Abstract:
A multi-stage refrigeration system is provided. The refrigeration system includes a first compression element which produces a first compressed refrigerant stream. A mixer combines the first compressed refrigerant stream with an auxiliary refrigerant stream. A second compression element is coupled to the mixer and produces a second compressed refrigerant stream. A first heat exchanger receives the second compressed refrigerant stream and generates a cooled stream. A stream splitter receives the cooled stream and provides first and second output streams. A first expansion valve receives the first output stream and controls the flow of the first output stream and a second expansion valve receives the second output stream and controls the flow of the second output stream. A second heat exchanger generates the auxiliary refrigerant stream provided to the mixer. An evaporator is coupled to the first expansion valve and the first compression element to evaporate the first output stream and provide an evaporated stream to the first compression element.

Description:
TECHNICAL FIELD  
       [0001]     This invention relates generally to refrigeration systems, and more particularly, to a multi-stage refrigeration system having main and auxiliary refrigerant streams regulated by control characteristics.  
       BACKGROUND  
       [0002]     A typical multi-stage refrigeration device includes a main refrigerant stream and one or more sub-cycle or auxiliary refrigerant streams. A multi-stage refrigeration device may have improved efficiency compared to a single-stage device because the auxiliary stream cools the main stream while maintaining the high pressure of the main stream (i.e., lower pressure on the suction side makes the compressor work harder). However, the effectiveness of the auxiliary stream in precooling the main stream depends on the performance of the intermediate heat exchanger. In this regard, what is needed is a control methodology to regulate the auxiliary expansion value that controls the flow rate intermediate heat exchanger.  
       SUMMARY  
       [0003]     In one aspect, a refrigerating apparatus includes a compression element, radiator, auxiliary expansion means, intermediate heat exchanger, main expansion means and evaporator constitute a refrigeration cycle, refrigerant flowing out of said radiator is branched into two streams. The first refrigerant stream is passed to the first flow path of the intermediate heat exchanger via said auxiliary expansion means, the second refrigerant stream is passed to the second flow path of the intermediate heat exchanger and then to the evaporator via said main expansion means. Heat exchange is performed between the two refrigerant stream within said intermediate heat exchanger, the refrigerant flowing out of said evaporator is sucked by low pressure part of said compression element, and the refrigerant flowing out of said intermediate heat exchanger is sucked by intermediate pressure part of said compression element. The pressure in said intermediate pressure part of said compression element is determined by controlling said auxiliary expansion means in accordance with the pressure of the suction side and the discharge side of said compression element.  
         [0004]     In another aspect, a refrigerating apparatus includes a compression element, radiator, auxiliary expansion means intermediate heat exchanger, main expansion means and evaporator constitute a refrigeration cycle, refrigerant flowing out of said radiator is branched into two streams. The first refrigerant stream is passed to the first flow path of the intermediate heat exchanger via said auxiliary expansion means, the second refrigerant stream is passed to the second flow path of the intermediate heat exchanger and then to the evaporator via said main expansion means. Heat exchange is performed between the two refrigerant stream within said intermediate heat exchanger, the refrigerant flowing out of said evaporator is sucked by low pressure part of said compression element, and the refrigerant flowing out of said intermediate heat exchanger is sucked by intermediate pressure part of said compression element. The the pressure in said intermediate pressure part of the compression element is controlled to an optimum intermediate pressure by controlling said auxiliary expansion means using an expression Pint,opt=Kint,opt*GMP=Kint,opt*(Psuc*Pdis) 0,5 , wherein, Pint,opt: Optimum intermediate pressure; Kint,opt: Optimum intermediate pressure coefficient; GMP: Geometric mean of the pressure of the high pressure side and the pressure of the low pressure side; Psuc: Pressure of the suction side of the compression element; and Pdis: Pressure of the discharge side of the compression element.  
         [0005]     In a further aspect, a refrigerating apparatus includes a compression element, radiator, auxiliary expansion means, intermediate heat exchanger, main expansion means and evaporator constitute a refrigeration cycle, refrigerant flowing out of said radiator is branched into two streams. The first refrigerant stream is passed to the first flow path of the intermediate heat exchanger via said auxiliary expansion means, the second refrigerant stream is passed to the second flow path of the intermediate heat exchanger and then to the evaporator via said main expansion means. Heat exchange is performed between the two refrigerant stream within said intermediate heat exchanger, the refrigerant flowing out of said evaporator is sucked by low pressure part of said compression element, and the refrigerant flowing out of said intermediate heat exchanger is sucked by intermediate pressure part of said compression element. The pressure in said intermediate pressure part of the compression element being set to an optimum intermediate pressure calculated using an expression Pint,opt=Kint,opt*GMP=Kint,opt*(Psuc*Pdis) 0,5 , wherein, Pint,opt: Optimum intermediate pressure; Kint,opt: Optimum intermediate pressure coefficient; GMP: Geometric mean of the pressure of the high pressure side and the pressure of the low pressure side; Psuc: Pressure of the suction side of the compression element; and Pdis: Pressure of the discharge side of the compression element.  
         [0006]     In another aspect, a refrigerating apparatus includes a compression element, radiator, auxiliary expansion means, intermediate heat exchanger, main expansion means and evaporator constitute a refrigeration cycle, refrigerant flowing out of said radiator is branched into two streams. The first refrigerant stream is passed to the first flow path of the intermediate heat exchanger via said auxiliary expansion means, the second refrigerant stream is passed to the second flow path of the intermediate heat exchanger and then to the evaporator via said main expansion means. Heat exchange is performed between the two refrigerant stream within said intermediate heat exchanger, the refrigerant flowing out of said evaporator is sucked by low pressure part of said compression element, and the refrigerant flowing out of said intermediate heat exchanger is sucked by intermediate pressure part of said compression element. The pressure in said intermediate pressure part of said compression element is determined by controlling said auxiliary expansion means in accordance with the ambient temperature and evaporator temperature.  
