Supersonic Cooling with a Pulsed Inlet

Abstract

A supersonic cooling system operates by pumping liquid without the need of a condenser. The compression system utilizes a compression wave in the generation of the cooling effect. An inlet of the system may be pulsed to reduce energy required of a pump. The evaporator of compression system operates in the critical flow regime where the pressure one or more evaporator tubes will remain almost constant and then ‘jump’ or ‘shock up’ to the ambient pressure.

Description

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention generally relates to cooling via a supersonic fluid flow cycle. More specifically, the present invention is related to a supersonic fluid flow cycle that utilizes a pulsed inlet.

2. Description of the Related Art

Vapor compression systems are used in many cooling applications such as air conditioning and industrial refrigeration. A vapor compression system generally includes a compressor, a condenser, an expansion device, and an evaporator. In a prior art vapor compression system, a gas in a saturated vapor state is compressed to raise the temperature of that gas, the gas then being in a superheated vapor state. The compressed gas is then run through a condenser and turned into a liquid, and heat is rejected from the system. The condensed and liquefied gas is then taken through an expansion device, which drops the pressure and the corresponding temperature. The resulting refrigerant is then boiled in an evaporator, with the refrigerant absorbing heat. The saturated vapor is then returned to the compressor.

FIG. 1 illustrates a vapor compression system 100 as might be found in the prior art. In the prior art vapor compression system 100 of FIG. 1, compressor 110 compresses the gas to (approximately) 238 pounds per square inch (PSI) and a temperature of 190° F. Condenser 120 then liquefies the heated and compressed gas to (approximately) 220 PSI and 117° F. The gas that was liquefied by the condenser 120 is then passed through the expansion valve 130 of FIG. 1. By passing the liquefied gas through expansion value 130, the pressure is dropped to (approximately) 20 PSI.

A corresponding drop in temperature accompanies the drop in pressure, which is reflected as a temperature drop to (approximately) 34° F. in FIG. 1. The refrigerant that results from dropping the pressure and temperature at the expansion value 130 is boiled at evaporator 140. Through boiling of the refrigerant by evaporator 140, a low temperature vapor results. The vapor is illustrated in FIG. 1 as having a temperature of (approximately) 39° F. and a corresponding pressure of 20 PSI.

The cycle carried out by the system 100 of FIG. 1 is an example of a vapor compression cycle. Such a cycle generally results in a coefficient of performance (COP) between 2.4 and 3.5. The COP, as illustrated in FIG. 1, is the evaporator cooling power or capacity divided by compressor power. It should be noted that the temperature and PSI references that are shown in FIG. 1 are exemplary and are for the purpose of illustration only.

FIG. 2 illustrates the performance that might be expected of a vapor compression system similar to that illustrated in FIG. 1. The COP illustrated in FIG. 2 corresponds to a typical home or automotive vapor compression system operating at an ambient temperature of (approximately) 90° F. The COP shown in FIG. 2 corresponds to a vapor compression system utilizing a fixed orifice tube system.

A system like that illustrated in FIG. 1 and FIG. 2 typically operates at an efficiency rate or COP that is far below that of system potential. To compress gas in a conventional vapor compression system like that illustrated in FIG. 1 (system 100) typically takes 1.75-2.50 kilowatts for every 5 kilowatts of cooling power. This exchange rate is less than optimal and directly correlates to the rise in pressure times the volumetric flow rate. Degraded performance is similarly and ultimately related to performance (or lack thereof) by the compressor 110.

Haloalkane refrigerants such as tetrafluoroethane (CH₂FCF₃) are inert gases that are commonly used as refrigerants in refrigerators and automobile air conditioners. Tetrafluoroethane has also been used to cool over-clocked computers. These gases are referred to as R-134 gases. The volume of an R-134 gas can be 600-1000 times greater than its corresponding liquid form. This multiplier shows that the theoretical efficiency of a system utilizing an R-134 gas is much higher than is currently being realized, and evidences the need for an improved cooling system that more fully recognizes system potential and overcomes technical barriers related to compressor performance.

