Patent Grant
7318325
The subject invention relates to microclimate cooling, and a miniature cooling system that can be used for any purpose that requires a compact cooling system. Such applications include, but are not limited to, microelectronics cooling such as computer processors and laser diodes, personal cooling systems, and portable cooling systems.
Clothing that protects soldiers, first responders, and other emergency personnel from chemical, biological, nuclear, and/or other similar threats can subject the individuals to heat stress. Certain hazardous environments can require the use of PPE (personal protective ensembles) with level A protection, which can place the working individual in an encapsulating micro-environment. These PPE can significantly diminish the ability of the body to reject heat to the external environment, leading to symptoms ranging from muscular weakness, dizziness and physical discomfort to more severe, life-threatening conditions such as heat exhaustion or heat stroke. In any case, the operational performance of the personnel wearing PPE can become severely impaired. The use of an auxiliary, portable microclimate cooling system can mitigate these effects, eliminate heat stress casualties, and reduce water consumption. At the present time, the efforts to develop a microclimate system have been limited to existing design concepts and use of a large number of commercial off-the-shelf components. The subject microclimate system can incorporate miniaturization and MEMS technology, in order to provide performance that cannot be matched simply by using smaller versions of currently available designs.
An effective compact cooling system (Holtzapple and Allen, 1983) should preferably satisfy the dual requirements of a high coefficient of performance and a light and compact design. One example of an effective and useful microclimate system preferably would be able to remove at least 120 W of heat while consuming no more than 50 W of electrical power for at least about 4 hours of operation. This would suggest that for this particular example the microclimate system would have a coefficient of performance, or heat removal to power input ratio, of 2.4. In conventional designs, the requirements of a high coefficient of performance and a light and compact design typically work against each other.
Current cooling methods, such as thermo-electric cooling and traditional refrigeration cycles, have a high coefficient of performance and efficient design size within certain cooling ranges. While thermo-electric coolers have a coefficient of performance close to 1.0 and a very small volumetric design relative to the cooling capacity when operating in the 10 to 100 watt range, the coefficient of performance of commercially available thermo-electric devices tend be at or below 0.6 when applied to higher cooling capacities. In personal or portable cooling units heat removal rates of this range are inadequate. An alternative to mitigating the lack of performance and increase cooling capacity would be to use more units in series or parallel, thus increasing the overall size and weight of the cooling unit to beyond the limits of portable, microclimate dimensions.
Commercially available refrigeration cycles also have difficulties in satisfying the heat load requirements of microclimate and portable systems while maintaining a light and compact design. Commercially available unit designs are typically optimized for operation above a minimum cooling load of 500 watts, which is too much or unnecessary for microclimate systems. At or above this minimum cooling load refrigeration cycles exhibit a high coefficient of performance of almost never less than two and increases significantly with increasing heat load designs. Furthermore, the size and weight relative to the cooling capacity also decrease with increasing heat load designs. Application of these units to microclimate systems however is difficult due to the large size and weight of such units when scaling down to lower cooling ranges that are suitable for microclimate systems. It is extremely difficult to find a commercially available compressor alone which is smaller than 1 liter and weighs less than several pounds, and which is rated for a cooling load near or below 500 watts. The cycle would then need additional components such a condenser and evaporator to become effective.
Accordingly, there is need for a cooling system having a high coefficient of performance and a light compact design.
The subject invention pertains to a method and apparatus for cooling. In a specific embodiment, the subject invention relates to a lightweight, compact, reliable, and efficient cooling system. The subject system can provide heat stress relief to individuals operating under, for example, hazardous conditions, or in elevated temperatures, while wearing protective clothing. The subject system can be utilized in other applications that can benefit from this type of cooling system. The performance of this system cannot be matched simply by using smaller versions of currently available designs. In a specific embodiment, the subject microclimate system can remove at least about 120 watts of heat while consuming less than about 50 watts of power, and weigh less than about 2.5 pounds while having less than about a 1000 cubic centimeter volume. In a further specific embodiment, the subject cooling system can remove at least about 300 Watts of heat while consuming less than about 100 Watts of electrical power, and can weigh less than about 3.5 pounds (not including the water jacket or the power source) within a volume of less than about 1500 cc or 1.5 L. In a specific embodiment, the subject system can run for at least about 4 hours or more with the use of batteries.