         [0007]     In a further aspect, a refrigerating apparatus includes a compression element, radiator, auxiliary expansion means, intermediate heat exchanger, main expansion means and evaporator constitute a refrigeration cycle, refrigerant flowing out of said radiator is branched, into two streams. The first refrigerant stream is passed to the first flow path of the intermediate heat exchanger via said auxiliary expression means, the second refrigerant stream is passed to the second flow path of the intermediate heat exchanger and then to the evaporator via said main expansion means. Heat exchange is performed between the two refrigerant stream within said intermediate heat exchanger, the refrigerant flowing out of said evaporator is sucked by low pressure part of said compression element, and the refrigerant flowing out of said intermediate heat exchanger is sucked by intermediate pressure part of said compression element. The intermediate pressure in the intermediate pressure part of the compression element is controlled to an optimum intermediate pressure by controlling said auxiliary expansion means using an expression z=a+bx+cy+dx2+ey2+fxy, wherein, z: The aimed optimum intermediate pressure; x: Ambient temperature; y: Evaporator temperature; a: coefficient; b: coefficient; c: coefficient; d: coefficient; e: coefficient; and f: coefficient.  
         [0008]     Further features of the invention, its nature and various advantages will be more apparent from the accompanying drawings and the following detailed description. 
     
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0009]     The accompanying drawings illustrate several embodiments of the invention and, together with the description, serve to explain the principles of the invention.  
         [0010]      FIG. 1  is a block diagram illustrating a two stage refrigeration cycle according to an embodiment of the present invention.  
         [0011]      FIG. 2  is a graph illustrating optimized control characteristics for the split cycle according to an embodiment of the present invention.  
         [0012]      FIG. 3  is a graph illustrating split cycle with variable and constant intermediate pressure according to an embodiment of the present invention.  
         [0013]      FIG. 4  is a graph illustrating a curve fit of the optimum intermediate pressure according to an embodiment of the present invention.  
         [0014]      FIG. 5  is a graph illustrating valve orifice area according to an embodiment of the present invention.  
         [0015]      FIG. 6  is a graph illustrating the valve orifice area shown in  FIG. 5  in two-dimensions.  
         [0016]      FIG. 7  is a graph illustrating optimum intermediate pressure Pint,opt according to an embodiment of the present invention.  
         [0017]      FIGS. 8 and 9  illustrate the range of the Optimum intermediate pressure coefficient Kint,opt.  
         [0018]      FIG. 10  illustrates the relationship between volume ratio and COP according to an embodiment of the present invention.  
         [0019]      FIG. 11  illustrates a control value incorporating two expansion valves in one body according to one embodiment of the present invention.  
         [0020]      FIG. 12  is a block diagram illustrating a split cycle configuration with multiple evaporators according to an embodiment of the present invention.  
         [0021]      FIG. 13  is a block diagram illustrating a split cycle configuration according to another embodiment of the present invention.  
         [0022]      FIGS. 14-18  illustrate a multi-stage rotary compressor according to an embodiment of the present invention. 
     
    
     DETAILED DESCRIPTION OF THE EMBODIMENTS  
       [0023]     The present invention is now described more fully with reference to the accompanying figures, in which several embodiments of the invention are shown. The present invention may be embodied in many different forms and should not be construed as limited to the embodiments set forth herein. Rather these embodiments are provided so that this disclosure will be thorough and complete and will fully convey the invention to those skilled in the art.  
         [0024]     A. Split Cycle System  
         [0025]      FIG. 1  is a block diagram illustrating a two stage refrigeration cycle according to an embodiment of the present invention. The split cycle includes a low stage compression element  101 , an intercooler  102 , a mixing device for two fluid streams  103 , a high stage compression element  104 , a gas cooler heat exchanger  105  that cools the fluid stream leaving the high stage compression element by rejecting heat to a second fluid such as air or water, a main expansion valve  106 , an intermediate heat exchanger  107 , an evaporator  108  that evaporates the fluid stream in evaporator in heat exchange with a third fluid such as air or water. The outlet of the evaporator is connected to the low stage compression element suction port. There is further an auxiliary expansion valve  109  that connects the outlet of the gas cooler via the stream splitter  110  to the second path of the intermediate heat exchanger and the outlet of that path to the mixing device  103 .  
         [0026]     In certain embodiments, the system illustrated in  FIG. 1  includes the following features: 
    1. The compression elements may be two separate compressors with separate motors, or may be combined into one unit with one motor or may be achieved by having one compression element with an intermediate suction port (and in that case no intercooler  102 ). In the case of a single compression element, the compressor has an intermediate suction port (intermediate pressure part) between the suction port (low pressure port) and the discharge port, and the refrigerant flowing out of the intermediate heat exchanger is sucked by the intermediate suction port. The preferred embodiment has two separate compression elements with an intercooler.     2. The intercooler may or may not be present. The preferred embodiment uses the intercooler.     3. The intermediate heat exchanger  107  may be arranged in a counter flow fashion or a parallel flow fashion or a mixed counter flow/parallel flow fashion. The preferred embodiment uses counter flow.    
 
         [0030]     The expansion valves are controlled as described below and can be two separate valves or be incorporated into one valve body. The control concepts apply independent of the application of the refrigeration system (e.g., water heating, air-conditioning, heat pumping and refrigeration application) over the entire range of evaporator temperature levels.  
         [0031]     B. Compressor Volume Ratio  
         [0032]     The ratio of the displacement volume of the high side compressor over that of the low side compressor is dependent on the relative mass flow rates and densities at the respective compressor suction ports. The preferred volume ratio is in the range of 0.3 to 1.0. In an another exemplary embodiment, the volume ratio is in the range of 0.5 to 0.8.  
         [0033]     System simulation has shown that the optimum displacement ratio is constant over a wide range of air-conditioning operating conditions. At equal speed of both compressor stages the optimum volume ratio of the stages is 0.76 for the component specifications assumed in the simulation.  FIG. 2  shows the change of the remaining control variables at optimized operating conditions for a range of ambient temperatures.  
