heat exchanger is expressed
through the following basic heat transfer equation:
Q = U A T
(1)
The overall heat transfer coefficient U depends on the fin efficiencies, individual side heat transfer coefficients and fouling
resistances, and the thermal resistance of the separating wall.
The flow inside a minichannel is expected to be laminar due to
the small channel hydraulic diameter [8]. Although the flow will
be under developing conditions, fully developed laminar flow is
considered in the analysis presented here to demonstrate the
enhancement effects due to channel size reduction. Under the
assumption of fully developed laminar flow, the Nusselt number is constant and the heat transfer coefficient in internal flow
varies inversely with channel hydraulic diameter as given by the
following equation:
k
h = Nu
(2)
Dh
heat transfer engineering
975
Figure 2 Variation of heat transfer coefficient with hydraulic diameter for
fully developed laminar flow under constant heat flux boundary condition for
water and R-134a.
Figure 2 shows the variation of h with channel diameter for
fully developed laminar flow in circular tubes under constant
heat flux boundary condition. The heat transfer coefficient with
R-134a in a plain, 10-mm diameter tube is only 35 W/m2◦ C,
while it increases to 350 W/m2◦ C for a 1-mm diameter tube and
700 W/m2◦ C for a 0.5-mm diameter tube. Although enhancement devices such as microfins are used in larger diameter tubes
to improve the heat transfer performance, similar enhancement
techniques are also employed in minichannels to improve their
heat transfer characteristics.
In addition to the heat transfer considerations, pressure drop
penalty also needs to be carefully evaluated when selecting
the channel size. The pressure gradient (pressure drop per unit
length) increases drastically with channel size reduction, but it
was also noted that minichannel heat exchangers need to be suitably designed to provide short flow lengths to limit the overall
pressure drop [8]. Today’s compact heat exchangers are successfully implemented in various applications, and higher pressure
gradients in minichannel heat exchangers simply become a design constraint in arriving at a suitable heat exchanger design.
Heat Exchanger Size Reduction with Minichannels
Another major advantage offered by minichannels is the reduction in the internal volume and the overall size of the heat
exchanger. Neglecting the header and intermediate pass connecting tubes, the surface area available per unit volume of a heat
exchanger employing circular tubes of diameter D and length L
is given by the following equation.
A/V =
πDL
= 4/D
(π/4)D 2 L
(3)
The surface area to volume ratio A/V is an indicator of the compactness of a heat exchanger, and is routinely used in the compact
heat exchanger literature [17]. Figure 3 shows the variation of
this ratio with the tube diameter. The channel sizes of around
10-mm are commonly employed in today’s stationary refrigeration applications, although it is not uncommon to find even larger
vol. 28 no. 12 2007
976
S. G. KANDLIKAR
the internal volume of the heat exchangers with a significant reduction in the refrigerant charge. Hrnjak [19] presents a study
focusing on charge reduction strategies in ammonia refrigeration
systems. He recommends using smaller diameter channels and
plate heat exchangers with small gaps. A higher void fraction
in the heat exchanger is desirable from this standpoint. Hrnjak
reports the following specific refrigerant charge per kW of heat
removal rate in ammonia systems:
•
•
air-cooled condensers: 18–159 g/kW
water-cooled condensers: 23–228 g/kW
• condensers with minichannels: 5–10 g/kW
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Figure 3 Variation of area to volume ratio with channel diameter.
channel sizes, such as in flooded evaporators. The area to volume
ratio for 10-mm diameter channels is found to be 400 m2 /m3 ,
while this ratio becomes 4000 m2 /m3 and 8000 m2 /m3 for 1-mm
and 0.5-mm diameter channels, respectively. In other words, for
the same internal flow volume, a 1-mm diameter tube provides
a 10 times larger surface area as compared to a 10-mm diameter tube. This, coupled with a ten-fold improvement in the heat
transfer coefficient (as seen from Figure 2) provides a total of
100-fold increase in the individual side hA product (heat transfer
coefficient times the surface area) for both water and R-134a in
single-phase flow.
