Patent Abstract:
A thermodynamic system that can approximate the Ericsson or Brayton cycles and operated in reverse or forward modes to implement a cooler or engine, respectively. The thermodynamic system includes a device for compressing a first fluid stream containing a first gas-liquid mixture having a sufficient liquid content so that compression of the gas within the first gas-liquid mixture by the compressing device is nearly isothermal, and a device for expanding a second fluid stream containing a second gas-liquid mixture having a sufficient liquid content so that expansion of the gas within the second gas-liquid mixture by the expanding device is nearly isothermal. A heat sink is in thermal communication with at least the liquid of the first gas-liquid mixture for transferring heat therefrom, and a heat source is in thermal communication with at least the liquid of the second gas-liquid mixture for transferring heat thereto. A device is provided for transferring heat between at least the gas of the first gas-liquid mixture after the first fluid stream exits the compressing device and at least the gas of the second gas-liquid mixture after the second fluid stream exits the expanding device. The compressing and expanding devices are not liquid-ring compressors or expanders, but instead are devices that tolerate liquid flooding, such as scroll-type compressors and expanders.

Full Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
   This application claims the benefit of U.S. Provisional Application No. 60/596,019, filed Aug. 24, 2005, the contents of which are incorporated herein by reference. 

   BACKGROUND OF THE INVENTION 
   The present invention generally relates to thermodynamic systems, and more particularly to thermodynamic systems operating according to the Ericsson or Brayton cycle and capable of achieving near-isothermal compression and expansion of a gas by mixing therewith a substantial quantity of liquid. 
   A refrigeration machine, heat pump, or cooler can be defined as any device that moves heat from a low temperature source to a high temperature sink. Operation of a refrigeration machine requires an input of energy, usually thermal, mechanical or electrical. Depending on the specific need, the heat absorbed in the low temperature source can be utilized to provide cooling, or the heat rejected to the high temperature sink can be used to provide heating, or both may be utilized simultaneously. As an example, for a typical household refrigerator the low temperature source is the space inside the refrigerator and the high temperature sink is the air in the room where the refrigerator is placed. Electrical energy is typically used to operate the system. 
   With the exception of a few niche applications, virtually all refrigeration machines operate on the vapor-compression (V-C) cycle. Common examples include home and automobile air conditioners, domestic and industrial food refrigeration, commercial comfort cooling, industrial process cooling, and many others. The traditional refrigerant fluids used in these machines contain compounds that result in ozone depletion if they escape into the upper atmosphere. These ozone depleting refrigerants are in the process of being phased out and eventually banned. However the new refrigerants, while not posing a risk to the ozone layer, are very potent greenhouse gasses. Other refrigerants that don&#39;t pose a substantial environmental risk have other drawbacks, such as being flammable or toxic. One such example is ammonia, which is an excellent refrigerant from a system performance perspective, but is highly toxic. There is a great need and much work is being done to develop and commercialize practical refrigeration systems that do not require the use of environmentally hazardous refrigerants. 
   The reverse Ericsson cycle is an alternative refrigeration cycle capable of operating with benign refrigerants, such as air, argon, xenon, and helium. The Ericsson cycle combines four thermodynamic processes. For an ideal cycle that uses a gas as the working material, the processes are isothermal (constant temperature) compression, constant pressure heat rejection from the high pressure stream to the low pressure stream, isothermal expansion, and constant pressure heat addition to the low pressure stream from the high pressure stream. A system that approximates these processes can be termed an Ericsson device or machine. The Ericsson cycle has several notable advantages. For example, the cycle is thermodynamically reversible, meaning that its coefficient of performance (COP) is theoretically the same as the Carnot COP, which is the maximum efficiency any refrigeration machine can achieve while operating between given temperatures. Another advantage of the Ericsson cycle is that it can use fluid refrigerants that pose no or low environmental risk. Virtually any gas can be used as the working fluid, including the aforementioned air, argon, xenon, and helium as well as other readily available gases such as carbon dioxide. 
