U.S. patent application number 13/115956 was filed with the patent office on 2011-09-22 for heat exchanger having winding channels.
This patent application is currently assigned to MIKROS MANUFACTURING, INC.. Invention is credited to Javier A. Valenzuela.
Application Number | 20110226448 13/115956 |
Document ID | / |
Family ID | 44646284 |
Filed Date | 2011-09-22 |
United States Patent
Application |
20110226448 |
Kind Code |
A1 |
Valenzuela; Javier A. |
September 22, 2011 |
HEAT EXCHANGER HAVING WINDING CHANNELS
Abstract
A winding channel heat exchanger includes a heat transfer member
having winding channels, a manifold, and a cover plate. The
channels' winding design is defined by a non-linear flow axis that
may include a plurality of short pitch and small amplitude
undulations, which cause the flow to change directions, and may
also or alternatively include two or more large amplitude bends
that cause the flow to reverse direction. In one embodiment, the
undulations have varying amplitudes to increase the heat transfer
coefficient along the length of the channel. The winding channels
allow a user to customize the pressure drop to promote good flow
distribution, to achieve improved heat transfer uniformity, and to
improve the heat transfer coefficient.
Inventors: |
Valenzuela; Javier A.;
(Portsmouth, RI) |
Assignee: |
MIKROS MANUFACTURING, INC.
Claremont
NH
|
Family ID: |
44646284 |
Appl. No.: |
13/115956 |
Filed: |
May 25, 2011 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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12188859 |
Aug 8, 2008 |
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13115956 |
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61347949 |
May 25, 2010 |
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Current U.S.
Class: |
165/109.1 ;
165/173; 165/174 |
Current CPC
Class: |
H01L 23/473 20130101;
H01L 2924/0002 20130101; H01L 2924/0002 20130101; F28F 3/12
20130101; H01L 2924/00 20130101 |
Class at
Publication: |
165/109.1 ;
165/173; 165/174 |
International
Class: |
F28F 13/12 20060101
F28F013/12; F28F 9/02 20060101 F28F009/02 |
Goverment Interests
GOVERNMENT LICENSE RIGHT
[0002] The U.S. Government has a paid-up license in this invention
and the right in limited circumstances to require the patent owner
to license others on reasonable terms as provided for by the terms
of Contract No. N65540-06-C-0015 awarded by the U.S. Navy.
Claims
1. A heat exchanger comprising: a heat transfer member including at
least one heat transfer layer; one or more inlet openings disposed
in the heat transfer member; one or more outlet openings disposed
in the heat transfer member; at least one winding channel disposed
in each of the at least one heat transfer layers and constructed
and arranged to carry a fluid, the at least one winding channel
having a length and a non-linear flow axis, the non-linear flow
axis defining a non-linear path between the one or more inlet
openings and the one or more outlet openings; at least one
undulation having an amplitude constructed and arranged to increase
the heat transfer coefficient of the fluid as it passes there
through and further being constructed and arranged to change the
direction of the flow of the fluid as it travels along the
non-linear flow path; wherein during use fluid flows through the
one or more inlet opening in the heat transfer member, into the at
least one winding channel, and flows into the at least one
undulation, the amplitude of the undulation increasing the heat
transfer coefficient as the fluid passes there through, and wherein
the fluid continues to move toward the outlet opening after
changing direction and passing through the at least one
undulation.
2. The heat exchanger of claim 1, wherein the at least one
undulation comprises a plurality of undulations, and wherein the
amplitude of the undulations are constant along the length of the
at least one winding channel.
3. The heat exchanger of claim 1, wherein the at least one
undulation comprises a plurality of undulations, and wherein the
amplitude of the undulations vary along the length of the at least
one winding channel.
4. The heat exchanger of claim 3, wherein the amplitude of the
plurality undulations increases along the length of the at least
one winding channel moving in a direction from the one or more
inlet openings toward the one or more outlet openings.
5. The heat exchanger of claim 1, wherein the at least one winding
channel comprises a plurality of winding channels, the plurality of
winding channels each having one or more undulations, and wherein
winding channels having undulations of similar amplitude are
grouped together, the amplitudes of the undulations varying between
groups of winding channels over the heat transfer member so as to
vary the thermal resistance over an active area of the heat
transfer member.
6. The heat exchanger of claim 1, further comprising a manifold
including an inlet port and an outlet port constructed and arranged
to distribute fluid to and collect fluid from the heat transfer
member.
7. The heat exchanger of claim 1, wherein the at least one winding
channel comprises mini-channels.
8. The heat exchanger of claim 1, wherein the at least one winding
channel comprises micro-channels.
9. The heat exchanger of claim 1, wherein the non-linear flow path
of the at least one winding channel includes an inlet side adjacent
a corresponding inlet opening and an outlet side adjacent a
corresponding outlet opening, the non-linear path further including
at least one pair of bends, each pair having: a) a first bend
constructed and arranged to reverse the direction of the flow of
the fluid as it travels between the corresponding inlet opening and
the corresponding outlet opening such that the fluid flows toward
the inlet side after passing through the first bend; b) a second
bend constructed and arranged to reverse the direction of the flow
of the fluid as it travels between the corresponding first opening
and the corresponding second opening such that the fluid flows
toward the outlet side after passing through the second bend; and
wherein during use fluid flows from the manifold, through the
corresponding inlet opening in the heat transfer member, into the
at least one winding channel and travels along the non-linear path
toward the outlet side of the non-linear channel and flows into the
first bend which reverses the direction of the fluid flow toward
the inlet side of the channel, the fluid thereafter flowing into
the second bend which reverses the direction of the fluid flow
toward the outlet side of the channel, the flow of fluid traveling
along the non-linear path to the outlet opening.
10. The heat exchanger of claim 9, wherein the first bend, the
second bend and the at least one undulation each have an arcuate
shape.
11. The heat exchanger of claim 1, wherein the depth of the at
least one winding channel is substantially equal to the width of
the at least one winding channel.