In a specific embodiment, the subject invention pertains to a cooling system having a total weight of less than about 3.5 pounds, a coefficient of performance of at least 2.4, and a volume of less than about 1500 cc with a cooling capacity between about 100 and about 500 watts. The subject cooling system can provide between 28 and 140 watts of cooling per pound and occupy between 3 and 15 cc of volume per watt of cooling. In comparison, commercially available units for cooling in this range would provide between 2.7 and 18.5 watts of cooling per pound and occupy a volume of between 48 and 240 cc per watt of cooling. Furthermore, commercially available units typically provide a coefficient of performance of 2 or less for this cooling range.
The subject system can be scaled to larger or smaller sizes for different applications. The subject system can incorporate a compressor and condenser design so as to achieve a high coefficient of performance and a light and compact design. A compressor can be a key component with respect to the overall performance of a vapor compression system, whereas a condenser can be a key component with respect to the overall weight and size. The subject cooling system can also utilize a miniaturized high efficiency motor design, along with integration of a compact heat exchanger for refrigerant evaporation and liquid pump.
A specific embodiment of the subject cooling system can involve the use of micro-fabrication techniques, an innovative rotary lobed compressor, a miniature high efficiency permanent magnet motor, a high efficiency condenser, a compact heat exchanger for refrigerant evaporation, and a liquid pump. In a specific embodiment, the subject system can provide approximately 200 watts of cooling for microclimate and other cooling environments.
The subject invention also relates to a condenser for transferring heat from a refrigerant to an external fluid in thermal contact with the condenser. The subject condenser can have a heat transfer surface and can be designed for an external fluid, such as air, to flow across the heat transfer surface and allow the transfer of heat from heat transfer surface to the external fluid. In a specific embodiment, the flow of the external fluid is parallel to the heat transfer surface. In another specific embodiment, the heat transfer surface can incorporate surface enhancements which enhance the transfer of heat from the heat transfer surface to the external fluid. In another specific embodiment, an outer layer can be positioned above the heat transfer surface to create a volume between the heat transfer surface and the outer layer through which the external fluid can flow. Such an outer layer can be thin to keep the weight of the system down. A portion, or all, of the outer layer can be thermally insulating, for example for use in cooling systems in contact with a person's skin or clothing. Alternatively, the outer layer can be thermally conducive to assist in thermal transfer to the environment. In an embodiment with the heat transfer surface incorporating surface enhancements, the surface enhancements can contact the outer layer to, for example, maintain the relative position of the heat transfer surface and the outer layer. The subject condenser can allow the flow of refrigerant in ducts or channels such that the refrigerant is in thermal contact with the heat transfer surface and the flow of the refrigerant is substantially parallel with the heat transfer surface. Accordingly, in a specific embodiment, the refrigerant flows substantially parallel to the curve of the heat transfer surface and the external fluid flows substantially parallel to the curve of the heat transfer surface, such that the refrigerant and the external fluid are flowing in substantially parallel curves. In a specific embodiment, while flowing in these substantially parallel curves, the refrigerant and external fluid can be flowing substantially perpendicular to each other. These embodiments of the subject condenser can be incorporated into the subject cooling system.
In a further specific embodiment, the subject condenser can be tubular in shape with the heat transfer surface being on the outside of the tubular condenser. The tubular shaped condenser can then have a first end and a second end. The condenser can have a second surface on the inside of the tubular condenser such that a volume is created by the second surface to the inside of the tubular condenser. This volume can, for example, house elements of a cooling system in accordance with the subject invention. The tubular shaped condenser can have a circular, square, rectangular, polygonal, hexagonal, oval, peanut, or other cross sectional shape. With respect to an embodiment of the tubular shaped condenser, a means for flowing an external fluid across the heat transfer surface can incorporate a fan located at a first end of the tubular shaped condenser which flows air from the first end to the second end, or vice versa, across the heat transfer surface. The fan can also flow air from the first end to the second end of the tubular condenser through the volume formed by the second surface of the condenser so as to, for example, cool other components of a cooling system housed in the volume surrounded by the second surface of the condenser. Such a flow of external fluid from the first end to the second end of the tubular condenser can also allow the transfer of heat from the second surface to the external fluid.
