- T_B=normal boiling point, K
- R=ideal gas constant, 8.3145 J·K⁻¹mol⁻¹
- P₀=vapor pressure at a given temperature, atm
- ΔH_vap=heat of vaporization of the coolant, J/mol
- T₀=given temperature, K
- Ln=that natural logarithm to the base e.

In the above equation, the given temperature (T₀) is the temperature of coolant in contact with—and heated by—the surface. The normal boiling point (T_B) is the boiling point of the coolant at one (1) atmosphere. The heat of vaporization (ΔH_vap) is the amount of energy required to convert or vaporize a saturated liquid (i.e., a liquid at its boiling point) into a vapor. As an alternative to determining the appropriate pressure theoretically, the appropriate pressure can be determined empirically by adjusting the pressure and detecting evaporation or bubble generation at a surface as shown in the examples.

Some versions of the invention comprise flowing a coolant through a chamber with the surface exposed therein at a pressure that promotes a phase change upon the coolant contacting—and being heated by—the surface. Anexemplary apparatus1 for performing such a method is shown inFIG. 1. Theexemplary apparatus1 includes achamber10 with asurface12 to be cooled exposed therein, apump20, and apressurizer30, all in fluid communication. As used herein, “fluid communication” between two or more elements refers to a configuration in which fluid can be communicated between or among the elements and does not preclude the possibility of having a filter, flow meter, or other devices disposed between such elements. The elements comprising theapparatus1 are preferably configured in a closed fluidic system, as shown inFIG. 1. However, in other versions, they may also be configured in an open system, wherein thepump20 is disposed upstream of thechamber10 and thepressurizer30 is disposed downstream of thechamber10. Theapparatus1 is capable of including one18 or more19 additional chambers in fluid communication with thepump20 andpressurizer30.

Thechamber10 includes one ormore inlets14 to permit fluid to inter thechamber10 and one ormore outlets15 to permit fluid to exit thechamber10. In this manner, thechamber10 is configured to permit fluid to flow therethrough. Theinlets14 are preferably configured to project astream16 of a fluid, such as acoolant50, against thesurface12. Thestream16 of fluid projected against thesurface12 is preferably a jet stream but may also be a spray stream. As used herein, a “jet” or “jet stream” refers to a substantially liquid fluid filament that is projected through a substantially liquid or fluid medium or a mixture thereof. “Jet” or “jet stream” is contrasted with “spray” or “spray stream,” wherein “spray” or “spray stream” refers to a substantially atomized liquid fluid projected through a substantially vapor medium.

Thesurface12 exposed within thechamber10 preferably comprises a surface portion of awork piece13, such that thestreams16 ofcoolant50 impinge directly on thework piece13 without thermal interference materials disposed between thework piece13 and thecoolant50. As used herein, “work piece” refers to any electronic or other device having a surface that generates heat and that is desired to be cooled. Non-limiting,exemplary work pieces13 include microprocessors, microelectronic circuit chips in supercomputers, or any other electronic circuits or devices requiring cooling such as diode laser packages. Thesurface12 can be exposed within thechamber10 by constructing thechamber10 around thework piece13 to include thesurface12 within thechamber10.

Thepump20 is in fluid communication at least with theinlets14 of thechamber10 and is configured to forceflow51 ofcoolant50 through theapparatus1 and into thechamber10 through theinlets14. Any pump capable of generating apositive coolant50 pressure to force thecoolant50 through theinlets14 and against thesurface12 is suitable for use in the present invention. Thepump20 is preferably a variable speed positive displacement pump, such as a gear pump. An example includes the “MICROPUMP”-brand gear pump (Cole-Parmer, Vernon Hills, Ill.). A variable speed pump enables theflow51 ofcoolant50 to be set at a rate required to meet the expected heat load at thesurface12. In place of or in addition to apump20, a reservoir ofpressurized coolant50 may be used.