   The principle difficulty of implementing a practical device that operates in a manner substantially similar to the Ericsson cycle is the requirement for isothermal or near isothermal compression and expansion of the working fluid to achieve a reasonable efficiency. When a gas is compressed, the temperature of the gas increases. To keep the temperature of the gas constant during compression, the gas must be cooled while it is compressed. In practice, isothermal compression of a gas is extremely difficult to achieve because, for practical compression machines, the area available for heat transfer is very small and the compression process occurs very quickly. Slowing down the compression process or increasing the surface area for heat transfer leads to very large, impractical, and expensive machinery. 
   U.S. Pat. No. 4,984,432 to Corey discloses an Ericsson cycle machine that uses liquid ring compressors to compress and expand a gas-liquid mixture. However, several disadvantages are believed to exist with this machine as disclosed. First, liquid ring compressors have difficulty producing large pressure differentials, which can result in small volumetric capacities and necessitate large equipment to achieve relatively small cooling capacities. Liquid ring compressors also exhibit low efficiencies due in part to high viscous (fluid friction) losses, resulting in tremendous degradation of performance. Furthermore, the power required to pump the liquid through the heat exchanger loops is substantial, with no means disclosed to recover this power. Another shortcoming is that the liquid ring is simultaneously in substantial thermal contact with both the inlet and outlet gas streams, which has the undesirable effect of preheating the suction gas on the compression side and precooling the inlet gas on the expander side and results in higher compression work and lower expander work recovery, respectively. In any event, a thermodynamic analysis of the cycle is not presented in the Corey patent, and attempts to test the disclosed Ericsson cycle machine have failed to achieve a net heat pumping effect. 
   BRIEF SUMMARY OF THE INVENTION 
   The invention pertains to a thermodynamic system that can approximate the Ericsson or Brayton cycles and operated in reverse or forward modes to implement a refrigeration device (e.g., a cooler or heat pump) or engine, respectively. 
   The thermodynamic system includes a device for compressing a first fluid stream containing a first gas-liquid mixture having a sufficient liquid content so that compression of the gas within the first gas-liquid mixture by the compressing device is nearly isothermal, and a device for expanding a second fluid stream containing a second gas-liquid mixture having a sufficient liquid content so that expansion of the gas within the second gas-liquid mixture by the expanding device is nearly isothermal. A heat sink is in thermal communication with at least the liquid of the first gas-liquid mixture for transferring heat therefrom, and a heat source is in thermal communication with at least the liquid of the second gas-liquid mixture for transferring heat thereto. Finally, a device is provided for transferring heat between at least the gas of the first gas-liquid mixture after the first fluid stream exits the compressing device and at least the gas of the second gas-liquid mixture after the second fluid stream exits the expanding device. According to the invention, the compressing and expanding devices are not liquid-ring compressors or expanders, but instead are devices that are very tolerant of liquid flooding, such as scroll-type compressors and expanders. 
   The current invention overcomes the difficulty of achieving isothermal compression and expansion in Ericsson and Brayton cycles (or approximations thereof) by mixing a substantial quantity of liquid into the gas during the compression and expansion processes. Since the liquid is in intimate contact with the gas, and can be injected in the form of a mist to promote contact, excellent heat transfer between the gas and liquid is able to occur. Because the liquid have a larger thermal mass compared to the gas being compressed, the liquid absorbs a large amount of the heat of compression. The temperature of the gas therefore remains nearly constant during the compression process. Benefits to the expansion process are analogous. 
   It should be noted that flooding with liquid will damage most gas compression and expansion machines because, unlike a gas, a liquid is substantially incompressible. Therefore very large forces are produced on compression and expansion machinery if an attempt is made to compress a liquid. However, scroll compressors and expanders have been shown to be very tolerant of liquid flooding when implemented with the thermodynamic system of this invention. Because the volume ratio of a scroll compressor is fixed and relatively small, a scroll compressor is able to accommodate liquid within compression pockets in the compressor. In addition to scroll-type compressors, other types of compressors are believed to be tolerant of liquid flooding, particularly screw compressors. In addition, vane-type rotary compressors can also be configured to accommodate liquid flooding to the extent necessary for use in the present invention. 