The subject invention pertains to a method and apparatus for cooling. In a specific embodiment, the subject invention relates to a lightweight, compact, reliable, and efficient cooling system. The subject system can provide heat stress relief to individuals operating under, for example, hazardous conditions, or in elevated temperatures while wearing protective clothing. The subject system can be utilized in other applications that can benefit from this type of cooling system. The performance of this system cannot be matched simply by using smaller versions of currently available designs.
The subject invention also relates to a condenser for transferring heat from a refrigerant to an external fluid in thermal contact with the condenser. The subject condenser can have a heat transfer surface and can be designed for an external fluid, such as air, to flow across the heat transfer surface and allow the transfer of heat from heat transfer surface to the external fluid. In a specific embodiment, the flow of the external fluid is parallel to the heat transfer surface. In another specific embodiment, the heat transfer surface can incorporate surface enhancements which enhance the transfer of heat from the heat transfer surface to the external fluid. In another specific embodiment, an outer layer can be positioned above the heat transfer surface to create a volume between the heat transfer surface and the outer layer through which the external fluid can flow. Such an outer layer can be thin to keep the weight of the system down. A portion, or all, of the outer layer can be thermally insulating, for example for use in cooling systems in contact with a person's skin or clothing. Alternatively, the outer layer can be thermally conducive to assist in thermal transfer to the environment. In an embodiment with the heat transfer surface incorporating surface enhancements, the surface enhancements can contact the outer layer to, for example, maintain the relative position of the heat transfer surface and the outer layer. The subject condenser can allow the flow of refrigerant in ducts or channels such that the refrigerant is in thermal contact with the heat transfer surface and the flow of the refrigerant is substantially parallel with the heat transfer surface. Accordingly, in a specific embodiment, the refrigerant flows substantially parallel to the curve of the heat transfer surface and the external fluid flows substantially parallel to the curve of the heat transfer surface, such that the refrigerant and the external fluid are flowing in substantially parallel curves. In a specific embodiment, while flowing in these substantially parallel curves, the refrigerant and external fluid can be flowing substantially perpendicular to each other. These embodiments of the subject condenser can be incorporated into the subject cooling system.
In a specific embodiment, the subject invention relates to a condenser having a tubular body. The subject tubular condenser can have a variety of cross sectional shapes, such as, but not limited to, circular, rectangular, square, polygonal, hexagonal, oval, peanut, or other shapes conducive to the specific use of the system. The tubular shape of the subject condenser can allow other components of a cooling system of which the condenser is part to be located, at least partially, within the volume created by the inner surface of the condenser. In this way, an external fluid such as flowing air can be brought in thermal contact with the condenser to remove heat from the condenser. Referring to
Such a flow path can allow a user to conveniently wear the subject cooling system on the user's body as the flowing air exits the subject cooling system to be directed parallel to the users body while allowing intake of air at the first end unobstructed by the user. In a specific embodiment, the tubular condenser can be contoured to lie against a users body and can house the remaining components of the cooling system within a volume created by an inner surface 800 of the condenser.
The use of cylindrical components as shown in
In a specific embodiment, the subject microclimate system can remove at least about 120 watts of heat while consuming less than about 50 watts of power, and weigh less than about 6 pounds while having less than about a 1000 cubic centimeter volume. In a further specific embodiment, the subject cooling system can remove at least about 300 Watts of heat while consuming less than about 100 Watts of electrical power, and can weigh less than about 3.5 pounds (not including the water jacket or the power source) within a volume of less than about 1500 cc or 1.5 L. In a specific embodiment, the subject system can run for at least about 4 hours or more with the use of batteries. In a specific embodiment, a cooling power to weight ratio of more than 28 W/lb. and/or a volume to cooling power ratio of less than 15 cc/W can be achieved utilizing a vapor compression cycle with cooling capacities lower than 500 W.