12. The heat exchanger of claim 1, wherein the at least one heat
transfer layer comprises a bonded stack of at least two
laminations.
13. The heat exchanger of claim 1, wherein the at least one heat
transfer layer comprises a single heat transfer layer.
14. A heat exchanger comprising: a heat transfer member including
at least one heat transfer layer, the at least one heat transfer
layer having a first surface and a second surface and including a
thickness extending between the first surface and the second
surface; a manifold including an inlet port and an outlet port
constructed and arranged to distribute fluid to and collect fluid
from the heat transfer member; one or more inlet openings disposed
in the heat transfer member and in fluid communication with the
manifold; one or more outlet openings disposed in the heat transfer
member and in fluid communication with the manifold; at least one
winding channel disposed in each of the at least one heat transfer
layers, the at least one winding channel having a non-linear flow
axis, the non-linear flow axis defining a non-linear path between
the one or more inlet openings and the one or more outlet openings,
the non-linear flow path having an inlet side adjacent a
corresponding inlet opening and an outlet side adjacent a
corresponding outlet opening, the non-linear path further including
at least one pair of bends, each pair having: a) a first bend
constructed and arranged to reverse the direction of the flow of
the fluid as it travels between the corresponding inlet opening and
the corresponding outlet opening such that the fluid flows toward
the inlet side after passing through the first bend; b) a second
bend constructed and arranged to reverse the direction of the flow
of the fluid as it travels between the corresponding first opening
and the corresponding second opening such that the fluid flows
toward the outlet side after passing through the second bend;
wherein during use fluid flows from the manifold, through the
corresponding inlet opening in the heat transfer member, into the
at least one winding channel and travels along the non-linear path
toward the outlet side of the non-linear channel and flows into a
first bend which reverses the direction of the fluid flow toward
the inlet side of the channel, the fluid thereafter flowing into
the second bend which reverses the direction of the fluid flow
toward the outlet side of the channel, the flow of fluid traveling
along the non-linear path to the outlet opening.
15. The heat exchanger of claim 14, wherein the at least one
winding channel comprises mini-channels.
16. The heat exchanger of claim 14, wherein the at least one
winding channel comprises micro-channels.
17. The heat exchanger of claim 14, wherein reversing the direction
of the fluid flow toward the inlet side and back toward the outlet
side of the non-linear path results in substantially uniform
thermal resistance throughout the heat transfer member.
18. The heat exchanger of claim 14, further comprising at least one
undulation constructed and arranged to change the direction of the
flow of the fluid as it travels along the non-linear flow path,
without reversing the direction of the flow of the fluid, wherein
if the fluid is moving toward the outlet side before passing
through the at least one undulation, the fluid continues to move
toward the outlet side after passing through the at least one
undulation and wherein if the fluid is moving toward the inlet side
before passing through the at least one undulation, the fluid
continues to move toward the inlet side after passing through the
at least one undulation.
19. The heat exchanger of claim 18, wherein the first bend, the
second bend and the at least one undulation each have an arcuate
shape.
20. The heat exchanger of claim 14, wherein the depth of the at
least one winding channels is substantially equal to the width of
the winding channel.
21. The heat exchanger of claim 14, wherein the at least one heat
transfer layer comprises a bonded stack of at least two
laminations.
22. The heat exchanger of claim 14, wherein the at least one heat
transfer layer comprises a single heat transfer layer.
23. The heat exchanger of claim 14, further comprising a flow
restrictor plate disposed between the manifold and the heat
transfer member, the flow restrictor plate including a body having
a plurality of openings disposed there through that are configured,
dimensioned and positioned to vary the flow to the winding channels
according to the heat transfer requirements.
24. The heat exchanger of claim 23, wherein the body of the flow
restrictor plate is constructed and arranged to selectively block
the fluid flow to the winding channels in sections of the heat
transfer member
25. The heat exchanger of claim 14, wherein the at least one
winding channel comprises a plurality of winding channels, the
plurality of winding channels being disposed over an active area of
the heat transfer member in grouping of two or more, the number of
winding channels per unit area varying between groups so as to vary
the thermal resistance over the active area of the heat transfer
member.
Description
RELATED APPLICATIONS
[0001] The present application claims priority under 35 USC
.sctn.119(e) to U.S. provisional application No. 61/347,949, filed
on May 25, 2010, and is a continuation-in-part under 37 CFR
.sctn.1.53(b) of U.S. application Ser. No. 12/188,859, filed Aug.
8, 2008, the entire contents of all the above applications are
herein incorporated by reference in their entirety.
TECHNICAL FIELD
[0003] This invention relates generally to an apparatus for cooling
a heat-producing device and, more specifically, to a liquid cooled
heat exchanger having winding, non-linear channels.
BACKGROUND
[0004] The use of heat exchangers for cooling a range of heat
producing devices, for example, electronic devices is known in the
art. Liquid cooled heat exchangers are generally characterized as
having macro-channels, mini-channels, or micro-channels, depending
on the size of the channels. The term `micro` is applied to devices
having the smallest hydraulic diameters, generally between ten to
several hundred micrometers, while `mini` refers to diameters on
the order of about 0.5 mm to about 2 millimeters, and `macro`
channels are the largest in size, generally greater than about 2
millimeters. An example of a typical macro channel design is the
conventional swaged-tube cold plate illustrated in FIG. 1a.
[0005] As shown in FIG. 1a, the prior art swaged-tube cold plate
includes a copper tube 11 swaged into grooves machined in an
aluminum plate 13. Swaged tubes are generally suitable for cooling
large-area devices, particularly when cost is a factor, and/or when
the cooling requirements do not require a very low thermal
resistance. The lowest thermal resistance that can generally be
achieved with a conventional swaged-tube cold plate is
approximately 2.degree. C./(W/cm.sup.2). Because of these
limitations, applications requiring lower thermal resistances often
use finned cold plates, such as the prior art finned cold plate
shown in FIG. 1b.