Thepressurizer30 is a device capable of ensuring that thechamber10 is maintained at a particular pressure. Thepressurizer30 is preferably in fluid communication with thechamber10, such as theoutlet15 of thechamber10, as well as thepump20. The pressurizer30 may be disposed anywhere on a low-pressure side of the apparatus, i.e., between thepump20 and the chamber10 (see below), but is preferably close to thepump20. In versions of the invention configured to project ajet stream16 ofcoolant50 against thesurface12, thepressurizer30 is configured to ensure that thechamber10 is filled withcoolant50 and to maintain the pressure of thecoolant50 in thechamber10. In versions of the invention configured to project aspray stream16 ofcoolant50 against thesurface12, thepressurizer30 is configured to maintain the vapor pressure in thechamber10. In addition to determining the pressure in thechamber10, thepressurizer30 preferably maintains pressure ofcoolant50 at aninlet21 of thepump20 at least slightly above the saturation pressure to prevent cavitation. As used herein, “saturation pressure” is the pressure for a corresponding saturation temperature at which a liquid evaporates into its vapor phase. The pressurizer30 further preferably maintains pressure in cases of significant vapor generation or “hot swapping” of components, the latter of which may add or decrease the amount ofcoolant50 in the system.

In the preferred version, thepressurizer30 comprises a pressurizing tank. The tank is preferably sized to accomplish any or each of the above functions. The tank may employ a bladder or may use a gas volume without a bladder. The pressurized gas volume, possibly separated from the liquid by a bladder, is maintained at a constant pressure by apressure regulator60. As the liquid volume in the tank changes, the gas volume changes to accommodate the change in liquid volume, with thepressure regulator60 allowing gas into or out of the tank as necessary to maintain a constant pressure. The pressurizer30 may also be implemented via a piston-cylinder apparatus. The piston in the piston-cylinder apparatus may be controlled by springs. Alternatively, the piston may be connected to a small, sealed reservoir containing vapor and saturated refrigerant at equilibrium. As the liquid level in the cylinder changes, the vapor volume in the small reservoir decreases or increases accordingly at constant pressure by evaporating or condensing. If the temperature of the refrigerant in the small reservoir changes, the pressure changes accordingly, which allows this device to be used as a temperature-controlled pressurizer. This concept is used in thermostatic valves for air conditioners and refrigerators. Yet other versions ofpressurizers30 include alternative tubing configurations. For example, valved bypass loops leading from thepump20 to a position downstream of thechamber10 may provide optional suction from theoutlet15 of thechamber10 to decrease pressure within thechamber10 when required. Alternatively or in addition, valved loops leading from a pump, such aspump20, to thechamber10 in a manner that bypasses the flow-restrictedinlets14 may provide increased delivery ofcoolant50 to thechamber10 to increase pressure when required.

Theapparatus1 preferably further includes aheat exchanger40 in fluid communication with thechamber10,pressurizer30, and pump20. Theheat exchanger40 is preferably disposed downstream of thechamber10 and upstream of thepressurizer30 and/or pump20. “Downstream” and “upstream” are used herein in relation to the direction offlow51 ofcoolant50 within theapparatus1. Inclusion of aheat exchanger40 ensures that coolant exiting thechamber10 is sufficiently cooled to below the saturation temperature established by the pressure within thechamber10 before being recycled back into thechamber10. Apreferred heat exchanger40 exchanges heat from thecoolant50 exiting thechamber10 to anexternal cooling fluid42,44 (reference42 refers to non-heat-exchanged external cooling fluid, andreference44 refers to heat-exchanged external cooling fluid). Any heat exchanger capable of cooling thecoolant50 to below the saturation temperature is acceptable. Non-limiting examples include shell-and-tube, plate, adiabatic-wheel, plate-fin, pillow-plate, fluid, dynamic-scraped-surface, phase-change, direct contact, and spiral heat exchangers. Theheat exchanger40 may operate by parallel flow or counter flow. The heat exchanger may also be designed to incorporate air as the second cooling fluid. This air-liquid heat exchanger may be a fin-and-tube or micro-channel implementation, among many other air-liquid heat exchangers known in the art. Yet another version of a heat exchanger includes piping providing fluid communication from theoutlet15 of thechamber10 to theinlet14 of the chamber, as well as the other components therebetween, wherein the piping or portions thereof is suitably configured to dissipate thermal energy of the fluid50 flowing therethrough.