A cooling cycle for an embodiment of a microclimate cooling system in accordance with the subject invention can incorporate a vapor compression cycle intended for use with compressible refrigerants. There are four basic features to such a vapor compression cycle. The cycle begins with a compressor that compresses refrigerant vapor to a pressure at which the corresponding vapor temperature is above the ambient temperature of the condenser. The compressed hot refrigerant vapor flows to a condenser that is typically a gas to vapor or liquid to vapor heat exchanger where the vapor is hotter than the gas or liquid. Heat is removed from the compressed refrigerant vapor by the ambient fluid on the other side of the heat exchanger. This causes the temperature of the compressed vaporized refrigerant to decrease below the saturation temperature of the refrigerant and the vapor condenses to liquid. The high pressure liquid can then be expanded through an expansion device, such as a throttling valve, which can cause a rapid decrease in refrigerant pressure after the valve. The lower pressure can cause the temperature of the liquid coolant to drop to, for example, the corresponding saturation temperature.
In a specific embodiment, the cool liquid refrigerant can then flow through an evaporator that allows the liquid refrigerant to absorb the heat from a fluid which is desired to be cooled. The evaporator can act as another heat exchanger with cool refrigerant on one side and the fluid, either liquid or gas, that is desired to be cooled on the other side of the heat exchanger. The absorption of heat in the evaporator causes the liquid refrigerant to boil. The vaporized refrigerant then flows back into the compressor to begin the cycle again. In an alternative embodiment, the evaporator can be in thermal contact with a heat source, such as a metal plate, so that as the refrigerant flows through the evaporator heat is transferred from the heat source to the refrigerant. In a specific embodiment, the embodiment shown in
In a specific embodiment, the subject invention can allow the use of the standard vapor compression cycle in a compact and lightweight design by utilizing specialized components that have been developed specifically for the subject system.
In a specific embodiment, the subject invention can incorporate compressor 515, shown in
Referring to
The shape of an epiterchoid chamber is determined by the following equations:
where MA is the major axis.
In a specific embodiment, a length of 49 mm can be utilized for the major axis of the epitrochoid with a height of 6 mm. Using the above equations, an epiterchoid shape, which is framed in a Cartesian coordinate system, is found to have the shape shown in
Using the equations relating to the shape of the epiterchoid chamber suggested above, the rotor size and shape can also be chosen. Finally, the geometric height of the epiterchoid and rotor can be determined by the amount of fluid that is desired to be displaced on each revolution. After having calculated these dimensions, the compressor's speed can be chosen to determine the displacement per unit time or volumetric flow rate. In a specific embodiment, incorporating an epiterchoidal chamber with a major axis of 49 mm and a height of 6 mm, a speed of 1200 rpm is chosen to provide a mass flow rate of approximately 1 g/s of vapor refrigerant 134a at an inlet pressure of 57 psia.
The flow through the compressor can be controlled by inlet port 517 (shown in
To reduce the vibrations caused by the mass of the rotor spinning eccentrically in the compressor, a counter balance 635 can be placed on the main shaft. A second rotor can be used to balance the compressor. In embodiment the second rotor can be positioned 180° out of phase with the first rotor so as to counter balance the rotating force. The addition of the second rotor adds complexity to the compressor, but can double the mass flow rate for a given RPM speed. Shaft seals and bearings can be used along the shaft to assist in sealing and to absorb the loads caused by the rotating parts. External sealing can be achieved by the shaft seals and gaskets 614 and 628 while internal sealing of the compression chambers can be accomplished using, for example, a sealing gasket 622 or o-ring.
To increase the efficiency and life of the compressor, referring to
Additional methods of sealing may be considered for the compressor as well. Rather than face sealing with gaskets and spring loaded plastics, sufficient sealing can be created by machining the parts with very high precision. In a specific embodiment, the gaps between the rotor and the upper or lower walls are machined to fit to within 0.0005 inches so that the fluid being pressurized has significant difficulty in leaking past the two surfaces.
End plates 612 and fasteners 610 can seal the compressor compartment. To aid in cooling the compressor, cooling fins 636 can be added to the outside housing of the compressor. Cooling fins 636 can be designed to increase the surface area of the outside housing to improve heat transfer out of the compressor housing. Cooling fins 636 can have a variety of shapes. In a specific embodiment, the cooling fins 636 can have long narrow channels running axially with the compressor. During operation of the subject cooling system, air can be blown past the compressor housing to help cool the internal components. In a specific embodiment, air flow provided by the condenser fan 570 can flow between the condenser inner wall surface 800 and the compressor 515 outer wall in space 900, for example as shown in
Where Re is the Reynolds number, Pr is the Prandtl number, w is the channel width, Dh is the hydraulic or effective diameter, μ is the bulk fluid viscosity, and μs is the fluid viscosity at the heat transfer surface.
