U.S. patent application number 11/256007 was filed with the patent office on 2006-07-27 for cooling electronics via two-phase tangential jet impingement in a semi-toroidal channel.
Invention is credited to Michael Anthony Brown, Triem T. Hoang.
Application Number | 20060162365 11/256007 |
Document ID | / |
Family ID | 36695233 |
Filed Date | 2006-07-27 |
United States Patent
Application |
20060162365 |
Kind Code |
A1 |
Hoang; Triem T. ; et
al. |
July 27, 2006 |
Cooling electronics via two-phase tangential jet impingement in a
semi-toroidal channel
Abstract
A two-fluid-phase cooling device for absorbing high thermal flux
from electronics devices and other thermally dissipating devices.
It consists of a thermally conductive plate with thermally
dissipating elements on one face and a semi-toroidal cavity in the
opposite face with the cavity's axis perpendicular to the face of
the plate, a liquid refrigerant supply tube ending in a
thermodynamic cycle's refrigeration expansion valve that directs
jets of liquid to impact the conical surface in the center region
of the semi-toroidal cavity in a direction along the cavity's axis
and tangent to the conical surface, a second plate with a
semi-toroidal protrusion extending into the semi-toroidal cavity to
form a thin, semi-toroidal channel between the two plates, and a
seal between the liquid supply tube and the second semi-toroidal
plate. In operation liquid refrigerant jets strike the conical
surface generally tangential to the surface and flow at high
velocity in a thin film on the surface of the semi-toroidal cavity
from its center radially to the outer edge of the toroidal channel,
absorbing heat and boiling as it does so. The high radial
acceleration forces caused by the liquid film moving at high
velocity on the cavity's concave surface force the liquid film
against the surface and create a pressure gradient that biases
evaporation toward the liquid/vapor interface. The vapor moves
parallel to the liquid flow radially outwards between the liquid
film and the surface of the semi-toroidal protrusion at very high
velocity, causing extreme turbulence in the liquid film and highly
augmented heat transfer between the heated plate and the liquid
film, while the liquid film nevertheless remains intact and forced
against the heated surface by radial acceleration and carried to a
distance significantly greater than in conventional jet impingement
systems. The device may also be composed of wedge-shaped sections
of the semi-toroidal plates. It may further have two expansion
valves in series in the liquid supply line, the first generating a
small amount of vapor (increase in quality) so the resulting
increase in flow volume greatly increases the velocity through the
second expansion valve toward the heated surface to further enhance
heat transfer.
Inventors: |
Hoang; Triem T.; (Clifton,
VA) ; Brown; Michael Anthony; (Columbia, MD) |
Correspondence
Address: |
Michael A. Brown;TTH Research Inc.
Suite 215
14300 Cherry Lane Court
Laurel
MD
20707
US
|
Family ID: |
36695233 |
Appl. No.: |
11/256007 |
Filed: |
February 21, 2006 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
|
60621894 |
Oct 26, 2004 |
|
|
|
Current U.S.
Class: |
62/259.2 ;
257/E23.1; 62/310 |
Current CPC
Class: |
H01L 23/4735 20130101;
F25B 39/02 20130101; F25B 2339/021 20130101; H01L 2924/00 20130101;
H01L 2924/0002 20130101; H01L 2924/0002 20130101 |
Class at
Publication: |
062/259.2 ;
062/310 |
International
Class: |
F25D 23/12 20060101
F25D023/12; F28D 5/00 20060101 F28D005/00 |
Claims
1. A cooling device comprising: a thermally conductive solid such
as a plate with thermally dissipating elements attached to one face
and a semi-toroidal cavity in the opposite face, the cavity formed
by cutting in the material face, and about an axis perpendicular to
the face, a groove of generally semi-circular shape of radius r1,
with the center of radius r1 at a radius C from the axis of the
circle to the center of the groove, so that setting the radius r1
equal to radius C causes the semi-toroidal cavity to form an apex
point that lies both on the center line of the cavity and in the
plane of the plate in which the cavity is cut, so the surface is,
to a distance from its centerline to the radius C, convex in a
direction circumferential to the axis of the cavity and concave in
a direction radial from the axis of the cavity, and past the radius
C is concave both in circumferential direction and radial
direction; and a second plate having a semi-toroidal protrusion in
the shape of a circular ridge of semi-circular cross-section of
radius r2<r1, with the center of radius r2 located at radius C
from the axis of the circular protrusion, and the protrusion
located concentric with and extending into the semi-toroidal cavity
so a semi-toroidal channel exists between the cavity wall surface
and the protrusion wall surface; and a nozzle on the centerline of
the circular protrusion and parallel to its axis, passing through
and sealed to the plate containing the protrusion, and containing
several orifices that direct streams of a volatile fluid in a
direction generally parallel to the axis of the semi-toroidal
cavity to strike the upper surface of the cavity immediately below
the centerline apex of its surface and in a direction generally
tangent to the cavity surface, and flows outward in the
semi-toroidal channel and exits the channel in a direction
generally opposite that of the fluid streams directed from the
nozzle.