[0006] Conventional finned cold plates have a number of closely
spaced fins 21 attached to the heat transfer surface 23. The fluid
flows through the channels 25 formed by the spaces between the
fins. The channels typically have a width between about 1 to 5 mm.
Conventional finned cold plates can achieve thermal resistances as
low as approximately 1.degree. C./(W/cm.sup.2).
[0007] The thermal resistance of macro channel cold plates
decreases as the flow rate is increased and approaches
asymptotically a minimum value at a flow of about 0.1 LPM/cm.sup.2.
Increasing the flow rate further has not been found to result in an
additional reduction in the thermal resistance.
[0008] For cooling high heat flux devices, such as solid-state
laser diodes, which dissipate heat at a rate of 500-1000
W/cm.sup.2, cold plates with substantially lower thermal resistance
than that of the swaged-tube cold plates or the machined fin cold
plates are needed. In these applications, micro-channel cold plates
are generally employed.
[0009] There are two primary types of prior art micro-channel cold
plates: parallel flow and normal flow. As the name implies,
parallel flow micro-channel cold plates have the liquid flowing
through the heat transfer passages in a direction parallel to the
surface being cooled. In contrast, normal flow micro-channel cold
plates (NCP) have the liquid flowing through the heat transfer
passages in direction normal to the surface being cooled. The
parallel flow cold plates have geometries similar to that of the
finned cold plate shown in FIG. 1b, except that the dimensions are
scaled down by an order of magnitude. For example, the channel
width in a micro-channel cold plate is typically less than 500
microns. Because of the high-pressure drop in the micro-channels,
the size of the parallel flow micro-channel cold plates is
typically less than about 1 cm to 2 cm on a side. Even at these
small sizes, the pressure drop can be too large for some
applications. The pressure drop can be reduced by subdividing the
micro-channel into several sections and providing alternating inlet
and outlet manifolds along the length of the cold plate, for
example as described in U.S. Pat. No. 6,986,382 to Upadhya.
[0010] One objective in the design of conventional micro-channel
cold plates is to minimize the pressure drop consistent with
achieving the target thermal performance. Minimizing the flow
length and maximizing the flow area of the micro-channels is most
often employed to achieve this objective. Conventionally, the flow
length is minimized by making the flow axis straight, while making
the micro-channel depth large compared to its width maximizes the
flow area. As such, prior art parallel-flow micro-channels have a
depth that is an order of magnitude larger than the width.
[0011] Normal flow cold plates invented by the present inventor,
Javier Valenzuela and as described in U.S. Pat. Nos. 5,145,001 and
6,935,411 among other patents, demonstrate excellent heat transfer
effectiveness, especially in high heat-flux applications. However,
for some systems the highly effective cooling provided by the
normal flow design is not required, and the cost of the heat
exchanger may not be warranted. FIG. 2 shows one example of a
cross-section of a prior-art normal flow micro-channel cold plate.
Normal-flow micro-channel cold plates incorporate a low-pressure
drop manifold structure 27 that distributes and collects the flow
over the active area of the cold plate. The micro-channels 29 are
embedded in a thin layer between the manifold structure and the
active surface 31. The micro-channels direct the fluid in a
direction substantially normal to the active surface: first from
the manifold structure towards the active surface, and then from
the active surface towards the manifold structure. The total length
of the micro-channels is very short, about twice the thickness of
the heat transfer layer and, therefore, the pressure drop in the
normal flow micro-channel cold plates is small, even at the high
flow rates per unit area required in these high heat flux
applications.
[0012] In spite of the order of magnitude lower thermal resistances
that can be obtained through the use of micro-channels, they are
seldom used in large-area cold plates. The principal objections to
the use of micro-channels in large area cold plates are: (1) the
large pressure drop associated with the flow through long,
small-hydraulic-diameter passages, and (2) the relatively high cost
of fabricating passages with such small dimensions. The cost of
fabrication and large pressure drops can also prevent
micro-channels from being used in other applications as well. When
micro-channels are not utilized, macro or mini-channels may be
utilized and performance is sacrificed for cost, ease of use or
other requirements.
[0013] There are also other methods of cooling that utilize fluid
flowing through channels in order to cool a device. For example,
U.S. Pat. No. 6,213,194 discloses the use of a hybrid cooling
system for an electronic module which includes refrigeration cooled
cold plate and an auxiliary air cooled heat sink. The '194 patent
also discloses the use two independent fluid passages embedded in
the same cold plate to provide redundancy. A single serpentine
passage, akin to that of a swaged tube cold plate, or multiple
straight passages feed by headers, akin to a finned cold plate, is
used for each one of the redundant systems.
SUMMARY
[0014] In accordance with the present disclosure, there is provided
a winding channel heat exchanger that includes a heat transfer
member having winding channels for cooling a heat-producing
device.
[0015] The channels' winding design is defined by a non-linear flow
axis that, in one embodiment, has a plurality of short pitch and
small amplitude undulations, which cause the flow of fluid in the
channels to change directions. The winding channels may also
include two or more large amplitude bends that cause the flow to
reverse direction. The winding channels advantageously allow a user
to customize the pressure drop to promote good flow distribution,
achieve improved heat transfer uniformity, and increase the heat
transfer coefficient.
[0016] The heat transfer member includes one or more heat transfer
layers, each layer having one or more inlet openings and
corresponding outlet openings. Each of the winding channels is in
fluid communication with at least one of the inlet openings and at
least one of the corresponding outlet openings, such that the
cooling fluid enters the inlet openings, flows along the channels,
and exits via the outlet openings. In one embodiment, the openings
are arranged in rows through each layer, each opening extending
from the first surface through to the second surface of each heat
transfer layer.
[0017] A manifold can supply fluid to each of the inlet openings of
the heat transfer member and receives fluid from each of the outlet
openings of the heat transfer member. The manifold may distribute
and collect the fluid throughout the active heat transfer area in
order to promote uniform heat transfer throughout the area.
[0018] In alternate embodiments disclosed herein the configuration
of the winding channels is modified according to the particular
application, but in all embodiments, the winding channel's axis
remains non-linear along at least a portion of the length of the
channel.