A comparison based on the heat transfer rate per unit volume
of the heat exchanger is illustrative in further recognizing the
size benefits of minichannel heat exchangers.
q
h AT
A
=
= hT
V
V
V
(4)
Equation (4) shows that q/V depends directly on A/V, indicating that this ratio would go up inversely with the diameter as
seen in Fig. 3. A ten-fold decrease in the channel diameter from
10 mm to 1 mm will yield a ten-fold increase in heat transfer
rate per unit volume, in addition to the increase derived from
the improvement in the n associated with the smaller diameter.
The actual gain in performance will be influenced by additional
factors, such as the heat transfer coefficient on the other side
and redesign to accommodate pressure drop constraints [18].
Nevertheless, such dramatic improvements have been realized
in many practical systems. For example, shell-and-tube heat exchangers employing tube diameters of only 1 mm and a total
length of 150 mm are currently available as standard commercial products from a number of manufacturers that outperform
the conventional shell-and-tube heat exchangers employing 10mm diameter tubes with approximately 20 times the volume.
In ammonia evaporators, Pearson [20] and Ayub [21] provide
the following values for typical specific charge in different heat
exchanger types:
•
shell-and-tube: 1000 g/kW
plate: 500 g/kW
• gravity-fed plate: 250 g/kW
• spray-type: 54–113 g/kW
•
In general, specific refrigerant charge is considerably higher in
the evaporators as compared to condensers. Ayub [22] provides
the values of the refrigerant charge in a DX evaporator employing 15-mm enhanced tubes of around 75 g/kW for a 100 ton
evaporator under ARI conditions. Significant refrigerant charge
reductions can be obtained by employing minichannel heat exchangers in these systems. Considering the surface area and
performance enhancements presented earlier with minichannels, a design goal of 10–20 g/kW for refrigerant charge seems
feasible in evaporators. The higher heat transfer rates possible with refrigerant-to-water heat exchangers (as compared to
refrigerant-to-air heat exchangers) will result in further reductions in the refrigerant charge.
Hrnjak [19] also suggests increasing the void fraction in condensers and evaporators to reduce the refrigerant charge in a
system. Withdrawal of refrigerant in condensers at several locations along its length and use of falling film type evaporators
are among several possibilities discussed. Reducing liquid subcooling prior to the expansion valve will provide a higher inlet
quality and higher void fraction in the inlet section of an evaporator. However, reducing the liquid subcooling may have an
adverse effect on the system COP or capacity. The refrigerant
charge penalty is considerably reduced if minichannels are used
in the evaporators.
CURRENT STATUS OF RESEARCH ON MINICHANNEL
CONDENSERS AND EVAPORATORS
Reduction in Refrigerant Charge
The refrigerant charge is another major consideration from
overall cost and environmental viewpoints. The amount of refrigerant in large industrial refrigeration equipment can be in the
range of several hundred kilograms. Using minichannels reduces
heat transfer engineering
This section reviews the fundamental information available on condensation and boiling modes of heat transfer in
minichannels.
vol. 28 no. 12 2007
S. G. KANDLIKAR
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Figure 4
[23].
977
Minichannels with and without microfins studied by Yang and Webb
Minichannel Condensers
Minichannels as small as 0.5 mm in hydraulic diameter have
been successfully employed in condensers for many years, and
the heat transfer, pressure drop, two-phase flow patterns, and system configuration in minichannel condensers have been studied
extensively in literature. Garimella [22a] provides an exhaustive literature survey on these topics; readers are referred to this
source for an in-depth review of the condensation phenomena in
minichannels. A brief overview of the current status is presented
here. The multi-port condenser geometries employed by Yang
and Webb [23], Koyama et al. [24, 25] and Garimella [22a] are
shown in Figures 4–6. Yang and Webb [23] conducted one of the
first studies utilizing a flat rectangular tube with four rectangular minichannels nominally 2 mm × 3.5 mm with and without
microfins. One of the interesting outcomes of their study was
the effect of microfins on condensation heat transfer in small
diameter tubes. The entire subcooled single-phase data, for both
plain and microfin tubes in the Re range from 4,000 to 21,000
was correlated using Petukhov’s single-phase correlation. This
indicates that the microfin effects in single-phase flow are well
accounted for by using the hydraulic diameter. During condensation experiments, the microfins were flooded in the low quality
Figure 6 Channel c/s geometry and test section details [22a].
