US20110094249A1

US20110094249A1 - Pressure Shock-Induced Cooling Cycle

Info

Publication number: US20110094249A1
Application number: US12/961,015
Authority: US
Inventors: Jayden Harman; Thomas Gielda
Original assignee: CAITIN Inc F/K/A NEW PAX Inc; Caitin Inc
Current assignee: Pax Scientific Inc
Priority date: 2009-03-25
Filing date: 2010-12-06
Publication date: 2011-04-28
Also published as: US20110088419A1; IL215350A0; GB201021925D0; GB201021892D0; US20140174113A1; US8353168B2; EP2411744A1; KR20120093060A; WO2010111560A1; GB2473981B; GB2472965A; GB2473981A; BRPI1012630A2; JP2012522204A; US20110088878A1; US20100287954A1; US8333080B2; CN102449413A; US8353169B2; AU2010229821A1

Abstract

A supersonic cooling system operates by pumping liquid. Because the supersonic cooling system pumps liquid, the compression system does not require the use of a condenser. The compression system utilizes a compression wave. An evaporator of the compression system 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.

Description

CROSS-REFERENCE TO RELATED APPLICATIONS

The present application is a continuation and claims the priority benefit of U.S. patent application Ser. No. 12/732,131 filed Mar. 25, 2010, which claims the priority benefit of U.S. provisional application No. 61/163,438 filed Mar. 25, 2009 and U.S. provisional application No. 61/228,557 filed Jul. 25, 2009. The disclosure of each of the aforementioned applications is incorporated herein by reference.

BACKGROUND OF THE INVENTION

1. Field of the Invention
The present invention generally relates to cooling systems. The present invention more specifically relates to supersonic cooling systems.
2. Description of the Related Art
A vapor compression system as known in the art generally includes a compressor, a condenser, and an evaporator. These systems also include an expansion device. In a prior art vapor compression system, a gas is compressed whereby the temperature of that gas is increased beyond that of the ambient temperature. The compressed gas is then run through a condenser and turned into a liquid. 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. This vapor compression cycle is generally known to those of skill in the art.
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, which is illustrated in FIG. 1 as having (approximately) a temperature of 39 F and a corresponding pressure of 20 PSI.
The cycle related to the system 100 of FIG. 1 is sometimes referred to as the vapor compression cycle. Such a cycle generally results in a coefficient of performance (COP) between 2.4 and 3.5. The coefficient of performance, as reflected 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 reflected in FIG. 1 are exemplary and illustrative.
A vapor compression system 100 like that shown in FIG. 1 is generally effective. FIG. 2 illustrates the performance of a vapor compression system like that illustrated in FIG. 1. The COP illustrated in FIG. 2 corresponds to a typical home or automotive vapor compression system—like that of FIG. 1—with an ambient temperature of (approximately) 90 F. The COP shown in FIG. 2 further corresponds to a vapor compression system utilizing a fixed orifice tube system.
Such a system 100, however, operates at an efficiency rate (e.g., coefficient of performance) that is far below that of system potential. To compress gas in a conventional vapor compression system (100) like that illustrated in FIG. 1 typically takes 1.75-2.5 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 high-temperature refrigerants in refrigerators and automobile air conditioners. Tetrafluoroethane have also been used to cool over-clocked computers. These inert, refrigerant gases are more commonly referred to as R-134 gases. The volume of an R-134 gas can be 600-1000 times greater than the corresponding liquid. As such, there is a need in the art for an improved cooling system that more fully recognizes system potential and overcomes technical barriers related to compressor performance.

SUMMARY OF THE CLAIMED INVENTION

In a first claimed embodiment of the present invention, a supersonic cooling system is disclosed. The supersonic cooling system includes a pump that maintains a circulatory fluid flow through a flow path and an evaporator. The evaporator operates in the critical flow regime and generates a compression wave. The compression wave shocks the maintained fluid flow thereby changing the PSI of the maintained fluid flow and exchanges heat introduced into the fluid flow.
In a specific implementation of the first claimed embodiment, the pump and evaporator are located within a housing. The housing may correspond to the shape of a pumpkin. An external surface of the housing may effectuate forced convection and a further exchange of heat introduced into the compression system.
The pump of the first claimed embodiment may maintain the circulatory fluid flow by using vortex flow rings. The pump may progressively introduce energy to the vortex flow rings such that the energy introduced corresponds to energy being lost through dissipation.
A second claimed embodiment of the present invention sets for a cooling method. Through the cooling method of the second claimed embodiment, a compression wave is established in a compressible fluid. The compressible liquid is transported from a high pressure region to a low pressure region and the corresponding velocity of the fluid is greater or equal to the speed of sound in the compressible fluid. Heat that has been introduced into the fluid flow is exchanged as a part of a phase change of the compressible fluid.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 illustrates a vapor compression system as might be found in the prior art.