For a specific embodiment of a compressor in accordance with the subject invention incorporating an epiterchoidal chamber with a major axis of 49 mm, a cross-sectional geometry shown in
This direct cooling of the compressor can aid in the thermodynamic cycle shown in
The motor 513, as shown in
The subject cooling system can be powered by, for example, batteries, AC power, and/or fuel cells. An embodiment powered by batteries can connect to external battery packs or can utilize a central power unit.
The compressed vapor refrigerant exiting outlets 630 of the compressor can flow into a condenser inlet port 820, shown in
The design of the ambient fluid portion of the heat exchanger can involve maximizing the heat transfer from the heat exchanger to the ambient fluid. A simple design of a heat exchanger can incorporate a smooth surface on the outside of the condenser, which can be, for example, flat or curved. In a specific embodiment, the heat exchanger, or condenser, can reject heat from the compressed refrigerant vapor to ambient air and can have a heat transfer surface 880 with enhanced surface geometry that, in conjunction with an air moving device 570 (shown in
q=hAΔT
where q[W] is the heat removal, h[W/m2K] is the heat transfer coefficient, A[m2] is the area of the heated surface, and ΔT [K] is the temperature difference between the heated surface and the ambient fluid such as air. An optimal design can, therefore, maximize h, A, and ΔT so that the product of the three will yield the largest q given space and power limitations.
The subject cooling system, in order to maintain a reduced size, can modify the surface of the condenser so as to increase A as much as possible without substantially increasing the volume of the cooling device. In a specific embodiment, a large number of small extended surface features 860 can be incorporated with the heat transfer surface 880 so as to increase the total heat transfer surface area without significantly increasing the volume of the cooling device. A variety of extended surfaces can be used in conjunction with the subject device. Examples of such extended surfaces are found in DeWitt, D. P. and Incropera, F. P., Fundamentals of Heat and Mass Transfer, John Wiley and Sons, Inc. (1996), which is hereby incorporated herein by reference.
An example of the many different shapes and sizes of extended surfaces 860 which can be utilized by the subject invention is shown in
As discussed, the heat transfer surface 880 can be a smooth, flat or curved, surface or can have extended surface features 860 to increase the surface area without significantly increasing the volume. In a specific embodiment, the extended surfaces can be round, elliptical, square, polygonal, or rectangular fins. For example the extended surfaces can be long fins positioned along the full length of the condenser. In a specific embodiment, the extended surfaces can be a porous material such as expanded copper, aluminum, or carbon. Extended surfaces can increase the surface area by, for example, 2 times more than the base surface area of the heat transfer surface 880. In a specific embodiment, the base surface area is between about 200 and about 500 square centimeters with a surface area increase due to extended surfaces of 2 to 5 times. A further specific embodiment having extended surfaces with respect to a base surface area between about 200 and about 500 square centimeters, with a surface area increase due to extended surfaces of 2 to 5 times, can provide up to 300 watts of cooling. In a further specific embodiment, the bases area is between about 300 and about 400 square centimeters with a surface area increase due to extended surfaces of 2.5 to 4 times and providing between 200 and 250 watts of cooling.
In a specific embodiment, extended surface features 860 can have an elliptical cross section. The elliptical cross section can provide a reduced pressure loss (allowing more air flow) so as to increase h. Examples of the utilization of extended surfaces having elliptical cross sections is given in Li, Q., Chen, Z., Flechtner, U., and Warnecke, H. J., “Heat Transfer and Pressure Drop Characteristics in Rectangular Channels with Elliptic Pin Fins,” Heat and Fluid Flow 19 (1998) 245-250, which is hereby incorporated by reference. These extended surfaces can then be placed on the outside of the cylindrical cooling device in, for example, a staggered arrangement. Referring to
Accordingly, heat can be transferred between the hot compressed vapor refrigerant and an external fluid. In a specific embodiment, heat is transferred from the hot compressed vapor refrigerant to an ambient fluid, such as air or water, on the refrigerant side of the heat exchanger. This heat transfer can involve, for example, a simple flat plate, straight tubing, or a coil of tube that flows the condensing fluid by an air-cooled or liquid-cooled surface. In specific embodiments, condensing fluid can flow through a simple annulus or cylindrical design with a open path from top to bottom, through a series of straight ducts created within the annulus or cylinder, or through one or more spiral wound ducts created around the inside of the annulus or cylinder. The heat removal from the coil can also be calculated by q=hAΔT where q [W] is the heat removal, h[W/m2K] is the heat transfer coefficient, A[m2] is the surface area of the cooled surface, and ΔT [K] is the temperature difference between the cooled surface and the refrigerant. The temperature of the refrigerant can drop until it begins to condense, at which point it can remain at a constant temperature until the refrigerant is fully condensed into liquid.