Figure 5 Minichannels studied by Koyama et al. [24, 25].
heat transfer engineering
region and the heat transfer was comparable to plain tubes using
the hydraulic diameter. In the high quality region, the microfins
showed a definite enhancement, indicating the effect of microfins
on the thin annular liquid film flow.
The work of Koyama et al. [24, 25] is instructive in explaining
the geometry effect on condensing flows in minichannels. For
their internally finned channels, the heat transfer performance
was similar to the plain channels after accounting for the area
enhancement. This work, coupled with Yang and Webb’s [23]
observations, indicates that the microfins are effective in the
high-quality region, and the fins in general provide a heat transfer
enhancement that can be attributed to the increase in surface
area.
vol. 28 no. 12 2007
978
S. G. KANDLIKAR
Table 1 Constant C in Eqs. (5) and (6) in Chisholm [26] correlation
where
Two-phase flow characteristics
Chisholm’s parameter C
Laminar liquid, laminar gas
Turbulent liquid, laminar gas
Laminar liquid, turbulent gas
Turbulent liquid, turbulent gas
5
10
12
21
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In predicting the pressure drop, a number of models are available in literature as summarized by Garimella [22a]. The homogeneous flow model is seen to be promising in a large number
of cases. Another correlation that is widely used in calculating
pressure drop is by Chisholm [26].
Pressure drop correlation by Chisholm [26]:
C
1
+ 2
X
X
(5)
φ2G = 1 + C X + X 2
(6)
φ2L = 1 +
h LO
kL
= 0.023
D
GD
µL
0.8
0.4
Pr
L
(10)
More accurate methods specifically designed for minichannels smaller than 1 mm in diameter have been proposed in literature. The analytical approach by Wang and Rose [34, 35]
and the empirical method of Brandhauer et al. [36] are considerably more complex, but are recommended for better accuracy. Minichannels with internal microfins are used to enhance
the heat transfer performance during condensation. A model
by Yang and Webb [30] takes into account the surface tension
forces explicitly in minichannels with internal microfins and is
recommended for such channels.
Minichannel Evaporators
or
where φ L and φG are the ratios of the two-phase to respective
liquid or vapor single-phase pressure drops. The values of the
constant C for different two-phase flow conditions are given
in Table 1. Mishima and Hibiki [27] modified the Chisholm
constant C as follows for small diameter tubes.
For circular tubes,
C = 21 (1 − e−333D )
(7)
For non-circular geometries,
C = 21 (1 − e−319Dh )
(8)
Some of the available condensation data show that this correlation overpredicts the pressure drop data. For example, Koyama
et al. [24] found that the constant C varied between 5 and
21, depending on the flow conditions. English and Kandlikar
[28] conducted experiments with air-water flow at lower flow
rates, corresponding to the laminar-laminar case in Table 1,
and found that using a value of C = 5 from Table 1—for this
case, in Eq. (8)—yielded very good agreement. Instead of using a constant value of C = 21 in Eqs. (7) and (8), English
and Kandlikar recommend a value of C from Table 1 depending on the laminar/turbulent flow conditions in individual
phases.
Heat transfer during condensation in minichannels has been
studied extensively, and a number of correlations are available in
literature for specific geometries. Some of the more successful
ones are by Akers et al. [29], Yang and Webb [30], Yan and Lin
[31], Wang et al. [32], Cavallini et al. [33], Koyama et al. [24],
Wang and Rose [34, 35], and Bandhauer et al. [36].
The model proposed by Shah [37] for condensation in large
diameter tubes is simple to use and is suggested for preliminary
estimates. The Shah correlation is given by:
h TP
3.8x 0.76 (1 − x)0.04
= (1 − x)0.8 +
h LO
(P/PCrit )0.38
(9)
heat transfer engineering
The implementation of minichannels in evaporators lagged
behind the condenser application mainly because of the instability concerns [38–41]. Recent attention to these problems by
researchers worldwide has resulted in some practical solutions
that are expected to pave the way for widespread implementation
of minichannel evaporators in various applications.
