Standing wave compressor

Abstract: A compressor for vapor-compression cooling systems, which exploits the properties of acoustic resonance in fluids for fluid compression, and provides a discharge pressure which can be varied during operation in response to changing operating conditions, thereby providing an oil-less compressor and reducing the compressor's energy consumption. The thermoacoustic properties of standing acoustic waves are exploited to provide a refrigerant subcooling system which is contained within the compressor. Refrigerant subcooling occurs when heat exchange is provided between the refrigerant and a heat pumping surface, which is exposed to the standing acoustic wave within the compressor. Acoustic energy can be provided by either a mechanical driver, or by direct exposure of the fluid to microwave and infrared energy, including solar energy. Inlets (4) and outlets (6) arranged along the chamber (2) provide for the intake and discharge of a fluid refrigerant, and can be provided with optional reed valve arrangements, so as to increase the compressor's compression ratio. The performance of the compressor can be optimised by a control circuit which holds the wavelength of the standing wave constant, by varying the driving frequency in response to changing operating conditions.

Description: This invention relates to apparatus for compressing and conveying fluids, and with regard to certain more specific features, to apparatus which are used as compressors in vapor-compression cooling equipment (also referred to as compression-evaporation cooling equipment). Heretofore, nearly all refrigeration and air-conditioning compressors which have found widespread and practical application, required many moving parts. Reciprocating, rotary, and centrifugal compressors, to name a few, all have numerous moving parts. Each of these compressors will consume a portion of energy which serves only to move its parts against their frictional forces, as well as to overcome their inertia. This energy is lost in overcoming the mechanical friction and inertia of the parts, and cannot contribute to the actual work of gas compression. Therefore, the compressor's efficiency suffers. Moving parts also reduce dependability and increase the cost of operation, since they are subject to mechanical failure and fatigue. Consequently, both the failure rate and the energy consumption of a compressor tend to increase as the number of moving parts increases. Typical refrigetation and air-conditioning compressors must use oils to reduce the friction and wear of moving parts. The presence of oils in contemporary compressors presents many disadvantages. Compressors which need oil for their operation will allow this oil to mix with the refrigerant. The circulation of this oil through the refrigeration cycle can lower the system's overall coefficient of performance in several ways, thus increasing the system's energy consumption. As such, the issue of oil-refrigerant mixtures places a restraint on ideal system design. Another disadvantage of oil-refrigerant mixtures relates to the development of refrigerants. Non-ozone depleting refrigerants must be developed to replace the CFC family of refrigerants. For a new refrigerant to be considered successful, it must be compatible with compressor oils. Oil compatibility is the subject of performance and toxicity tests which add long delays to the commercial release of new refrigerants. Hence, the presence of oils in refrigeration and air-conditioning compressors slows the development of new refrigetants. Mechanical compressors which are employed in compressionevaporation systems provide a fixed displacement, which is difficult to vary during operation. Thus, their discharge pressure is also difficult to vary. For vapor-compression systems, the compressor's discharge pressure must be high enough to provide condensation at the highest temperature of the condensing medium. As such, the design choise of the compressor's discharge pressure must be made on a worstcase basis. During periods when the condensing medium's temperature is below this worst-case temperature, the discharge pressure of the compressor is larger than the minimum pressure required for condensation to occur. Therefore, during normal operation conditions, energy is wasted by producing excessive discharge pressures. For example, a compressor's discharge pressure for a typical residential refrigerator might be designed to sustain condensation at room air temperatures of up to 100 degrees Fahrenheit. During periods when the room's air temperature is below 100 degrees Fahrenheit, a lower discharge pressure could sustain condensation. Thus, during periods of average room air temperatures, the compressor wastes energy by producing discharge pressures which are higher than necessary. Also, the selection of electric motors is made on this same worst case basis. The electric motor must be capable of startup and pulldown of a warm refrigerator, during periods of high room temperature. Consequently, a motor must be used whose power consumption is greater than the minimum required for normal operation. In short, any compression-evaporation system where the condensing medium's temperature changes, will suffer from these inefficiencies. These fixed discharge pressure considerations can also be applied to heat pumps and air-conditioners. During periods when the indoor-outdoor temperature difference is small, the minimum pressure differential needed is reduced. since mechanical compressors cannot easily vary their displacement, compression-evaporation systems are unable to exploit the increased efficiency of a variable discharge pressure. The design of mechanical compressors with variable displacement, has always led to the addition of many more moving parts. These extra moving parts decrease the compressor's efficiency and dependability. Consequently, the advantages offered by a variable discharge pressure, remain unexploited. In general, much effort has been exerted to design compressors which lack these traditional moving parts and their associated disadvantages. Some of these efforts have produced pumps which seek to operate on the pumped medium, using non-mechanical means. Typically, these pumps operate by pressurizing the pumped medium using heat, or by exciting the pumped medium by inertia-liquid-piston effects. Of particular interest is the inertia-liquid-piston type pump of US-A-3,743,446, which claims to provide a pump whose pumping action is due to the properties of standing acoustical waves. Although the above patent can provide a pumping action, it does not exploit certain modes of operation which can provide large pressure differentials and high efficiency. As such, this patent does not provide a practical compressor for vapor-compression systems, where large compression ratios and high efficiencies are required. Another example is shown in US-A-3,397,648. Therein is disclosed a chamber in which a gas is heated and subsequentially expelled through an egress check valve. As the chamber's remaining gas cools the resulting pressure differential causes more gas to be drawn into the chamber through an ingress check valve. This same method is employed in US-A-3,898,017. A further example of an acoustic cooling engine is described in US-A-4,722,201. Seldom have any of the above mentioned pumping methods been applied to the field of refrigeration and air-conditioning. One such attempt is seen in patent 2,050,391 to Spencer, 1936 August 11. In the Spencer patent, a chamber is provided in which a gaseous refrigerant is heated by spark discharge, and subsequently expelled through an egress check valve, due to the resulting pressure increase. As the chamber's remaining gas cools, the resulting pressure differential causes more gas to be drawn into the chamber through an ingress check valve. This approach results in ionization of the refrigerant, and could cause highly undesirable chemical reactions within the refrigeration equipment. For a practical refrigeration system, such chemical reactions would be quite unsatisfactory. It is apparent that oil-free refrigeration and air-conditioning compressors, which can provide variable discharge pressure and require few moving parts, have not been satisfactorily developed. Furthermore, it is apparent that the application of acoustic energy to compress gases for vapor-compression cooling systems, has not been achieved. If such compressors were available, they could simplify the development of new refrigerants, and offer improved dependability and efficiency, thereby promoting energy conservation. In accordance with a first aspect of the present invention, we provide a method of compressing a fluid refrigerant comprising introducing a fluid refrigerant to be compressed into a chamber having at least one inlet and at least one outlet, characterized in that said method further comprises establishing a standing wave in the fluid refrigerant in the chamber, so that the fluid refrigerant is compressed. In accordance with a second aspect of the present invention, we provide a compression-evaporation system comprising means for subjecting a compressed refrigerant to a heat exchange operation, characterized in that said compression-evaporation system further comprises a standing wave compressor for compressing a refrigerant. In accordance with a third aspect of the present invention, we provide a refrigerant compressor comprising a chamber for holding a fluid refrigerant to be compressed, said chamber including at least one inlet and at least one outlet, characterised in that said compressor further comprises: driving means for establishing a travelling wave in the fluid refrigerant in said chamber, said driving means and said chamber converting the travelling wave into a standing wave in the fluid refrigerant in said chamber, so that the fluid refrigerant is compressed. The present invention typically provides a compressor, for vapor-compression cooling systems, which uses acoustic energy for fluid compression, and does so by exploiting the properties of acoustic resonance in fluids. Typically, the present invention provides a compressor with a discharge pressure, that can be varied during operation in response to a vapor-compression system's operating conditions, thereby increasing the system's efficiency by reducing the compressor's energy consumption. Typically, the present invention provides a non-mechanical acoustical driver, Which exploits a fluid's absorption of electromagnetic enerny, and provides an oil-less compressor. Typically, the present invention provides a compressor which can subcool a refrigerant by employing the thermoacoustic properties of standing acoustic waves, thereby providing subcooling of the refrigerant without the addition of any moving parts. The present invention is directed to a refrigerant compressor which exploits the properties of acoustic resonance in fluids for fluid compression, and provides a discharge pressure which can be varied during operation. The variable dicharge pressure is regulated by a control circuit which varies the power of an acoustic driver as a function of changes in the operating conditions of a vapor-compression system. In another aspect, the present invention is directed to a refrigerant compressor for compressing a refrigerant fluid by creating a standing accoustic wave. The standing wave creates a temperature differential along a resonant chamber, so that a first portion of the chamber is at a temperature wich is higher than a second portion of the chamber. A heat exchanger is coupled to the chamber adjacent the second portion of the chamber so that the heat exchanger provides thermal contact between the refrigerant and the second portion of the chamber. Within the heat exchanger, the refrigerant can be sub-cooled before being to the evaporator, thereby enhancing cooling efficiency. The subcooling capacity can be further enhanced by providing heat pumping surfaces within the chamber. The heat pumping surfaces are exposed to the standing acoustic wave, so that a temperature differential is created along the heat pumping surface. According to a fourth aspect of the present invention, we provide a refrigerant compressor characterised in that said refrigerant compressor comprises a standing wave compressor which receives, acoustically compresses, and discharges a refrigerant, said standing wave compressor including an acoustic chamber having at least first and second different cross sectional areas at at least first and second positions, respectively, along said acoustic chamber. These and other objects and advantages of the invention will become apparent from the accompanying specifications and drawings, wherein like reference numerals refer to like parts throughout. Fig. 1 is a partly schematic, partly sectional view of a mechanically driven embodiment of the present invention; Fig. 2 is an embodiment which is functionally the same as the embodiment of Fig. 1, but has a minimum of pressure nodes and antinodes; Fig. 3 is an embodiment which is functionally the same as the embodiment of Fig. 1, but provides additional pressure nodes and antinodes as well as additional inlets and outlets; Fig. 4 is an embodiment that reduces the total number of outlet check valves needed for a full-wave discharge cycle to a maximum of two; Fig. 5 is an embodiment of the present invention which limits the number of outlet check valves needed for a half-wave discharge cycle to one; Fig. 6 is an embodiment of the invention, which locates both inlets and outlets at the pressure antinodes; Fig. 7 is an embodiment that reduces the total number of inlet and outlet check valves needed for a full-wave suction and discharge cycle to a maximum of four; Fig. 8 is an embodiment that reduces the total number of inlet and outlet check valves needed for a half-wave suction and discharge cycle to a maximum of two; Fig. 9 is an amplitude vrs. time plot, Which illustrates the demodulation of high frequency ultrasonic energy into lower frequency pulses; Fig. 10 is a valveless embodiment of the invention, which provides an ultrasonic driver; Fig. 11 is one type of check valve which could be used by the invention; Fig. 12 is a control circuit which can be used to maintain the proper driving frequency under changing conditions ; Fig. 13 is