2. The cooling device of claim 1 wherein the nozzle acts as the
expansion valve in a thermodynamic cooling cycle, so that a
volatile compressed liquid at subcooled temperature forced through
the nozzle drops to a saturated temperature and pressure causing a
fraction of the liquid to flash to vapor so a mix of cooled liquid
and vapor enters the semi-toroidal channel.
3. The cooling device of claim 1, wherein the fluid striking the
upper surface of the cavity directly below its centerline apex in a
direction generally tangent to the surface does so virtually
without loss of momentum and velocity.
4. The cooling device of claim 1 where the velocity of the liquid
on the cavity surface immediately below the impact point has a
relatively small component in a direction perpendicular to the axis
of the cavity, so the liquid film has negligible axial velocity and
so thins rapidly in this region giving the film time to coalesce
into an even thickness in the circumferential dimension about the
axis of the cavity by flowing in the convex direction of the
surface under the Coanda effect.
5. The cooling device of claim 1 wherein the volatile liquid film
flowing in the semi-toroidal channel experiences very high
centripetal acceleration that forces the film against the concave
cavity wall of the channel, so the liquid and vapor phases of the
fluid are separated with liquid against the cavity wall and vapor
between the liquid film and the semi-toroidal protrusion wall
surface above it, creating a high pressure gradient in the liquid
film with the minimum value equal to the saturation pressure at the
liquid/vapor interface, so that when the liquid is heated by the
thermally dissipating elements the boiling is biased toward the
vapor/liquid interface that is at saturation pressure, and any
bubbles forming on the cavity surface are immediately pushed to the
liquid/vapor interface thereby increasing the efficiency of the
heat transfer process.
6. The cooling device of claim 1 wherein the rate of thinning of
the liquid film as it expands outwards in the channel is increased
by the evaporation of liquid from the film, so the heat transfer
coefficient between cavity wall and liquid film increases more
rapidly and generally in inverse proportion to the thinness of the
liquid film.
7. The cooling device of claim 1 wherein the vapor formed by the
expansion valve process combines with the vapor formed by the
energy input from the dissipating elements to expand radially
outwards in the semi-toroidal channel at a great velocity relative
to the velocity of the liquid film moving in the same direction, so
that the vapor creates extreme turbulence in the liquid film to
increase the heat transfer from the cavity wall to the liquid film,
and the vapor pushes the liquid film to overcome the flow friction
between wall and film so the film increases its velocity, over that
rate of increase naturally afforded by the effect of centripetal
acceleration on the thinning film, even when the film becomes
extremely thin, and maintains this velocity to provide cooling over
an area much larger than state of art jet impingement, while the
centripetal acceleration forces on the liquid film prevent the film
from being broken up by the vapor into mist flow that would
decrease the heat transfer rate as occurs in annular pipe flow.
9. The cooling device of claim 1 wherein the expansion valve is
formed by extending the refrigerant liquid supply tube over the
apex of the semi-toroidal cavity until a narrow annulus is created
between tube and cavity surface, at which point the tube expands
orthogonally to create irreversible flow conditions in the fluid
and then blends into the semi-toroidal protruding surface
10. The cooling device of claim 1 wherein the semi-toroidal
protruding surface is removed so the vapor is not channeled above
the semi-toroidal cavity, but the vapor emitted from the nozzle
flows above and in a general direction tangential to the liquid
film flowing on the surface of the semi-toroidal cavity so that
some increased turbulence is created in the liquid film on the
surface of the semi-toroidal cavity and heat transfer is
enhanced.
11. The cooling device of claim 1 wherein the liquid refrigerant
supply tube has an orifice upstream of the nozzle, acting to create
irreversible flow to cause a small amount of liquid refrigerant to
flash to vapor so that the volume flow rate through nozzle directed
at the semi-toroidal cavity is greatly increases to increase the
heat transfer of the system.
