CROSS-REFERENCE TO RELATED APPLICATIONS
FEDERALLY SPONSORED RESEARCH
This application claims the benefit of provisional patent application Ser. No. 60/621,894, filed 2004 Oct. 22 by the present inventors.
- SEQUENCE LISTING OF PROGRAM
- BACKGROUND OF THE INVENTION
1. Field of Invention
This invention relates to cooling electronics, specifically to spray-cooling of two-phase fluid on a heated surface contained within a conventional refrigeration loop.
2. Prior Art
The problem addressed in this invention is removal of high thermal dissipation flux from electronic devices such as amplifier gate arrays, laser diodes, etc.
Heat flux from electronics is now in the range of 100 to 1,000 Watts per square centimeter (W/cm2). Thermal literature refers to this as the high-flux range, and ultra-high flux being from 103 to 105 W/cm2 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.
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/cm2-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/cm2-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/cm2-C and water 10 W/cm2-C, and in boiling heat transfer refrigerants reach about 10 W/cm2-C and water over 100 W/cm2-C. In single-phase flow, water would require a temperature difference of 100C to carry away 1 kW/cm2, 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.
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/cm2-C. Conventional Jet impingement cooling (FIG. 2 a) 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/cm2-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.
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.
The following prior-art patents describe specific attempts to solve he problem of high thermal flux removal.
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. 2 b). 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.
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.
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.
- OBJECTS AND ADVANTAGES
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.
- BRIEF SUMMARY OF THE INVENTION
Several objects and advantages of the present invention are:
- 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;
- 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;
- 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;
- 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;
- to create an equivalent of circular pipe annular flow over the heated surface to increase the heat flux removal rate.
BRIEF DESCRIPTION OF SEVERAL VIEWS OF THE DRAWINGS
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.
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.
FIG. 2 a shows prior art conventional jet impingement assemblies where a fluid jet impacts a flat plate, and 2 b prior art where the jet impacts a concave conic section, in both cases in a direction perpendicular to the surface at the impact point.
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.
FIG. 4 a defines the geometry of a torus. FIG. 4 b shows an elevation view of a torus with specific relations among dimensions, and 4 c, 4 d sections thereof to show how the doubly-curved surface is created.
FIG. 5 shows a comprehensive view of the present invention.
FIG. 6 a 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. 6 b.
FIG. 7 shows a partial cross-section detail view of another design of an expansion valve formed by a continuous circumferential gap.
FIG. 8 shows a cross-section view of another embodiment of a two phase cooling device having no convex surface to control vapor flow.
FIG. 9 shows application of the invention to a rectangular slot.
FIG. 10 shows how circumferential segments of the toroidal surface can be arranged to absorb heat from a large, rectangular surface.
DETAILED DESCRIPTION OF THE INVENTION
FIG. 11 shows a refrigeration expansion valve with a flow orifice upstream of it to increase the velocity of the fluid jet.
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/cm2 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).
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.
FIG. 2 a 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.
FIG. 2 b 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.
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.
FIG. 4 a 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 4 b is a elevation view of a torus with C=r. FIG. 4 c shows a view of a semi-toroidal surface of 4 b cut by plane 43. FIG. 4 d shows a cross-section of the semi-toroidal surface in 4 c 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.
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.
FIG. 6 a 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.
FIG. 6 b 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.
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. 6 a 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.
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.
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-2C.
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.
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.