In a specific embodiment, a condenser in accordance with the subject invention can incorporate one or more helical ducts created, for example, by a spiral wound wire tube 890 (shown in
where d is the diameter of the cylinder. Since Lchannel=f(P,n)=f(w,y,n), therefore, N=f(w,y,n), where n is the number of parallel channels wrapping around the cylinder such that refrigerant flows through each of the parallel channels, simultaneously, from the first end of the condenser to the second end of the condenser. Therefore, the length of the coil, assuming 1 mm thickness between passes, will be
Lcoil(w, y, n)=N(w, y, n)·(y+1 mm)·n
Lcoil(w, y, n) is set equal to the length of the condenser in order to maximize contact with the air cooled surface. Doing so and solving for w for varying values of y and n and setting a design limit of ΔP=1 psi, in a specific embodiment, the final design is found to be
for a cycle load of 200 W.
Further design parameters can take into account the pressure losses from refrigerant flowing through the helical channels. The pressure loss, ΔP, of the internal flow can be calculated to check that the design does not induce excessive inefficiencies to the thermodynamic cycle of the cooling device. Similarly to the heat transfer coefficient, ΔP can be a function of the flow conditions, the cross sectional geometry, and the length of the tube. Correlations to model the pressure loss may be found in McDonald, A. T., and Fox, R. W., Introduction to Fluid Mechanics, John Wiley and Sons, Inc. (2000), which is hereby incorporated herein by reference. Pressure loss can be reduced by reducing the length of the duct, since pressure loss and length can be directly proportional. The length of the duct may be reduced by dividing the flow into multiple ducts. In a specific embodiment, the number of ducts is one continuous channel. In a further embodiment, the number of ducts is 2 or more ducts flowing in parallel.
The fluid that the heat is rejected to can flow through the condenser due to the forces generated by, for example, wind, natural convection, fans, blowers, or compressors. In a specific embodiment, referring to
Cooling the components in this way can increase the performance efficiency of the subject cooling unit as compared with standard vapor compression cycles. The stand and cycle typically involves a compressor held within a housing. The compressor's inefficiency can add heat to the cycle, so as to lower the cooling capacity of the standard unit or necessitate an increase in the amount of power required to achieve a given cooling capacity. Referring to
Referring to
Cool high pressure liquid refrigerant can flow from the condenser 880 via exit port 830 (shown in
The subject evaporator can exchange heat between a coolant and the refrigerant. While the refrigerant passes through the evaporative heat exchanger, it can experience a phase change from liquid to vapor as it picks up heat from the coolant on the opposing side. This atypical heat exchanger can utilize non-traditional methods for predicting the performance of and designing such a device. The liquid side can adhere to well established heat transfer correlations, which suggest that the total heat transfer between two substances at different temperatures is equal to a heat transfer coefficient constant times the total area that it is acting on and the temperature gradient.
Heat transfer characterization and prediction on the refrigerant side, however, is more complicated due to the phase change process that occurs while the refrigerant is passing through the heat exchanger. Approximate correlations, which include experimental correction factors, have been recently determined and are discussed in detail in Carey, Van P., Liquid-Vapor Phase Change Phenomena, Taylor and Francis, New York (1992), which is hereby incorporated by reference. A specific embodiment of the subject invention can utilize a heat exchanger geometry which is based on correlation predictions from Carey (1992) that maximize the possible amount of heat transfer on the refrigerant side from the coolant on the other side.