As shown inFIG. 1, anyadditional chambers18 may be added to the device in parallel such that they are serviced by thesame pressurizer30, pump20, andheat exchanger40. Alternatively or in addition, the device may includeadditional pressurizers30, pumps20, and/orheat exchangers40 in parallel for the purpose of redundancy or reliability. As used herein, an additional component “in parallel” refers to a component in fluid communication with the other components in a manner that bypasses only components of the same type without bypassing different types of components. An example of an additional component added in parallel is shown with theadditional chamber18 inFIG. 1, wherein theadditional chamber18 bypasses thechamber10 without bypassing thepressurizer30, pump20, orheat exchanger40.

Anapparatus1 as described above and shown inFIG. 1 has distinct, steady-statezones comprising coolant50 at different temperatures and pressures. A zone comprising high-temperature coolant52 includes the coolant surrounding thesurface12 within the chamber10 (excluding thestreams16 ofcoolant50 projected into the chamber10) and extends downstream to the heat exchanger40 (seeFIG. 1 for direction of flow51). The high-temperature coolant52 is preferably at a temperature approximately equal to its saturation temperature. A zone of low-temperature coolant53 extends from downstream of theheat exchanger40 to at least theinlets14 of thechamber10 and includes thestreams16 ofcoolant50 injected into thechamber10. The low-temperature coolant53 is preferably at a temperature slightly below the saturation temperature of thecoolant50 surrounding thesurface12, wherein “slightly below” may comprise about 0.5° C., about 1° C., about 3° C., about 5° C., about 7° C., about 10° C., about 15° C., about 20° C., or about 30° C. or more below the saturation temperature ofcoolant50 surrounding thesurface12. Theheat exchanger40 serves to transition the high-temperature coolant52 to low-temperature coolant53, wherein theheat exchanger40 transfers the difference in thermal energy between the high-temperature coolant52 and the low-temperature coolant53 to the non-heat-exchangedexternal cooling fluid42, thereby generating heat-exchangedexternal cooling fluid44. Thesurface12 serves to transition the low-temperature coolant53 to high-temperature coolant52, wherein thesurface12 heats thecoolant50 contacting thesurface12 to the saturation temperature, thereby promoting evaporation.

A zone of low-pressure coolant55 includes the coolant surrounding thesurface12 within the chamber10 (which excludes thestreams16 ofcoolant50 projecting into the chamber10) and extends downstream to an inlet of thepump21. The low-pressure coolant55 is preferably at a pressure that promotes evaporation of coolant when heated at thesurface12. Therefore, the pressure of the low-pressure coolant55 preferably determines a saturation temperature to be about equal to the temperature of the high-temperature coolant52. A zone of high-pressure coolant54 includes a portion downstream of thepump20 and extends to at least theinlets14 of thechamber10. The high-pressure coolant54 is preferably at a pressure suitable for generating streams offluid16 that contact thesurface12. Thepump20 serves to transition the low-pressure coolant55 to high-pressure coolant54, wherein thepump20 applies a positive pressure against the flow-restrictedinlets14. Theinlets14 and thesurface12 serve to transition the high-pressure coolant54 to low-pressure coolant55, wherein thecoolant50 equilibrates to the pressure of the low-pressure coolant55 after contacting thesurface12.

With theapparatus1 as described above, a flow rate is set by thepump20 to meet the expected heat load produced by thesurface12. A specific pressure for the low-pressure coolant55 is set and maintained by the pressurizer30 to establish a saturation temperature for the coolant surrounding thesurface12 to be slightly above the temperature of the low-temperature coolant53. High-pressure54, low-temperature53 coolant is projected from theinlets14 of thechamber10 against thesurface12, whereby thecoolant50 undergoes a pressure drop upon equilibrating with fluid in thechamber10 and also heats to the saturation temperature upon contacting thesurface12. The heated coolant undergoes a partial phase transition at thesurface12, which causes efficient cooling of thesurface12. The resulting low-pressure55, high-temperature52 coolant flows through theheat exchanger40, where it is cooled to below the saturation temperature. This generates low-pressure55, low-temperature53 coolant. The low-pressure55, low-temperature53 coolant is then transitioned to high-pressure54, low-temperature53 coolant by virtue of thepump20. The high-pressure54, low-temperature53 coolant is recycled back into thechamber10.