   Another advantage of the invention is the ability to use many different liquids in the thermodynamic system, including water, mineral oil, or natural biodegradable oils such as rapeseed oil. One advantage of using an oil as the heat transfer fluid is that it can also be used as the lubricant for mechanical components in the system. In addition, because oils are generally strong dielectrics, their use can be combined in a hermetic system that encloses mechanical components of the system, such as electric motors used to drive the compressor. 
   Other objects and advantages of this invention will be better appreciated from the following detailed description. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
       FIG. 1  is a schematic of a thermodynamic system operating as an Ericsson cycle cooler with liquid flooding in accordance with an embodiment of this invention. 
       FIG. 2  is a schematic of a modified thermodynamic system operating as an Ericsson cycle cooler with liquid flooding in accordance with another embodiment of this invention. 
       FIG. 3  is a schematic of a thermodynamic system similar to that of  FIG. 1 , but modified to operate as an Ericsson cycle engine in accordance with an embodiment of this invention. 
       FIG. 4  is a schematic of a thermodynamic system operating as a Brayton cycle cooler with liquid flooding in accordance with an embodiment of this invention. 
       FIG. 5  is a schematic sectional view of integral compression, expansion, and separation machinery for use with an Ericsson cycle system of this invention. 
       FIG. 6  is a schematic of a modified thermodynamic system operating as an Ericsson cycle cooler with liquid flooding in accordance with an embodiment of this invention. 
       FIGS. 7 and 8  are schematics of Ericsson cycle coolers similar to FIG.  1 , but further equipped with means for equalizing the amount of flooding liquid within two liquid circuits of the system in accordance with an embodiment of this invention. 
       FIG. 9  is a schematic of an Ericsson cycle cooler similar to  FIG. 1 , but further equipped with additional heat exchangers to improve the performance of the system as a heat pump or heat engine in accordance with an additional embodiment of this invention. 
       FIG. 10  is a schematic of an Ericsson cycle cooler similar to  FIG. 1 , but with heat exchangers relocated. 
       FIG. 11  is a schematic of an Ericsson cycle cooler similar to  FIG. 1 , but modified to use ejectors. 
       FIG. 12  is a schematic of a modified Brayton cycle cooler similar to  FIG. 4 . 
   

   DETAILED DESCRIPTION OF THE INVENTION 
   The invention is described in reference to thermodynamic systems that employ Ericsson or Brayton cycles in combination with liquid flooding during compression and expansion of a compressible fluid so that the compression and expansion processes are nearly isothermal. As will be evident from the following, numerous liquids can be used as the flooding liquid and numerous gases can be used as the compressible fluid. Particular examples of suitable compressible fluids include air, argon, xenon, helium, etc., though others could also be used, with a preference for fluids that are not toxic, flammable, ozone-depleting, or potent greenhouse gases. Particular examples of suitable liquids include water, mineral oil, natural biodegradable oils such as rapeseed oil, etc. In some cases, nonvolatile liquids will likely be preferred, though it is believed that the use of a liquid (e.g., water) that partially vaporizes and condenses as it goes through compression and expansion, respectively, would result in more isothermal compression and expansion at lower liquid flooding rates, which has potential advantages. Those skilled in the art will appreciate that suitable temperatures, pressures, etc., for the operation of the systems will depend on the particular liquids and gases used. 
   Those skilled in the art will also appreciate that compressors and expanders suitable for use with the invention must be tolerant of liquid flooding. While scroll-type compressors and expanders will be described in reference to the multiple embodiments of this invention, the thermodynamic system can be implemented using other types of compressors and expanders that are relatively tolerant of liquid flooding, including but not limited to screw compressors, rotary vane compressors, diaphragm compressors, and even rotary and reciprocating piston compressors if sufficient clearance volume is introduced to prevent damage to components. Furthermore, it is foreseeable that centrifugal machine could be modified or designed for this purpose. The construction and operation of such compressors and expanders are well documented in the art, and therefore will not be repeated here. 