         [0034]     While simulation results show that the maximum coefficient of performance (COP) for the Split cycle is reached when the intermediate pressure is adjusted with ambient conditions, the system can be operated close to optimum conditions when the intermediate pressure is constant at an appropriate value. The difference in performance is illustrated in  FIG. 3 .  FIG. 4  shows a curve fit of the optimum intermediate pressure as a function of evaporator and ambient temperatures.  
         [0035]     C. Control Options  
         [0036]     The mass flow rate through the intermediate heat exchanger  107  is controlled in one of the following ways:  
         [0037]     1. First Option  
         [0038]     The auxiliary expansion valve  109  is adjusted such that the intermediate pressure is maintained at a constant value within +/−50% of the value described by the equation shown in  FIG. 4 . In the preferred embodiment, the intermediate pressure may have a value of +/−20% of the one specified in the above equation. It should be noted that the preferred value will depend on the actual design of the system and is a function of other variables such as displacement volume ratio. The above equation serves as an example and covers the entire range of operating conditions.  
         [0039]     The relationship between the operating pressures is expressed as follows: Control the high-side pressure while using the second order linear 6 coefficients equation below, which is a result of curve fitting of high-side pressure. This correlation has a confidence level of 98.9. 
 
 P   dis   =a+b T   amb   +c T   evap   +dT   amb   2   +e T   evap   2   +f T   amb    T   evap   (1) 
 
         [0040]     Where  
         [0041]     a: −1854.91508 b: 334.4838095 c: −98.3269048  
         [0042]     d:−0.60666667 E: 0.932619048 f: 3.522285714  
         [0043]     Then determined the intermediate pressure from Equation 2 with constant value of optimum intermediate pressure coefficient (1.26) such as: 
 
 P   int,opt   =K   INT.OPT   *GMP= 1.26*( P   suc   *P   dis ) 0.5   (2) 
 
         [0044]     The optimum intermediate pressure coefficient is given as 1.26 as the preferred value. Depending on operating conditions and system design, such as compressor displacement volume ratio, the value may vary from 1.1 to 1.6.  
         [0045]     2. Second Option  
         [0046]     The auxiliary expansion valve  109  is a thermostatic expansion valve for the following reason: In the conventional single-stage cycle the refrigerant entering the evaporator has been cooled from the high temperature of the gas cater outlet to the evaporator temperature by evaporating a portion of that refrigerant stream itself. Thus the entering vapor quality is quite high. The portion of refrigerant that was evaporated just of cool itself down is no compressed from the evaporator pressure level all the way to the high side pressure level. However, in the two-stage split cycle, the intermediate heat exchanger  107  has the purpose of precooling the main stream with the aid of the auxiliary stream. The inherent advantage is that the auxiliary stream cools the main stream by providing this cooling at a pressure level that is much higher than the evaporator pressure level and the resulting compressor work for this portion of the overall refrigerant flowrate is reduced considerably, leading to net savings. Thus, the more heat the auxiliary stream removes from the main stream, the better its effectiveness. Since the effectiveness of the auxiliary stream in precooling the main stream depends on the performance of the intermediate heat exchanger  107 , the following control options are described. The auxiliary expansion valve  109  is a thermostatic expansion valve that adjusts the intermediate now rate such that one or more of the following temperatures are maintained constant as described below:  
         [0047]     A. The intermediate heat exchanger  107  is a counter flow heat exchanger: 
        1. The temperature of the auxiliary stream leaving the intermediate heat exchanger  107  is within a certain range of the temperature of the incoming main stream. The actual value depends on whether or not the intermediate heat exchanger  107  is a counter flow heat exchanger and on its size relative to the other system components and the operating conditions of the system. In a preferred embodiment, the temperature is controlled within 5K of the incoming stream. In a second preferred embodiment, the temperature is controlled within 2K of the incoming stream.     2. The temperature of the main stream leaving the intermediate heat exchanger  107  is controlled within a certain range of the temperature of the incoming auxiliary stream. The actual value depends on whether or not the intermediate heat exchanger  107  is a counter flow heat exchanger and on its size relative to the other system components and the operating conditions of the system. In a preferred embodiment, the temperature is controlled within 5K of the incoming stream. In a second preferred embodiment, the temperature is controlled within 2K of the incoming stream.     3. The temperature of the auxiliary stream leaving the intermediate heat exchanger  107  is controlled within a certain range of the temperature of the incoming secondary stream to the gas cooler. The actual value depends on whether or not the intermediate heat exchanger  107  is a counter flow heat exchanger and on its size relative to the other system components and the operating conditions of the system. In a preferred embodiment, the temperature is controlled within 8K of the incoming stream. In a second preferred embodiment, the temperature is controlled within 4K of the incoming stream.     4. The temperature difference between the auxiliary stream leaving the intermediate heat exchanger  107  and the main stream entering that heat exchanger is controlled within a certain predetermined range. The actual value depends on whether or not the intermediate heat exchanger  107  is a counter flow heat exchanger and on its size relative to the other system components and the operating conditions of the system. In a preferred embodiment, the temperature is controlled within 5K of the incoming stream. In a second preferred embodiment, the temperature is controlled within 2K of the incoming stream.     5. The temperature difference between the auxiliary stream entering the intermediate heat exchanger  107  and the main stream leaving that heat exchanger is within a certain predetermined range. The actual value depends on whether or not the intermediate heat exchanger  107  is a counter flow heat exchanger and on its size relative to the other system components and the operating conditions of the system. In a preferred embodiment, the temperature is controlled within 5K of the incoming stream. In a second preferred embodiment, the temperature is controlled within 2K of the incoming stream.        