In many practical applications, the

external cooling fluid

42,44 flowing through theheat exchanger40 is not held at a constant temperature due to varying heat load produced by thesurface12 or varying temperatures of ambient that exchanges heat with cooling

fluid

42,44. With the above-describedapparatus1, the pressure produced by thepressurizer30 would therefore need to be set to a point corresponding to the warmest temperature expected for the high-temperature coolant52 to prevent overheating of thesurface12 due to critical heat flux. However, setting the pressure according to the warmest temperature expected for the high-temperature coolant52 would result in a phase change not occurring when the heat load of thesurface12 is not sufficient to heat the high-temperature coolant52 to the warmest expected temperature. This leads to non-optimal cooling.

As shown inFIG. 1, a more efficient arrangement is to include apressure regulator60 that responds to the temperature of the heat-exchangedexternal cooling fluid44 by modulating the pressure maintained by thepressurizer30. The temperature or change in temperature of the heat-exchangedexternal cooling fluid44 serves as an indicator of the temperature of the high-temperature fluid52 and, thus, of the heat load of thesurface12. Thepressure regulator60 may include a device that detects the temperature or change in temperature of the heat-exchangedexternal cooling fluid44. Thepressure regulator60 then communicates with the pressurizer30 to adjust the pressure of the low-pressure coolant55 to re-establish the saturation temperature. As the temperature of the heat-exchangedexternal cooling fluid44 rises, thepressure regulator60 increases the pressure set and maintained by the pressurizer30 to increase the saturation temperature. Conversely, as the temperature of the heat-exchangedexternal cooling fluid44 lowers, thepressure regulator60 decreases the pressure set by the pressurizer30 to decrease the saturation temperature. This regulation ensures that a phase change occurs at a variety ofcoolant50 temperatures and prevents reaching the critical heat flux as a result of having the saturation temperature set at too low a value. Thepressure regulator60 may be aided by anair compressor62. Theair compressor62 serves to create a reservoir of constant high pressure as a source of pressure for thepressure regulator60. In some versions of the invention, thepressure regulator60 may directly detect the temperature of the high-temperature coolant52.

Theapparatus1 may further include a controller and variable speed drive for thepump20. See, e.g., U.S. Pat. Pub. 2006/0196627 to Shedd et al., incorporated herein by reference. These elements enable thepump20 to operate at a lower power when the thermal load falls. The ability to operate at a lower power further conserves energy.

As shown inFIG. 2, the one ormore inlets14 of thechamber10 preferably comprises one or moretubular nozzles70 that extend into thechamber10. The tubular nozzles provide adrainage path72 through thechamber10 to prevent exiting coolant from substantially interfering with thestreams16 projected from theinlet14. This protects theincoming streams16 from the exiting, warm coolant. Thetubular nozzles70 and associatedinlet manifold71 may be made from a variety of materials selected for ease of manufacture and compatibility with the chosen coolant. They may even be injection molded to cut manufacturing costs significantly. Eachtubular nozzle70 comprises a central axis74 defined by the extended dimension of thetubular nozzle70. The central axis of the tubular nozzle74 may either be angled perpendicularly with respect to thesurface12 or angled non-perpendicularly with respect to thesurface12, the latter of which is shown inFIG. 2. If angled non-perpendicularly with respect to thesurface12, the central axis74 may define any angle between 0° and 90° with respect to thesurface12, such as about 5°, about 10°, about 15°, about 20°, about 25°, about 30°, about 35°, about 40°, about 45°, about 50°, about 55°, about 60°, about 65°, about 70°, about 75°, about 80° or about 85° or any range therebetween. Thetubular nozzles70 may comprise any cross-sectional shape when viewed along the central axis. Various versions include a circular shape, an oval shape (to generate a fan-shaped nozzle), and virtually any other cross-sectional shape.