The flow boiling phenomenon in narrow channels was studied
using high-speed imaging techniques by Steinke and Kandlikar
[42] and Kandlikar and Balasubramanian [43]. Figure 7 shows
a high-speed image obtained at 2000 fps of water boiling inside a set of six parallel microchannels, each of 1054 × 197
µm rectangular cross-section. Although the overall flow direction is from left to right, instantaneous directions of liquid vapor
interfaces in adjacent channels are seen to be moving in both
directions. The resulting reverse flow in some of the channels
is detrimental to the operation of an evaporator. It also leads to
backflow of vapor into the inlet manifold, causing flow maldistribution and instabilities. The flow patterns in minichannels
are dominated by surface tension effects, and significant differences can be expected in diabatic flow patterns due to nucleating
bubbles.
The basic mechanisms leading to unstable operation during
flow boiling are discussed by Steinke and Kandlikar [42]. Flow
Figure 7 High-speed images of flow boiling of water in 1054 × 197 µm
parallel microchannels showing backflow in a few channels (main flow from
left to right) [43].
vol. 28 no. 12 2007
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S. G. KANDLIKAR
instabilities were found to be the result of rapid growth of vapor
bubbles nucleating near the inlet region of the microchannels
and minichannels. Kosar et al. [44, 45], and Kandlikar et al. [9]
introduced inlet restrictors and artificial nucleation cavities to
suppress the instabilities. This method seems to have promise,
and further research is warranted in this area.
Another topic that is of great importance is the predictive
ability for heat transfer and pressure drop in minichannels. Qu
and Mudawar [46] and Jacobi and Thome [47] proposed flow
pattern based models to predict the heat transfer. The original
flow boiling correlation by Kandlikar [48] was very successful in predicting the heat transfer coefficient and its trends with
quality for water and refrigerants. The same correlation is extended to minichannels, mainly by recognizing that the all-liquid
flow heat transfer coefficient needed in the correlation may be
in the laminar region. The turbulent single-phase flow correlation is therefore replaced by the appropriate laminar flow correlation. The details are provided by Kandlikar and Balasubramanian [49]. The recommended correlation scheme is given
below.
For ReLO >410:
h TP = larger of
h TP,NBD
h TP,CBD
(11)
h TP,NBD = 0.6683Co−0.2 (1 − x)0.8 h LO
(13)
Co = [(1 − x)/x]0.8 [ρV /ρ L ]0.5
Bo = q/(Gh LV ).
The single-phase all-liquid flow heat transfer coefficient h LO
is given by
ReLO Pr L ( f /2)(k L /D)
2/3
1 + 12.7(Pr L −1)( f /2)0.5
(14)
for 104 < ReLO ≤ 5 × 106
h LO =
(ReLO − 1000) Pr L ( f /2) (k L /D)
2/3
1 + 12.7 (Pr L −1) ( f /2)0.5
FFl
Water
R-11
R-12
R-13B1
R-22
R-113
R-114
R-134a
R-152a
R-32/R-132
R-141b
R-124
Kerosene
FFl =1 for stainless steel tubes
for 3000 ≤ ReLO
1.00
1.30
1.50
1.31
2.20
1.30
1.24
1.63
1.10
3.30
1.80
1.00
0.488
Any refrigerant
≤ 104
h LO =
Nu k
Dh
(16)
for 410 ≤ ReLO ≤ 1600
The friction factor f in Eqs. (14) and (15) is given by
(17)
In the transition region between Reynolds numbers of 1600
and 3000, a linear interpolation is suggested for h LO . The fluiddependent factors FFl for various refrigerants on copper or aluminum are listed in Table 2. For stainless-steel tubes, a value of
1.0 is recommended for all refrigerants.
MINICHANNELS IN ADVANCED REFRIGERATION
SYSTEMS
where
h LO =
Fluid
(12)
h TP,CBD = 1.136Co−0.9 (1 − x)0.8 h LO
+ 667.2Bo0.7 (1 − x)0.8 FFl h LO
Table 2 Recommended FFl (fluid-surface parameter) values in flow boiling
correlation by Kandlikar [48]
f = [1.58 ln(ReLO ) − 3.28]−2
where NBD and CBD refer to the nucleate boiling-dominant and
convective boiling-dominant regions, respectively. Expressions
for the respective h TP in these regions are given below.