an acoustic chamber having a varying cross-sectional area, which provides greater pressure differentials and suppresses unwanted resonant acoustic modes; Fig. 14 is an embodiment of the present invention which provides as the driving means, a standing microwave being spatially coincident with the standing acoustic wave; Fig. 15 is another microwave driving arrangement which provides a standing microwave being spatially coincident with the standing acoustic wave; Fig. 16 is a graph which shows an electric energy distribution curve in relation to the chambers of Fig. 14 and Fig. 15; Fig. 17 is another method for using microwave energy as a means for driving the standing acoustic wave; Fig. 18 is another method for using microwave energy as a means for driving the standing acoustic wave; Fig. 19 is another embodiment of the present invention which provides a LASER as a means for maintaining a standing acoustic wave; Fig. 20 illustrates the present invention as part of a typical vapor-compression cooling system; Fig. 21 is a block diagram of a control circuit which maintains the minimal discharge pressure needed for condensation to occur, and also maintains the proper driving frequency under changing conditions; Fig. 22 is a perspective view of the present invention as part of a vapor-compression system having a refrigerant subcooling system; Fig. 23 is a section on line 3-3 of Fig. 22; Fig. 24 is a sectional view of Fig. 22, which provides a detailed view of the heat pump plate stack; Fig. 25 is a microwave driven version of the embodiment of Fig. 23. Embodiments Having Valves Fig. 1 illustrates an embodiment of the present invention. A chamber 2 is provided which has an inlet 4 and an outlet 6. Outlet 6 has a check valve 8 attached thereto, such that any gas/liquid (hereinafter called medium) passing through outlet 6 must also pass through check valve 8 in order to reach channel 36. Check valve 8 allows flow out of but not into the chamber 2. Forming one wall of chamber 2 is a driver 10 comprising a flexible diaphragm 16, which has a coil 22 attached thereto. Coil 22 encircles the end of a stationary cylindrical magnet 18. Cylindrical magnet 18 is press fitted into the body 20 of driver 10. Coil 22 of driver 10 is energized through wires 14 by a generator 12, such as an oscillating circuit. Driver 10 or Fig. 1 provides ease of illustration for the present invention, and is arbitrarily depicted as a low-impedance driver (i.e. small force, large displacement), thus being located near a pressure node. In general, the materials, construction, and placement of driver 10 will depend on the requirements of a particular application. Further details on drivers are given in the section "Drivers" below. In operation, generator 12 causes coil 22 to be driven by a periodic waveform of predetermined frequency, which in turn sets up an oscillating magnetic field about coil 22. Due to the alternating polarity of this oscillating field, the coil-diaphragm assembly is alternately repulsed and attracted by the cylindrical magnet 18. Thus, diaphragm 16 vibrates at a predetermined frequency which causes a traveling wave 26 to be generated in the medium in chamber 2. When this traveling wave 26 hits far wall 30 of chamber 2, it is reflected back as wave 28, 180° out of phase with the initial wave. If the length of chamber 2 is made to be equal to an integer times a quarter wavelength of the traveling wave in the medium (i.e. ¼nℓ, where ℓ is the wavelength, and n is an integer), then chamber 2 will act as a resonant cavity and will have a standing wave pattern set up in it. This resonance condition allows the pressure amplitude of the standing acoustic wave to rise to the levels needed for large compression ratios. Thus, a standing wave pattern is set up in chamber 2, which has pressure antinodes, or displacement nodes, at end wall 30 and at point 34, and pressure nodes, or displacement antinodes, at diaphragm 16 and at point 32. The placement of inlet 4 and outlet 6 is described as follows. Outlet 6 is located at pressure antinode 34. The pressure at pressure antinode 34 oscillates above and below the undisturbed pressure of the medium. Also, at high acoustic amplitudes, the average pressure at the pressure antinode can rise above the undisturbed pressure of the medium, due to nonlinearities. Inlet 4 is located at pressure node 32. At high acoustic amplitudes, the minimum pressure which exists at pressure node 32, can be less than the undisturbed pressure of the medium, due to nonlinearities. Check valve 8 provides a rectification of the oscillating pressure at pressure antinode 34. When the pressure at antinode 34 reaches a predetermined value, which is higher than the undisturbed pressure of the medium, check valve 8 opens. Thus, some of the medium is allowed to escape chamber 2 by passing in turn through outlet 6, check valve 8, and then into channel 36. When the pressure at antinode 34 drops below the predetermined value, checkvalve 8 closes and prevents the discharged medium from flowing back into chamber 2. Resultingly, the quantity of medium in chamber 2 is continually reduced, and the pressure at node 32 drops even lower than its normal minimum value, which in turn causes additional medium to be drawn through inlet 4 into chamber 2. Thus, when the medium in chamber 2 is excited by the action of driver 10 and a standing wave pattern is set up therein consisting of pressure nodes and antinodes, some of the medium inside chamber 2 at antinode 34 will be periodically forced out of chamber 2, due in part to check valve 8's rectification of the oscillating pressure at outlet 6. In addition, the medium immediately outside chamber 2 at inlet 4 will be drawn into chamber 2. In this way, the embodiment of Fig. 1 compresses and conveys the medium introduced at inlet 4. It should be noted that none of the embodiments of the present invention are limited to a chamber of only one length. Thus, for a given wavelength ℓ and assuming the proper impedance of driver 10, the length of chamber 2 in Fig. 1 can be any length which equals nℓ/4, so that chamber 2 is not limited to the length 3ℓ/4. For instance, the embodiment of Fig. 1 could be reduced to that of Fig. 2, and still function in exactly the same manner. Fig. 2 shows an embodiment for the case of n = 1, which gives a chamber length of nℓ/4 = ℓ/4. In this case there is only one pressure node and one pressure antinode, located at the diaphragm 16 and end wall 30 respectively. The location of inlet 4 is coincident with said node, and the location of outlet 6 is coincident with said antinode. In short, there are any number of possible chambers with lengths that are greater than ℓ/4. Fig. 3 shows an embodiment of the invention which provides a chamber 2 having multiple inlets 4a, 4b, 4c and multiple outlets 6a, 6b, 6c. Channel 40 has inlets 4a, 4b, 4c all attached thereto by respective conduits 5a, 5b, 5c, such that any medium passing from inlets 4a, 4b, 4c into chamber 2, must first pass through channel 40. Outlets 6a, 6b, 6c have check valves 8a, 8b, 8c attached respectively thereto, and said checkvalves are attached to channel 36 by respective conduits 3a, 3b, 3c, such that any medium passing through the outlets 6a, 6b, 6c must also pass through respective check valves 8a, 8b, 8c in order to reach channel 36. Check valves 8a, 8b, 8c allow flow out of but not into the chamber 2. Forming one wall of the chamber 2 is driver 10, said driver being the same in form and function as driver 10 of Fig. 1. Driver 10 is energized by a generator 12, such as an oscillating circuit. The embodiment of Fig. 3 operates in exactly the same manner and according to the same theory and principles as the embodiment of Fig. 1. This can be seen by realizing that the acoustic processes which occur between the single inlet 4 and checkvalve 8 of Fig. 1, can also occur between multiple inlets 4a, 4b, 4c and multiple checkvalves 8a, 8b, 8c of Fig. 3. The number of inlets in Fig. 3 could be reduced to one if so desired. In Fig. 4 an embodiment of the invention is shown, which limits The number of outlet check valves needed to two, regardless of the number of outlets. In general, each consecutive pressure antinode is 180° out of pressure-phase with its neighboring pressure antinodes. For example, if antinode n has pressure +P, then antinode n+1 has pressure -P, and antinode n+2 has pressure +P, and so on. In other words, if at a certain time "t" a given antinode's pressure is high, then at that same instant its neighboring antinode's pressure will be low, and the next will be high, and so on. Consequently, since only two pressure-phases exist, all outlets of one phase can be routed through one check valve, and all outlets of the other phase can be routed through another check valve. Fig. 4 shows channel 40 with inlets 4a, 4b, 4c, 4d all attached thereto by respective conduits 5a, 5b, 5c, 5d such that any medium passing from inlets 4a, 4b, 4c, 4d into chamber 2, must first pass through channel 40. Outlets 6a and 6c are attached by respective conduits 3a and 3c to check valve 8b, such that any medium passing through outlets 6a and 6c must also pass through check valve 8b in order to reach channel 36. Outlets 6b and 6d are attached by respective conduits 3b and 3d to check valve 8a, such that any medium passing through outlets 6b and 6d must also pass through check valve 8a in order to reach channel 36. This arrangement can be extended to any number of outlets, such that two check valves will be sufficient regardless of the number of outlets, as long as the two groups of like-pressure-phase outlets are routed through their two respective check valves. This matching of like-pressure-phase outlets is necessary, because if two or more outlets of unlike-pressure-phase were connected together, the medium would tend to flow back and forth between the alternating high and low pressure outlets. Thus, the medium would be allowed to shunt the outlet check valve and reenter the chamber, so that no discharge would occur. With the exception of this new outlet check valve arrangement, the embodiment of Fig. 4 operates in the same manner and according to the same theory and principles as the embodiment of Fig. 3. The number of inlets in Fig. 4 could be reduced to one if so desired. In Fig. 5 an embodiment of the invention is shown, which limits the number of outlet check valves needed to one, regardless of the number of outlets. Channel 40 has inlets 4a and 4b attached thereto by respective conduits 5a and 5b such that any medium passing from inlets 4a and 4b into chamber 2, must first pass through channel 40. Outlets 6a and 6b are attached by respective conduits 3a and 3b to check valve 8, such that any medium passing through outlets 6a and 6b must also pass through check valve 8 in order to reach channel 36. This grouping of outlets through a single check valve, is again due to the matching of like-pressure-phase antinodes. This arrangement can be extended to any number of outlets, such that one check valve will be sufficient regardless of the number of outlets, as long as like-pressure-phase outlets are routed through a single check valve. With the exception of this new outlet check valve arrangement, the embodiment of Fig. 5 operates in the same manner and according to the same theory and principles as the embodiment of Fig. 3. The number of inlets in Fig. 5 could be reduced to one if so desired. The embodiments of Fig. 3 and Fig. 4 will discharge the medium twice in one period of the standing wave. This full-wave discharge is due to the fact that the outlets are connected to pressure antinodes of both pressure phases. The embodiments of Figs.1, 2, and 5 will discharge the medium once in one period of the standing wave. This half-wave discharge is due to the fact that the outlets are connected to pressure antinodes of only one pressure phase. Fig. 6 shows an embodiment of the invention which relocates the inlets to the pressure antinodes. A chamber 2 has multiple inlets 4a, 4b, 4c and multiple outlets 6a, 6b, 6c. Outlets 6a, 6b, 6c have check valves 8a, 8b, 8c attached respectively thereto, and said checkvalves are attached by respective conduits 3a, 3b, 3c to channel 36, such that any medium passing through the outlets 6a, 6b, 6c must also pass through respective check valves 8a, 8b, 8c in order to reach channel 36. Inlets 4a, 4b, 4c have check valves 38a, 38b, 38c attached respectively thereto, and said checkvalves are attached by respective conduits 5a, 5b, 5c to channel 40, such that any medium passing into channel 40, must first pass through respective check valves 38a, 38b, 38c in order to reach respective inlets 4a, 4b, 4c. Check valves 38a, 38b, 38c allow flow into but not out of the chamber 2. Check valves 8a, 8b, 8c allow flow out of but not into the chamber 2. Forming one wall of the chamber 2 is driver 10, said driver being the same in form and function as driver 10 of Fig. 1. Driver 10 is energized by a generator 12, such as an oscillating circuit. In operation, driver 10 of Fig. 6 maintains a standing wave of given wavelength "ℓ" in the chamber 2, resulting in multiple pressure nodes 32a, 32b, 32c and antinodes 34a, 34b, 34c. Inlets 4a, 4b, 4c and outlets 6a, 6b, 6c are all coincident with respective pressure antinodes 34a, 34b, 34c. When the pressure at any one of the antinodes 34a, 34b, 34c reaches a predetermined value, which is higher than the undisturbed pressure of the medium, its corresponding inlet check valve closes, and its corresponding outlet check valve opens. Hence, when the pressure of an antinode goes high, the medium is prevented from leaving the chamber 2 through that antinode's inlet, but is allowed to flow out of the chamber 2 by passing through that antinode's outlet, then through its outlet checkvalve, and then through channel 36. When the pressure at any one of antinodes 34a, 34b, 34c drops below a predetermined value, which is lower than the undisturbed pressure of the medium, its corresponding inlet check valve opens, and its corresponding outlet check valve closes. Hence, when the pressure of an antinode goes low, the medium is prevented from reentering the chamber 2 through that antinode's outlet, but is allowed to flow into chamber 2 by passing first through channel 40, then through the antinode's inlet check valve, and then through its