 
         [0053]     B. The intermediate heat exchanger  107  is a parallel flow heat exchanger: 
        1. The temperature of the auxiliary stream leaving the intermediate heat exchanger  107  is controlled within a certain range of the temperature of the incoming main stream. The actual value depends on whether or not the intermediate heat exchanger  107  is a counter flow heat exchanger and on its size relative to the other system components and the operating conditions of the system. In a preferred embodiment, the temperature is controlled within 12K of the incoming stream. In a second preferred embodiment, the temperature is controlled within 6K of the incoming stream.     2. The temperature of the main stream leaving the intermediate heat exchanger  107  is controlled within a certain range of the temperature of the incoming auxiliary stream. The actual value depends on whether or not the intermediate heat exchanger  107  is a counter flow heat exchanger and on its size relative to the other system components and the operating conditions of the system. In a preferred embodiment, the temperature is controlled within 12K of the incoming stream. In a second preferred embodiment, the temperature is controlled within 6K of the incoming stream.     3. The temperature of the auxiliary stream leaving the intermediate heat exchanger  107  is controlled within a certain range of the temperature of the incoming secondary stream to the gas cooler. The actual value depends on whether or not the intermediate heat exchanger  107  is a counter flow heat exchanger and on its size relative to the other system components and the operating conditions of the system. In a preferred embodiment, the temperature is controlled within 15K of the incoming stream. In a second preferred embodiment, the temperature is controlled within 8K of the incoming stream.     4. The temperature difference between the auxiliary stream leaving the intermediate heat exchanger  107  and the main stream leaving that heat exchanger is controlled within a certain predetermined range. The actual value depends on whether or not the intermediate heat exchanger  107  is a counter flow heat exchanger and on its size relative to the other system components and the operating conditions of the system. In a preferred embodiment, the temperature is controlled within 10K of the incoming stream. In a second preferred embodiment, the temperature is controlled within 5K of the incoming stream. In a third preferred embodiment, the temperature difference is controlled within 2K or less.        
 
         [0058]     3. Third Option  
         [0059]     Constant Orifice Expansion Device for Auxiliary Stream: As one skilled in the art will appreciate, the description above is based on the assumption that the split cycle can be controlled at or close to optimum COP with only 2 active control devices. To investigate the feasibility of replacing the expansion valve by a constant orifice device, the following tasks were conducted. It should be noted that the following analysis has been conducted for a commercially available compressor manufactured by SANYO Electric Co., Ltd. (Osaka, Japan) having a displacement volume ratio 0.576.  
         [0060]     a) Area of Constant Orifice Device  
         [0061]     Area of the constant orifice device was calculated by using Equation 3 for a control valve (ASHRAE Handbook, Fundamentals, 1997, p. 2.11).  
               m   =       C   d     ⁢     A   o     ⁢       C   1     ⁡     (           P   in                 T   in             )       ⁢       1   -       (       P   out       P   in       )         (     k   -   1     )     /   k               ⁢     
     ⁢   Where   ⁢     
     ⁢           Cd   =   0.8           {             discharge   ⁢           ⁢   coefficient   ⁢           ⁢   for     ⁢                       chamfered   ⁢           ⁢   orifice           }               Ao   =       pi   /   4     *     Do   ^   2               {     orifice   ⁢           ⁢   area     }               k   =     CP1   /   CVI             {     ratio   ⁢           ⁢   of   ⁢           ⁢   specific   ⁢           ⁢   heats     }               R   =       8314.41   /   44     ⁢     {       J   /   kg     -   K     }               {     Gas   ⁢           ⁢   constant     }               C1   =       (       (     2   *   k     )     /     (     R   *     (     k   -   1     )       )       )     ^   0.5             {   constant   }                   (   3   )             
 
         [0062]     By using properties of each state point and mass flow rate calculated from the above description, the orifice area is calculated for both sub- and main-cycle at various operating conditions. As shown in Table 1 below, the sub-cycle shows similar orifice area for various conditions: standard deviation is 7.9% of the average value. While the main-cycle shows the orifice area varying over a wide range: standard deviation is 22.6% of the average value. These behaviors are also shown in  FIG. 5 , which indicates that the valve area of the main-cycle decreases linearly with increasing ambient temperature and increasing evaporating temperature, and the valve area of the sub-cycle is approximately constant. The observation shows that it is possible to use a capillary tube or short tube for the sub-cycle expansion device.  
                                                   TABLE 1                           Orifice Area            Tamb[C.]   Tevap [C.]   A orifice subc  [mm 2 ]   A orifice mainc  [mm 2 ]                    35   −20   0.287   0.456       40   −20   0.267   0.413       45   −20   0.292   0.390       35   −15   0.273   0.512       40   −15   0.297   0.474       45   −15   0.311   0.442       35   −10   0.278   0.579       40   −10   0.290   0.531       45   −10   0.309   0.493       35   −5   0.302   0.673       40   −5   0.270   0.591       45   −5   0.256   0.528       35   0   0.284   0.766       40   0   0.266   0.668       45   0   0.270   0.599       35   5   0.223   0.849       40   5   0.256   0.747       45   5   0.276   0.672       Average   [mm 2 ]   0.278   0.577       St. Dev   [%]   7.9   22.6                  
 
         [0063]     b) COP Changes by Using Constant Orifice Device for the Sub-Cycle:  
         [0064]     COP changes by using the constant orifice device for the sub-cycle were investigated. Results are summarized in the following Table. As shown in Table 2, the optimized COPs of the two cases are essentially the same.  