+ 1058.0Bo0.7 (1 − x)0.8 FFl h LO
979
(15)
heat transfer engineering
Minichannels in CO2 and Ammonia Systems
Although CO2 was one of the earliest refrigerants used, it
was replaced by the advent of man-made refrigerants. It was not
until 1993 that the interest in transcritical CO2 was revived after
the work of Lorentzen and Pettersen [50] and Lorentzen [51].
Minichannels are well suited for containing the high pressures
encountered in the CO2 systems. Pettersen et al. [52] developed
new designs of MCHX with 0.8-mm diameter tubes placed together forming a 42-mm wide flat plate arrangement (similar
to Figures 4 or 5). Cross-conduction of heat between different
plates is of greater concern in a gas cooler than in a condensing
application. In the design of a gas cooler, the use of single-phase
heat transfer correlations incorporating variable property effects
was recommended because of the large variations in the properties of transcritical CO2 along the length of the heat exchanger.
CO2 systems are also being considered for residential heat
pump application. Richter et al. [53] compared the performance
of an R404A system with a CO2 system designed to fit the
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S. G. KANDLIKAR
same chassis. Garimella and Bansal [54] and Zhao and Ohadi
[55] developed gas coolers for automotive CO2 systems. Flow
boiling in minichannels and overall system performance were
studied by a number of investigators, including Kim and Bullard
[56], Pettersen [57], and Elbel and Hrnjak [58], who showed that
flash gas bypass around the evaporator resulted in an increase in
capacity and a 7–9 percent improvement in the COP. A review of
the flow boiling heat transfer with CO2 is presented in a recent
article by Thome and Ribatski [59]. The backflow described
earlier is also seen to be present, and needs to be addressed in
developing commercial CO2 systems.
Worldwide research in CO2 systems indicates that these are
strong contenders in mobile as well as stationary refrigeration
applications. The use of minichannels enables these systems
to handle high pressures while maintaining a compact design.
However, these systems are still in experimental stages, and
further research is needed before considering their practical implementation. Pearson [60] provides a very succinct overview
of the historical development of the CO2 systems and identifies compressor technology as one of the major challenges. Use
of minichannel heat exchangers is recognized as important in
developing leak-proof, low-charge CO2 systems.
Large ammonia systems, however, are currently in operation
for industrial and commercial refrigeration applications. As discussed in the earlier sections, using minichannels reduces the
component sizes and the amount of ammonia charge in the system. It is therefore reasonable to expect widespread use and
acceptance of minichannels in ammonia systems in the coming
decade.
Minichannels in Absorption System Components
The high surface area-to-volume ratio provided by narrow
channels is also attractive for mass transfer applications. In fact,
some of the very high mass transfer rates within a compact geometry are accomplished in natural systems such as the kidney,
lung, and brain. The absorber in an absorption refrigeration system provides a unique opportunity for utilizing minichannels.
Garimella [61] provides a comprehensive overview of the heat
exchangers in an absorption system that could potentially benefit
with the use of minichannels. He developed an absorber utilizing
short lengths of minichannels in vertical stacks with intermediate distribution trays to promote uniform flow distribution and
wetting, the major areas of concern. Meacham and Garimella
[62] presented a model to predict the local variation in heat and
mass transfer coefficients in an ammonia-water absorber utilizing minichannel absorber passages. The compact design offers
several unique features, such as the ease of modifying pass arrangement, and can be utilized in other heat and mass transfer
components in the system—the desorber, condenser, and evaporator. Employing the same heat exchanger geometry in the entire
system makes it easier from a maintenance viewpoint. The high
cost of system components utilizing minichannel tubes can be
brought down considerably through the mass production of these
components.
heat transfer engineering
RESEARCH NEEDS AND ECONOMIC
CONSIDERATIONS
Secondary Refrigerant and Water-Cooled Systems
In order to fully realize the benefits of minichannel passages for refrigerant system size reduction, the air streams in
the condenser and evaporator need to be replaced with water
or a secondary refrigerant providing a better match of thermal
resistances on the two sides of the respective heat exchangers.