   In reference to  FIG. 1 , a thermodynamic system is represented as a reverse Ericsson cycle system  10  that employs a scroll compressor  12  and scroll expander  44 . In  FIG. 1 , a low-pressure high-temperature gas-liquid mixture  14  enters the compressor  12  and is compressed to a higher pressure. (In the Figures, liquid streams are represented by a solid line, gas streams are represented by a dashed line, and gas-liquid mixtures are represented by combined solid and dashed lines.) Because the liquid and gas are in intimate thermal contact during compression and the thermal capacitance of liquids is typically significantly greater than that of the gas within the mixture  14 , the system  10  is operated so that the heat of compression of the gas within the mixture  14  is absorbed by the liquid within the mixture  14 , preferably to the extent that the gas is compressed nearly isothermally. This relationship can be quantified as the capacitance rate ratio (C ratio ), defined as
 
 C   ratio   =m   l   c   l   /m   g   c   p.g. 
 
where m l  and c l  are respectively the mass flow rate and thermal capacitance of the liquid, the product of m l  and c l  is the thermal capacitance rate of the liquid, m g  and c p.g.  are respectively the mass flow rate and thermal capacitance of the gas, and the product of m g.  and c p.g.  is the thermal capacitance rate of the gas. From the above equation, it is evident that the C ratio  of the system  10  will depend on the particular gas and liquid used and their relative amounts. In some cases, the thermal capacitance rate of the liquid may be much greater than that of the gas (i.e., Cratio&gt;&gt;1). However, it is believed that in some cases the system  10  can operate with a C ratio  of approximately 1. In all cases, the relative amount of liquid in the gas-liquid mixture  14  (in other words, the liquid flooding during compression and, as discussed below, during expansion) should be substantial enough to significantly reduce the temperature change of the gas during the compression process (as well as during the expansion process).
 
   The compressor  12  is indicated as being powered (P c ) by an electric motor or other suitable device (not shown). The high-pressure high-temperature gas-liquid mixture  16  exiting the compressor  12  enters a high temperature gas-liquid separator  18 , which can be of a type well known in the art. A high-pressure high-temperature liquid stream  20  and a high-pressure high-temperature gas stream  22  separately exit the separator  18 , from which the liquid stream  20  enters a liquid circuit containing a liquid motor  24  that reduces the pressure of the liquid stream  20  to a relatively low level. Work (P m ) from the liquid motor  24  can be recovered and used to drive the compressor  12  and/or other devices within the system, such as a liquid pump  42  within a second liquid circuit of the system  10 . The resulting low-pressure high-temperature liquid stream  26  exits the liquid motor  24  and enters a high-temperature heat exchanger  28 , where heat from the low-pressure high temperature liquid stream  26  is rejected to a high-temperature sink (Q out ) The resulting low-pressure high-temperature liquid stream  30  exiting the heat exchanger  28  is subsequently mixed with a low-pressure high-temperature gas stream  32  to reform the low-pressure high-temperature gas-liquid mixture  14  delivered to the compressor  12 . 
   The high-pressure high-temperature gas stream  22  separated by the separator  18  enters a regenerator  34 , where heat (Q R ) from the gas stream  22  is rejected to a low-pressure low-temperature gas stream  36  (discussed below). The resulting high-pressure low-temperature gas stream  38  that exits the regenerator  34  preferably has a temperature near that of a refrigerated space  56  cooled by the second liquid circuit of the system  10 . The gas stream  38  mixes with a high-pressure low-temperature liquid stream  40  from the liquid pump  42 , forming a high-pressure low-temperature gas-liquid mixture  46  that enters the scroll expander  44 . Within the expander  44 , the gas-liquid mixture  46  is expanded nearly isothermal as a result of intimate thermal contact between the liquid and gas during expansion and the significantly greater thermal capacitance of the liquid. The expander  44  produces work (P e ) that can be used to provide power for other components of the system  10 , including the compressor  12  and liquid pump  42 , through various known arrangements such as direct shaft coupling. The resulting low-pressure low-temperature gas-liquid mixture  48  then enters a low-temperature gas-liquid separator  50 , which separates the gas-liquid mixture  48  into a low-pressure low-temperature liquid stream  52  and the aforementioned low-pressure low-temperature gas stream  36 . 