Thechamber10 preferably includes anarray76 oftubular nozzles70. The central axes74 of thetubular nozzles70 in thearray76 may define different angles with respect to thesurface12. A preferred arrangement is wherein the central axis74 of eachtubular nozzle70 in thearray76 comprises the same angle with respect tosurface12, as shown inFIG. 2.

The array oftubular nozzles70 may be arranged in any configuration suitable for cooling thesurface12. In a version of the invention depicted inFIG. 3, thearrays76 are organized into staggeredcolumns77 androws78. The staggering oftubular nozzles70 in thearray76 is such that a giventubular nozzle70 in a givencolumn77 androw78 does not have a correspondingtubular nozzle70 in a neighboringrow78 in the givencolumn77 or a correspondingtubular nozzle70 in a neighboringcolumn77 in the givenrow78. If thetubular nozzles70 are configured to induce a substantially same direction offlow90 along the surface12 (see below), either thecolumns77 or therows78 are preferably oriented substantially perpendicularly to the substantially same direction offlow90. Arrays oftubular nozzles70 in a non-staggered arrangement can also be used in the present invention.

In a preferred version of the invention, thetubular nozzles70 are configured to project astream16 that impinges on thesurface12 such that at least thecentral axis17 of thestream16, and more preferably theentire stream16, impinges non-perpendicularly on the surface12 (i.e, at an angle other than 90° with respect to the surface). As a non-limiting example, thecentral axis17 of thestream16 may impinge on thesurface12 at any angle between 0° and 90°, such as about 1°, about 2°, about 3°, about 4°, about 5°, about 7°, about 10°, about 15°, about 20°, about 25°, about 30°, about 35°, about 40°, about 45°, about 50°, about 55°, about 60°, about 65°, about 70°, about 75°, or about 80° or any range therebetween.FIG. 4 depicts a top plan view of asurface12 on which eachstream16 of an array of streams impinges non-perpendicularly on thesurface12. Such impingement creates a flow pattern in which all thecoolant50 flows along thesurface12 in the substantiallysame direction90. In some versions of patterns flowing in the substantiallysame direction90, flow ofcoolant50 at each portion of thesurface12 comprises a common directional vector component along a plane defined by thesurface12. In other versions,coolant50 at no two points on thesurface12 flows in opposite directions. In yet other versions,coolant50 at no two points on thesurface12 flows in opposite directions or flows in perpendicular directions. Flowingcoolant50 in the substantially same direction eliminates stagnant regions on the surface.

As further shown inFIG. 4,tubular nozzles70 in thearray76 are preferably configured to impingestreams16 on thesurface12 in anarray96 of contact points91 comprisingstaggered columns97 androws98. The staggering is such that a givencontact point91 in a givencolumn97 androw98 does not have acorresponding contact point91 in a neighboringcolumn97 in the givenrow98 or acorresponding contact point91 in a neighboringrow98 in the givencolumn97. If thecoolant50 is induced to flow across thesurface12 in a substantiallysame direction90, as inFIG. 4, either thecolumns97 or therows98 are preferably oriented substantially perpendicularly to the substantiallysame direction90 of flow.Arrays96 of contact points91 arranged in this manner permitcoolant50 emanating from eachcontact point91 in a givencolumn97 orrow98 to flow substantially between contact points91 in a neighboringcolumn97 orrow98, respectively, as shown inFIG. 4. Even, consistent flow ofcoolant50 over asurface12 without stagnant regions94 as provided by this configuration encourages bubble generation and evaporation whereby the heat transfer performance increases significantly.