12. A cooling device wherein a flow channel with a rectangular
cross-section is located below a surface on which is mounted a
generally elongated thermally dissipating device, with the entrance
and exits of the channel in directions generally orthogonal to the
plane on which the dissipating devices are located and the
connecting channel curved in a convex direction toward the
dissipating elements, the wall of the channel in opposition to the
dissipating devices being initially convex in the cross-sectional
view at the entrance, changing to flat and then concave with
increasing distance along the channel; and a nozzle sealed around
the entrance to the channel and directing a jet of volatile liquid
into entrance of the channel.
13. The cooling device of claim 12 wherein the a jet of volatile
liquid strikes the convex surface at the entrance to the channel in
a direction generally tangential to the surface, spreads out in the
convex direction of the surface to form a film of even thickness
and maintains an even thickness by means of the gradual change of
the cross-sectional shape of the surface from convex to concave,
the liquid film being held against the wall by the centripetal
force caused by the concave shape of the shape in the flow
direction.
14. A cooling device wherein a number of semi-toroidal channels and
nozzles are located in a face of a conductive plate to increase the
effective heat-gathering area of the apparatus.
15. The cooling device of claim 14 wherein cooling channels are
composed of a number of alternately placed wedge-shaped sections of
the semi-toroidal surface so that heat may be evenly absorbed from
a large rectangular surface, and the dividing walls between the
alternately placed channels act as structural support ribs in the
plate to allow the thickness of the plate between the thermally
dissipating elements and the cavity surfaces to be minimized to
enhance heat transfer.
Description
CROSS-REFERENCE TO RELATED APPLICATIONS
[0001] This application claims the benefit of provisional patent
application Ser. No. 60/621,894, filed 2004 Oct. 22 by the present
inventors.
FEDERALLY SPONSORED RESEARCH
[0002] Not Applicable
SEQUENCE LISTING OF PROGRAM
[0003] Not Applicable
BACKGROUND OF THE INVENTION
[0004] 1. Field of Invention
[0005] This invention relates to cooling electronics, specifically
to spray-cooling of two-phase fluid on a heated surface contained
within a conventional refrigeration loop.
[0006] 2. Prior Art
[0007] The problem addressed in this invention is removal of high
thermal dissipation flux from electronic devices such as amplifier
gate arrays, laser diodes, etc.
[0008] Heat flux from electronics is now in the range of 100 to
1,000 Watts per square centimeter (W/cm.sup.2). Thermal literature
refers to this as the high-flux range, and ultra-high flux being
from 10.sup.3 to 10.sup.5 W/cm.sup.2 and describes a number of ways
to remove the heat. If the heated surface is in the interior of an
electronics package it can be removed only by circulation of a
fluid against the heated surface.
[0009] Fluids commonly available for this are air, water and
fluorochemicals (generally called "refrigerants", although they may
be used in high temperature applications), and the means of
circulation can be natural convection, single-phase forced
(mechanically pumped) convection, and boiling (2-phase pumped
flow). The heat transfer coefficient Watts per centimeter-squared
and degree centigrade (W/cm.sup.2-C) defines the rate of heat
removal from a surface for a given temperature difference between
the surface and the cooling liquid, and is highly dependant on the
type of fluid and the means of circulation. Air is a poor choice
for any type of circulation because of its low mass and low thermal
conductivity. Water will have a coefficient about an order of
magnitude greater than a refrigerant. Natural convection with water
reaches only about 0.1 W/cm.sup.2-C, so this process cannot be
considered for use with a refrigerant for high flux needs. In
single-phase forced convection flow refrigerants reach about 1
W/cm.sup.2-C and water 10 W/cm.sup.2-C, and in boiling heat
transfer refrigerants reach about 10 W/cm.sup.2-C and water over
100 W/cm.sup.2-C. In single-phase flow, water would require a
temperature difference of 100C to carry away 1 kW/cm.sup.2,
limiting the practical approach in most cases to boiling heat
transfer. A further, key advantage of phase change flow is that
only a modest increase in heated surface temperature results in a
large increase in heat flux, and in certain situations such as
freezing environments only a refrigerant can be used in the
two-phase system.