BRIEF DESCRIPTION OF THE DRAWINGS
[0019] The foregoing and other objects, features and advantages
will be apparent from the following description of particular
embodiments of the invention, as illustrated in the accompanying
drawings in which like reference characters refer to the same parts
throughout the different views. The drawings are not necessarily to
scale, emphasis instead being placed upon illustrating the
principles of various embodiments of the invention.
[0020] FIG. 1a is a perspective view of a prior art swaged-tube
cold plate;
[0021] FIG. 1b is a perspective view of a prior art finned cold
plate with its cover raised;
[0022] FIG. 2 is a cross-sectional view of a prior art normal flow
heat exchanger;
[0023] FIG. 3a is an exploded perspective view of a winding channel
heat exchanger according to a first embodiment of the present
invention;
[0024] FIG. 3b is an enlarged view of a winding channel of FIG.
3a;
[0025] FIG. 4a is a top plan view of the heat exchanger of FIG. 3a,
with its cover removed;
[0026] FIG. 4b is a cross sectional view taken along lines 4b-4b of
FIG. 4a;
[0027] FIG. 5 is a comparison of the winding channel design
according to the present application with prior art linear designs
where:
[0028] FIG. 5a is a schematic view of the winding channel of FIG.
3b with numerals 0-5 representing the average fluid temperature in
the winding channel segment according to the present invention;
and
[0029] FIG. 5b is a schematic view of a prior art linear channels
with numerals 0-5 representing the average fluid temperature in the
channels segment;
[0030] FIG. 6 is a comparison of the winding channel design
according to the present application with prior art linear designs
where:
[0031] FIG. 6a is a schematic view of the winding channel of FIG.
3b according to the present invention;
[0032] FIG. 6b is a schematic view a second prior art linear
channel configuration in which the distance between the inlet and
exit openings remains the same as in the winding channel
configuration of FIG. 6a;
[0033] FIG. 7 is a partial top plan view in cross-section of an
increasing amplitude mini-channel heat transfer member according to
one embodiment of the present invention;
[0034] FIG. 7a is a graph showing the heat transfer coefficient
enhancement and pressure drop of exemplary winding channel geometry
according to the present invention;
[0035] FIG. 7b is a graph showing test data comparing performance
of a prior art swage tube and normal flow heat exchangers to two
exemplary winding channel heat exchangers according to the present
invention;
[0036] FIG. 7c is a graph comparing the pressure drop of a straight
channel to that of two winding mini-channels each having different
amplitude undulations according to the present invention;
[0037] FIG. 7d is a graph comparing the pressure drop of a straight
channel to that of two winding mini-channels each having different
wavelength undulations according to the present invention;
[0038] FIG. 7e is a graph comparing the heat transfer coefficient
of a straight channel to that of two winding mini-channels having
different amplitude undulations according to the present
invention;
[0039] FIG. 7f is a graph comparing the heat transfer coefficient
of a straight channel to that of two winding mini-channels having
different wavelength undulations according to the present
invention;
[0040] FIG. 8 is an exploded perspective view of the winding
channel heat exchanger of FIG. 3a including a flow restrictor
plate.
[0041] FIGS. 9a-9e are schematic views of alternate winding channel
designs according to the present invention; and
[0042] FIG. 10 is a top plan view of an alternate embodiment of a
heat transfer member including winding channels according to the
present invention.
DETAILED DESCRIPTION
[0043] The embodiments disclosed herein relate to a heat exchanger
having winding channels. As used herein, the term "winding" is used
to mean a twisting, serpentine, sinuous path, or the like, which
may have a curvature or be angular, and which creates a non-linear
path between an inlet and an outlet. As also used herein the term
micro-channel is used in the conventional sense with respect to
liquid-cooled heat exchanger technology and does not have specific
dimensional constraints, but is generally understood to mean
channels having the smallest hydraulic diameters, generally between
ten to several hundred micrometers, although it is understood that
as industry standards change, so too may the dimensions of
micro-channels. Likewise, as also used herein the term mini-channel
is used in the conventional sense with respect to liquid-cooled
heat exchanger technology, and does not have specific dimensional
constraints, but is generally understood to mean channels that have
diameters on the order of about 0.5 mm to about 2 millimeters,
although it is understood that as industry standards change, so too
may the dimensions of mini-channels.
[0044] Referring initially to FIGS. 3a-4b, a first, exemplary
winding channel heat exchanger 10 including a manifold 12, a heat
transfer member 14 having winding channels 30, and a cover plate
15, is illustrated. In use, heat is transferred to or from the heat
exchanger 10 over the portion that is enclosed by the dashed line
17, otherwise referred to as the "active area", which corresponds
to the portion of the heat exchanger 10 that includes winding
channels 30, as described in more detail below. During use, the
heat transfer member 14 is placed into thermal contact with a
device to be cooled, as is known in the art, and fluid flows
through the winding channels 30 in order to cool the device. In the
heat exchanger 10 of the present embodiment, the winding channels
are illustrated as winding micro-channels formed in one or more
layers 28 of heat transfer member 14. Alternately, the channels may
be larger, mini-channels and may be formed in a single layer or
multiple layers. The description that follows, while referring to
micro-channels is not limited to such channels, and applies to
other size channels as well, in particular mini-channels. The
plurality of channels may be identical to the exemplary winding
channel, as in the present embodiment, or may be varied. For
example, the plurality of channels formed in the one or more layers
may be mirror images, may be symmetrical or non-symmetrical, or may
have different geometries from each other, and may be varied from
the exemplary winding channel described herein, as described in
greater detail below.