Similar to the coolant side, however, the two phase heat transfer phenomenon is highly dependent upon the amount of area available for heat transfer to take place. In a specific embodiment, the design of the subject evaporative heat exchanger can, in general, maximize heat transfer area, while minimizing overall weight and dimensions and minimizing the liquid pressure drop through the heat exchanger. Preferably, the two fluids pass as close to each other as possible in order to minimize conduction heat transfer resistance through the separating medium. In a specific embodiment, a parallel channel configuration can be utilized. In a further specific embodiment, the parallel channel configuration can have a separation wall of 1 mm and can follow the path of an Archemedian spiral. An archemidian spiral is defined in a parametric coordinate system as:
x(t)=A·t·cos(B·t)
y(t)=A·t·sin(B·t)
where the constants A and B govern the number of spiral revolutions and the overall diameter of the geometry. One example yields a spiral path as is seen in
In a specific embodiment, the path for both fluids can begin on the outer edge of the cylinder and terminate in the center, where both fluids can exit perpendicular to the plane that they are flowing parallel on. Such a design can eliminate abrupt fluid turning points, thus minimizing pressure drop. Thin separation walls can be used to provide a sufficient length of, for example, approximately 25 inches within the limited area of the evaporator having a diameter of 53 mm. The channel depth can be chosen, using two-phase heat transfer correlations as a guide, to maximize the heat transfer area available for both fluids and meet the heat exchange rate requirements of the evaporator. In a further specific embodiment, a channel depth of about 8 mm can be used with an evaporator having 25 inch long fluid path with an evaporator diameter of 53 mm.
A specific embodiment of the subject compact vapor compression cooling system, shown in
It should be understood that the examples and embodiments described herein are for illustrative purposes only and that various modifications or changes in light thereof will be suggested to persons skilled in the art and are to be included within the spirit and purview of this application.
All patents, patent applications, provisional applications, and publications referred to or cited herein are incorporated by reference in their entirety, including all figures and tables, to the extent they are not inconsistent with the explicit teachings of this specification.
The present application is a divisional application of U.S. application Ser. No. 10/625,014, filed Jul. 22, 2003 now U.S. Pat. No. 7,010,936, which claims the benefit of U.S. Provisional Application Ser. No. 60/413,056, filed Sep. 24, 2002, both of which are hereby incorporated by reference herein in their entirety, including any figures, tables, or drawings.
| Number | Name | Date | Kind |
|---|---|---|---|
| 1896081 | Hampson | Feb 1933 | A |
| 1896953 | Hassell | Feb 1933 | A |
| 1974317 | Steenstrup | Sep 1934 | A |
| 2566865 | Wingerter | Sep 1951 | A |
| 2768508 | Guyton | Oct 1956 | A |
| 2920463 | Gould | Jan 1960 | A |
| 3200480 | Heuer | Aug 1965 | A |
| 3214087 | Luck | Oct 1965 | A |
| 3555848 | Johnson | Jan 1971 | A |
| 3926008 | Webber | Dec 1975 | A |
| 4287724 | Clark | Sep 1981 | A |
| 4300630 | Trojani | Nov 1981 | A |
| 4630669 | Kessler et al. | Dec 1986 | A |
| 5009262 | Halstead et al. | Apr 1991 | A |
| 5097897 | Watanabe et al. | Mar 1992 | A |
| 5178209 | Aoki et al. | Jan 1993 | A |
| 5372188 | Dudley et al. | Dec 1994 | A |
| 5377500 | Yang | Jan 1995 | A |
| 5511384 | Likitcheva | Apr 1996 | A |
| 6499534 | Tawney et al. | Dec 2002 | B1 |
| Number | Date | Country |
|---|---|---|
| 106 312 | Jan 1925 | CH |
| 160 351 | May 1905 | DE |
| 848 656 | Sep 1952 | DE |
| 971 287 | Jan 1951 | FR |
| 1 158 943 | Jun 1958 | FR |
| 1 590 923 | Jun 1981 | GB |
| 2 107 852 | May 1983 | GB |
| Number | Date | Country | |
|---|---|---|---|
| 20060150666 A1 | Jul 2006 | US |
| Number | Date | Country | |
|---|---|---|---|
| 60413056 | Sep 2002 | US |
| Number | Date | Country | |
|---|---|---|---|
| Parent | 10625014 | Jul 2003 | US |
| Child | 11343431 | US |