[0010] There are several phase change cooling schemes available:
micro- and mini-channel cooling, jet impingement cooling and spray
cooling. In all of these the upper limit of heat transfer is set by
critical heat flux (CHF) which is the point at which liquid cannot
reach the heated surface fast enough to prevent dryout of the
surface. Micro-channel and mini-channel refer to flow devices
having hydraulic diameters of 10 to several hundred micro-meters,
and one to a few millimeters, respectively. Typically the channels
are rectangular grooves cut in a metal plate on which the thermally
dissipating element is mounted. High heat transfer coefficients,
inversely proportional to the Reynolds Number, are achieved by the
thinness of the liquid channel in laminar flow. Drawbacks include
the limitations of the minimum size of the hydraulic diameter
necessary to avoid flow clogging, and high streamwise pressure
drops that can cause flow choking as the fluid suddenly evaporates.
This latter problem limits the size of the cooling device. In
addition, there will be thermal resistance to the flow of heat
through the fins to the heated baseplate. Typical values for heat
transfer coefficient with refrigerant fluids are 3 to 5
W/cm.sup.2-C. Conventional Jet impingement cooling (FIG. 2a) is
done by directing a stream of liquid orthogonally against a heated
flat plate. Heat transfer from plate to liquid is enhanced by the
thinness of the boundary layer at the jet's small area of
impingement, and then by the high velocity of the liquid moving
tangential for two or three jet radii along the heated surface.
Problems here are first the smallness of the effective cooling area
and the necessity for very high jet velocities. In particular for
two phase flows, the vapor bubbles formed on the plate's surface
tend to push the liquid film away from the surface. There is also a
loss in liquid momentum by the orthogonal impact on the plate, and
areas of sub-saturated pressure directly under and near the
impinging jet that may cause surface bubbles at this region. Heat
transfer coefficients are in the range of 2 to 3 W/cm.sup.2-C.
Spray cooling produces a peak heat transfer rate about half that of
jet impingement, but cools a larger area. A problem with spray
cooling is maintenance of the nozzles.
[0011] There are three other relevant two-phase phenomenon that
must be listed. The first is flow in a curved channel where the
concave surface is heated. Here the g-forces generated by the flow
velocity on the curved heated surface tend to force bubbles to move
away from the heated surface and so prevent the bubbles from
blocking access of liquid to the surface. Another flow regime of
interest is annular flow in a pipe (FIG. 3). Heat transfer texts
show this can produce the highest heat transfer rates in pipe flow
boiling. In annular flow, there is a thin liquid film moving along
the pipe wall, with the vapor moving down the center of the pipe at
very high velocity. The high velocity of vapor relative to that of
the liquid creates turbulence in the liquid film much higher than
that created by flow of the liquid against the pipe wall. This can
increase the heat transfer coefficient by more than an order of
magnitude. However, the high turbulence quickly causes the liquid
film to break up into what is called mist flow, so CHF is exceeded
and the heat transfer coefficient falls back sharply. A third flow
phenomenon is called the Coanda effect. This is the tendency for a
flowing liquid to remain attached to a convex surface, with the
result that unevenness in film thickness is eliminated as the
pressure head in the thicker film areas pushed the fluid toward
thinner areas. This is seen in water flowing over an apple held
under a faucet.
[0012] The following prior-art patents describe specific attempts
to solve he problem of high thermal flux removal.
[0013] Chu (U.S. Pat. No. 6,519,151) discloses a jet impingement
thermal control device consisting of a nozzle that directs a fluid
to strike perpendicular to, and at the bottom center of, a
(bowl-shaped) concave conic-sectioned heated surface, so the liquid
flows radially outwards along the surface of the bowl and exits the
apparatus in a direction generally opposite to the incoming jet
(FIG. 2b). Several such assemblies may be located in parallel to
cool a large surface. The liquid film thins as it expands from the
point of impact, and, combined with a high-g centrifugal force,
this causes the fluid velocity to increase while the velocity in
conventional flat plate jet impingement rapidly decreases by flow
friction as it moves from the impact point. The combination of high
velocity and thin, stable liquid film in Chu's invention causes an
increase in efficiency over conventional jet impingement cooling.
However, the perpendicular impact will cause momentum and velocity
loss in the liquid stream as it turns a right angle to flow along
the curved surface. Further, the radial velocity of the liquid is
highest where it moves away in a direction perpendicular from the
jet, so if there is any initial circumferential difference in film
thickness there will not be sufficient time for the film to come to
even thickness. The extent of the radial flow is limited because
eventually the flow friction overcomes the momentum in the liquid
film when the film becomes very thin. Hocker (Application
2002/0062945 A1) shows the same concept as Chu cited above.
[0014] Rini et al. (U.S. Pat. No. 6,571,569) shows a design of an
evaporative cooling system wherein the refrigeration expansion
valve (nozzle) directs fluid directly against the flat plate having
the heat dissipating elements on its opposite side. This approach
suffers from the same problems described above in spray cooling.