                                                                                                             TABLE 2                           Comparison of Two Control Schemes for Sub-Cycle                TXV Control   ST Control   COP            T   —     amb     T   —     evap     P int     P dis, 2nd         P int     P dis, 2nd         change       [° C.]   [° C.]   [kPa]   [kPa]   COP opt, TXV     [kPa]   [kPa]   COP opt, TXV     [%]                    35   −20   5391   8883   1.695   5362   8968   1.692   −0.2       40   −20   5708   10216   1.419   5778   9921   1.462   3.0       45   −20   5990   11187   1.293   6195   10805   1.287   −0.5       35   −15   5797   8998   1.9   5834   8945   1.898   −0.1       40   −15   6195   10060   1.63   6230   9999   1.629   −0.1       45   −15   6580   11137   1.424   6615   11068   1.423   −0.1       35   −10   6146   9082   2.098   6235   9051   2.132   1.6       40   −10   6646   10182   1.811   6638   10199   1.811   0.0       45   −10   7075   11282   1.569   7050   11341   1.569   0.0       35   −5   6760   8920   2.397   6623   9184   2.397   0.0       40   −5   7050   10405   2.013   7053   10396   2.013   0.0       45   −5   7388   12004   1.715   7496   11625   1.728   0.8       40   0   7497   10507   2.251   7469   10602   2.245   −0.3       45   0   7941   12005   1.905   7952   11959   1.907   0.1       35   5   7388   9369   3.101   7379   9413   3.096   −0.2                  
 
         [0065]     Thus, one skilled in the art will appreciate that an appropriately designed constant orifice expansion device can be applied for the auxiliary stream in a split cycle.  
         [0066]      FIG. 6  illustrates a two-dimensional figure of  FIG. 5 . Main cycle refers to the main expansion valve and the evaporator circuit, and sub cycle refers to the auxiliary expansion circuit.  
         [0067]      FIG. 7  illustrates the Optimum intermediate pressure Pint,opt according to the temperature of the evaporator obtained by simulation.  
         [0068]      FIGS. 8 and 9  illustrate the range of the Optimum intermediate pressure coefficient Kint,opt.  FIG. 8  shows the optimized intermediate pressure coefficient for various conditions. In the illustrated embodiment, the figure indicates that the optimized intermediate pressure coefficient ranges between 1.2 and 1.3.  FIG. 9  shows the relationship between the optimized intermediate pressure coefficient and COP.  
         [0069]      FIG. 10  illustrates the relationship of the ratio of the displacement volume of the high stage compression element  104  to the displacement volume of the low stage compression element  101  and the COP of the present refrigerating apparatus.  
         [0070]     D. Expansion Valve Designs  
         [0071]     Traditionally, two separate Parallel Control Valve expansion valves are used to control the two fluid streams.  FIG. 11  illustrates a control value incorporating two expansion valves in one body according to one embodiment of the present invention. This implies that the auxiliary stream braches off after the intermediate heat exchanger  107 . In  FIG. 11 , both the main and auxiliary streams share the same inlet stream  203 , the high pressure fluid from the intermediate heat exchanger  107  outlet. The valve on the left  201  controls the intermediate mass flow rate using the intermediate pressure  204  or the temperature reading through the bulb  205  as input parameters as described above. The valve on the right  202  controls the high side pressure using its value at port  206  as input.  
         [0072]     E. Other Cycle Configurations  
         [0073]     The control concepts described herein are applicable independently of how many evaporator or gascoolers the cycle employs.  FIG. 12  illustrates an example multiple evaporator system. The system can be used for air conditioning, heating and/or hot water preparation. It employs the split cycle design. For the portion of the split cycle, the same control considerations apply as described above with two added capabilities: (i) The expansion valve for the intermediate pressure EXP.V 2  has a shut-off function built in for those cases where the intermediate flow rate is intended to be zero. (ii) Depending on the operating mode, the intermediate heat exchanger is operated in parallel or counter flow configuration. Thus the control mode and specifications of the valve EXP.V 2  have to be adjusted according to the control algorithms specified above. In particular, the operating modes are as follows: 
        1. Air-conditioning mode: The intermediate heat exchanger  107  is operated in counter flow and the expansion valve EXP.V 2  operated in counter flow mode.     2. Heating mode: The intermediate heat exchanger is operated in parallel mode and the expansion valve EXP.V 2  is operated in parallel mode.     3. Water heating mode: The intermediate heat exchanger is not utilized and the expansion valve EXP.V 2  is shut off.        
 
         [0077]      FIG. 13  illustrates a split cycle system having two evaporators, two main expansion devices and a suction line heat exchanger according to another embodiment of the present invention. This embodiment is suitable for a refrigeration system having two or more compartments which are maintained at different temperatures. For example, this system can be applied to a household refrigerator. Also, this exemplary embodiment can be used for commercial refrigeration systems (e.g., restaurants and stores).  
         [0078]     One evaporator can be higher temperature, for example, suitable for fresh foods, and the other can be lower temperature suitable for frozen foods. The two main expansion devices have a shut-off function so that the refrigerant flows through the two evaporators alternately. When the main expansion valve for high temperature evaporator is closed, the refrigerant flows through the low temperature evaporator. On the contrary, when the main expansion valve for low temperature evaporator is closed, the refrigerant flows through the high temperature evaporator.  
         [0079]     As one skilled in the art will appreciate, the control options described above are also applicable to this embodiment. The openings of the valves are determined by the same algorithm. Using a constant opening expansion device such as a capillary tube is especially suitable for domestic refrigerators because it is a simple method and low cost.  
         [0080]     F. Compressor  
         [0081]     1. Structure  
         [0082]      FIGS. 14-18  illustrate a rotary compressor  10 . The rotary compressor  10  is an internal intermediate pressure type multi-stage compression rotary compressor that uses carbon dioxide (CO 2 ) as its refrigerant. The rotary compressor  10  is constructed of a cylindrical hermetic vessel  12  made of a steel plate, an electromotive unit  14  disposed and accommodated at the upper side of the internal space of the hermetic vessel  12 , and a rotary compression mechanism  18  that is disposed under the electromotive unit  14  and constituted by a low stage compression element  101  and a high stage compression element  104  that are driven by a rotary shaft  16  of the electromotive unit  14 . The height of the rotary compressor  10  of the embodiment 220 mm (outside diameter being 120 mm), the height of the electromotive unit  14  is about 80 mm (the outside diameter thereof being 110 mm), and the height of the rotary compression mechanism  18  is about 70 mm (the outside diameter thereof being 110 mm). The gap between the electromotive unit  14  and the rotary compression mechanism  18  is about 5 mm. The excluded volume of the high stage compression element  104  is set to be smaller than the excluded volume of the low stage compression element  101 .  