Using water for condenser heat removal and water or secondary
refrigerant in the evaporator is highly desirable. The resulting
size reductions in the heat exchangers will reduce the refrigerant charge further. Pumping power will be lower to move the
liquid streams as compared to the air streams. The lower temperature differences in the heat exchangers due to higher overall
heat transfer coefficients will also increase the system COP. The
design issues and related research needs are well documented
in the literature (e.g., [63]). The major issues include selection
of a proper secondary coolant, component (solenoid valves and
actuators) failures due to continuous low temperature operation,
corrosion, and bacterial activity.
Condenser Design Data
Available research on condensation indicates that minichannels can be effectively utilized in commercial and industrial
refrigeration systems. Further research is needed, however, to
provide specific design information related to heat transfer and
pressure drop in different channel geometries. The effectiveness of enhancement features such as internal fins and microfins needs to be studied further. Design information regarding large capacity condensers with minichannels is also
needed.
Flow Boiling Stabilization
Minichannel flow passages are not as commonly employed
in the evaporator application as in the condenser applications
because of the flow instabilities during the boiling process.
Further research and testing in this area is needed to achieve
stable operation, especially with refrigerants, ammonia, and
CO2 .
Water-to-Refrigerant Minichannel Heat Exchangers
The available minichannel heat exchanger configurations
for water-to-refrigerant heat transfer are somewhat limited. Although some shell-and-tube heat exchangers are developed with
1-mm ID tubes, the manufacturing costs are very high due to
tighter tolerances needed for the cover plates and baffles. The
plate heat exchangers with smaller hydraulic diameter passages
are perhaps ideal among the available choices. Plate-fin heat
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S. G. KANDLIKAR
981
forming a cross-hatch or cross-rib pattern on the outside surfaces
of the plates, similar to that employed in plate heat exchangers.
The passage dimensions, width of the plate, groove arrangement, and other geometrical parameters need to be optimized
for specific applications. Further research on thermal performance and manufacturing aspects of these heat exchangers is
recommended.
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Economic Considerations
Figure 8 Refrigerant-to-water heat exchanger employing enhanced microchannel passages on both sides for refrigeration chiller or condenser application.
exchangers with evaporating and condensing streams are commonly employed in air-liquefaction industry. This arrangement
is not suited for the current applications, as the serrated fin design
will result in rather large pressure drops that may be detrimental to the overall system COP. On the other hand, the extruded
tubes shown in Figures 4–6 are ideally suited for the condensing or evaporating refrigerant streams. New heat exchanger developments specifically addressing the system requirements are
needed. For example, Neksa et al. [64] propose a fan-less gas
cooler that employs 1.1-mm ID tubes attached to vertical plates.
Rane and Tandale [65] employed a tube-tube heat exchanger
(TTHE) arrangement in which minichannels are placed side by
side and bonded thermally using a thermal bonding material for
effective heat transfer. Hot and cold streams flow through the
alternate minichannels, which are bent periodically along the
flow length to enhance heat transfer.
Fundamental understanding of the transport processes is another area of great importance. For example, use of superhydrophobic walls with transverse ribs was shown to reduce the
pressure drop during liquid flow [66]. However, the combined
effect of such methods on heat and momentum transfer needs to
be investigated.
The minichannel heat exchangers that are currently being
used in refrigeration applications are mainly derived as extensions of automotive heat exchangers. The flat tubes containing minichannels are assembled in a plate-fin type arrangement to serve as condensers or evaporators. These geometries
are particularly suited for air-to-refrigerant heat transfer. In the
case of secondary refrigerant or water-cooled condenser applications, heat exchangers are needed to handle the refrigerantto-water or refrigerant-to-secondary refrigerant streams. A new
arrangement is presented in Figure 8 that shows a unit cell with
minichannel flow passages on both refrigerant and water sides.