This patent further describes a means for a mechanical pump to
force a high velocity vapor steam into the stream of liquid
refrigerant to increase its velocity and cooling effectiveness.
This approach adds to the weight and complexity of the cooling
system.
[0015] Remsburg (U.S. Pat. Nos. 5,864,466 and 6,064,572) shows a
conic-sectioned plate in a heat exchange apparatus. However, the
function of the curved piece is to create a themosyphon action to
direct liquid flow against a heated flat plate. The flow is then
convectional to that heat transfer coefficients will be very low.
Searight (U.S. Pat. No. 4,108,242) shows a means to inject fluid
jets into a cylindrical cavity to induce swirling flow in general
flow along the axis of the cavity. Here the heated surface has a
single axis of curvature so the flow is not accelerated by motion
along the curved surface nor is a thin flow film created. Lynch
(U.S. Pat. No. 4,140,302) shows a water-cooled blast-furnace
tuyeres nozzle having a number of liquid jets at high speed
directed against the contoured inner surface of the nozzle. The jet
impinges the surface at low angle to avoid momentum loss, but the
curved surface shown is only to direct flow against a heated
surface that is flat. Further, in this design the water passages
are filled with liquid, so this arrangement does not produce a thin
film liquid flow nor does the single-axis curved surface provide an
acceleration of flow. Bemisderfer (U.S. Pat. No. 5,056,586) shows a
spray system whereby the liquid is directed against cusp-shaped
surfaces to increase turbulence. This does not produce a thin film
nor accelerate flow. Tilton (20030172669) shows transverse
thin-film evaporative spray cooling. The spray nozzle directs
droplets down a narrow channel on whose side(s) are electronic
devices to be cooled. This does not create a continuous liquid
film, nor does it provide uniform cooling of the devices.
[0016] Niggeman (U.S. Pat. No. 4,643,250) shows a heat exchanger
whereby a conical surface is used as a means to separate cryogenic
liquid from vapor phase, and then to condense the vapor phase in a
heat exchanger wherein the liquid phase is the heat sink. This is
not possible since the two phases will be at the same temperature
at the entrance to the apparatus.
OBJECTS AND ADVANTAGES
[0017] Several objects and advantages of the present invention are:
[0018] to provide a means to direct a coolant fluid jet against a
heated surface without loss of velocity or momentum after contact
with the surface; [0019] to provide a means to enable the coolant
fluid jet to form into a thin, high-velocity liquid film of
consistent thickness on the heated surface to create a high value
of convection heat transfer coefficient and significant increase in
critical flux; [0020] to provide a means to maintain velocity of a
cooling liquid film over a heated surface to a distance
significantly longer than conventional jet impingement devices;
[0021] to increase the effectiveness of the coolant fluid jet
beyond what is available from the thermodynamic refrigeration
system's expansion valve jet but without addition of a mechanical
system; [0022] to create an equivalent of circular pipe annular
flow over the heated surface to increase the heat flux removal
rate.
BRIEF SUMMARY OF THE INVENTION
[0023] In accordance with the present invention a coolant fluid jet
directed against a doubly-curved, semi toroidal surface located in
a conductive plate on whose the opposite face are thermally
dissipating electronic devices.
BRIEF DESCRIPTION OF SEVERAL VIEWS OF THE DRAWINGS
[0024] FIG. 1 shows a schematic of a conventional refrigeration
loop cooling system that employs an evaporative cooling plate in
accordance with the principles of the present invention.
[0025] FIG. 2a shows prior art conventional jet impingement
assemblies where a fluid jet impacts a flat plate, and 2b prior art
where the jet impacts a concave conic section, in both cases in a
direction perpendicular to the surface at the impact point.
[0026] FIG. 3 describes annular two-phase evaporative cooling flow
in a conventional circular pipe whose geometry shall be modified
for application in the present invention.
[0027] FIG. 4a defines the geometry of a torus. FIG. 4b shows an
elevation view of a torus with specific relations among dimensions,
and 4c, 4d sections thereof to show how the doubly-curved surface
is created.
[0028] FIG. 5 shows a comprehensive view of the present
invention.
[0029] FIG. 6a shows a cross-section view of the preferred
embodiment of the two-phase cooling device invention, with an
expansion valve having multiple orifices shown in detail in FIG.
6b.
[0030] FIG. 7 shows a partial cross-section detail view of another
design of an expansion valve formed by a continuous circumferential
gap.