   The low-pressure low-temperature liquid stream  52  enters a cold heat exchanger  54 , where the liquid stream  52  absorbs heat from the refrigerated space  56 . The resulting low-pressure low-temperature liquid stream  58  exiting the cold heat exchanger  54  enters the liquid pump  42 , where its pressure is increased to the high system pressure. The low-pressure low-temperature gas stream  36  from the low-temperature gas-liquid separator  50  enters the regenerator  34 , where it absorbs heat from the high-pressure high-temperature gas stream  22  separated by the high-temperature gas-liquid separator  18 . The resulting low-pressure high-temperature gas stream  32  exiting the regenerator  34  is at a temperature near the hot side temperature of the system  10 , i.e., near that of the low-pressure high-temperature liquid stream  30 . 
   The reverse Ericsson cycle of the system  10  operates in a continuous fashion as described. The locations of the liquid motor  24  and liquid pump  42  can be on either side of the heat exchangers  28  and  54 , respectively. Furthermore, the liquid motor  24  can be replaced with a throttling valve (not shown), though with a loss in system performance. If so desired, different liquids can be used in the hot side of the system  10  (to the left of the regenerator  34  in  FIG. 1 ) and cold side of the system  10  (to the right of the regenerator  34  in  FIG. 1 ). As an example, the use of different liquids can be advantageous if the system  10  operates under extreme temperature differentials where a single liquid would either vaporize or solidify at the hot and cold sides, respectively. For instance, a cryogenic application for the system  10  could use water or oil on the hot side of the system  10  and liquid nitrogen on the cold side of the system  10 , with helium or nitrogen used as the gas for the gas loop. 
     FIGS. 2 through 12  depict additional thermodynamic systems in accordance with further embodiments of this invention. For convenience, in these Figures consistent reference numbers are used to identify functionally similar structures. 
     FIG. 2  is an alternative reverse Ericsson cycle system  100  to that of  FIG. 1 , with the primary difference being the elimination of the liquid motor  24  and liquid pump  42 . In this embodiment, the gas-liquid mixture  14  enters the scroll compressor  12 , where it is compressed nearly isothermally and exits the compressor  12  before entering the separator  18 . As before, the separator  18  separates the liquid and gas of the mixture  16 , and the resulting high-pressure high-temperature liquid stream  20  enters the high temperature heat exchanger  28  where heat (Q out ) is rejected from the liquid stream  20  to the high-temperature heat sink. In contrast to the embodiment of  FIG. 1 , the high-pressure high-temperature liquid stream  20  then enters a liquid regenerator  60 , where heat (Q L ) is rejected to the low-pressure low-temperature liquid stream  58  from the low temperature heat exchanger  54 . The high-pressure liquid stream  40  exiting the liquid regenerator  60  is now at a low temperature, and is then mixed with the high-pressure low-temperature gas stream  38  before entering the scroll expander  44 . As before, the resulting high-pressure low-temperature gas-liquid mixture  46  is expanded nearly isothermally by the expander  44 , after which the liquid and gas constituents of the now low-pressure low temperature gas-liquid mixture  48  are separated by the separator  50 . The low-pressure low-temperature liquid stream  52  enters the cold heat exchanger  54 , where the liquid absorbs heat from the refrigerated space  56 . The low-pressure low-temperature liquid stream  58  then enters the liquid regenerator  60 , where it absorbs heat from the high-pressure high-temperature liquid stream  20 . The resulting low-pressure high-temperature liquid stream  30  exits the liquid regenerator  60  and is subsequently mixed with the low-pressure high-temperature gas stream  32  before entering the compressor  12 . In view of the foregoing, the functions of the gas streams  22 ,  32 ,  36 , and  38  are essentially the same as in the embodiment of  FIG. 1 . Work produced by the expander  44  can be used to offset some of the power (P c ) required by the compressor  12 . 