inlet into chamber 2. Thus, when the medium in chamber 2 is excited by the action of driver 10, a standing wave pattern is set up therein consisting of pressure nodes and antinodes. As a result, the medium at pressure antinodes 34a, 34b, 34c will be periodically forced out of chamber 2 due to check valve's 8a, 8b, 8c rectification of the oscillating pressure at the outlets 6a, 6b, 6c. In addition, the medium outside chamber 2 at channel 40 will be periodically drawn into chamber 2 due to check valve's 38a, 38b, 38c rectification of the oscillating pressure at the inlets 4a, 4b, 4c. Thus, the embodiment of Fig. 6 compresses and conveys the medium introduced at channel 40. The number of inlets and outlets in Fig. 6 could be reduced to one each, or extended to many. In Fig. 7 an embodiment of the invention is shown which limits the number of inlet check valves needed to two, and the number of outlet check valves needed to two, regardless of the number of inlets and outlets. Fig. 7 shows outlets 6a and 6c attached by respective conduits 3a and 3c to check valve 8b, such that any medium passing through outlets 6a and 6c must also pass through check valve 8b in order to reach channel 36. Outlets 6b and 6d are attached by respective conduits 3b and 3d to check valve 8a, such that any medium passing through outlets 6b and 6d must also pass through check valve 8a in order to reach channel 36. Inlets 4a and 4c are attached by respective conduits 5a and 5c to check valve 38a, such that any medium passing through channel 40, must pass first through check valve 38a in order to reach inlets 4a and 4c. Inlets 4b and 4d are attached by respective conduits 5b and 5d to check valve 38b , such that any medium passing through channel 40, must pass first through check valve 38b in order to reach inlets 4b and 4d. This grouping of inlets and outlets with their respective check valves, is again due to the matching of like-pressure-phase antinodes. The arrangement of Fig. 7 can be extended to any number of inlets and outlets, such that only two inlet check valves and two outlet check valves will be sufficient regardless of the number of inlets and outlets, as long as the two groups of like-pressure-phase outlets and the two groups of like-pressure-phase inlets are routed through their four respective check valve. With the exception of this new inlet and outlet check valve arrangement, the embodiment of Fig. 7 operates in the same manner and according to the same theory and principles as the embodiment of Fig. 6. In Fig. 8 an embodiment of the invention is shown which limits the number of inlet check valves needed to one, and number of outlet check valves needed to one, regardless of the number of inlets and outlets. Fig. 8 shows outlets 6a and 6b attached by respective conduits 3a and 3b to check valve 8, such that any medium passing through outlets 6a and 6b must also pass through check valve 8 in order to reach channel 36. Inlets 4a and 4b are attached by respective conduits 5a and 5b to check valve 38, such that any medium passing through channel 40, must pass first through check valve 38 in order to reach inlets 4a and 4b. This grouping of inlets and outlets with their respective check valves, is again due to the matching of like-pressure-phase antinodes. In Fig. 8, the inlets and outlets are located at different like-pressure-phase antinodes, but the inlets and outlets could also be located at the same like-pressure-phase antinodes. This arrangement can be extended to any number of inlets and outlets, such that one inlet check valve and one outlet check valve will be sufficient regardless of the number of inlets and outlets, as long as the like-pressure-phase outlets and the like-pressure-phase inlets are routed through their two respective check valves. With the exception of this new inlet and outlet check valve arrangement, the embodiment of Fig. 8 operates in the same manner and according to the same theory and principles as the embodiment of Fig. 6. The embodiments of Fig. 6 and Fig. 7 will draw in medium twice during one period of the standing wave, and will also discharge the medium twice in one period of the standing wave. This full-wave intake and discharge is due to the fact that the inlets and outlets are connected to pressure antinodes of both pressure phases. The embodiment of Fig. 8 will draw in medium once during one period of the standing wave, and will also discharge the medium once in one period of the standing wave. This half-wave intake and discharge is due to the fact that the inlets are connected to pressure antinodes of only one pressure phase and the outlets are connected to pressure antinodes of only one pressure phase. Mechanically Driven Embodiments Without Valves It has long been known, that a standing acoustical wave in a chamber can produce a discernable pressure differential between nodes and antinodes, without the use of valves. Kundt's tube, which uses this effect to measure acoustic wavelengths, has been used since the early 19th century. However, this valveless arrangement would not appear to be a candidate as a refrigeration compressor. As mentioned previously, the pressure differential which occurs between pressure nodes and antinodes is a nonlinear effect. As such, the magnitude of this node-antinode pressure differential, relative to the peak pressure amplitude, becomes increasingly large at higher acoustic pressures amplitudes. At the large acoustic pressure amplitudes of the present invention, this node-antinode pressure differential can provide a practical source of gaseous compression for vapor-compression cooling systems. For vapor-compression systems which do not require high compresssion ratios, this valveless embodiment offers the elimination of all moving parts, except for the driver. For applications which require higher compression ratios, the valved embodiments can be employed. Fig. 13 shows a valveless embodiment of the present invention. A chamber 73 is comprised of a variable cross section segment 74, a variable cross section segment 75, and a cylindrical center section 71. Chamber 73 is terminated by a discharge plate 76 and a discharge chamber 77. Discharge plate 76 is sandwiched between acoustic chamber 73 and discharge chamber 77, and held together by common flange bolts. A multiplicity of discharge ports 78 are drilled through discharge plate 76. A multiplicity of suction ports 79 are drilled through center section 71. Suction chamber 80 forms an outer chamber around suction ports 79. The use of variable cross section segments is discussed in the section "Chambers" below. An acoustic driver 84 applies acoustic energy to the fluid within chamber 73, at the proper frequency to establish acoustic wave 81 in chamber 73. Acoustic wave 81 represents the first resonant mode of chamber 73, which is a half wavelength standing wave. A pressure node exists at center section 71, and pressure antinodes exist at discharge plate 76 and driver 84. Due to the nonlinearities of large amplitude standing acoustic waves, a low average pressure will exist at a pressure node and a high average pressure will exist at a pressure antinode. Thus, center section 71 will experience a low average pressure, and discharge plate 76 will experience a high average pressure. Once the acoustic wave 81 is established in acoustic chamber 73, gaseous refrigerant is drawn in turn through suction tube 82, into suction chamber 80, through suction ports 79, and into acoustic chamber 73. Having been acoustically compressed by acoustic wave 81, the gaseous refrigerant escapes in turn through discharge ports 78, into discharge chamber 77, and through discharge tube 83. In terms of efficiency, the ratio of the node-antinode pressure to the peak-to peak acoustic pressure, increases as the acoustic pressure amplitude is increased. Consequently, the valveless embodiment's efficiency improves as it is driven further into the nonlinear region (i.e. higher pressure amplitudes). There will of course be a practical limit, where dissipative forces will offset further efficiency gains. This behavior is most advantageous for compressor applications, due to the requisite high acoustic pressure amplitudes. Acoustic Pressure Amplitude The ability of the present invention to provide significant compression ratios, depends primarily on obtaining large acoustic pressure amplitudes. In order in obtain large acoustic pressure amplitudes, it is necessary to have a high degree of resonance in the acoustic chamber. As with most resonant systems, the degree, or quality, of resonance can be designated by the parameter "Q." In the present invention, Q will be a function of both the fluid and the geometry of the acoustic chamber. The thermoacoustic properties of many refrigerants are favorable for providing large Q values. Consequently, large acoustic pressure amplitudes can be achieved in the present invention for vapor-compression applications. Drivers Driver 10 of Fig. 1 provides ease of illustration for the present invention. In general, there are a number of different drivers which can be used. The preferred driver for refrigeration applications is commonly referred to as a "linear motor". Such devices work along the same principles as electric motors, except that the motion is one dimensional rather than rotational. Typically, a moving piston is driven back and forth by an oscillating magnetic field. The piston can be a "free piston" which actually floats on a thin cushion of gas between the piston and the chamber wall. For the present invention, this layer of gas would consist of the working fluid. Due to this gas bearing, no contact occurs between the chamber wall and the piston, thus no lubricating oil is required. Linear motors have been designed with efficiencies up to 95%. Examples of linear motors can be seen in U.S. patent 4,602,174 to Sunpower July 22, 1986, and U.S. patent 4,924,675 to Helix Technology Corporation May 15,1990. In some applications it may be desireable to use ultrasonic sources. An ultrasonic driver can be used in a nonresonant or resonant mode. In the nonresonant mode, the frequency of the driver is higher than the frequency of the acoustic resonance, and transduction depends on the acoustic absorption of the fluid. In this mode, the ultrasonic driver operates at a frequency which is much higher than the frequency of the acoustic resonance, and is pulsed off and on at a repetition rate equal to the frequency of the acoustic resonance. As the driver is switched rapidly off and on, a succession of short pulses is created; each pulse consisting of a short train of high frequency oscillations. Fig. 9 shows the acoustic waveform of a single "high frequency pulse", just after it leaves the driver. After traversing a short distance through the medium, the "high frequency pulse" evolves into the "demodulated pulse". This demodulation occurs when the high frequency acoustic waves are absorbed by the medium, leaving only pulses behind. The desired mode of the standing acoustic wave can be excited by the demodulated pulses. One or more ultrasonic drivers could be placed in contact with the gas at one or more pressure antinodes. As an alternative to pulsing, the output of the ultrasonic driver could be modulated by a lower frequency waveform. Thus a standing acoustical wave could be excited whose frequency would be equal to the modulating frequency, since one positive demodulated pulse is produced per period of the modulating waveform. In the resonant mode, the ultrasonic driver's frequency is equal to the acoustic resonance, and provides a continuous acoustic output. Fig. 10 shows a valveless embodiment of the present invention which operates in the resonant ultrasonic mode. A chamber 134 is provided whose one end is terminated by end wall 152, and whose other end is provided with a chamber flange 140. The length of chamber 134 in Fig. 10 is exaggerated, and would be much shorter at ultrasonic wavelengths. Inlet 136 is located at a pressure node, and outlet 138 is located at a pressure antinode. End flange 142 is fastened to chamber flange 140 by common flange bolts 154. Ultrasonic driver 148 has ultrasonic horn 146 attached thereto. Horn flange 144 is affixed to a nodal plane of ultrasonic horn 146, and horn flange 144 is sandwiched between end flange 142 and chamber flange 140. Coaxial cable 150 supplies R.F. energy to ultrasonic driver 148. An ultrasonic driver and flanged horn arrangement, as shown in Fig. 10, are available from Sonic Systems, Inc. Newtown, PA. In operation, ultrasonic driver 148 and ultrasonic horn 146 create a high pressure ultrasonic wave which propagates through the gas in chamber 134. Said ultrasonic wave is reflected from end wall 152. As described in previous embodiments, the frequency of ultrasonic driver 148 and the length of chamber 134 are chosen so that a standing acoustic wave is established as shown in Fig. 10. Due to the nonlinear effects described above, a pressure differential will be established between pressure nodes and pressure antinodes. Consequently, low pressure gas will be drawn in at inlet 136 and high pressure gas will be discharged at outlet 138. It should be noted that any number of inlets and outlets could be used in Fig. 10. Also, like-pressure-phase matching is not essential for valveless embodiments, since operation depends more on average pressures, than on instantaneous pressures. A further discussion of driver types and methods is disclosed in U.S. patent application Ser. No. 07/493,380, filed March 14, 1990, the contents of which is hereby incorporated by reference, and is not reproduced herein. The above discussion of drivers will suggest many other ways to design efficient high power acoustic drivers. This discussion of specific drivers is not intended as a limit on the scope of the invention, but rather to indicate the variety of acoustic drivers which can be employed by the present invention. Valve Types As described above, some of the embodiments of the present invention employ check valves. It is understood that the term "check valve" refers to a function rather than to a specific type of valve. Many different types of these rectifying components could be used; the exact choice of which depends on the particular design requirements of a given application. In a practical system operating in the low to sub