[0031] FIG. 8 shows a cross-section view of another embodiment of a
two phase cooling device having no convex surface to control vapor
flow.
[0032] FIG. 9 shows application of the invention to a rectangular
slot.
[0033] FIG. 10 shows how circumferential segments of the toroidal
surface can be arranged to absorb heat from a large, rectangular
surface.
[0034] FIG. 11 shows a refrigeration expansion valve with a flow
orifice upstream of it to increase the velocity of the fluid
jet.
DETAILED DESCRIPTION OF THE INVENTION
[0035] The present invention is designed to use a two-phase cooling
fluid to remove high heat flux from electronics systems over a
surface area that is relatively large compared with state-of-art
cooling systems. Electronic system designers are now seeking
cooling system for thermal fluxes greater than 1 kW/cm.sup.2 over
areas of tens of square centimeters. Thermal research shows the
highest heat removal rate is achieved by a two-phase fluid system
wherein heat dissipating devices mounted on a conductive plate
evaporate a liquid directed against the opposite side of the plate.
The highest flux rates are achieved with water. However, in some
cases, e.g., when the system must be dormant in freezing
temperatures, it is necessary to use a volatile fluid referred to
as a refrigerant (although the operating temperature of the system
may be above that normally thought of as refrigeration).
[0036] FIG. 1 is a schematic of a conventional refrigeration cycle
as shown in textbooks on thermodynamics. Heat from a thermally
dissipating element 10 transfers to an evaporator 11 where it
evaporates a liquid entering the evaporator through pipe 12. The
resulting vapor is transported by differential pressure to a
compressor 13 that raises both the temperature and pressure of the
vapor. A condenser 14 then removes heat from the vapor, forming a
high pressure, subcooled liquid. When the liquid flows through an
orifice 15 called an expansion valve, its pressure drops so that a
fraction of the liquid evaporates producing a cold saturated
liquid-vapor mix that flows to the evaporator 11 to complete the
cycle.
[0037] FIG. 2a prior art shows conventional jet impingement within
an evaporator. The liquid jet 20 enters the evaporator 11 through
pipe 12 and vapor leaves the evaporator through pipe 24. The jet
impacts the evaporator wall at very high velocity directly below
the dissipating element 10, and in a direction perpendicular to the
surface of the wall. This creates a high pressure stagnation region
21 at the point of impact. The flow, which has now lost some
momentum, turns in radial direction 22 along the flat wall,
initially creating a high turbulent heat transfer coefficient that
is generally proportional to the flow velocity and inversely
proportional to liquid film thickness. In a few jet radii, however,
the momentum of the thinned film is overcome by flow friction that
slows the flow and thickens the liquid film 23, sharply decreasing
the heat transfer coefficient. The turbulence directly adjacent to
the stagnation area 21 can also cause pockets of sub-stagnation
pressure leading to bubble formation on the surface that can block
heat transfer. Bubbles will also form on the evaporator wall where
the liquid film 23 slows and thickens, producing the same effect.
This blockage of liquid from the heated surface creates what is
called the critical heat flux that is the limit of the heat flux
that the evaporator can absorb.
[0038] FIG. 2b prior art shows jet 20 impacting the bottom center
25 of a semi-spherical concave surface 26 adjacent to a dissipating
element 10. At the stagnation region the liquid behaves as
described above, but the flow radially along the concave surface is
very different. If the flow velocity is high enough, the force of
centrifugal acceleration will cause the thinning film to overcome
flow friction and accelerate, create a desirable thin, high
velocity film that is shown to increase heat transfer coefficient
by about 65% in single-phase flow. Eventually, however, the liquid
film will become very thin and so its reduced momentum will be
overcome by flow friction and the heat transfer coefficient will
decrease. Since the highest radial velocity (orthogonal to the
center axis of the apparatus) is at the point where the liquid film
leaves the jet impact point, any substantial circumferential
differences in film thickness may not have time to even out.
[0039] FIG. 3 is a heat transfer textbook diagram showing a flow
regime called annular flow that exists in boiling heat transfer
flow in circular pipes as in a powerplant. The figure shows fluid
flow in a pipe 30 heated along its outer surface. Here the liquid
flows as a continuous film 31 along the wall, and the vapor 32
moves at a very high velocity down the center of the liquid
annulus. The high velocity of the vapor relative to that of the
liquid creates turbulence in the liquid film much greater than that
created by the interface between pipe wall and liquid, and the heat
transfer coefficient can increase one or more orders of magnitude
greater than that in turbulent liquid flow. But the vapor friction
quickly creates ripples and waves 33 in the liquid film which is
then forced off the wall as droplets 34 to create regimes called
annular mist and then mist flow, both having greatly reduced heat
transfer coefficients. The roughened liquid/vapor interface surface
also creates a very high pressure drop in the fluid flow direction.