     FIG. 3  is an embodiment of the Ericsson cycle set forth in  FIG. 1 , but operated in a forward mode as a heat engine  200 . Operated as an engine, the system  200  uses the expander-side heat exchanger  54  to absorb heat (Q in ) from a high temperature heat source, with the result that the gas-liquid mixture  48  downstream of the expander is at low pressure but high temperature. On the compressor side, heat (Q out ) is rejected from the liquid stream  26  to a low temperature heat sink, with the result that the gas-liquid mixture  16  downstream of the compressor  12  is at high pressure but low temperature. Because the temperatures of the gas and liquid streams on the expander-side of the system  200  are elevated relative to the gas and liquid streams on the compressor-side of the system  200  (opposite that of  FIG. 1 ), the regenerator  34  operates to transfer heat from the low-pressure high-temperature gas stream  36  downstream of the expander  44  to the high-pressure low-temperature gas stream  22  downstream of the compressor  12 . The expander work (P e ) is greater than the compressor work (P w ) and a net power output is achieved. A portion of the expander work (P e ) can be delivered to the compressor  12  through a shaft (not shown). Otherwise, the individual components of the system  200  in  FIG. 3  operate in a very similar manner to the identical components of the system  10  in  FIG. 1 . 
     FIG. 4  is a schematic of a reverse Brayton cycle system  300  utilizing a scroll compressor  12  and scroll expander  44  with liquid flooding, similar to  FIGS. 1 and 2 . Most notably, the entire system  300  operates on a mixture of gas and liquid, with essentially only pressure and temperature being variable. A low-pressure high-temperature gas-liquid mixture  62  enters the compressor  12 , where it is compressed nearly isothermally. The resulting high-pressure high-temperature gas-liquid mixture  64  enters the hot heat exchanger  28 , where the mixture  64  rejects heat to a high-temperature heat sink. The resulting high-pressure high-temperature gas-liquid mixture  66  exits the heat exchanger  28  and enters the regenerator  34 , where heat (Q R ) is reject to a low-pressure low-temperature gas-liquid mixture  68 . The resulting high-pressure gas-liquid mixture  70 , now at a low temperature, enters the expander  44  where it is expanded nearly isothermally. The resulting low-pressure low-temperature gas-liquid mixture  72  exits the expander  44  then enters the cold heat exchanger  54 , where heat is absorbed from the refrigerated space  56 . The resulting low-pressure low-temperature gas-liquid mixture  68  exits the cold heat exchanger  54  and enters the regenerator  34 , where heat is absorbed from the high-pressure, high-temperature gas-liquid mixture  66 . The work (P e ) produced by the expander  14  can be used to offset some of the work (P c ) required by the compressor  12 . 
   The embodiment of  FIG. 4  is more efficient than a simple reverse Brayton gas cycle because the compression and expansion processes occur nearly isothermally. As with the Ericsson cycle systems  10  and  100  of  FIGS. 1 and 2 , the system  300  of  FIG. 4  can be operated as a heat engine by replacing the refrigerated space  56  with a high-temperature heat source. In this case, the heat exchanger  28  becomes a low temperature heat sink. Notably, the flow direction of the gas-liquid mixture is also reversed. 
     FIG. 5  represents a sectional view of a portion of any one of the Ericsson systems of  FIGS. 1 and 3 , in which the compressor  12 , expander  44 , liquid motor  24 , liquid pump  42 , and separators  18  and  50  are part of an integral unit  74 . The scroll compressor  12  and scroll expander  44  are axially opposed in a common shaft  76 . A motor rotor  78  is located between the compressor  12  and expander  44  on the shaft  76 . The liquid pump  42  and liquid motor  24  are also driven by shaft  76 . The separators  18  and  50  are located on opposite ends of the unit  74 . The remaining components attach to the unit  74  through open-ended connections as shown. The integral unit  74  greatly simplifies the system designs shown schematically in  FIGS. 1 and 3 . By eliminating the liquid motor  24  and liquid pump  42 , the integral unit  74  is further simplified for use with the system  200  shown schematically in  FIG. 2 , and could also be used with additional modifications for implementation of the Brayton cycle system  300  of  FIG. 3 . 