The minichannels are incorporated within flattened plates. Water or secondary refrigerant flows in the narrow passages formed
between the two adjacent plates. The heat transfer coefficient of
the water is high due to the flow in narrow passages. Further enhancements can be achieved by introducing microfins or grooves
heat transfer engineering
The economic advantages offered by minichannel heat exchangers include the following:
• use of lightweight and low-cost aluminum,
• mass-production capability through advances in brazing technology,
• cost savings through performance enhancement,
• cost savings through low refrigeration charge, and
• cost savings associated with lower space requirements
Detailed analyses of cost savings through refrigerant charge reductions or process variable optimizations are presented in the
literature by a number of investigators. For example, Walker
and Baxter [67] showed that the operating costs are reduced
by 5–15 percent by using a low-charge, distributed, secondary
loop, water-cooled evaporative condenser. Using minichannels
is expected to provide further savings. Selbas et al. [68] and
Mishra et al. [69] conducted thermoeconomic optimization of
vapor compression and vapor-absorption refrigeration systems
by varying the design variables that affect heat transfer rate, pressure drop, and cost factors. It is recommended that a detailed economic analysis be performed (using computerized techniques,
such as by Usta and Ileri [70]) to show the potential benefits
associated with minichannels in large-scale commercial refrigeration and air conditioning systems. Although it is difficult to
assign the heat exchanger costs in such economic analyses, it
is expected that significant cost reductions in minichannel heat
exchangers will be realized as they become more widely used.
The use of aluminum and advances in brazing technology offer
significant benefits in producing minichannel heat exchangers.
ROADMAP FOR MINICHANNEL IMPLEMENTATION
Based on the literature review presented in this paper, it can
be concluded that minichannels provide a great opportunity to
reduce the overall size and cost and improve the performance
of the heat exchangers and the entire system. Further research
and development work is needed for commercial utilization of
this technology. At the same time, innovations are needed in
implementing the minichannels. Some of the opportunities in
arriving at energy-efficient and environmentally friendly systems are outlined by Garimella [71].
Table 3 presents a roadmap showing the implementation
status and specific research needs in residential, commercial,
and industrial refrigeration and air conditioning systems. It is
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S. G. KANDLIKAR
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Table 3 Roadmap for implementing minichannels in refrigeration and air-conditioning applications
Application
Implementation status
Research needs
Residential room air-conditioners
I1: Evaporator technology not ready
I2: Extension of current refrigerant-to-air
condenser technology. Some efforts needed.
Commercial unitary air-conditioners
I1
I3:
Refrigerant-to-water condenser development
needed.
Significant efforts needed.
Industrial and commercial refrigeration, a/c
I1, I3
I4: Use water-to-air minichannel HX (plate-fin
arrangement), technology available from
compact HX and air-liquefaction industry.
I5: Refrigerant-to-water (or secondary
refrigerant) evaporator development needed.
Major efforts needed.
I1, I3, I5
Major efforts needed.
R1: Fundamental stability concerns
during boiling need to be
addressed.
R2: Performance evaluation and
optimization of evaporators and
condensers for specific products.
R3: Water retention and drainage
during dehumidification.
Fundamental research as well as
product design and development
needed.
R1–R3
R4: Design and development of
water-to-refrigerant condensers
with minichannels.
Fundamental research as well as
product design and development
needed.
R1–R4
R5: Design and development of
water-to-refrigerant evaporators
with minichannels.
Fundamental research as well as
product design and development
needed.
R1–R5
Fundamental research as well as
product design and development
needed.
Absorption refrigeration, CO2 , and ammonia
systems
Note: Explanations for terms I1–I5 and R1–R5 appear at their first appearance in the table.
expected that this roadmap will guide us in undertaking targeted research in academia and industry to realize the overall
energy savings that are possible through the implementation of
minichannel heat exchangers.
CONCLUSIONS
A comprehensive overview of the current status of minichannel research as related to refrigeration and air conditioning application is presented. As the minichannels are defined as channels
in the diameter ranges of 200 µm–3 mm, most of the refrigeration heat exchangers can be classified as minichannel heat
exchangers (abbreviated as MCHX).