[0045] The winding micro-channels 30 according to the present
application each include a non-linear flow axis 36, as best shown
in FIG. 3b. The non-linear flow axis 36 may include one or more
undulations 38 that cause the flow to change directions, as well as
one or more pairs of bends 40a, 40b that cause the flow to reverse
direction. A reference line "CL" bisects the length of the channel
30 through the approximate center between the inlet opening 32 and
outlet opening 34. In the following description, the portion of the
channel 30 disposed between the line "CL" and the inlet opening 32
is referred to as the inlet side (I.sub.S), and the portion of the
channel disposed between the line "CL" and the outlet opening 34 is
referred to as the outlet side (O.sub.S). It will be understood
that the line "CL" is provided for reference purposes only and is
not part of the actual heat exchanger. The fluid flow is reversed
in that the first bend 40a reverses the direction of the fluid flow
from traveling from the inlet side (I.sub.S) toward the outlet side
as represented by arrow A, to a direction traveling from the outlet
side (O.sub.S) toward the inlet side of the channel as represented
by arrow B. Likewise, the second bend 40b reverses the direction of
the fluid that is now flowing toward the inlet side (arrow B), and
re-directs the fluid flow back toward the outlet side (arrow C) of
the channel. In order that the flow of fluid ultimately reaches the
outlet side and outlet opening 34, for each bend that changes the
direction of flow toward the inlet side I.sub.S of the channel
there is preferably a corresponding bend that changes the flow back
toward the outlet side (O.sub.S) of the channel.
[0046] In addition to the one or more pair of reversing bends 40a,
40b, the winding micro-channel 30 may also include one or more
undulations 38 that change the direction of the fluid flow, but
which do not reverse the direction of the fluid flow. In the
present embodiment, the undulations 38 have a smaller amplitude "a"
than that of the bends 40a, 40b. These smaller amplitude
undulations 38 change the local direction of the fluid flow without
reversing the overall direction so that the flow continues in the
same overall direction the fluid was traveling before reaching the
undulation. It will be appreciated that the number and size of the
bends and undulations can be varied depending upon the particular
application, and the micro-channels may include both bends and
undulations or include only bends or only undulations.
[0047] In the present embodiment, a bonded stack of three layers 28
is illustrated and may be varied according the needs of the
particular application, as would be known to those of skill in the
art. Each one or more layer 28 is generally planar and includes a
first surface 10a and a second surface 10b, opposite the first
surface. The micro-channels 30 may be formed in the second surface,
for example by etching, such that the micro-channels have a depth
that is less than the thickness "t" of their corresponding layer.
The micro-channels may be formed by alternate methods as would be
known to those of skill in the art, and may have a depth equal to
the corresponding layer in some embodiments. Unlike the prior art,
parallel flow micro-channels; the winding micro-channels 30 of the
present invention are characterized by having an aspect ratio close
to unity (i.e., the depth of the micro-channels is comparable to
the width). The depth of the micro-channels may preferably be
between about 1/2 the size of that of the width to about 11/2 the
size of that of the width for the depth and width to be comparable.
For example, if the width of the channel is 2 mm, the depth may be
from about 1 mm to about 3 mm.
[0048] As best shown in FIG. 4a, the heat transfer layers 28 each
have a plurality of inlet openings 32 and corresponding outlet
openings 34 arranged in substantially parallel rows a, b, c, d, e,
and f through each layer, each opening extending from the first
surface 10a through to the second surface 10b. Each winding
micro-channel 30 is in fluid communication with at least one of the
inlet openings 32 and at least one of the corresponding outlet
openings 34. Each of the inlet openings 32 and outlet openings 34,
in turn, is in fluid communication with corresponding inlet
channels 22 and outlet channels 24 of the manifold 12 (FIG. 4b).
Each winding micro-channel 30 may share their inlet openings and/or
outlet openings, although the winding micro-channels may
alternately have independent inlet openings and outlet openings.
During use, fluid is provided from the manifold 12 and flows into
the inlet openings 32 and into each of the micro-channels 30. The
fluid then flows through the micro-channels 30 and out of each of
the outlet openings 34 which return the fluid to the manifold
12.
[0049] In the present embodiment, the functions of distributing and
collecting the fluid over the active heat transfer area 17, and
transferring the heat between the fluid and the active heat
transfer area 17 may be achieved by two separate components: the
manifold 12 and the heat transfer member 14, respectively. This
separation in functions allows the selection of the flow passage
geometry in each component to the benefit of their respective
functions.
[0050] The manifold 12 may distribute and collect the fluid over
the entire heat transfer surface 17 in order to promote uniform
heat transfer over the surface. In the present embodiment, the
manifold has an interdigitated design as described below. However,
alternate manifold designs may be utilized, such as traditional
linear manifolds that are not interdigitated and which may include
only a single channel. As best shown in FIGS. 3a and 4a, in the
interdigitated manifold, the fluid enters the heat exchanger 10
through inlet port 16 of the manifold, which may be disposed near a
first edge of the manifold 12. Inlet port 16 is fluidly connected
to the inlet header 20, such that fluid is fed from the inlet port
16 to the header 20. The header 20 distributes the inlet fluid
along the y-axis of the manifold and is fluidly connected to the
inlet channels 22, such that the fluid is fed from the header 20 to
the inlet channels 22 which distribute the fluid along the x-axis
of the manifold. The fluid is then directed to the heat transfer
member 14 where it enters inlet openings 32, is directed through
fluidly connected winding micro-channels 30, and exits the
micro-channels through corresponding outlet openings 34. The fluid
then flows into a plurality of outlet channels 24 that are
interdigitated, i.e. alternating, with the inlet channels 22, along
the x-axis of the manifold 12. The outlet header 26, in turn,
collects the fluid exiting the outlet channels 24 along the y-axis
of the manifold. Fluid exits the heat exchanger 10 through outlet
port 18 that is disposed near the opposite edge of the manifold in
the present embodiment.
[0051] In the winding micro-channel heat exchanger 10, as the fluid
is being distributed and collected it is desirable to minimize the
pressure drop in the manifold to promote good flow distribution. As
discussed in greater detail below, it is also desirable to keep a
suitably small separation between the heat transfer member inlet
openings 32 and the respective outlet openings 34 to promote
uniform heat transfer over the heat transfer area 17. These
requirements are conflicting since a small pressure drop would
favor large dimensions for the manifold channels 22 and 24, whereas
a small separation between the heat transfer member inlet and
outlet openings would favor small dimensions for the channels.