         [0083]     The hermetic vessel  12  according to this embodiment is formed of a steel plate having a thickness of 4.5 mm, and has an oil reservoir at its bottom, a vessel main body  12 A for housing the electromotive unit  14  and the rotary compression mechanism  18 , and a substantially bowl-shaped end cap (cover)  12 B for closing the upper opening of the vessel main body  12 A. A round mounting hole  12 D is formed at the center of the top surface of the end cap  12 B, and a terminal (the wire being omitted)  20  for supply power to the electromotive unit  14  is installed to the mounting hole  12 D.  
         [0084]     In this case, the end cap  12 B surrounding the terminal  20  is provided with an annular stepped portion  12 C having a predetermined curvature that is formed by molding. The terminal  20  is constructed of a round glass portion  20 A having electrical terminals  139  penetrating it, and a metallic mounting portion  20 B formed around the glass portion  20 A and extends like a jaw aslant downward and outward. The thickness of the mounting portion  20 B is set to 2.4+0.5 mm. The terminal  20  is secured to the end cap  12 B by inserting the glass portion  20 A from below into the mounting hole  12 D to jut it out to the upper side, and abutting the mounting portion  20 B against the periphery of the mounting hole  12 D, then welding the mounting portion  20 B to the periphery of the mounting hole  12 D of the end cap  12 B.  
         [0085]     The electromotive unit  14  is formed of a stator  22  annularly installed along the inner peripheral surface of the upper space of the hermetic vessel  12  and a rotor  24  inserted in the stator  22  with a slight gap provided therebetween. The rotor  24  is secured to the rotary shaft  16  that passes through the center thereof and extends in the perpendicular direction.  
         [0086]     The stator  22  has a laminate  26  formed of stacked donut-shaped electromagnetic steel plates, and a stator coil  28  wound around the teeth of the laminate  26  by series winding or concentrated winding. As in the case of the stator  22 , the rotor  24  is formed also of a laminate  30  made of electromagnetic steel plates, and a permanent magnet MG is inserted in the laminate  30 .  
         [0087]     An intermediate partitioner  36  is sandwiched between the low stage compression element  101  and the high stage compression element  104 . More specifically, the low stage compression element  101  and the high stage compression element  104  are constructed of the intermediate partitioner  36 , a cylinder  38  and a cylinder  40  disposed on and under the intermediate partitioner  36 , upper and lower rollers  46  and  48  that eccentrically rotate in the upper and lower cylinders  38  and  40  with a 180-degree phase difference by being fitted to upper and lower eccentric portions  42  and  44  provided on the rotary shaft  16 , upper and lower vanes  50  (the lower vane being not shown) that abut against the upper and lower rollers  46  and  48  to partition the interiors of the upper and lower cylinders  38  and  40  into low-pressure chambers and high-pressure chambers, as it will be discussed hereinafter, and an upper supporting member  54  and a lower supporting member  56  serving also as the bearings of the rotary shaft  16  by closing the upper open surface of the upper cylinder  38  and the bottom open surface of the lower cylinder  40 .  
         [0088]     The upper supporting member  54  and the lower supporting member  56  are provided with suction passages  58  and  60  in communication with the interiors of the upper and lower cylinders  38  and  40 , respectively, through suction ports  161  and  162 , and recessed discharge muffling chambers  62  and  64 . The open portions of the two discharge muffling chambers  62  and  64  are closed by covers. More specifically, the discharge muffling chamber  62  is closed by an upper cover  66 , and the discharge muffling chamber  64  is closed by a lower cover  68 .  
         [0089]     In this case, a bearing  54 A is formed upright at the center of the upper supporting member  54 , and a cylindrical bush  122  is installed to the inner surface of the bearing  54 A. Furthermore, a bearing  56 A is formed in a penetrating fashion at the center of the lower supporting member  56 . A cylindrical bush  123  is attached to the inner surface of the bearing  56 A also. These bushes  122  and  123  are made of a material exhibiting good slidability, as it will be discussed hereinafter, and the rotary shaft  16  is retained by a bearing  54 A of the upper supporting member  54  and a bearing  56 A of the lower supporting member  56  through the intermediary of the bushes  122  and  123 .  
         [0090]     In this case, the lower cover  68  is formed of a donut-shaped round steel plate, and secured to the lower supporting member  56  from below by main bolts  129  at four points on its peripheral portion. The lower cover  68  closes the bottom open portion of the discharge muffling chamber  64  in communication with the interior of the lower cylinder  40  of the low stage compression element  101  through a discharge port  41 . The distal ends of the main bolts  129  are screwed to the upper supporting members  54 . The inner periphery of the lower cover  68  projects inward beyond the inner surface of the bearing  56 A of the lower supporting member  56  so as to retain the bottom end surface of the bush  123  by the lower cover  68  to prevent it from coming off.  
         [0091]     The lower supporting member  56  is formed of a ferrous sintered material (or castings), and its surface (lower surface) to which the lower cover  68  is attached is machined to have a flatness of 0.1 mm or less, then subjected to steaming treatment. The steaming treatment causes the ferrous surface to which the lower cover  68  is attached to an iron oxide surface, so that the pores inside the sintered material are closed, leading to improved sealing performance. This obviates the need for providing a gasket between the lower cover  68  and the lower supporting member  56 .  
         [0092]     The discharge muffling chamber  64  and the upper cover  66  at the side adjacent to the electromotive unit  14  in the interior of the hermetic vessel  12  are in communication with each other through a communicating passage  63 , which is a hole passing through the upper and lower cylinders  38  and  40  and the intermediate partitioner  36  ( FIG. 17 ). In this case, an intermediate discharge pipe  121  is provided upright at the upper end of the communicating passage  63 . The intermediate discharge pipe  121  is directed to the gap between adjoining stator coils  28  and  28  wound around the stator  22  of the electromotive unit  14  located above.  