SUMMARY OF THE CLAIMED INVENTION

A first claimed embodiment of the present invention is a system that includes a fluid flow path with a high pressure region and a low pressure region. The fluid flow path transports a working fluid at a velocity that is greater than or equal to the speed of sound in the fluid as the fluid travels from the high pressure region of the fluid flow path to the low pressure region of the fluid flow path. A pump is used to facilitate the flow of fluid in the fluid flow path. Installing a pulsing valve in the fluid flow path to create a pulsed inlet to the high pressure region and reduce the energy required to generate a given cooling capacity. An optional resonance chamber may be installed in line with the pulsing valve to further reduce the demand on the pump of the system to assist in the formation of a compression wave.

A second claimed embodiment of the present invention is a method that includes pumping a working fluid through a fluid flow path. The fluid flow path includes a low pressure region in which the fluid flows at a critical flow rate. A pulsing valve may be installed downstream from the pump and upstream from the low pressure region of the fluid flow path to reduce the mass flow requirement of the system. An optional resonance chamber may be installed in line with the pulsing valve to assist in the formation of a compression wave to further reduce the demand on the pump.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic diagram of a vapor compression cooling system as may be found in the prior art.

FIG. 2 is a pressure—enthalpy graph for a vapor compression cooling system like that illustrated in FIG. 1.

FIG. 3 is a schematic diagram of an embodiment of a supersonic cooling system with a pulsed inlet.

FIG. 4 is a sectional view of an evaporator tube/nozzle.

FIG. 5 is a graphical representation of the effect of the pulsed inlet on the mass flow rate of a cooling system.

FIG. 6 shows the inlet pressure as a function of time.

DETAILED DESCRIPTION

FIG. 3 illustrates an exemplary supersonic cooling system 300 in accordance with an embodiment of the present invention. The supersonic cooling system 300 does not need to compress a gas as otherwise occurs at compressor 110 in a prior art vapor compression system 100 like that shown in FIG. 1. Supersonic cooling system 300 operates by pumping liquid. Because supersonic cooling system 300 pumps liquid, the compression system 300 does not require the use a condenser 120 as does the prior art compression system 100 of FIG. 1. Compression system 300 instead utilizes a compression wave. The evaporator of compression system 300 operates in the critical flow regime where the pressure in an evaporator tube will remain almost constant and then ‘jump’ or ‘shock up’ to the ambient pressure.

The supersonic cooling system 300 of FIG. 3 recognizes a improved efficiencies in that the pump 310 of the system 300 does not, nor does it need to, draw as much power as the compressor 110 in a prior art compression system 100 like that shown in FIG. 1. A compression system designed according to an embodiment of the presently disclosed invention may recognize exponentially improved pumping efficiencies. For example, where a prior art compression system 100 may require 1.75-2.5 kilowatts for every 5 kilowatts of cooling power, a system 300 like that illustrated in FIG. 3 may pump liquid from 14.7 to 120 PSI with a pump 310 drawing power at approximately 500 W.

The power demands on the pump 310 may be further reduced by adding an accumulator 320 to the fluid flow path. Still further efficiency gains may be realized by installing a pulsing valve 330 and a resonance chamber 340 in the fluid flow path. As a result of these improvements in efficiency, system 300 may utilize many working fluids, including but not limited to water.

The cycle of the supersonic cooling system 300 of FIG. 3 may begin with the pump 310 pumping the working fluid into the accumulator 320. The flow out of the accumulator 320 may be controlled by a pulsing valve 330. An exemplary pulsed mass flow is illustrated in FIG. 5. An exemplary pressure profile resulting from a pulsed flow is illustrated in FIG. 6.

The pulsing valve 330 opens and closes the fluid flow path from the accumulator 320 and operates at a frequency that may be determined by the characteristics of a given installation of the supersonic cooling cycle 300. The pulsing valve 330 may operate at a frequency of from approximately 10 Hz to approximately 100 Hz. It will be recognized by those skilled in the art that many types of fluid control valves may be utilized as the pulsing valve 330. The pulsing valve 330 may have a solenoid or other control mechanism that opens and closes the pulsing valve 330 at the desired frequency.

The utilization of the pulsing valve 330 establishes pressure waves in the working fluid in the cooling system 300. The pulsing valve 330 may reduce the mass flow rate in the system 300 by approximately 50%, and may allow the system 300 to operate at a reduced pressure relative to systems not utilizing a pulsing valve, thereby reducing the demands on the pump 310.