     FIG. 6  shows a schematic representation of an open Ericsson cooler system  400  that operates in the same manner as the system of  FIG. 1 , with the exception that the low-pressure low-temperature gas stream  36  exiting the separator  50  is in fluid communication with the refrigerated space  56 , such that the gas stream  36 A that is passed through the regenerator  34  is drawn from the refrigerated space  56 . 
   In principle, the liquids in the hot and cold loops of the system  10  represented in  FIG. 1  are isolated from each other and do not intermix. In practice, however, the hot-side and cold-side separators  18  and  50  will not be able to entirely remove the liquids from each gas-liquid mixture  16  and  48 , so that the gas streams  22 ,  32 ,  36 , and  38  flowing between both sides of the system  10  will transport liquid from one side to the other. Due to normal manufacturing variations and differences in flow velocities and other conditions between the separators  18  and  50 , it will likely always be the case that gas flowing from one side of the system  10  will contain more liquid than gas flowing from the other side of the system  10 . Under this assumption, after many hours of running, liquid will accumulate on one side of the system  10 . To prevent this, means can be provided for equalizing the liquid between the two sides of the system  10 . While various techniques can be devised to accomplish this, a passive equalization system and an active equalization system are shown for this purpose in  FIGS. 7 and 8 . 
   In  FIG. 7 , flow paths  80  and  82  can be formed by tubes that pierce the separators  18  and  50 , respectively, at the approximate levels that the liquids are to be maintained in the separators  18  and  50 . At the downstream end of the flow path  80 , the tube is in fluid communication with, for example, the gas-liquid mixture  46  upstream of the expander  44 , while the downstream end of the flow path  82  fluidically communicates with, for example, the gas-liquid mixture  14  upstream of the compressor  12 . Because of flow losses, the pressures at the downstream ends of the flow paths  80  and  82  are slightly lower than at the separators  18  and  50 , such that pumps are not required for equalization.  FIG. 8  addresses a situation in which there is a tendency for liquid to accumulate in the separator  18 . The flow path  80  is equipped with a float-type metering valve that opens and allows liquid to flow to the separator  50  when the liquid level in the separator  18  reaches a predetermined level. 
   As previously noted, the compression and expansion processes of the various systems shown in the Figures will be nearly isothermal if sufficient liquid is mixed with the gas during compression and expansion. In practice, however, there will still likely be a temperature rise or drop during flooded compression and expansion, respectively, in which case it can be advantageous to place additional heat exchangers  86  and  88  as shown in  FIG. 9 . The additional heat exchangers  86  and  88  are in an arrangement similar to a Brayton cycle cooler, and serve to improve the performance of an Ericsson cycle system, whether for a heat pump (e.g.,  FIG. 1 ) or a heat engine (e.g.,  FIG. 3 ). 
     FIG. 10  represents a further modification of  FIG. 1  in which the heat exchangers  28  and  54  are relocated directly downstream of the compressor  12  and expander  44 , respectively. These locations allow for additional heat to be rejected (Q out ) and additional heat to be absorbed (Q in ), with the net effect that the coefficient of performance (COP) can be improved for the system  10 . The improvement in COP is possibly such that the system of  FIG. 10  is believed to be a preferred configuration for a reverse Ericsson cycle system of this invention. 
   In another embodiment shown in  FIG. 11 , the liquid motor  24  and pump  42  are replaced with ejectors  90  and  92 , respectively. As known in the art, ejectors use a high pressure fluid stream to compress a low pressure fluid stream to an intermediate pressure. Therefore, in the embodiment of  FIG. 11 , the ejector  90  is employed to reduce the pressure of the liquid stream  30  entering the compressor  12 , and the ejector  92  is employed to increase the pressure of the liquid stream  40  entering the expander  44 . 