kHz acoustic range, reed valves can be employed. Reed valves are commonly used on reciprocating type compressors. Reed valve assemblies, complete with suction and discharge reed valves, are typically sandwiched between the cylinder and head of a reciprocating compressor. A reedvalve assembly, as used on the present invention, is shown in Fig. 20, and described in the section "Description of Refrigeration and Air-conditioning Applications" below. Care must be taken to make the suction and discharge openings small compared to the total area of the end wall, so as not to disturb the chamber's resonance. Another possibility is illustrated in Fig. 11 which shows a series connected restrictive orifice valve 151. This valve will provide a greater resistance to flow in one direction than in the other. Since the pressure at a pressure antinode is oscillating, the resulting oscillatory flow could be rectified by this orifice valve, thus giving a net flow in one direction. In some applications, it may be desireable to drive a valved embodiment of the present invention at an acoustic frequency which is higher than the response time of most standard valves. In such a case, the compressor's performance would suffer if the valves could not open fast enough to allow the medium to pass through. The orifice valve offers one solution to this problem. Another solution would be to employ an actuated valve, which would open and close in response to an electrical signal. These activated valves would be operated by a control circuit, which would maintain a constant synchronization with the pressure oscillations of the standing wave. Actuated valves could be made to open once per cycle, or once during a plurality of cycles. Such a valve could be activated by a piezoelectric element, which could provide high speed operation. Many other rectifying components will suggest themselves to one skilled in the art. Electronic Controls In all of the mechanically driven embodiments of the present invention, an automatic frequency control of the driving system will assure optimal performance under changing conditions. An acoustic wave's velocity through a fluid, changes as a function of conditions such as temperature and pressure. As seen from the relationship ℓ = v/ƒ, if the velocity "v" of the wave changes, then the frequency "ƒ" can be changed to keep the wavelength "ℓ" constant. As described previously, there are certain preferred alignments between the standing wave's position and the inlets and outlets, which result in the optimal performance of the present invention. To preserve these alignments during operation, the wavelength must be held constant by varying the frequency in response to changing conditions inside the compressor. Fig. 12 illustrates an exemplary circuit, which could be used to maintain the required wavelength of the standing wave. In operation, the voltage drop across driver 10 at point 70 is rectified and integrated by integrator 62. This integrated voltage is then compared with the reference voltage at point 72 by comparator 64. The output voltage of comparator 64, serves to adjust the frequency of voltage controlled oscillator 66. Driver 10 is driven by amplifier 68, which amplifies the output of voltage controlled oscillator 66. For a given compression ratio, the voltage drop across driver 10 will be a minimum when the frequency of driver 10 is equal to the resonance frequency of chamber 2. This voltage minimum is due to a minimum in the displacement of driver 10 at resonance, which minimizes the back emf of driver 10. If the resonance frequency of chamber 2 begins to change, then the back emf of driver 10 changes. This change is seen at comparator 64 and causes the comparator's output voltage to change, thus shifting the voltage controlled oscillator's frequency back to the resonance of chamber 2. Many other control circuits could be designed by those skilled in the art. A phase-locked-loop circuit could be provided which would compare the phase of a pressure signal from within chamber 2, with the waveform of driver 10. Alternatively, a microprocessor could monitor various system parameters in order to keep the driver's frequency locked to the resonance of chamber 2. Control systems can also be adapted to the electromagnetically driven embodiments described below, In this case, the pulse repetition rate or the modulation frequency could be varied in response to changes in the chamber's resonance. Inlet and Outlet Placement The following considerations are pointed out, concerning the various inlet/outlet arrangements of the present invention. It is clear that the points of highest obtainable pressure in the chamber, for valved or valveless arrangements, will be the pressure antinodes, which includes the end walls. As such it is desirable to place both valved and valveless outlets at these positions. It is also clear that the point of lowest pressure in the chamber, for valveless arrangements, will be the pressure nodes. As such it is desirable to place valveless inlets at these points. For valved inlets, a lower pressure may be obtained at the pressure antinodes, including the end walls. Thus, the pressure nodes and antinodes provide ideal locations for inlets and outlets. However, it is understood that the invention is not limited to a precise placement of inlets and outlets with respect to the pressure nodes and antinodes. Many valve and inlet/outlet arrangements have been described above which make efficient use of the pressure effects associated with standing acoustic waves. These pressure effects are minimized or maximized at the pressure nodes and antinodes, but do not exist only at the pressure nodes and antinodes. Rather they can exist, although at reduced levels, at points removed from the pressure nodes and antinodes. In fact, any number of intermediate positions for inlets and outlets are possible. Although these intermediate positions can result in reduced pressure differentials and efficiencies, they can still provide an operable form of the present invention. Since both inlets and outlets can be operably moved to many intermediate locations, the exact location of inlets and outlets is not intended as a limitation on the scope of the present invention. For all of the valved embodiments, attention must be given to conduit lengths, if valves are to be located some distance from the chamber 2. It is pressure pulses which travel in these conduits. For optimal performance, these pulses should arrive at any common check valve at the same instant. Therefore, conduit lengths should be matched to this end. Chambers Chambers which provide a varying cross sectional area, offer certain advantages. Fig. 13 shows an acoustic chamber 73 with a varying cross sectional area, the construction of which was described in the section "Mechanically Driven Embodiments Without Valves" above. Acoustic chamber 73 offers the following three advantages. First, by properly designing the relative lengths of variable cross section segment 74, variable cross section segment 75, and center section 71, unwanted higher ordered resonant modes can be suppressed. These higher modes can diminish the desirable characteristics of the standing wave, thereby reducing the efficiency of the present invention. Secondly, acoustic chamber 73 provides a higher pressure differential between suction and discharge ports, than the pressure differential of a standard cylindrical chamber. This is due to the venturi effect produced by the varying cross sectional area of acoustic chamber 73. Thirdly, by providing a multiplicity of small diameter suction and discharge ports, turbulence is reduced. Larger ports would tend to create turbulence which dissipates acoustic energy, thereby reducing efficiency. The valveless chamber of Fig. 13 can be easily converted to the valved chamber of Fig. 20, by eliminating suction ports 79, and providing discharge plate 76 with a suction reed valve 131, a discharge reed valve 133, a suction plenum 137, and a discharge plenum 135. Many variations on the acoustic chamber shown in Fig. 13 are possible, and can provide these same advantages. A varying cross section can be achieved by providing a cylindrical chamber with inserts, rather than actually machining the chamber itself. However, a varying cross section is the common feature which allows any such chamber to attenuate unwanted resonant modes. Accordingly, it is the use of a varying cross sectional area, rather than any specific design features, which is the subject of this chamber design. Electromagnetically Driven Embodiments Figs. 14, 15, 17, 18, and 19 illustrate several arrangements for driving the standing acoustical wave with electromagnetic energy. These arrangements differ from each other only in the way in which the electromagnetic energy is directed to the pressure antinodes. For simplicity, Figs. 14, 15, 17, 18, and 19 omit details of the various inlets, outlets, and valve arrangements described above, and thus are only intended to illustrate how electromagnetic energy can be used to establish a standing acoustical wave. It is understood that any of the electromagnetic drive arrangements of figs.14, 15, 17, 18, and 19 can be used with any of the valved or valveless arrangements described above. When used in conjunction with the valveless embodiment, the following electromagnetic drive arrangements can provide a compressor which requires no moving parts. In Fig. 14 a flanged chamber 2 is provided which has flanged waveguide 102 attached thereto by common flange bolts, such that the microwave energy in waveguide 102 is coupled to chamber 2 through port 104. Port 104 can be pressure sealed with a microwave window made of microwave transparent materials such as Pryex, mica, or certain ceramics. This microwave window would allow microwave energy to pass from waveguide 102 into chamber 2, but would prevent the medium in chamber 2 from entering waveguide 102. In operation a microwave source generates microwaves which are guided by waveguide 102 to chamber 2. This microwave energy then enters chamber 2 through port 104. The frequency of the microwave radiation is chosen so that a standing microwave is set up along the length of chamber 2, having an energy distribution similar to the curve of Fig. 16. Thus, the chamber 2 acts as a resonant cavity for the microwave radiation. The standing microwave's frequency is also chosen so that the regions of maximum energy accumulation, coincide with like-pressure-phase antinodes 34a, 34b, 34c of standing acoustical wave 101. Whether or not these regions of maximum energy accumulation are electric or magnetic, depends on the mode of the standing microwave in chamber 2. The choice of which microwave mode to use, depends on the microwave molecular absorption characteristics of the medium being present in chamber 2. For example, if the molecules have electric dipole moments, then the regions of maximum energy accumulation should be electric rather than magnetic. Hereinafter, it will be assumed for the sake of example, that these regions of maximum energy accumulation are electric. When the standing microwave is present in chamber 2, microwave energy is absorbed by the medium primarily within the regions of maximum electric energy accumulation. The microwave source is pulsed or modulated at a rate which excites the desired acoustical mode. By pulsing or modulating the microwave source, the intensity of the standing microwave pattern is caused to vary periodically. This periodic variation of the microwave's intensity, causes a periodic pressure increase primarily at like-pressure-phase antinodes 34a, 34b, 34c, since these are the points of maximum electric energy accumulation. These periodic pressurizations, create pressure wavefronts which emanate from each of the pressure antinodes 34a, 34b, 34c, thus forming longitudinal waves which propagate bi-directionally along the length of chamber 2. The microwave source is pulsed or modulated at a rate which keeps the periodic thermal excitation of the medium in phase with the pressure oscillations of like-pressure-phase antinodes 34a, 34b, 34c. In other words, the intensity of the microwave field will be greatest when pressure antinodes 34a, 34b, 34c are at their peak positive pressure, thereby providing the correct reinforcement needed to sustain standing acoustical wave 101. This method could be extended to any number of pressure antinodes, as long as these antinodes are all of like-pressure-phase. Fig. 15 illustrates another method for establishing a standing microwave in chamber 2. This embodiment provides flanged chamber 2 with a center conductor 106, which is located along the axis of chamber 2. Center conductor 106 is electrically connected to the center conductor of coaxial cable 107. The shield of coaxial cable 107, is electrically connected to chamber 2. Center conductor 106 is also electrically connected to the end wall of chamber 2. This arrangement is basically a shorted co-axial cable whose outer shield forms chamber 2. A pulsed or modulated microwave source applies microwave energy of a chosen frequency, by way of coaxial cable 107, to the center conductor 106 and to chamber 2. Said microwave energy causes a standing microwave to be established along the length of chamber 2 between center conductor 106 and chamber 2. This arrangement is similar to that of a standing microwave in a co-axial cable. The standing microwave has an electric energy distribution along the length of chamber 2, which resembles the curve of Fig. 16. Once the standing microwave is so established, this embodiment induces a standing acoustical wave in the same manner and according to the same theory and principles as the embodiment of Fig. 14. The embodiment of Fig. 15 lends itself easily to miniaturization. The embodiment of Fig. 17 provides a circular microwave cavity 112, which transversely intersects chamber 2 at the pressure antinode 34. To avoid disrupting the standing acoustic wave, microwave cavity 112 can be filled with a solid dielectric material. Such a solid dielectric would fill only the volume of microwave cavity 112 which is not common to chamber 2. Alternatively, the radius of circular microwave cavity 112 could be made equal to the radius of chamber 2. Screen 111 and screen 113 are placed transversely across chamber 2, thus extending the