The surface area and heat transfer coefficient enhancements
associated with minichannel flow can be exploited effectively
to make the refrigeration equipment smaller in size. Significant
refrigerant charge reductions are also possible due to the higher
surface area-to-volume ratio for the minichannels. Additional
advantages include capital cost reductions, reduced environmental impact due to lower refrigerant inventory, and possible improvements in COP of the system.
The minichannel heat exchangers currently being considered
for refrigeration applications are modeled after the existing heat
exchangers in automotive applications. A new initiative is recommended to develop new types of heat exchangers that fully
heat transfer engineering
exploit the advantages of minichannel flow passages on both
sides of refrigerant-to-water or refrigerant-to-secondary refrigerant heat exchangers. Such balanced coupling will enable even
more compact units with associated savings in handling costs
of the fluid streams as well. As an example, a heat exchanger
geometry that is suitable for water-to-refrigerant heat transfer is
presented in Figure 8.
A roadmap for the implementation of minichannels in refrigeration and air-conditioning application is presented, and associated research and developmental efforts needed in academia
and industry are highlighted.
NOMENCLATURE
A
Bo
C
cp
Co
D
Dh
dP/dx
f
FFl
surface area, m2
boiling number, = q/(Gh LV )
constant in Eqs. (5) and (6)
specific heat, J/kg◦ C
convection number, = [(1 − x)/x]0.8 [ρV /ρ L ]0.5
minimum dimension of a channel, or diameter, m
hydraulic diameter, m
pressure gradient, Pa/m
friction factor
fluid-surface dependent parameter in Kandlikar correlation
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S. G. KANDLIKAR
G
h
h LV
k
Kn
L
Nu
P
Pr
Q
q
Re
T
U
V
X
x
mass flux, kg/m2 s
heat transfer coefficient, W/m2◦ C
latent heat, J/kg
thermal conductivity, W/m◦ C
Knudsen number, = λ/D
length, m
Nusselt number (= hDh /k)
pressure, Pa
Prandtl number, = c p µ/k
heat transfer rate, W
heat flux, W/m2
Reynolds number, = GD/µ
temperature difference, ◦ C
overall heat transfer coefficient, W/m2◦ C
volume, m3
Martinelli parameter, = [(dP/dx) L /(dP/dx)G ]1/2
quality or gas mass fraction
Greek Symbols
λ
µ
ρ
σ
φG
φL
mean free path between gas molecules, m
viscosity, Pa/s
density, m3 /kg
surface tension, N/m
= [(dP/dx)TP /(dP/dx)G ]1/2
= [(dP/dx)TP /(dP/dx) L ]1/2
Subscripts
CBD
crit
G
L
LO
NBD
TP
V
convective boiling dominant
properties at critical state
gas
liquid, only liquid fraction flowing
with all flow as liquid
nucleate boiling dominant
two-phase
vapor, only vapor fraction flowing
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Satish Kandlikar is the Gleason professor of mechanical engineering at Rochester Institute of Technology (RIT), Rochester, New York. He received
his Ph.D. from the Indian Institute of Technology
in Bombay in 1975 and has been a faculty member
there before coming to RIT in 1980. His research is
mainly focused in the area of flow boiling and advanced single-phase chip cooling. After investigating
the flow boiling phenomena from an empirical standpoint, which resulted in widely accepted correlations
for different geometries, he started to look at the problem from a fundamental perspective. Using high-speed photography techniques, he demonstrated that small
bubbles are released at a high frequency under flow conditions. His current work
involves stabilizing flow boiling in microchannels, interface mechanics during
rapid evaporation, and advanced chip cooling with single-phase liquid flow. He
has published more than 180 journal and conference papers and presented over
20 keynote papers. He is a fellow member of ASME and has been the organizer of
the five international conferences on microchannels and minichannels sponsored
by ASME. He is the founder of the ASME Heat Transfer chapter in Rochester
and founder and first Chairman of the E-cubed fair (http://www.e3fair.org) science and engineering fair for middle school students in celebration of Engineers’
Week. He is the Heat and History Editor for Heat Transfer Engineering and an
Associate Editor for Journal of Heat Transfer and Journal of Nanofluidics and
Microfluidics.
vol. 28 no. 12 2007