FIG. 2 illustrates the performance of a vapor compression system like that illustrated in FIG. 1.

FIG. 3 illustrates an exemplary supersonic cooling system in accordance with an embodiment of the present invention.

FIG. 4 illustrates performance of a supersonic cooling system like that illustrated in FIG. 3.

FIG. 5 illustrates a method of operation for the supersonic cooling system of FIG. 3.

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 certain degree of efficiency in that the pump (320) 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 exponential 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, an system (300) like that illustrated in FIG. 3 may pump liquid from 14.7 to 120 PSI with the pump drawing power at approximately 500 W. As a result of these efficiencies, system 300 may utilize many working fluids, including but not limited to water.
The supersonic cooling system 300 of FIG. 3 includes housing 310. Housing 310 of FIG. 3 is akin to that of a pumpkin. The particular shape or other design of housing 310 may be a matter of aesthetics with respect to where or how the system 300 is installed relative a facility or coupled equipment or machinery. Functionally, housing 310 encloses pump 330, evaporator 350, and accessory equipment or flow paths corresponding to the same (e.g., pump inlet 340 and evaporator tube 360). Housing 310 also maintains (internally) the cooling liquid to be used by the system 300.
Housing 310, in an alternative embodiment, may also encompass a secondary heat exchanger (not illustrated). A secondary heat exchanger may be excluded from being contained within the housing 310 and system 300. In such an embodiment, the surface area of the system 300—that is, the housing 310—may be utilized in a cooling process through forced convection on the external surface of the housing 310.
Pump 330 may be powered by a motor 320, which is external to the system 300 and located outside the housing 310 in FIG. 3. Motor 320 may alternatively be contained within the housing 310 of system 300. Motor 320 may drive the pump 330 of FIG. 3 through a rotor drive shaft with a corresponding bearing and seal or magnetic induction, whereby penetration of the housing 310 is not required. Other motor designs may be utilized with respect to motor 320 and corresponding pump 330 including synchronous, alternating (AC), and direct current (DC) motors. Other electric motors that may be used with system 300 include induction motors; brushed and brushless DC motors; stepper, linear, unipolar, and reluctance motors; and ball bearing, homopolar, piezoelectric, ultrasonic, and electrostatic motors.
Pump 330 establishes circulation of a liquid through the interior fluid flow paths of system 300 and that are otherwise contained within housing 310. Pump 330 may circulate fluid throughout system 300 through use of vortex flow rings. Vortex rings operate as energy reservoirs whereby added energy is stored in the vortex ring. The progressive introduction of energy to a vortex ring via pump 330 causes the corresponding ring vortex to function at a level such that energy lost through dissipation corresponds to energy being input.
Pump 330 also operates to raise the pressure of a liquid being used by system 300 from, for example, 20 PSI to 100 PSI or more. Pump inlet 340 introduces a liquid to be used in cooling and otherwise resident in system 300 (and contained within housing 310) into pump 330. Fluid temperature may, at this point in the system 300, be approximately 95 F.
The fluid introduced to pump 330 by inlet 340 traverses a primary flow path to nozzle/evaporator 350. Evaporator 350 induces a pressure drop (e.g., to approximately 5.5 PSI) and phase change that results in a low temperature. The cooling fluid further ‘boils off’ at evaporator 350, whereby the resident liquid may be used as a coolant. For example, the liquid coolant may be water cooled to 35-45 F (approximately 37 F as illustrated in FIG. 3). As noted above, the system 300 (specifically evaporator 350) operates in the critical flow regime thereby allowing for establishment of a compression wave. The coolant fluid exits the evaporator 350 via evaporator tube 360 where the fluid is ‘shocked up’ to approximately 20 PSI because the flow in the evaporator tube 360 is in the critical regime. In some embodiments of system 300, the nozzle/evaporator 350 and evaporator tube 360 may be integrated and/or collectively referred to as an evaporator.
The coolant fluid of system 300 (having now absorbed heat for dissipation) may be cooled at a heat exchanger to assist in dissipating heat once the coolant has absorbed the same (approximately 90-100 F after having exited evaporator 350). Instead of an actual heat exchanger, however, the housing 310 of the system 300 (as was noted above) may be used to cool via forced convection. FIG. 4 illustrates performance of a supersonic cooling system like that illustrated in FIG. 3.
FIG. 5 illustrates a method of operation 500 for the supersonic cooling system 300 of FIG. 3. In step 510, a gear pump 330 raises the pressure of a liquid. The pressure may, for example, be raised from 20 PSI to in excess of 100 PSI. In step 520, fluid flows through the nozzle/evaporator 350. Pressure drop and phase change result in a lower temperature in the tube. Fluid is boiled off in step 530.
Critical flow rate, which is the maximum flow rate that can be attained by a compressible fluid as that fluid passes from a high pressure region to a low pressure region (i.e., the critical flow regime), allows for a compression wave to be established and utilized in the critical flow regime. Critical flow occurs when the velocity of the fluid is greater or equal to the speed of sound in the fluid. In critical flow, the pressure in the channel will not be influenced by the exit pressure and at the channel exit, the fluid will ‘shock up’ to the ambient condition. In critical flow the fluid will also stay at the low pressure and temperature corresponding to the saturation pressures. In step 540, after exiting the evaporator tube 360, the fluid “shocks” up to 20 PSI. A secondary heat exchanger may be used in optional step 550. Secondary cooling may also occur via convection on the surface of the system 300 housing 310.
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 pressure shock-induced cooling method, comprising:

a first isenthalpic step, the first isenthalpic step comprising a shock-induced increase in pressure of a working fluid.

2. The method of claim 1, further comprising a second isenthalpic step, the second isenthalpic step comprising a decrease in pressure of the working fluid.

3. The method of claim 2, wherein the second isenthalpic step further comprises an increase in pressure of the working fluid prior to the decrease in pressure of the working fluid.

4. The method of claim 3, wherein the increase in pressure of the working fluid is to a pressure of about 1 bar or higher.

5. The method of claim 4, wherein the increase in pressure of the working fluid is to a pressure of about 10 bar or higher.

6. The method of claim 3, wherein the increase in pressure of the working fluid is facilitated by a pump.

7. The method of claim 2, wherein the decrease in pressure of the working fluid is to a pressure less than about 0.2 bar.

8. The method of claim 2, wherein the decrease in pressure of the working fluid is facilitated by an evaporator.

9. The method of claim 2, further comprising heating the working fluid between the first and second isenthalpic steps.

10. The method of claim 9, wherein the heating of the working fluid includes heat transfer from a heat exchanger to the working fluid.

11. The method of claim 2, further comprising cooling the working fluid between the first and second isenthalpic steps.

12. The method of claim 11, wherein the cooling of the working fluid includes heat transfer from the working fluid to a heat exchanger.

13. The method of claim 1, wherein the working fluid is a liquid.

14. A method for cooling and heating a fluid circulated through a fluid flow path, the method comprising:

isenthalpically decreasing the pressure of the fluid;

increasing the temperature of the fluid;

insenthalpically increasing the pressure of the fluid, wherein the increase in pressure is shock-induced; and

decreasing the temperature of the fluid.

15. The method of claim 14, wherein the fluid is circulated using a pump.

16. The method of claim 14, wherein the temperature is increased in an evaporator.

17. The method of claim 14, wherein the fluid undergoes an increase in pressure prior to the isenthalpic decrease in pressure.

18. The method of claim 17, wherein the increase in pressure is isenthalpic.

19. The method of claim 17, wherein the increase in pressure is to a pressure of about 1 bar or higher, and the decrease in pressure is to a pressure below about 0.2 bar.

20. The method of claim 14, wherein the isenthalpic decrease in pressure of the fluid occurs at a critical flow rate.