boundaries of microwave cavity 112 through chamber 2. Screen 111 and screen 113 help restrict the microwave energy to the area between them, while still allowing axial oscillations of the medium along chamber 2. Microwave cavity 112 has coaxial cable 110 attached thereto, such that the microwave energy carried by coaxial cable 110 is coupled to microwave cavity 112 through microwave radiator 114. In operation, a microwave source generates microwaves which travel in turn through coaxial cable 110, through microwave radiator 114, and into microwave cavity 112. Screen 111 and screen 113 restrict the microwave energy to a region within chamber 2 corresponding to pressure antinode 34. The microwave source is pulsed or modulated at a rate which excites the desired acoustical mode. This causes a periodic pressure increase which is localized at pressure antinode 34, since this is the region in chamber 2 which is exposed to the microwave energy inside microwave cavity 112. Hence, the periodic microwave pulses create pressure wavefronts which emanate from pressure antinode 34, thus forming longitudinal waves which propagate along the length of chamber 2. The microwave source is pulsed at a rate that will keep the thermal excitation of the medium in phase with the pressure oscillations of the pressure antinode 34. This arrangement could be extended to any number of pressure antinodes by adding additional microwave cavities. Such additional cavities could be driven by a single microwave source, as long as these cavities are located at antinodes of like-pressure-phase. In Fig. 18 an embodiment is shown which employs the radiation leakage from a co-axial cable. A flanged chamber 2 is provided with a co-axial cable 116 located along the chamber's axis. Microwave energy is applied to coaxial cable 116, by external coaxial cable 119. Co-axial cable 116 is filled with a solid dielectric 117 between center conductor 118 and shield 120. The co-axial cable's shield 120 is open at points corresponding to like-pressure-phase antinodes 34a and 34b. In operation a microwave source applies microwave energy of a chosen frequency to center conductor 118 and the co-axial cable shield 120, by way of external coaxial cable 119. This microwave energy travels through co-axial cable 116 and leaks out, due to the open shield, at like-pressure-phase antinodes 34a and 34b. In this way like-pressure-phase antinodes 34a and 34b, will be exposed to microwave energy. The microwave source can be pulsed or modulated at a rate which excites the desired acoustical mode. This causes a periodic pressure increase which is localized at pressure antinodes 34a and 34b, since these are regions in chamber 2 which are exposed to microwave energy. Hence, the pulsed or modulated microwave energy creates pressure wavefronts which emanate from pressure antinodes 34a and 34b, thus forming longitudinal waves which propagate along the length of chamber 2. The microwave source is pulsed or modulated at a rate which will keep the thermal excitation of the medium in phase with the pressure oscillations of pressure antinodes 34a and 34b. The pulses occur when pressure antinodes 34a and 34b are at their peak positive pressure, thus providing the correct reinforcement needed to sustain the standing acoustical wave. This method could be extended to any number of pressure antinodes by adding additional leakage points in the co-axial cable 116, as long as all of these leakage points are located at antinodes of like-pressure-phase. The microwave source for the embodiments of figs.14, 15, 17, and 18 could be any microwave generator, such as a MAGNETRON or KLYSTRON tube, or could comprise solid state devices such as GUNN or IMPATT diodes, as long as they supply enough power for a given application. Magnetrons are available with pulse repetition rates ranging from 1 kHz to 100 kHz. Therefore, magnetrons could be used to drive standing acoustic waves having frequencies well into the ultrasonic range if so desired. It should be noted that the absorption properties of a gas may be enhanced, by applying static electric or magnetic fields across the gas in the region of electromagnetic absorption. Fig. 19 illustrates an embodiment of the invention which provides a LASER driving means for maintaining a standing wave. A chamber 2 is provided which is transversely intersected at its alternate pressure antinodes by LASER beam guides 90a, 90b, 90c, 90d, 90e. The beam guides are equipped with reflective surfaces a,b,c,d,e,f which reflect the LASER beam at 90° angles, so that the LASER beam follows the beam guide. Identical optical windows 98, provide pressure seals between each of the beam guides and the interior of chamber 2. Beam spreader 100 provides control of the LASER beam's cross sectional geometry so as to maximize the medium's exposure to the beam at the pressure antinodes. A LASER 92 emits LASER beam 94, so that LASER beam 94 passes in turn through beam spreader 100 then through optical window 98, and then is directed along the interior of beam guide 90a. The beam 94 then experiences multiple reflections due to reflective surfaces a,b,c,d,e,f and therefore propagates in turn through beam guides 90a, 90b, 90c, 90d, 90e. Beam guide 90e is terminated by reflective surface 96, which reflects the beam through 180° causing it to return along the same path. Alternatively, beam guide 90e could be terminated by an absorber, which would absorb the beam's energy and prevent the beam's reflection. In operation, the LASER beam 94 is pulsed, and so causes a periodic highly localized pressure increase of the medium. Hence, the periodic LASER pulses create pressure wavefronts which emanate from pressure antinodes 34a, 34b, 34c, 34d and propagate as longitudinal waves along the length of chamber 2. The LASER pulses will have a repetition rate that will keep the instantaneous thermal excitation of the medium in phase with the pressure oscillations of the like-pressure-phase antinodes 34a, 34b, 34c, 34d. The pulses occur when said pressure antinodes are at their peak positive pressure, thus providing the correct reinforcement needed to sustain the standing wave. This method could be extended to any number of pressure antinodes, as long as these antinodes are all of like-pressure-phase. Alternatively, this present embodiment could be reduced to a single beam-chamber intersection, as long as said intersection is located at a pressure antinode, and excites the medium in phase with its pressure oscillations, as described above. LASER 92 could be a CO 2 LASER or an infrared LASER which could directly excite the medium's molecular vibrational states. An alternative driving means would be to locate individual IRLEDs at each of the like-pressure-phase antinodes, as long as they could provide enough power for a particular application. Also, solar energy could provide an abundant source of infrared radiation for driving the embodiment of Fig. 19. The following considerations will be common to all of the microwave driven embodiments, and to some degree, the infrared driven embodiments. In each of these embodiments, the electromagnetically induced pressure increase of the medium, is due to the electromagnetic excitation of the medium's molecular energy states. Molecular collisions serve to convert the energy of these excited molecular states into the increased kinetic energy of the gas, and thus a pressure increase. Most often, these molecular energy states which are responsible for microwave absorption will be rotational, but in some cases may include hindered motions. An example of a hindered motion is the inversion transition of ammonia at 24 GHz. Due to the broadening of the molecular absorption lines at high pressures (i.e. fractions of an atmosphere or higher), a broad range of frequencies can be used. In short, any frequency of electromagnetic radiation can be used, as long as its absorption results in a change of pressure in the gas. In the case of gases, the microwave absorption at a pressure antinode will be much higher than would be expected from the undisturbed pressure of the gas. In general, the microwave absorption of gases increases with the pressure and density of the gas. During operation, the microwave field is turned on when the pressure at the pressure antinode is at its maximum value, which is higher than the undisturbed pressure of the gas. Therefore, the microwave absorption coefficient of the gas at this instant will be greater than the absorption coefficient for the gas at its undisturbed pressure. If the microwave power is increased during operation, more microwave power is absorbed, and the gas pressure and density at the pressure antinode will increase for subsequent acoustic cycles. This increase in pressure and density in turn increases the microwave absorption of the gas, and even more power can be absorbed during subsequent microwave pulses. Also, if the microwave driven embodiments are operated at high pressure amplitudes, shock wave formation will begin to occur. Shock wave formation, due to the nonlinear effects of large amplitude pressure waves, can cause the pressure and density at the pressure antinodes to increase dramatically. Thus, shock wave formation can further increase the microwave absorption of the gas. In a paper by W. D. Hershberger (Thermal and Acoustic Effects Attending Absorption of Microwaves by Gases, RCA Review, Vol. 7 No. 3 Sept. 1946), it was shown experimentally that the acoustic power developed, due to microwave absorption, varies with the square of the microwave power. This square law behavior, indicates that the electro-acoustic efficiency improves as the microwave input power is increased. Such a behavior is most opportune, for achieving the large acoustic amplitudes needed for significant compression ratios. It is also possible to drive a standing acoustic wave by applying electromagnetic energy of constant intensity to the pressure antinodes, as long as the desired acoustical mode is initially excited. The electromagnetic absorption of a gas varies with the pressure and density of the gas. Since the pressure and density of the gas at the pressure antinodes both vary in phase with the resonant mode, microwave absorption will automatically vary in the proper phase to drive the resonant mode. Thus, energy will be added to the acoustic wave from a constant intensity electromagnetic field, as long as the desired acoustic mode is initially excited. In some cases, the sudden application of the constant intensity field may be enough to provide initial excitation of the desired acoustical mode. A constant field arrangement has the added advantage of not requiring timing circuitry, for keeping a pulsed or modulated electromagnetic source in phase with the pressure oscillations of the acoustic wave. The infrared and microwave driving arrangements described above, demonstrate how electromagnetic radiation throughout a broad frequency range, can be used to establish a standing acoustical wave inside an acoustically resonant chamber. However, these arrangements are not presented as being exhaustive techniques in optical, infrared, and microwave technology, since these technologies are rich in many alternative methods for directing electromagnetic radiation to regions inside a chamber. Rather, these examples illustrate how electromagnetic energy, once directed to the proper regions of a chamber, can be used to establish a standing acoustical wave in the chamber. Description of Refrigeration and Air-conditioning Applications Fig. 20 illustrates the use of the present invention as a compressor, in a vapor-compression refrigeration system. In Fig. 20 the present invention is connected in a closed loop, consisting of condenser 124, capillary tube 126, and evaporator 130. Alternatively, capillary 126 could be replaced by any number of well known refrigerant expansion devices. This arrangement constitutes a typical vapor-compression system, which can be used for refrigeration, air-conditioning, heat pumps, water coolers, dehumidifiers, and many other applications. The valved embodiment of the present invention shown in Fig. 20, is provided with a suction reed valve 131, a discharge reed valve 133, a suction plenum 137, and a discharge plenum 135. In operation, a pressurized liquid refrigerant flows into evaporator 130 from capillary tube 126, therein experiencing a drop in pressure. This low pressure liquid refrigerant inside evaporator 130 then absorbs its heat of vaporization from the refrigerated space 128, thereby becoming a low pressure vapor. The low pressure vapor is drawn out of the evaporator and into suction plenum 137, and then through suction reed valve 131 into chamber 132. After entering chamber 132, the low pressure vapor is then acoustically compressed and discharged at a higher temperature and pressure through discharge reed valve 133, into discharge plenum 135, and then into condenser 124. As the high pressure gaseous refrigerant passes through condenser 124, it gives up heat and condenses into a pressurized liquid once again. This pressurized liquid refrigerant then flows through capillary tube 126, and the vapor-compression cycle is repeated. Any of the various embodiments of the present invention can be used in the system of Fig. 20; the description and operation of which has been given above. The embodiment which is chosen, will depend on the design needs of a particular application. For some applications, it may be desirable to enclose the present invention, including the driving means, in a hermetic vessel. When designing a system like that of Fig. 20, some advantage will be found in the proper choice of the average pressure within the acoustic chamber. This average pressure is the undisturbed pressure which exists inside the acoustic chamber in the absence of an acoustic wave. During operation the standing wave creates a pressure differential whose suction pressure is lower than the average pressure, and whose discharge pressure is higher than the average pressure. Thus, to make the suction pressure equal to the evaporator pressure, the average chamber pressure should lie somewhere between the evaporator and condenser pressures. To provide added control over the average chamber pressure, various regulating valves and devices can be used, which would manipulate suction or discharge flows. In some