         [0093]     The upper cover  66  closes the upper surface opening of the discharge muffling chamber  62  in communication with the interior of the upper cylinder  38  of the high stage compression element  104  through a discharge port  39 , and partitions the interior of the hermetic vessel  12  to the discharge muffling chamber  62  and a chamber adjacent to the electromotive unit  14 . The upper cover  66  has a thickness of 2 mm or more and 10 mm or less (the thickness being set to the most preferable value, 6 mm, in this embodiment), and is formed of a substantially donut-shaped, circular steel plate having a hole through which the bearing  54 A of the upper supporting member  54  penetrates. With a gasket  124  sandwiched between the upper cover  66  and the upper supporting member  54 , the peripheral portion of the upper cover  66  is secured from above to the upper supporting member  54  by four main bolts  78  through the intermediary of the gasket  124 . The distal ends of the main bolts  78  are screwed to the lower supporting member  56 .  
         [0094]     Setting the thickness of the upper cover  66  to such a dimensional range makes it possible to achieve a reduced size, durability that is sufficiently high to survive the pressure of the discharge muffling chamber  62  that becomes higher than that of the interior of the hermetic vessel  12 , and a secured insulating distance from the electromotive unit  14 .  
         [0095]     The intermediate partitioner  36  that closes the lower open surface of the upper cylinder  38  and the upper open surface of the lower cylinder  40  has a through hole  131  that is located at the position corresponding to the suction side in the upper cylinder  38  and extends from the outer peripheral surface to the inner peripheral surface to establish communication between the outer peripheral surface and the inner peripheral surface thereby to constitute an oil feeding passage. A sealing member  132  is press-fitted to the outer peripheral surface of the through hole  131  to seal the opening in the outer peripheral surface. Furthermore, a communication hole  133  extending upward is formed in the middle of the through hole  131 .  
         [0096]     In addition, a communication hole  134  linked to the communication hole  133  of the intermediate partitioner  36  is opened in the suction port  161  (suction side) of the upper cylinder  38 . The rotary shaft  16  has an oil hole oriented perpendicularly to the axial center and horizontal oil feeding holes  82  and  84  (being also formed in the upper and lower eccentric portions  42  and  44  of the rotary shaft  16 ) in communication with the oil hole. The opening at the inner peripheral surface side of the through hole  131  of the intermediate partitioner  36  is in communication with the oil hole through the intermediary of the oil feeding holes  82  and  84 .  
         [0097]     As it will be discussed hereinafter, the pressure inside the hermetic vessel  12  will be an intermediate pressure, so that it will be difficult to supply oil into the upper cylinder  38  that will have a high pressure due to the second stage. However, the construction of the intermediate partitioner  36  makes it possible to draw up the oil from the oil reservoir at the bottom in the hermetic vessel  12 , lead it up through the oil hole to the oil feeding holes  82  and  84  into the through hole  131  of the intermediate petitioner  36 , and supply the oil to the suction side of the upper cylinder  38  (the suction port  161 ) through the communication holes  133  and  134 .  
         [0098]     As described above, the upper and lower cylinders  38 ,  40 , the intermediate partitioners  36 , the upper and lower supporting members  54 ,  56 , and the upper and lower covers  66 ,  68  are vertically fastened by four main bolts  78  and the main bolts  129 . Furthermore, the upper and lower cylinders  38 ,  40 , the intermediate partitioner  36 , and the upper and lower supporting members  54 ,  56  are fastened by auxiliary bolts  136 ,  136  located outside the main bolts  78 ,  129  ( FIG. 17 ). The auxiliary bolts  136  are inserted from the upper supporting member  54 , and the distal ends thereof are screwed to the lower supporting member  56 .  
         [0099]     The auxiliary bolts  136  are positioned in the vicinity of a guide groove  70  (to be discussed later) of the foregoing vane  50 . The addition of the auxiliary bolts  136 ,  136  to integrate the rotary compression mechanism  18  secures the sealing performance against an extremely high internal pressure. Moreover, the fastening is effected in the vicinity of the guide groove  70  of the vane  50 , thus making it possible to also prevent the leakage of the high back pressure (the pressure in a back pressure chamber  201 ) applied to the vane  50 , as it will be discussed hereinafter.  
         [0100]     The upper cylinder  38  incorporates a guide groove  70  accommodating the vane  50 , and an housing portion  70 A for housing a spring  76  positioned outside the guide groove  70 , the housing portion  70 A being opened to the guide groove  70  and the hermetic vessel  12  or the vessel main body  12 A. The spring  76  abuts against the outer end portion of the vane  50  to constantly urge the vane  50  toward the roller  46 . A metallic plug  137  is press-fitted through the opening at the outer side (adjacent to the hermetic vessel  12 ) of the housing portion  70 A into the housing portion  70 A for the spring  76  at the end adjacent to the hermetic vessel  12 . The plug  137  functions to prevent the spring  76  from coming off.  
         [0101]     In this case, the outside diameter of the plug  137  is set to value that does not cause the upper cylinder  38  to deform when the plug  137  is press-fitted into the housing portion  70 A, while the value is larger than the inside diameter of the housing portion  70 A at the same time. More specifically, in the embodiment, the outside diameter of the plug  137  is designed to be larger than the inside diameter of the housing portion  70 A by 4 μm to 23 μm. An O-ring  138  for sealing the gap between the plug  137  and the inner surface of the housing portion  70 A is attached to the peripheral surface of the plug  137 .  
         [0102]     In this case, as the refrigerant, the foregoing carbon dioxide (CO 2 ), an example of carbonic acid gas, which is a natural refrigerant is used primarily because it is gentle to the earth and less flammable and toxic. For the oil functioning as a lubricant, an existing oil, such as mineral oil, alkylbenaene oil, ether oil, or ester oil is used.  