[0052] The distance between the inlet 32 and outlet openings 34 of
the heat transfer member 14 determines the minimum length, and
thereby the minimum pressure drop, of the micro-channel passages;
and the distance also determines the degree of temperature
uniformity (or heat transfer uniformity) which can be achieved
throughout the heat transfer member. To make best use of the flow
heat transport capacity, and thereby minimize the flow and pressure
drop requirements for a given application, it is desirable that the
fluid exit temperature be close to the temperature of the surface
of the heat transfer member (i.e. high heat exchanger
effectiveness). The temperature difference between the fluid and
the micro-channel walls is greater near the inlet openings 32 than
the outlet openings 34, thereby providing greater heat transfer
capability near the inlet openings than near the outlet openings.
The variation in heat transfer capability is mitigated by heat
conduction in the heat transfer member 14 along a plane parallel to
the heat transfer surface. For high thermal conductivity materials,
such as copper, this mitigation is most effective when the distance
between the inlet and outlet openings is no more than a factor of 5
to 10 times larger than the thickness of the heat transfer member
14. Hence for a member 0.5 mm thick, the distance between inlet and
outlet ports should be between about 2.5 to 5 mm.
[0053] In some applications it may be desirable to increase the
pressure drop if it is unacceptably low. For example, in large area
cold plates (larger than about 2.times.2 cm), it has been
determined that at certain flow rates, air bubbles can block linear
micro-channels. It has been determined that at the intended water
flow rate, the pressure drop through the micro-channel heat
transfer member was lower than the bubble point of the
micro-channels, and hence insufficient to drive the bubbles out of
the micro-channels, an unexpected result of the use of the linear
micro-channels in a large-area application. As such, any gas
present in the system could block areas of the heat exchanger,
resulting in undesirable hot spots. Moreover, since good flow
distribution requires that the pressure drop in the manifold be an
order of magnitude smaller that in the heat transfer member, such a
low pressure drop in the heat transfer member would place undue
constraints on the manifold pressure drop, requiring the use of
much larger manifold. Thus, contrary to expectations and to the
common perception that micro-channels are not desirable because
they have too large a pressure drop, the Applicant determined that
particularly in large-area applications the opposite was actually
true. In particular, that the desired inlet-to-outlet channel
spacing, the pressure drop of conventional micro-channels is too
low at the typical flow rates employed in large-area cold plates to
achieve acceptable performance.
[0054] As discussed in greater detail below, the winding channel
configuration disclosed herein provides a way to regulate the
pressure drop in the heat transfer member to a desired value, while
at the same time improving both the heat transfer capability and
the heat transfer uniformity of the heat exchanger. At the flow
rates per unit area typical of large area cold plates, the flow in
the micro-channels is laminar (Reynolds number typically less than
about 500) and the pressure drop in the micro-channels is
proportional to the product of the velocity and the micro-channel
length, and inversely proportional to the micro-channel hydraulic
diameter. Therefore, increasing the length, increasing the
velocity, or decreasing the diameter can all increase the pressure
drop. The performance of micro-channel heat exchangers improves as
the diameter of the micro-channels is decreased. Hence the diameter
is often selected as the minimum diameter consistent with other
considerations, such as ease of manufacture, or filtration level
requirements. Therefore, for the purpose of comparing different
micro-channel configurations, the diameter is not considered a
design variable.
[0055] FIGS. 5 and 6 compare two channel configurations (which
could be micro or mini) designed to have nominally the same
pressure drop when operating at the same flow per unit area of the
cold plate. FIGS. 5a and 6a illustrate schematic diagrams including
winding channels according to the present invention. FIG. 5b
illustrates a first, conventional, prior-art linear channel
configuration in which the desired pressure drop has been achieved
by making the overall length of the linear channel the same as that
of the winding channel. Equal length is achieved by increasing the
distance between the inlet and the exit openings. FIG. 6b
illustrates a second conventional, linear channel configuration in
which the distance between the inlet and exit openings remains the
same as in the winding channel configuration, and desired pressure
drop has been achieved by increasing the flow velocity in the
linear channel. Increased velocity is achieved by decreasing the
number of channels per unit area of the cold plate.
[0056] As discussed below, the winding channel configuration is
advantageous relative to the straight (i.e. linear) channel
configurations on two counts: (1) improved heat transfer
uniformity; and (2), greater average heat transfer capability. To
illustrate the two points, the total length of each channel
depicted in FIGS. 5 and 6 has been divided into 6 segments of equal
length. The numerals 0-5 are assigned to each segment to represent
the average fluid temperature in that segment, the lowest number,
"0", near the inlet opening representing the coolest temperature,
and the highest number, "5", near the exit opening representing the
highest temperature. The heat transfer surface in FIG. 5 is further
divided by the dashed lines into six vertical columns, or strips,
labeled A-F. Similarly, the heat transfer surface in FIGS. 6a and
6b is divided by the dashed lines into two vertical columns, or
strips, labeled A-B and A'-B'. For each channel configuration in
FIGS. 5 and 6, the average fluid temperature for each strip is then
computed as numerical average of the fluid temperatures within that
strip. The average strip temperatures are indicated with an
underscore.
[0057] For the winding configurations 5a and 6a the average fluid
temperatures alternate between values of 2.3 to 2.7 between the
strips. For the linear channel configuration 5b with increased
distance between the inlet and exit ports, the average fluid
temperatures range from 0 near the inlet port to 5 near the exit
port. For linear channel configuration 6b with increased velocity
in the channels, the average fluid temperatures alternate between 1
and 4. It will be readily appreciated that the winding
configuration disclosed in the present invention aids in providing
greater uniformity in fluid temperature, and hence greater
uniformity in heat transfer, over the heat transfer surface.