applications, a vapor compression system could be driven by solar energy. Solar energy comprises an excellent infrared source for electromagnetic driving of the present invention. For example, an embodiment like that of Fig. 19 could be solar driven. A simple solar driving arrangement could comprise a mirror for intensifying the sun's radiation, and a beam chopper to provide a pulsed beam. Alternatively, the present invention can be driven by constant intensity electromagnetic energy, although the desired acoustical mode may need to be initially excited. Initial excitation of the desired acoustical mode, could be accomplished by a mechanical driver. In some cases, the sudden exposure to the constant intensity electromagnetic energy may be enough to initiate the desired acoustical mode. Self initiation of the desired acoustic mode becomes more reliable if more than one pressure antinode is driven by the constant intensity source. Multiple antinode driving would tend to lock in the desired mode. Constant intensity driving provides great simplicity for the solar driven embodiments, since the pulsing means can be eliminated. In general, a pulsed source would represent greater efficiency. However, since solar energy is free, the added simplicity of a constant source becomes more desirable. Several solar driven embodiments of the present invention could be placed in series to provide higher pressure differentials, or in parallel to provide higher net flow rates. The solar driven embodiments could also find applications in outer space, where intense infrared energy from the sun is plentiful. A mechanical drive could be combined with a solar drive to provide a hydrid drive vapor-compression system. For example, the present invention could be driven by both a mechanical driver, and by solar energy. In the absence of sunlight the mechanical driver would provide the energy needed to drive the present invention. On sunny days, the energy consumption of the mechanical driver could be supplemented by solar energy. The solar infrared energy would be directed to the pressure antinodes as described above. Depending on changing conditions, this hybrid drive system could operate in three modes: (1) all mechanically driven, (2) all solar driven, (3) both mechanically and solar driven at the same time. Mode selection could be varied automatically in response to ongoing operating conditions. Alternatively, a solar driven embodiment could serve as a pre-compressor for other conventional compressors, thereby reducing the compression ratio, which must be provided by the conventional compressor, during sunlit hours. For cooling applications like air-conditioning, this would reduce the system's power consumption during daylight hours when heat loads are high and power consumption is at its peak. Since the present invention eliminates all moving parts which require oil, a vapor-compression system can be operated with an oil-free refrigerant. Thus, the many system design problems associated with oils can be eliminated, and a vapor-compression system could approach more closely the efficiency of an ideal refrigeration cycle. Variable Discharge Pressure A variable discharge pressure can be achieved with the present invention by simply varying the acoustic amplitude of the standing acoustic wave. A simple control circuit can be provided to vary the acoustic amplitude as a function of the condensing medium's temperature, or other system variables. In this way, the discharge pressure of the present invention would never be any larger than the minimum discharge pressure needed for condensation to occur at the existing operating conditions. Therefore, no energy is wasted by generating discharge pressures which are in excess of the minimum pressure required for condensation to occur. In general, when the acoustic amplitude is increased to provide a higher discharge pressure at the pressure antinodes, the suction pressure will tend to decrease. Therefore, care must be taken if the evaporator pressure is to be kept constant as the acoustic amplitude varies. To provide added control over the evaporator pressure, various regulating valves and devices can be used, which would regulate the pressure between the evaporator and the inlet of the present invention. Fig. 21 shows an example of a control circuit which provides automatic control of the discharge pressure, and also provides an automatic frequency control, which keeps the frequency of acoustic driver 10 tuned to the acoustic resonance of chamber 132. In operation, the circuit of Fig. 21 acts to maintain the liquid level of condensed refrigerant in condenser 124, to a level between conductivity transducers T1 and T2. Transducers T1 and T2 indicate the vapor or liquid state of the refrigerant, by detecting associated changes in the refrigerant's conductivity. Microprocessor 153 monitors transducers T1 and T2 for changes, via dual analog-to-digital converter 155. If operating conditions cause the liquid refrigerant level in condenser 124 to drop below transducer T2, then microprocessor 153 responds by sending a control signal to amplifier 160 via digital-to-analog convertor 156. The control signal increases the gain of amplifier 160, thus boosting the power of acoustic driver 10, which in turn increases the amplitude of standing acoustic wave 157. This increased amplitude, provides a higher discharge pressure which increases the rate of condensation in condenser 124. Once the level of liquid refrigerant in condenser 124 rises past transducer T2, microprocessor 153 responds by maintaining a constant acoustic amplitude and thus a constant discharge pressure. If operating conditions cause the liquid refrigerant level to rise above transducer T1, then microprocessor 153 responds by reducing the power of acoustic driver 10, which in turn reduces the discharge pressure. This reduced discharge pressure decreases the rate of condensation in condenser 124. When the level of liquid refrigerant in condenser 124 drops below transducer T1, microprocessor 153 responds by maintaining a constant acoustic amplitude, and thus a constant discharge pressure. In this way, the control circuit maintains the liquid refrigerant level in condenser 10, between transducers T2 and T1. The frequency control circuit of Fig. 21 has an integrator 62, a comparator 64, a voltage controlled oscillator 66, and an amplifier 160. The components of this frequency control circuit, and their function, are identical to the circuit of Fig. 12. Many different operating conditions are apt to change and can cause the level of liquid refrigerant in condenser 124 to vary. However, each will be treated equally by the control system. Thus, for any given set of operating conditions, the control circuit will maintain the minimum discharge pressure which is required for condensation to occur in the lower part of condenser 124. Other parameters of the system could be monitored by the control circuit to provide addition control and optimization of the cooling system. For lower cost applications, the microprocessor control circuit of Fig. 21 could be replaced by a simple switching network. Such a switching network would select a number of fixed power levels for acoustic driver 10, in response to signals from transducers T1 and T2. This switching control circuit would provide a limited number of fixed discharge pressures, rather then the continuously variable discharge pressure of the microprocessor control circuit. A control circuit, like that of Fig. 21, can be easily adapted to the microwave driving systems described above. First, an appropriate frequency locking control must be provided. For optimal operation, the microwave source should be pulsed on when the pressure antinode is at its point of highest pressure during an acoustic period. A single pressure transducer located in the acoustic chamber, can provide a reference signal for triggering the pulses of a microwave generator. Since the microwave energy is pulsed on only when the antinode pressure is at its peak, the system will naturally remain in resonance. This simple arrangement eliminates the need for more complex frequency locking circuits. Second, a means to vary the microwave power must be obtained to permit a variable discharge pressure. One easy method to vary the average microwave power, is to vary the duty cycle of the microwave pulses. Alternatively, varying the high voltage on a microwave generating tube, could be used to control the microwave power. Refrigerant Subcooling System The present invention provides a secondary cooling system which can be utilized for refrigerant subcooling or as a condensing medium. This secondary cooling system is contained inside the resonant acoustic chamber, and requires no moving parts. Fig. 23 is a sectional view on line 3-3 of Fig. 22, and shows an embodiment of a refrigerant compressor in accordance with the present invention, which provides a refrigerant subcooling system. The compressor for this vapor-compression system is formed by a chamber 2, a driver 10, a suction reed valve 164, a discharge reed valve 162, a suction plenum 168, and a discharge plenum 166. In Fig. 23, discharge tubing 170 connects discharge plenum 166 to a condenser 172. The outlet of condenser 172 is connected to a heat exchanger coil 174. Heat exchanger coil 174 forms a coil of tubing which is wound around and welded to chamber 2, so as to provide good thermal contact between chamber 2 and heat exchanger coil 174. The output of heat exchanger coil 174 is connected to a capillary tube 176. Capillary tube 176 is connected to an evaporator 178 which is located inside the refrigerated space 180. Suction tubing 182 connects the output of evaporator 178 to suction plenum 168. The midsection of chamber 2 is thermally isolated from the environment by optional insulation 184. Capillary tube 176 is thermally isolated from the environment by optional insulation 186. Suction tubing 182 is thermally isolated from the environment by optional insulation 188. A heat pump plate stack 190 is provided inside of chamber 2. Heat pump plate stack 190 includes a stack of evenly spaced parallel stainless steel plates which are placed longitudinally along the length of chamber 2. Alternatively, the plates can be made of other materials such as fiberglass of wire screens. A more detailed view of heat pump plate stack 190 is provided in Fig. 24. Heat pump plate stack 190 is everywhere thermally isolated from chamber 2, except at opposite ends T C an T H . At opposite ends T C and T H of each individual plate in heat pump plate stack 190, are attached respective copper strips 192C and 192H. As seen in Fig. 24, copper strips 192C and 192H extend along the ends of each individual plate of heat pump plate stack 190 and are soldered thereto. None of the plates in heat pump plate stack 190 come in contact with the inner surface of chamber 2. Thermal contact between heat pump plate stack 190 and chamber 2 is provided by copper strips 192H and 192C. The two ends of each copper strip extend beyond the plates to meet the inner surface of chamber 2, and are soldered thereto. Copper strips 192C provide good thermal contact between end T C of heat pump plate stack 190 and the wall of chamber 2. Copper strips 192H provide good thermal contact between end T H of heat pump plate stack 190 and the wall of chamber 2. This arrangement allows heat conduction, between heat pump plate stack 190 and chamber 2 to occur only at ends T C and T H . Chamber 2 is also provided with heat fins 194, for the dissipation of heat from the walls of chamber 2 to the surrounding air. In operation, acoustic driver 10 emits acoustic waves into the gaseous refrigerant inside chamber 2. The frequency of acoustic driver 10 is controlled in such a way as to maintain a standing acoustical wave, which is depicted as displacement waveform 198. As disclosed above, the gaseous refrigerant inside chamber 2 is acoustically compressed and discharged into discharge tubing 170. This high pressure gaseous refrigerant then passes into condenser 172, where it condenses to a liquid by giving up heat to the surrounding air. Liquid refrigerant then flows from condenser 172 into heat exchanger coil 174, wherein it is subcooled to below its previous condenser temperature. The basis for this cooling is treated separately below. Subcooled liquid refrigerant then flows out of heat exchanger coil 174 and into capillary tube 176, which meters the flow of liquid refrigerant into evaporator 178. Insulation 186 minimizes the heating of the subcooled liquid refrigerant in capillary tube 176, as the refrigerant passes from heat exchanger coil 174 to evaporator 178. Once in evaporator 178, the liquid refrigerant absorbs its heat of vaporization from refrigerated space 180. This low temperature low pressure vapor is then drawn out of evaporator 178 and into chamber 2, via suction tubing 182. The gaseous refrigerant within chamber 2 is acoustically compressed and discharged into discharge tubing 170, and the cycle is repeated. The liquid refrigerant subcooling which occurs in heat exchanger coil 174 is explained as follows. The presence of a standing acoustical wave in a chamber, will cause heat to be pumped along the walls of the chamber; being pumped away from the pressure nodes and towards the pressure antinodes. Consequently, the chamber walls grow colder at the pressure nodes and warmer at the pressure antinodes. The heat flow capacity is proportional to the surface area exposed to the standing acoustic wave, and can be increased by providing heat pump plate stack 190 inside chamber 2. In the presence of the standing acoustic wave 198, heat will be pumped away from the cold side T C and towards the hot side T H of heat pump plate stack 190. When the temperature of side T C drops below the temperature of the adjacent wall of chamber 2, heat flows in turn from heat exchanger coil 174, through the wall of chamber 2, through the copper strips 192C, and into cold side T C of heat pump plate stack 190. As heat accumulates at the hot end T H of heat pump plate stack 190, the temperature of hot end T H rises above the wall temperature of chamber 2. When the temperature of side T H rises above the temperature of the adjacent wall of chamber 2, heat flows in turn from hot side T H of heat pump plate stack 190, through