Operation of the pump 310 establishes circulation of the working fluid in the fluid flow path of system 300. Pump 310 may raise the pressure of the working fluid being used by system 300 from, for example, 20 PSI to 100 PSI or more, to establish a high pressure region of the system 300. The temperature of the working fluid may, at this inlet section of the system 300, be approximately 95° F.

To further reduce the energy required to operate the pump 310 to generate a given cooling capacity, a resonance chamber 340 may be installed in line with the pulsing valve 330. The resonance chamber 340 assists in the formation of a compression wave in the system 300. The resonance chamber 340 may be situated between the pulsing valve 330 and a manifold 350 that feeds the evaporator 360. The manifold 350 may supply one or more evaporator tubes or nozzles 400 (see FIG. 4) in the evaporator 360.

As the working fluid is introduced to the evaporator 360, the evaporator 360 induces a pressure drop e.g., to approximately 5.5 PSI, to establish a low pressure region and a concurrent phase change that result in a lowered temperature. The evaporator 360 of system 300 operates in the critical flow regime of the working fluid, thereby establishing a compression wave that assist in the acceleration of the working fluid.

The evaporator 360 may also induce cavitation in the working fluid as part of the phase change. The cavitation also serves to reduce the speed of sound in the working fluid. Further explanation of the cavitation effect is provided below in the description of the evaporator nozzle 400.

As the working fluid is accelerated and undergoes a pressure drop and phase change, the working fluid further ‘boils off’ in evaporator 360, providing the cold sink desired in the system 300. In embodiments in which the working fluid is water, the water may be cooled to 35-45° F., or approximately 37° F. The working fluid exits the evaporator 360 via evaporator tube 360 where the fluid is ‘shocked up’ to approximately 20 PSI.

To facilitate the dissipation of heat in the system 300, the evaporator 360 may be thermally coupled with a heat exchanger 370. The heat exchanger 370 may be thermally coupled with a coolant fluid used in the system 300, the coolant fluid being circulated around or through an area or an object to be cooled. The working fluid of the system 300 may be at a temperature of approximately 90-100° F. after the working fluid exits evaporator 360 and returns to the inlet of pump 310.

FIG. 4 illustrates a structure that may be used in one or more embodiments of the evaporator tube or nozzle 400. A main body 410 of the evaporator nozzle 400 of FIG. 4 includes an inlet portion 420, a throat portion 430, and an expansion portion 440.

The inlet portion 420 receives the working fluid from the inlet section of the cooling system 300. The working fluid is directed into the throat portion 430. The throat portion 430 provides a duct of substantially constant profile (normally circular) through its length through which the working fluid is forced. The expansion portion 440 provides an expanding tube-like member wherein the diameter of the fluid flow path progressively increases between the throat portion 430 and the outlet of the expansion portion 440. The actual profile of the expansion portion 440 may depend upon the specific working fluid to be used in the system 300.

In operation, when the working fluid enters the throat portion 430, the working fluid is accelerated to high speed. The inlet pressure and the diameter of the throat orifice may be selected so that the speed of the working fluid at the entry of the throat portion 430 is approximately the speed of sound (Mach 1).

As the working fluid travels through the nozzle 400, the acceleration of the working fluid causes a sudden drop in pressure which results in cavitation that commences at the boundary between the exit of the inlet portion 420 and the entry to the throat portion 430. Cavitation is also triggered along the wall of the throat portion 430. Cavitation results in bubbles of the working fluid in the vapor phase being present within the fluid in the liquid phase, thereby providing a multi-phase working fluid. The creation of such vapor bubbles requires the input of energy for the input of latent heat of vaporization and as a result the temperature falls. At the same time, the reduction in pressure together with the working fluid achieving a multi-phase state causes the local speed of sound in the working fluid to be lowered, with the result that the working fluid exits the throat portion 430 at a supersonic speed of, for example, Mach 1.1 or higher. It is noted that the reduction in the localized speed of sound changes the character of the flow from traditional incompressible flow to a regime more in character with high speed nozzle flow.

As the working fluid travels within the expansion portion 440, the pressure remains at a low level and the fluid expands. As a result of the expansion, the flow accelerates further, reaching a speed on the order of for example approximately Mach 3.