The objective of the present invention is to create the annular
flow effect described in FIG. 3 above, but without the problems
experienced in pipe flow, on the evaporative cooling plate in a
two-phase cooling system. This is achieved in the present invention
by creating flow on a semi-toroidal cavity surface within an
evaporator, and the creation of this surface is now explained.
[0040] FIG. 4a shows a definition of a hollow torus formed by
rotating a circle of radius r about a central axis 41 at radius C,
with a sectional view cut by plane 42 for the general case of
C>r. The sectional view with C=r is a special case, and 4b is a
elevation view of a torus with C=r. FIG. 4c shows a view of a
semi-toroidal surface of 4b cut by plane 43. FIG. 4d shows a
cross-section of the semi-toroidal surface in 4c cut by plane 44.
This shape can be likened to a hemispherical bowl with a central,
curved cone. With radii C and r equal, the apex 45 of the cone lies
on the plane 44. If C>r, the conical surface is truncated at
plane 44. The surface from the apex of the cone to the radius C is
a doubly curved saddle shape, convex in the direction
circumferential to the axis 41 and concave in the direction of
curves extending from the apex to the outer diameter of the
toroidal surface and orthogonal to the circumferential curves.
Beyond radius C the curves circumferential about the axis 41 form a
concave surface.
[0041] FIG. 5 shows an isometric view 50 of the preferred
embodiment of the invention and a cross-section with arrows
indicating liquid entrance and vapor exit passages from the
semi-toroidal channel. Here the concave semi-toroidal cavity
surface 51 is part of a thermally conductive, cylindrical solid 52
on whose circular external end face are mounted the dissipating
devices 10 to be cooled. The semi-toroidal protrusion surface 53 on
cylindrical body 55 is coaxial with surface 51 and the two form a
channel 54 whose width may vary with distance radially from center
axis of body 52 in the flow direction. A refrigerant supply tube 12
is shown concentric with and sealed to cylindrical body 55.
[0042] FIG. 6a shows details of the preferred embodiment of the
invention applicable to a closed loop refrigeration system. Here
the cylindrical solid 52 containing semi-toroidal cavity surface 51
blends into the cylindrical pipe 61 that forms the pressure
containment vessel for this part of the two-phase cooling system.
Axially symmetric solid shape 55 containing semi-toroidal
protrusion surface 53 is shown with a conical shape to allow smooth
expansion of the vapor exiting semi-toroidal channel 54.
Refrigerant supply tube 12 terminates in a number of orifices 62
spaced around the center axis of tube 12.
[0043] FIG. 6b is a detail of the nozzle exit area showing that, in
operation, liquid refrigerant 63 is forced through the orifices 62
and exits the expansion valve at high velocity, so a certain
fraction of liquid evaporates as the pressure drops and the liquid
and vapor phases of the fluid are cooled to the same temperature.
The liquid jets are aimed so as to strike the semi-toroidal cavity
surface 51 directly below apex 45 and tangential to the surface so
that momentum and jet velocity are not lost. The individual liquid
films then coalesce under the Coanda effect into a continuous film
64 achieving constant circumferential thickness as they flow on the
convex curved surface about the center axis of the assembly. This
process is aided by the very small velocity of the film in a radial
direction perpendicular to the assembly's center axis in the region
directly below the apex of the cone. The film continues to flow and
thin as it moves radially outwards from the central axis. The high
g-forces acting on the thinning film accelerate the flow. The high
centrifugal force on the liquid film also creates a pressure
gradient in the liquid film, with highest pressure against the
torodial surface 51 decreasing to saturation pressure at the
liquid/vapor interface 65. This biases vapor formation toward the
liquid/vapor interface and forces any vapor bubble forming on
surface 51 to go to interface 65. The vapor formed as the liquid
jets exit the nozzles, and as the liquid evaporates as it flows
around the curved toroidal channel 54, are constrained by the
semi-toroidal protrusion surface 53 to flow in a direction parallel
to the liquid film flow. The high velocity of the vapor 66 with
respect to the liquid film 65 creates very high turbulence in the
liquid and so creates a very high heat transfer coefficient.