copper strips 192H, through the wall of chamber 2, through heat fins 194, and into the ambient air. Thus, as the liquid refrigerant flows through heat exchanger coil 174, it is subcooled to a temperature below that of the ambient air surrounding air-cooled condenser 172. A detailed theoretical and experimental description of the acoustical heat pumping effect described above, is provided in the following publications. (1) John Wheatley, T. Hofler, G. W. Swift, and A. Migliori, An Intrinsically Irreversible Thermoacoustic Heat Engine, J. Acoust. Soc. Am., Vol. 74, No. 1, p. 153 July 1983 (2) John Wheatley, T. Hofler, G. W. Swift, and A. Migliori, Understanding Some Simple Phenomena In Thermoacoustics With Applications To Acoustical Heat Engines, Am. J. Phys., Vol. 53, No. 2, p. 147 February 1985 (3) John Wheatley, G. W. Swift, and A. Migliori, The Natural Heat Engine, Los Alamos Science, No. 14, Fall 1986 (4) G.W. Swift, Thermoacoustic Engines, J. Acoust. Soc. Am., Vol. 84, No. 4, p. 1145 October 1988. These papers teach how to design and predict the performance of an acoustic heat pumping system in quantitative detail, and the disclosures of these publications are hereby incorporated by reference. To maximize the cooling capacity of the acoustic heat pumping system, all non-refrigerant heat loads on heat pump plate stack 190 should be kept to a minimum. The following considerations help to achieve this minimization. Insulation 184 reduces the heat absorbed by the walls of chamber 2 from the surrounding air, thereby promoting the refrigerant within heat exchanger coil 12 as the primary heat load of heat pump plate stack 190. Insulation 188 minimizes the superheating of the refrigerant vapor in suction tubing 182, which reduces the heat load on the heat pump plate stack 190. Locating driver 10 at the hot side T H of heat pump plate stack 190, allows the heat generated by driver 10 to be rejected to the environment through heat fins 194. However, heat pump plate stack 190 could be moved to the valved end of chamber 2, and subcooling would still continue at reduced capacity. Therefore, the exact placement of heat pump plates 190 in Fig. 23 is not intended as a limitation on the scope of the present invention. Thus, the acoustic cooling system described above, will serve to reduce the temperature of the liquid refrigerant before it enters evaporator 178, thereby minimizing evaporative flashing and increasing the refrigerating effect per mass of refrigerant circulated. The acoustic cooling system of Fig. 23 can be operated in another mode. Heat exchanger coil 174 can provide a condensing medium whose temperature is lower than the temperature of condenser 172. This lower temperature condensing medium allows the use of lower discharge pressures, while still maintaining condensation. Lower discharge pressures will reduce the energy consumption needed for gaseous compression. Discharge pressure will largely determine whether the refrigerant will enter heat exchanger coil 174 as a gas, liquid, or liquid-vapor mixture. If the discharge pressure is high enough for condensation to occur in air-cooled condenser 172, then the refrigerant will enter heat exchanger coil 174 as a liquid. If the discharge pressure is not high enough for condensation to occur in air-cooled condenser 172, then the refrigerant will enter heat exchanger coil 174 as a gas, and condense therein. For pressures between these two extremes, the refrigerant would enter heat exchanger coil 174 as a liquid-vapor mixture. For low discharge pressures, the "effective" condenser would be thought of as the combination of air-cooled condenser 172 and heat exchanger coil 174. Thus, the discharge pressure would be chosen on the basis of the temperature within heat exchanger coil 174, which is much lower than the temperature of air-cooled condenser 172. In this mode, the discharge pressure need not be any higher than is necessary for condensation to occur in heat exchanger coil 174. Therefore, discharge pressures can be used which are lower than would be possible if only air-cooled condenser 172 were present. It can be seen then, that there are efficiency advantages associated with each of these two extremes of discharge pressure. In general, a control circuit can be built to exploit this entire range of discharge pressures in response to such changing conditions as the cooling load and the condensing medium's temperature. Fig. 25 shows a refrigeration system similar to that described for Fig. 23, except that the acoustical standing wave is driven by electomagnetic-gas interactions. Acoustical driver 10 of Fig. 23 is replaced in Fig. 25 by a microwave resonant cavity 200. The boundaries of microwave resonant cavity 200 are defined by the walls of chamber 2 and the hot end T H of heat pump plate stack 190. Heat pump plate stack 190 forms a metal mesh which prevents the microwave energy from leaving microwave cavity 200. Pulsed or modulated microwave energy is delivered to microwave resonant cavity 200 from microwave generator 206, via coaxial cable 202 and microwave radiator 204. As described previously, the presence of microwave energy in microwave resonant cavity 200, causes a standing acoustical wave to be established in chamber 2. Once the standing acoustical wave is established in chamber 2 the refrigeration system of Fig. 25 operates in the same manner and according to same theory and principles as the refrigeration system of Fig. 23. The circuits described above, for controlling driver frequency and discharge pressure, can be directly applied to the embodiments of Fig. 23 and Fig. 25. In addition, several different configurations of the cooling system, and corresponding control circuits, are possible. For example, the system could be designed to run at low discharge pressures, by moving the transducers T1 and T2 of Fig. 21 to the inlet and outlet respectively of heat exchanger coil 174. For this configuration, the "effective" condenser would be the combination of condenser 172 and heat exchanger coil 174. The control circuit would perform in exactly the same manner, except that the discharge pressure would be maintained at the level needed for condensation to occur in heat exchanger coil 12. Many other parameters of the system could be monitored by the control circuit to provide additional control and optimization of the cooling system. Vapor-compression cooling equipment can take many forms and is found in many different applications and industries. As a compressor, the present invention is not limited only to those cooling applications described above, but can be adapted to any number of applications. Consequently, the vapor-compression cooling system described herein, can be employed in many different cooling applications, including air-conditioners, heat pumps, chillers, water coolers, refrigerators, and many more. The present invention provides a new compressor for vapor-compression cooling systems, which exploits the properties of acoustic resonance in fluids for fluid compression, and provides a discharge pressure which can be varied during operation in response to changing operating conditions, thereby reducing the compressor's energy consumption. Also, the present invention eliminates all moving drive parts, by exploiting a fluid's absorption of electromagnetic energy. Furthermore, the present invention provides an oil-less compressor. The present invention further provides a new compressor wherein a fluid refrigerant is compressed and then subcooled, by means of an acoustical heat pumping system. Alternatively, this acoustical heat pumping system can provide a low temperature condensing medium, which allows condensation to occur at lower than normal discharge pressures. While the above description contains many specifications, these should not be construed as limitations on the scope of the invention, but rather as an exemplification of one preferred embodiment thereof. Many other variations are possible, and may readily occur to those skilled in art. For example, more than one driver could be used in any of the mechanically driven embodiments, and the frequency and phase of these multiple drivers could be manipulated with respect to each other to create greater pressure differentials, beat frequency phenomenon, combinations of standing and travelling waves, and other effects which may prove useful in various applications. Also, a single driver could be placed in an intermediate position in the chamber, such that standing acoustic waves could be set up on both sides of the driver; the resulting two units being connected in series or parallel. Also, the waveforms that drive either single or multiple drivers need not be sinusoidal, but could be sawtooth, square wave, pulsed, or any waveform that satisfies a given design need. Furthermore, the size of the present invention will be determined by the required capacity of a given application, and not by the present invention itself. By choosing the proper acoustic wavelength, the present invention can be scaled up or miniaturized. In addition, a resonant acoustic chamber need not be cylindrical, but can be any geometry which will support a standing acoustical wave. A toroidal chamber could be used for the electromagnetically driven embodiments. For a toroidal chamber, the waves would propagate bi-directionally around the torus, thus forming a standing wave, and the electromagnetic energy would be directed to the pressure antinodes. Unlike a cylindrical chamber, there would be no reflections, since the toroidal chamber is continuous. Additional chamber geometries, and other acoustic modes can be used to support a standing wave pattern. For example, a cylindrical chamber whose radius is large relative to its length, can oscillate in a radial mode. Also, spherical chambers can be made to oscillate in a radial mode. Radial mode oscillations have the advantage of concentrating the acoustic wave's pressure at the center of the chamber. Different chamber geometries, such as cylinders and spheres, can be combined to form Helmholtz resonators. In short, any chamber which will support a standing acoustic wave can be used. Electromagnetic-gas interactions can be used to acoustically drive any of these chambers. Inlets an outlets may also be formed in different geometries, and thus could define openings in chamber 2 such as a series of circular holes, slits, indentations, or separate adjoining chambers. Alternatively, coaxial tubes, with periodic openings at the nodes and antinodes, could be used to locate inlets and outlets along the axis of the chamber 2. For the valveless embodiments, several inlets can be provided at a pressure node, with each inlet having a different distance from the pressure node. Each inlet would provide a slightly different suction pressure, and could be selected during operation to keep suction pressure constant, while the acoustic amplitude is varied to change the discharge pressure. Several units can be connected so that their inlets and outlets form series and/or parallel combinations, or separate chambers could intersect at common pressure antinodes, all of which can provide greater pressure differentials and improve capacity. Many variations are also possible for the acoustic heat pumping system. For example, in higher evaporator temperature applications, where smaller compression ratios can be used, the discharge gas will not be as hot, and the air-cooled condenser 172 of Figs.23 and 25 could be reduced in size or removed altogether. In this case, heat exchanger coil 174 would serve as the primary condensing medium, and the number of loops in heat exchanger coil 174 and the number of plates in heat pump plate stack 190, could be increased as necessary to improve the capacity. Moreover, more than one heat pump plate stack 190 can be used in chamber 2. Since plate stacks are placed between pressure nodes and antinodes, a standing acoustical wave with several nodes and antinodes could support more than one plate stack. Such additional plate stacks would increase the heat load capacity of the subcooling system. Alternatively, one plate stack could be used for condensing, and the other plate stack could be used for subcooling. Also, other features which are common to refrigeration technology, could be added. For example, capillary tube 176 of Figs.23 and 25 could be replaced with many different types of common refrigerant controls which are more responsive to changing operating conditions. In addition, other heat transfer devices besides heat fins 194 in Fig. 23 could be used to carry away excess heat. A forced air fan could be added to heat fins 194, thereby improving the rate of heat exchange with the surrounding air. Another alternative would be to provide a closed loop liquid coolant circulation system, in which a coolant would flow through a heat exchanger, with the heat exchanger being in thermal contact with the hot side T H of heat pump plate stack 28. The coolant would flow in turn through the heat exchanger, and then into an air-cooled radiator. Furthermore, plate stack 190 can be constructed of many different materials, such as fiberglass, plastics, or wire screens. These various plate stacks can be arranged longitudinally or transversely along chamber 2. Other geometries besides plates can be used, such as a continuous spirals or concentric cylindrical plates placed longitudinally along chamber 2. It should be noted that the acoustic heat pumping will occur along the chamber walls without any plates at all, although at a reduced level. Additionally, the heat exchanger coil 174 could be replaced with other types of heat exchangers. One such heat exchanger could be formed by replacing copper strips 30C with small channels which would be in thermal contact with the cold end T c of heat pump plate stack 190. The refrigerant could then pass through these channels, thereby giving up heat to the plates inside chamber 2. This arrangement would provide a more direct heat exchange between the refrigerant and the plates, than would be provided by heat exchanger coil 174. Finally, by virtue of the present invention's manifold features, such as diverse acoustic drivers, variable discharge pressure, acoustic heat pumping, and electromagnetic driving, there are naturally many embodiments which represent various combinations of these features.