As the fluid accelerates and pressure is reduced, the local static pressure drops, so that more vapor is generated from the surrounding liquid working fluid. As the working fluid moves below the saturation line, the cold sink required for the cooling method is generated and the flow behaves as if it was in an over expanded jet. Once the working fluid has picked up sufficient heat, and due to frictional losses, the fluid shocks back to a subsonic condition and returns to ambient conditions.

While various embodiments have been described above, it should be understood that they have been presented by way of example only, and not limitation. The descriptions are not intended to limit the scope of the invention to the particular forms set forth herein. Thus, the breadth and scope of a preferred embodiment should not be limited by any of the above-described exemplary embodiments. It should be understood that the above description is illustrative and not restrictive. To the contrary, the present descriptions are intended to cover such alternatives, modifications, and equivalents as may be included within the spirit and scope of the invention as defined by the appended claims and otherwise appreciated by one of ordinary skill in the art. The scope of the invention should, therefore, be determined not with reference to the above description, but instead should be determined with reference to the appended claims along with their full scope of equivalents.

Claims

1. A supersonic cooling system, comprising:

a pump facilitating a flow of a fluid through a fluid flow path, the fluid flow path having a high pressure region and a low pressure region, the pump transporting the fluid at a velocity that is greater than or equal to the speed of sound in the fluid as the fluid travels from the high pressure region to the low pressure region; and

a pulsing valve creating a pulsed inlet to the high pressure region of the fluid flow path, thereby reducing the mass flow rate of the fluid and corresponding energy required for a given cooling capacity.

2. The supersonic cooling system of claim 1, further including a resonance chamber situated downstream of the pulsing valve to assist in the formation of a compression wave.

3. The supersonic cooling system of claim 1, further comprising an evaporator at the low pressure region of the fluid flow path, the evaporator facilitating a phase change of the fluid.

4. The supersonic cooling system of claim 3, wherein the evaporator includes an evaporator tube that maintains a constant pressure of the fluid.

5. The supersonic cooling system of claim 3, wherein fluid flow in the evaporator is in a critical flow regime of the fluid.

6. The supersonic cooling system of claim 3, wherein the evaporator facilitates a fluid shock up to an elevated pressure as the fluid exits the evaporator.

7. The supersonic cooling system of claim 6, wherein the evaporator facilitates the fluid shock up to the elevated pressure at substantially constant enthalpy.

8. The supersonic cooling system of claim 1, wherein the fluid flow path decreases a pressure of the fluid at substantially constant enthalpy.

9. The supersonic cooling system of claim 1, wherein the fluid includes water.

10. The supersonic cooling system of claim 1, further comprising a heat exchanger to transfer heat to the fluid.

11. A supersonic cooling method, comprising:

pumping a fluid through a fluid flow path with the aid of a pump, the fluid flow path including a low pressure region wherein the fluid flows at a critical flow rate; and

pulsing a fluid input to the fluid flow path through a pulsing valve situated downstream from the pump and upstream from the low pressure region of the fluid flow path to reduce the mass flow rate of the fluid and corresponding power required for a given cooling capacity.

12. The supersonic cooling method of claim 11, further comprising generating a compression wave, wherein generation of the compression wave includes the use of a resonance chamber situated downstream of the pulsing valve.

13. The supersonic cooling method of claim 11, further comprising generating a phase change in the fluid, wherein generating the phase change includes the use of an evaporator that operates in the low pressure region of the fluid flow path.

14. The supersonic cooling method of claim 13, wherein the phase change occurs at least in part due to fluid flow within the evaporator being in a critical flow regime of the fluid.

15. The supersonic cooling method of claim 13, wherein the fluid shocks up to an elevated pressure as the fluid exits the evaporator.

16. The supersonic cooling method of claim 15, wherein the fluid shocks up to the elevated pressure at substantially constant enthalpy.

17. The supersonic cooling method of claim 11, further comprising transferring heat to the fluid, the transfer of heat accompanying a phase change of the fluid.

18. The supersonic cooling method of claim 11, further comprising transferring heat to the fluid via a heat exchanger.

19. The supersonic cooling method of claim 11, wherein the fluid flows from a high pressure region to the low pressure region of the fluid flow path at substantially constant enthalpy.

20. The supersonic cooling method of claim 11, wherein the fluid flows at a velocity greater than or equal to the speed of sound in at least a portion of the fluid flow path between a high pressure region and the low pressure region.