   The liquid motor  24  can also be replaced in any of the embodiments with a throttle valve  94  (or other suitable type of flow restriction), as represented in  FIG. 12  with the Brayton cycle system of  FIG. 4 . A flow restriction is a much simpler and lower cost approach than the liquid motor  24 , such as a hydraulic motor. However, system efficiency will decrease since no work is being recovered from the liquid stream as its pressure is reduced with the restriction. 
   In an investigation leading up to this invention, an experimental liquid-flooded Ericsson cooler system corresponding to the system  10  represented in  FIG. 1  was constructed and used to perform tests. In the construction of the system, primarily off-the-shelf parts with very little modifications were used. The sizing of components used in the system was based on preliminary modeling results reported in Hugenroth et al., “Liquid-Flooded Ericsson Cycle Cooler: Part 1-Thermodynamic Analysis,” Proceedings of the 2006 International Refrigeration and Air Conditioning Conference at Purdue, R168, the contents of which are incorporated herein by reference. Open drive scroll compressors were chosen for the compressor and expander in the experimental system. In addition to being readily available at low cost and at the approximate displacement volume desired, a scroll compressor can be operated as an expander by simply reversing the fluid flow through the machine. 
   The experimental system contained the following major components: compressor, expander, hydraulic motor, pump, hot and cold separators, hot and cold mixers, hot and cold heat exchangers, and a regenerator. The heat exchangers were commercially-available units that exchanged heat with an aqueous ethylene glycol coolant supplied by a chiller system. The regenerator was also a commercially available heat exchanger. The separators were custom-built units having a first stage for simple gravity separation of the liquid from the gas, and commercially-available centrifugal type oil separators formed a second stage to separate remaining oil from the gas. Mixing of the liquid and gas streams was accomplished simply by bringing the two streams together at a tee in the lines. Nitrogen and alkyl-benzene oil were used as the refrigerant and flooding liquid, respectively, in the experimental system. 
   The compressor, expander, hydraulic motor, and pump were coupled to electric motors to allow for independent speed control of each component. The expander and hydraulic motor produced power and the electric motors coupled to these components operated regeneratively. Torque cells were placed between the motor shafts and the shaft of each piece of rotating machinery to allow for torque measurements by the power produced or consumed by each component was calculated. Pressure transducers and thermocouples were located between each component in the system, flow in the liquid loops and gas loop were measured, and temperatures and flow rates of coolant flows were measured. 
   Approximately seventy tests were run with the experimental system under a number of conditions. The flooding liquid and compression fluid used in the experiments were alkyl-benzene oil and nitrogen, respectively, and the system was operated to evaluate C ratio  values of about 3.5, 5, 10, and 15. Volumetric capacities of over 110 kJ/m 3  were measured. Though the best second law efficiency was a little over 3%, the low performance for the experimental system was anticipated due to a number of factors, including the large physical size of the system compared to its cooling capacity, and various sources of pressure drops. Details of the results of the experiments are reported in Hugenroth et al., “Liquid-Flooded Ericsson Cycle Cooler: Part 2-Experimental Results,” Proceedings of the 2006 International Refrigeration and Air Conditioning Conference at Purdue, R169, the contents of which are incorporated herein by reference. 
   From the above, it was concluded that scroll compressors could tolerate the necessary amount of liquid flooding required for operation of a reverse Ericsson cycle according to the present invention. In addition, it was concluded that the scroll-type compressor and expander operated reliably under the flooding conditions, and that the adiabatic efficiency of both the compressor and expander were very satisfactory. 
   While the invention has been described in terms of specific embodiments, it is apparent that other forms could be adopted by one skilled in the art. For example, the physical configuration of the thermodynamic systems could differ from that shown in the Figures, and materials and processes other than those noted could be use. Therefore, the scope of the invention is to be limited only by the following claims.

Technology Classification (CPC): 5