Claims: A method of compressing a fluid refrigerant comprising introducing a fluid refrigerant to be compressed into a chamber (2) having at least one inlet and at least one outlet, characterized in that said method further comprises establishing a standing wave in the fluid refrigerant in the chamber, so that the fluid refrigerant is compressed. A method as set forth in claim 1, wherein said step of establishing a standing wave comprises the substeps of: establishing a travelling wave in the fluid refrigerant in the chamber; and converting the travelling wave into the standing wave in the fluid refrigerant in the chamber. A method as set forth in claim 1, wherein said step of establishing a standing wave comprises the step of applying electromagnetic energy to the refrigerant to set up the standing wave in the refrigerant in the chamber. A compression-evaporation system comprising means for subjecting a compressed refrigerant to a heat exchange operation, characterized in that said compression-evaporation system further comprises a standing wave compressor for compressing a refrigerant. A compression-evaporation system according to claim 4 comprising a condenser (172), a pressure reduction device (174) coupled to said condenser and an evaporator (178) coupled to said pressure reduction device, wherein said standing wave compressor is coupled to said condenser and said evaporator. A refrigerant compressor characterised in that said refrigerant compressor comprises a standing wave compressor which receives, acoustically compresses, and discharges a refrigerant, said standing wave compressor including an acoustic chamber (2) having at least first and second different cross sectional areas at at least first and second positions, respectively, along said acoustic chamber. A refrigerant compressor as set forth in claim 6, wherein said acoustic chamber (2) suppresses selected acoustic modes. A compression-evaporation system comprising a compressor according to claim 6 or 7, further comprising means for subjecting a discharged refrigerant to a heat exchange operation. A compression-evaporation system according to claim 8, wherein said acoustic chamber has one or more segments of varying cross sectional area. A compression-evaporation system according to claim 4 or claim 5 comprising a chamber (2), having at least first and second openings, for holding a fluid refrigerant to be compressed, and means, coupled to said at least first and second openings in said chamber, for subjecting a compressed fluid refrigerant to a heat exchange operation, wherein said standing wave compressor comprises driving means (10) for establishing a travelling wave in the fluid refrigerant in said chamber, said driving means and said chamber converting the travelling wave into a standing wave in the fluid refrigerant in said chamber, so that the fluid refrigerant is compressed. A compression-evaporation method according to claim 1 or 2, further comprising subjecting a compressed refrigerant to a heat exchange operation, wherein said compression-evaporation method further comprises directing acoustic energy into a fluid refrigerant in a chamber (2) having at least first and second openings; and selecting a frequency for said acoustic energy to establish a standing acoustic wave in the fluid refrigerant in the chamber so as to compress the fluid refrigerant. A compression-evaporation method as set forth in claim 11, further comprising suppressing selected acoustic modes within the chamber. A compression-evaporation method as set forth in claim 12, wherein said suppressing step comprises varying a cross sectional area of the chamber. A compression-evaporation method as set forth in claim 12, further comprising the step of controlling the flow of the fluid refrigerant through the at least first and second openings by using at least one valve (8). A refrigerant compressor comprising a chamber (2) for holding a fluid refrigerant to be compressed, said chamber including at least one inlet (4) and at least one outlet (6), characterised in that said compressor further comprises: driving means (10) for establishing a travelling wave in the fluid refrigerant in said chamber, said driving means and said chamber converting the travelling wave into a standing wave in the fluid refrigerant in said chamber, so that the fluid refrigerant is compressed. A refrigerant compressor as set forth in claim 15, wherein the standing wave in the fluid refrigerant has at least one pressure node and at least one pressure antinode. A refrigerant compressor as set forth in claim 16, wherein said at least one inlet (4) of said chamber (2) is located at said at least one pressure node, and said at least one outlet (6) of said chamber is located at said at least one pressure antinode. A refrigerant compressor according to any of claims 15 to 17, further comprising rectifying means (8) for allowing the fluid refrigerant to exit said chamber (2) via said at least one outlet (6) but for preventing the fluid from entering said chamber via said at least one outlet. A refrigerant compressor as set forth in claim 18, wherein said rectifying means (6) comprises at least one reed valve. The refrigerant compressor of claim 15, wherein the standing wave in the fluid refrigerant has at least one pressure node and at least one pressure antinode, wherein each said at least one inlet (4) is located at said at least one pressure antinode, and wherein each said at least one outlet (6) is located at said at least one pressure antinode, said refrigerant compressor further comprising: inlet rectifying means (38a-38c) for allowing the fluid refrigerant to enter but not exit said chamber; and outlet rectifying means (8a-8c) for allowing the fluid refrigerant to exit but not enter said chamber. A refrigerant compressor as set forth in claim 20, wherein said inlet rectifying means (38a-38c) comprises at least one reed valve, and said outlet rectifying means (8a-8c) comprises at least one reed valve. The refrigerant compressor according to any of the preceding claims, wherein said driving means (10) comprises at least one ultrasonic driver which emits ultrasonic energy of time varying intensity, the ultrasonic energy of time varying intensity has an acoustic frequency which is higher than the acoustic frequency of the standing wave, the ultrasonic energy of time varying intensity is demodulated into pulses by the fluid refrigerant, and the pulses have a repetition rate which is equal to the acoustic frequency of the standing wave, so that the pulses drive the standing wave. The refrigerant compressor according to any of claims 15 to 21, wherein said driving means (10) comprises at least one ultrasonic driver, which emits ultrasonic energy having an acoustic frequency which is equal to the acoustic frequency of the standing wave. A refrigerant compressor according to any of the preceding claims, wherein the fluid refrigerant comprises a gaseous refrigerant. A refrigerant compressor according to any of claims 15 to 21, wherein said driving means (10) comprises a linear motor. A refrigerant compressor according to any of claims 15 to 21, wherein said driving means (10) comprises a variable power acoustic driver, and wherein said refrigerant compressor further comprises control means for varying the power of said variable power acoustic driver based on changes in operation conditions, so that a discharge pressure is varied as a function of changing operating conditions. A refrigerant according to any of the preceding claims, wherein said chamber (73) includes variable cross section means for varying the cross sectional area of said chamber in the direction of acoustic displacement, so that undesired acoustic resonances are prevented from forming in said chamber. A refrigerant compressor as set forth in claim 27, wherein the standing wave has pressure nodes and pressure antinodes, and wherein said variable cross section means further acts to increase a pressure differential which exists between the pressure nodes and the pressure antinodes of the standing wave. A compressor according to claim 15, wherein the driving means applies electromagnetic energy to the refrigerant. A compressor as set forth in claim 29, wherein the applied electromagnetic energy has a time varying intensity. A compressor as set forth in claim 29, wherein the applied electromagnetic energy has a time constant intensity. A refrigerant compressor as set forth in claim 29, wherein said driving means comprises a variable power driver to vary the average power of the electromagnetic energy, and wherein said refrigerant compressor further comprises control means for varying the average power of said variable power driver based on changes in operating conditions, so that a discharge pressure is varied as a function of changing operating conditions. A compression-evaporation system comprising a refrigerant compressor according to claim 15, further comprising a condenser (172) coupled to said at least one outlet of said chamber, and an evaporator (178) coupled to said condenser and to said at least one inlet of said chamber. A compression-evaporation system as set forth in claim 33, wherein said driving means (10) comprises a variable power acoustic driver, and wherein said compression-evaporation system further comprises control means for varying the power of said variable power acoustic driver based on changes in operating conditions, so that a discharge pressure is varied as a function of changing operating conditions. A compression-evaporation method according to claim 1 or 2, further comprising the steps of condensing a compressed fluid refrigerant to produce a condensed fluid refrigerant which emits heat, reducing the pressure of the condensed fluid refrigerant to produce a reduced pressure fluid refrigerant which absorbs heat, and introducing the reduced pressure fluid refrigerant into the at least one inlet of a chamber having at least one inlet and at least one outlet, wherein said compression-evaporation method further comprises: directing acoustic energy into the fluid refrigerant in the chamber; selecting a frequency for said acoustic energy to establish a standing acoustic wave in the fluid refrigerant in the chamber; the pressure characteristics of the standing acoustic wave causing the reduced pressure fluid refrigerant to be drawn into the at least one inlet of the chamber, and causing the compressed fluid refrigerant to exit the chamber through the at least one outlet, so that the fluid refrigerant is moved through a compression-evaporation cycle. A compression-evaporation method as set forth in claim 35, wherein said directing step comprises varying the power of the acoustic energy directed into the fluid refrigerant based on changes in operating conditions, so that a discharge pressure is varied as a function of changing operating conditions. A refrigerant compressor according to claim 15, wherein the standing wave produces a temperature differential along said chamber, so that a first portion of said chamber is at a temperature which is higher than a second portion of said chamber; further comprising a heat exchanger (174) coupled to said chamber adjacent said second portion of said chamber and carrying the fluid refrigerant, said heat exchanger providing thermal contact between the fluid refrigerant and said second portion of said chamber. A refrigerant compressor as set forth in claim 37, further comprising a heat pumping surface (190) positioned in said chamber (2) and exposed to the standing wave existing within said chamber, wherein said heat pumping surface has first and second ends, wherein the second end of said heat pumping surface is adjacent the second portion of said chamber. A refrigerant compressor as set forth in claim 38, further comprising an additional heat exchanger coupled to said chamber, for providing thermal contact between a heat sink and the first end of said heat pumping surface. A refrigerant compressor as set forth in claim 37, wherein said driving means comprises a variable power acoustic driver, and wherein said refrigerant compressor further comprises control means for varying the power of said variable power acoustic driver based on changes in operating conditions, so that a discharge pressure is varied as a function of changing operating conditions. A compression-evaporation system according to claim 33, wherein the standing wave produces a temperature differential along said chamber, so that a first portion of said chamber is at a temperature which is higher than a second portion of said chamber; further comprising a heat exchanger (174) coupled to said refrigerant condenser and to said chamber adjacent said portion of said chamber, said heat exchanger providing thermal contact between the fluid refrigerant and the second portion of said chamber, said evaporator evaporating the condensed refrigerant provided by said heat exchanger and providing the evaporated refrigerant to said chamber.

Inventors: MACROSONIX CORP, LUCAS, TIMOTHY S., MACROSONIX CORP.
Applicants: MACROSONIX CORP, MACROSONIX CORP.
Application year: 1991
Main classification: F04F7
IPC codes: F04F7, F04B15, F25B1, F04B17, F02G1, F25B49