[0058] The winding channels of FIGS. 5a and 6a are also
advantageous over the prior art linear channels because they
provide a higher average heat transfer capability. The average heat
transfer capability of the heat transfer member is a function of
the product of the total channel wall area and the channel heat
transfer coefficient: the higher this product, the higher the
average heat transfer capability of the heat exchanger. For a
conventional, linear micro-channel operating in the laminar flow
regime, the heat transfer coefficient depends only on the
micro-channel geometry and fluid properties, and is independent of
the fluid velocity. For the winding channels, in contrast, the
frequent changes in flow direction result in a heat transfer
coefficient larger than that of a linear channel of equal
cross-section, and the magnitude of the enhancement increases with
increasing velocity. Hence, the average heat transfer capability of
the winding channel configuration 5a will be greater than that of
the linear channel configuration 5b, even though the channel wall
areas are comparable. The average heat transfer capability of the
winding channel configuration 6a will also be substantially greater
than that of the linear channel configuration 6b because in
addition to a higher heat transfer coefficient, the winding channel
has a wall area about three times larger than that of the linear
channel. These advantages would be found when used in
micro-channels or larger channels, such as mini-channels. Thus, it
will be readily appreciated that the winding channel configuration
disclosed in the present invention aids in providing greater
average heat transfer capability.
[0059] The following examples are provided as comparisons, are
intended to be illustrative in nature, and are not to be considered
as limiting the scope of the invention.
Example 1
[0060] To illustrate the magnitude of the heat transfer coefficient
enhancement and pressure drop increase, the ratio of the heat
transfer coefficient and pressure drop of a winding micro-channel
with a topology similar to that depicted in FIG. 3.b to that of a
straight micro-channel of equal length and cross-section were
computed using ANSYS, a commercial computational fluid dynamics
(CFD) software. The results of this computation are shown in FIG.
7a. At a Reynolds number of 300, the winding micro-channel heat
transfer coefficient is about three times larger than that of a
linear micro-channel and the pressure drop is about 1.5 times
larger. Similar results would be anticipated with a winding
mini-channel design.
Example 2
[0061] A heat exchanger according to the first embodiment described
above was fabricated and tested. The heat exchanger had a
60.times.60 mm transfer area and the heat transfer member consisted
of a stack of three heat transfer layers fabricated out of 0.25 mm
thick copper foil. Winding micro-channels with a width of 0.25 mm
and a depth of 0.17 mm were chemically etched into one surface of
the heat transfer layers. Inlet and outlet opening with a diameter
of 0.75 mm were etched through the heat transfer layers. The
distance between the inlet and outlet openings was 4.8 mm. The
winding micro-channel topology was similar to that depicted in FIG.
3b. It had six small-scale undulations with a pitch of 2.4 mm and
amplitude of 0.75 mm and three large scale bends spanning the
distance between the inlet and outlet openings.
[0062] A 20.times.20 mm heat exchanger with winding mini channels
having a width of 1 mm and a depth of 0.65 mm was fabricated and
tested. The mini channels had amplitude of 0.86 mm and a wavelength
of 3 mm. The total length of the mini channels was 20 mm.
[0063] The thermal resistance of the winding micro-channel cold
plate and winding mini-channel cold plate was measured as a
function of the water flow rate per unit area. The measured
resistance is shown in FIG. 7b. Also shown in FIG. 7b is the vendor
provided performance for a swaged-tube cold plate commercially
available from Lytron, Inc, product number CP15, a normal-flow cold
plate, and the theoretically minimum resistance that can be
achieved at a given flow rate per unit area. The theoretically
minimum resistance corresponds to that of an ideal cold plate whose
surface temperature was equal to the fluid exit temperature. As
evidenced from these measurements, the winding channel cold plates
(both mini and micro) can achieve a thermal resistance an order of
magnitude lower than that of a prior art, exemplary swaged-tube
cold plates. For the given flow rates, the thermal resistance of
both the winding micro-channel and mini-channel cold plates
approaches the theoretical lower bound.
[0064] It has also been found that in utilizing winding
mini-channels heat transfer uniformity can be improved by varying
the amplitude of the undulations of the winding channels along the
length, "l" thereof, as illustrated in FIG. 7. By increasing the
amplitude along the length, the heat transfer coefficient is
likewise increased. As the fluid travels along the length of the
winding channel, the temperature change (.DELTA.T) is greatest at
the beginning of the winding channel and decreases along the length
toward the exit. By varying the amplitude to increase the heat
transfer coefficient, the overall heat transfer along the length of
the channels remains largely consistent from the beginning to the
end of the channel. The amplitude may vary between individual
undulations as illustrated in the present embodiment, or the
undulations may be grouped together with two more undulations
forming a group. Each group of undulations may include undulations
having the same amplitude, the amplitude of the groups of
undulations increasing along the length of the winding channel,
moving in a direction toward the outlet openings. The heat transfer
coefficient may also be increased by creating undulations having
shorter wavelengths "w" along the length, "l" of the channel. For
example, for any given length "l" of a channel, fewer undulations
having longer wavelengths each will result in a lower heat transfer
coefficient whereas for the same length channel, and same amplitude
undulations, smaller wavelengths and a corresponding increase in
the number of undulations will increase the heat transfer
coefficient. In one embodiment, the ration of the amplitude of the
undulation to the width of the channel may vary from about 0 to
about 1.5, although other ratios may be utilized.
[0065] FIGS. 7c-7f graphically illustrate the effects of the
undulations amplitude and wavelength on the mini-channel pressure
drop and heat transfer coefficient. Referring initially to FIG. 7c,
a graph comparing the pressure drop of a straight channel to that
of two winding mini-channels each having different amplitude
undulations is shown. FIG. 7d is a graph comparing the pressure
drop of a straight channel to that of two winding mini-channels
each having different wavelength undulations. FIG. 7e is a graph
comparing the heat transfer coefficient of a straight channel to
that of two winding mini-channels each having different amplitude
undulations while FIG. 7f is a graph comparing the heat transfer
coefficient of a straight channel to that of two winding
mini-channels each having different wavelength undulations. It
should be noted, that while these graphs illustrate mini-channels,
they would be equally applicable to any size channels.