A preferred version of the invention includes anarray76 oftubular nozzles70 with eachtubular nozzle70 having a central axis74 angled non-perpendicularly with respect to thesurface12, wherein eachtubular nozzle70 projects astream16 having acentral axis17 collinear with the central axis74 of thetubular nozzle70, and wherein all the tubular nozzles80 have central axes74 oriented at the same angle, project streams16 having the same trajectory and shape, and impinge against thesurface12 at the same angle of impingement. Thearray76 oftubular nozzles70 is preferably further included in anapparatus1 as illustrated and described with respect toFIG. 1. Such a preferred device optimally promotes bubble generation and evaporation at the surface, thereby achieving higher heat transfer performance than conventional impingement cooling systems, whether those systems are direct, using a dielectric fluid, or indirect, using a liquid coolant within a cold plate. Other implementations may promote bubble generation using structures within thetubular nozzles70, such as structures that encourage cavitation or degassing of non-condensable gasses absorbed in the liquid.

The elements and method steps described herein can be used in any combination whether explicitly described or not. All combinations of method steps as described herein can be performed in any order, unless otherwise specified or clearly implied to the contrary by the context in which the referenced combination is made.

As used herein, the singular forms “a,” “an,” and “the” include plural referents unless the content clearly dictates otherwise.

Numerical ranges as used herein are intended to include every number and subset of numbers contained within that range, whether specifically disclosed or not. Further, these numerical ranges should be construed as providing support for a claim directed to any number or subset of numbers in that range. For example, a disclosure of from 1 to 10 should be construed as supporting a range of from 2 to 8, from 3 to 7, from 5 to 6, from 1 to 9, from 3.6 to 4.6, from 3.5 to 9.9, and so forth.

All patents, patent publications, and peer-reviewed publications (i.e., “references”) cited herein are expressly incorporated by reference to the same extent as if each individual reference were specifically and individually indicated as being incorporated by reference. In case of conflict between the present disclosure and the incorporated references, the present disclosure controls.

The methods and compositions of the present invention can comprise, consist of, or consist essentially of the essential elements and limitations described herein, as well as any additional or optional steps, ingredients, components, or limitations described herein or otherwise useful in the art.

The present disclosure is filed simultaneously with U.S. application Ser. No. ______ to Timothy A. Shedd, filed April X,2011 under Attorney Docket Number 09820.420, and entitled “Dual-Loop Cooling System,” the entirety of which is incorporated herein by reference.

It is understood that the invention is not confined to the particular construction and arrangement of parts herein illustrated and described, but embraces such modified forms thereof as come within the scope of the claims.

EXAMPLE

Many impingement technologies exist, but few have shown commercial promise and none have gained wide-scale commercial acceptance to date due to generally high flow rate requirements and limitations on scalability.

An improved impinging jet array apparatus has been developed and described herein. As described above, the current work has identified that angling the tubular nozzles and impinging the stream at a non-perpendicular angle with respect to the surface significantly improves the scalability of arrays of jets.

In addition, laboratory tests have demonstrated that two-phase impinging jets can perform 80% to 100% better than single-phase jets with the same flow rate. A chamber comprising a tubular nozzle configured to project a jet impinging on a work piece surface was configured. The pressure in the chamber was set to establish a saturation temperature of either 95° C. (FIG. 5A) or 74° C. (FIG. 5B). The latter saturation temperature (74° C.) was chosen to substantially match the mean temperature of the heater surface in the test. The same flow rate was used for each saturation temperature. As shown inFIG. 5B, bubbles were generated in the chamber with the fluid having the lower saturation temperature. Such a phase change did not occur in the chamber with fluid having the higher saturation temperature (FIG. 5A). The heat transfer performance was increased by 80% with the lower saturation temperature compared to the higher saturation temperature.

A thermal resistance of 0.4 K/(W/cm^{2) can be attained using HFE-}7100 and pressure drops across the inlets of less than 3 psi. It is anticipated that a thermal resistance of 0.3 K(W/cm²) or lower will be achieved with optimization of impinging jet arrays and fluid management.

A benefit of this technology is the management of local hot spots, even down to the processor level. If just one core of a given processor is more heavily used than the others in the same processor, more vapor will be generated in the fluid impinging on that region of the surface, and the temperature will be maintained much more uniformly than what is possible with a single-phase cooling system.