         [0103]     On a side surface of the vessel main body  12 A of the hermetic vessel  12 , sleeves  141 ,  142 ,  143 , and  144  are respectively fixed by welding at the positions corresponding to the positions of the suction passages  58  and  60  of the upper supporting member  54  and the lower supporting member  56 , the discharge muffling chamber  62 , and the upper side of the upper cover  66  (the position substantially corresponding to the bottom end of the electromotive unit  14 ). The sleeves  141  and  142  are vertically adjacent, and the sleeve  143  is located on a substantially diagonal line of the sleeve  141 . The sleeve  144  is located at a position shifted substantially 90 degrees from the sleeve  141 .  
         [0104]     One end of a refrigerant introducing pipe  92  for leading a refrigerant gas into the upper cylinder  38  is inserted into the sleeve  141 , and the one end of the refrigerant introducing pipe  92  is in communication with the suction passage  58  of the upper cylinder  38 . The other end of the refrigerant introducing pipe  92  is connected to the bottom end of a flow combiner  146 . The one end of the pipe  95  and  100  are connected to the upper end of the flow combiner  146 . And the other end of the pipe  95  connected to the sleeve  144  via the intercooler  102  ( FIG. 1 ) to be in communication with the interior of the hermetic vessel  12 .  
         [0105]     Furthermore, one end of a refrigerant introducing pipe  94  for leading a refrigerant gas into the lower cylinder  40  is inserted in and connected to the sleeve  142 , and the one end of the refrigerant introducing pipe  94  is in communication with the suction passage  60  of the lower cylinder  40 . The other end of the pipe  94  is connected to the evaporator  108  ( FIG. 1 ). A refrigerant discharge pipe  96  is inserted in and connected to the sleeve  143 , and one end of the refrigerant discharge pipe  96  is in communication with the discharge muffling chamber  62 . The other end of the pipe  96  is connected to the gas cooler heat exchanger  105  ( FIG. 1 ).  
         [0106]     Furthermore, collars  151  with which couplers for pipe connection can be engaged are disposed around the outer surfaces of the sleeves  141 ,  143 , and  144 . The inner surface of the sleeve,  142  is provided with a thread groove  152  for pipe connection. This allows the couplers for test pipes to be easily connected to the collars  151  of the sleeves  141 ,  143 , and  144  to carry out an airtightness test in the final inspection in the manufacturing process of the compressor  10 . In addition, the thread groove  152  allows a test pipe to be easily screwed into the sleeve  142 . Especially in the case of the vertically adjoining sleeves  141  and  142 , the sleeve  141  has the collar  151 , while the sleeve  142  has a thread groove  152 , so that test pipes can be connected to the sleeves  141  and  142  in a small space.  
         [0107]     2. Operation  
         [0108]     The descriptions will now be given of the operation. A controller controls the number of revolutions of the electromotive unit  14  of the rotary compressor  10 . The moment the stator coil  28  of the electromotive unit  14  is energized through the intermediary of the terminal  20  and a wire (not shown) by the controller, the electromotive unit  14  is started and the rotor  24  rotates. This causes the upper and lower rollers  46  and  48  fitted to the upper and lower eccentric portions  42  and  44  provided integrally with the rotary shaft  16  to eccentrically rotate in the upper and lower cylinders  38  and  40 .  
         [0109]     Thus, a low-pressure refrigerant gas (1st-stage suction pressure LP: 4 MPaG) that has been introduced into a low-pressure chamber of the lower cylinder  40  from a suction port  162  via the refrigerant introducing pipe  94  and the suction passage  60  formed in the lower supporting member  56  is compressed by the roller  48  and the vane in operation to obtain an intermediate pressure (MP 1 :8 MPaG). The refrigerant gas of the intermediate pressure leaves the high-pressure chamber of the lower cylinder  40 , passes through the discharge port  41 , the discharge muffling chamber  64  provided in the lower supporting member  56 , and the communication passage  63 , and is discharged into the hermetic vessel  12  from the intermediate discharge pipe  121 .  
         [0110]     At this time, the intermediate discharge pipe  121  is directed toward the gap between the adjoining stator coils  28  and  28  wound around the stator  22  of the electromotive unit  14  thereabove; hence, the refrigerant gas still having a relatively low temperature can be positively supplied toward the electromotive unit  14 , thus restraining a temperature rise in the electromotive unit  14 . At the same time, the pressure inside the hermetic vessel  12  reaches the intermediate pressure (MP 1 ).  
         [0111]     The intermediate-pressure refrigerant gas in the hermetic vessel  12  comes out of the sleeve  144  at the above intermediate pressure (MP 1 ), passes through the pipe  95  and the intercooler  102  ( FIG. 1 ), and is combined with the refrigerant from the intermediate heat exchanger  107  ( FIG. 1 ) through the pipe  100 .  
         [0112]     The combined refrigerant in the flow combiner  146  flow out from the bottom end, passes through the pipe  92  and the suction passage  58  formed in the upper supporting member  54 , and is drawn into the low-pressure chamber (2nd-stage suction pressure being MP 2 ) of the upper cylinder  38  through a suction port  161 . The intermediate-pressure refrigerant gas that has been drawn in is subjected to a second-stage compression by the roller  46  and the vane  50  in operation so as to be turned into a hot high-pressure refrigerant gas (2nd-stage discharge pressure HP: 12 MPaG). The hot high-pressure refrigerant gas leaves the high-pressure chamber, passes through the discharge port  39 , the discharge muffling chamber  62  provided in the upper supporting member  54 , and the refrigerant discharge pipe  96 .  
         [0113]     Having described embodiments of multi-stage refrigeration system including sub-cycle control characteristics (which are intended to be illustrative and not limiting), it is noted that modifications and variations can be made by persons skilled in the art in light of the above teachings. It is therefore to be understood that changes may be made in the particular embodiments of the invention disclosed that are within the scope and spirit of the invention as defined by the appended claims and equivalents.