[0066] The substantial increase in heat transfer coefficient
resulting from the winding channel geometry allows the use of
mini-channels in some applications formerly requiring the use of
micro-channels to achieve the desired heat transfer capacity. The
mini-channel design also has the added advantage of wider channels
that generally do not require filtration of the cooling liquid
prior to entering the winding mini-channels. Filtration is
generally required to reduce particulates that can block the
small-scale micro-channel passages, impeding performance and
sometimes leading to failure of the micro-channel device. As will
be appreciated, the elimination of filtration that is generally
required with micro-channels reduces the cost, labor and
possibility of break down for the mini-channel device as compared
to a similar device utilizing micro-channels.
[0067] In addition to the potential for improved heat exchanger
performance discussed above, the present disclosure provides an
inexpensive approach for fabricating winding channel heat
exchangers to meet a wide range of applications. For example, the
winding channels can be fabricated inexpensively by laser
machining, chemical milling, or the like. In the chemical milling
process, photosensitive resist layers are laminated to both sides
of a metal foil and a photomask is employed to pattern the winding
channel geometry onto the resist. After development, the resist is
removed from the areas that will be etched. The winding channels
may be made by etching the metal from only one side; thereby
obtaining a partially etched feature that does not extend through
the thickness of the layer 28. The inlet and outlet openings 32, 34
may be made by etching the metal from both sides, until all the
metal is dissolved and a through feature is obtained that connects
the first surface 10a to the second surface 10b.
[0068] The present construction also simplifies the fabrication of
heat exchangers having a range of heat transfer capabilities. The
heat transfer capability of the heat exchanger is proportional to
the flow rate, and to maintain the same thermal effectiveness, the
product of the winding channel wall area and the winding channel
heat transfer coefficient must be proportional to the flow rate. In
the heat exchanger of the present disclosure, this can be easily
accomplished by increasing the number of layers in the heat
transfer member in proportion to the required flow for the target
application. In addition, the present embodiment also allows for
inexpensive tailoring of the heat transfer capability over the
surface of the heat exchanger. In some applications it may be
desirable to provide a greater heat transfer capability (lower
thermal resistance) in one area of the heat exchanger and a smaller
heat transfer capability (higher thermal resistance) in another.
For example, if hot spots are disposed in one area greater heat
transfer capability in that area would be desired. This can be
easily accomplished by using the heat exchanger of the present
disclosure. For example, a flow restrictor plate 44 can be inserted
between the manifold 12 and the heat transfer member 14. As
illustrated in FIG. 8, the flow restrictor plate 44 includes
openings 46 that allow fluid flow through the micro-channels 30 in
the areas that require maximum heat transfer capacity, while the
body 48 of the plate 44 selectively blocks the flow from some of
the micro-channels 30 in areas that require a smaller heat transfer
capacity. The pattern, size and shape of the openings 46 can be
tailored according to the particular application while the winding
channel design of the heat transfer member 14 remains the same. In
this manner the same winding channel configuration could be used
for heat exchangers having different heat transfer patterns.
[0069] Alternatively, the same effect could be achieved by grouping
the winding channels so as to vary the heat transfer capability in
certain areas of the heat transfer member. For example, the number
of winding channels per unit area could be varied over the active
area of the heat transfer member, with some groupings being denser
than others so as to increase the heat transfer capability in the
areas with a greater density of winding channels. Likewise, the
winding channels could also be grouped according to the size of the
undulations and/or bends, with similar amplitude winding channels
being grouped together. In this manner, the amplitudes of the
undulations could be varied between groups of winding channels over
the heat transfer member so as to vary the thermal resistance over
an active area of the heat transfer member.
[0070] Yet another fabrication advantage of the heat exchanger of
the present disclosure is that the flow distribution and heat
transfer functions are confined to different components. The heat
transfer capacity depends primarily on the geometry and material
properties of the heat transfer member and high thermal
conductivity materials, such as copper or aluminum, need only be
used in the fabrication of the heat transfer member. The manifold
could be fabricated out of lower cost materials such as a
temperature resistant polymer. The manifold could also be a
stamping made out of a lower cost metal. Accordingly, the present
invention provides for a device that can be readily tailored to a
variety of needs in an inexpensive and readily achievable
manner.
[0071] It will be apparent to those skilled in the art, that there
are many variations in the winding channel geometry that can be
used to advantage to meet the requirements of different
applications. As shown in FIGS. 9a-9e and FIG. 10, various
embodiments of designs for alternate winding channels 30 are
illustrated, for example, FIGS. 9b and 9c illustrate non-arcuate
embodiments; FIG. 9a includes undulations 38 that change the
direction of the fluid flow, but does not include bends that
reverse the direction of the fluid flow; FIG. 9d includes a single,
arcuate undulation; FIG. 9e includes winding micro-channels 30 are
etched on both sides of layer 28 to increase the micro-channel
density; and FIG. 10 where each winding micro-channel 30 is
separated from every other micro-channel 30. All of these designs
are variations of the first embodiment and operate under the same
principals.
[0072] It will be understood that various modifications may be made
to the embodiments disclosed herein. For example, the dimensions
and geometric shapes may be modified, as would be known to those of
skill in the art. In addition, the winding channels may find use in
normal flow cold plates as well as parallel flow cold plates in
which case the directional examples would be modified. In addition,
the number and size of the small amplitude undulations and reverse
bends can be varied depending upon the application, and some
applications may only have bends that reverse the direction of the
fluid flow, while others may only have undulations that change the
direction of the fluid flow and some may have both. In addition,
the winding channel's axis may remain non-linear along only a
portion of the length of the winding channel. Likewise, the
examples provided are not to be construed as limiting, but as
projected outcomes of exemplary embodiments. Therefore, the above
description should not be construed as limiting, but merely as
exemplifications of preferred embodiments. Those skilled in the art
will envision other modifications within the scope, spirit and
intent of the invention.
* * * * *