However, the centrifugal force on the liquid film prevents it from
being broken up into inefficient mist flow as in annular pipe flow
and further prevents an increase in pressure drop in the flow
direction caused by a roughened liquid/vapor interface. The vapor
velocity continues to accelerate the liquid film against flow
friction forces so that the flow remains turbulent even when the
liquid film becomes very thin and CHF is never reached, so 100%
evaporative efficiency is achieved if desired. This parallel flow
of vapor also allows the liquid flow to continue over very large
diameters of the heat absorbing surface, without risk of exceeding
critical heat flux. The vapor exits the evaporator assembly through
the expansion volume formed by body 55 and returns to the
compressor. The shape of the exit from the gap between the toroidal
surfaces may be designed to use the gap's vapor exit velocity to
entrain any liquid and force it to the compressor for applications
such as in zero-g space environment.
[0044] FIG. 7 shows a detail of the expansion valve area. Here the
apex 45 of the semi-toroidal cavity surface 51 extends
concentrically up into the liquid supply tube 12 so the expansion
valve has the form of a very thin, circumferential gap 71. The
semi-toroidal protrusion surface 53 is shaped at the interface with
the end of the supply tube 12 to form a step expansion 72 of the
semi-toroidal channel 54. Since the liquid flow is already in a
high-g mode at this point, the vapor of expansion forms on the top
of the liquid film as shown in FIG. 6a and the heated surface
remains covered with liquid from the apex point onwards. In this
design and the previous design with a number of orifices, the
refrigerant supply tube 12 and the apex 45 of the semi-toroidal
cavity surface 51 my be mechanically in contact or joined to
provide exact alignment between liquid jets and apex.
[0045] FIG. 8 shows a further embodiment of the design, applicable
to both single phase and two-phase flow. Here the axially symmetric
shape 55 and semi-toroidal protrusion surface 53 have been
eliminated. The liquid jets impact the semi-toroidal cavity surface
51 as before. In two-phase flow, vapor formed at the liquid/vapor
interface does not contribute directly to the liquid film velocity,
but the vapor 81 formed exiting the nozzle is directed generally
parallel with the liquid jets and follows the liquid film moving on
the cavity surface so there is some turbulence created in the
liquid film. Further, the vapor 82 flowing off the liquid film
converges with the vapor flowing from the jets orifices to create a
swirling flow 81 to increase turbulence as a function of
evaporation rate from the toroidal surface. This design will also
benefit single-phase liquid flow by creating a thin, high velocity
and accelerating film without any loss in velocity at the jet
impact point on surface 51.
[0046] FIG. 9 shows how this concept can be used to create a flow
surface is a rectangular slot, for example to cool a very small
rectangular area as under an amplifier gate array 91. The nozzle 92
directs a jet of liquid refrigerant to impact tangential to surface
93 that is convex in one dimension at one end of the channel. The
convex surface causes the liquid jet to spread out quickly into a
thin film. The bottom of the channel gradually changes to flat as
shown as 93 in section A-A and then slightly concave 94 so the
high-g forces on the liquid do not force it to the edges. The high
heat transfer coefficient directly under the gate array eliminates
the need for heat spreaders seen in conventional spray or jet
cooling with heat transfer coefficients of only a few
W/cm-.sup.2C.
[0047] FIG. 10 shows an arrangement for cooling very large
rectangular areas by an arrangement of alternating cooling surfaces
101 that are circumferential segments of a semi-toroidal surface.
The dividing walls 102 between the segments act as structural
supports to contain the high pressure of the refrigeration system
and to minimize the thickness of the conductive plate 103 between
flow passages 54 and dissipating source 10. Vapor and liquid are
transferred to and from the segments by plenums 104 and 105.
Surface 51 is shown elliptical instead of circular as in previous
figures to reduce the thickness between thermally dissipating
elements and the evaporative surface 51.
[0048] FIG. 11 shows a sequential arrangement of orifices in the
liquid refrigerant supply line to create vapor in the line to
increase fluid jet velocity and/or turbulence. Refrigerant liquid
in supply tube 12 passes through an orifice 111 causing a small
fraction of the liquid to flash to vapor and create a liquid/vapor
mix 112. When this mix passes through the nozzle orifices 62 to
strike toroidal surface apex 45, the resulting fluid exit velocity
is increased over that possible with only liquid passing through
orifice 62. This increases turbulence in the liquid film on surface
51 in either thin channel flow configuration FIG. 5, or in open
cavity flow FIG. 8.
* * * * *