Patent Publication Number: US-8985985-B2

Title: Rotary compressor and refrigeration cycle apparatus

Description:
TECHNICAL FIELD 
     The present invention relates to a rotary compressor and a refrigeration cycle apparatus. 
     BACKGROUND ART 
     It is known that the efficiency of a refrigeration cycle apparatus is increased by injecting a gas phase refrigerant having an intermediate pressure into a compressor (see Patent Literature 1). With this technique, since the work of the compressor and the pressure loss of the refrigerant in an evaporator can be reduced, the coefficient of performance (COP) of the refrigeration cycle is improved. 
     As a compressor that can be applied to the injection technique, a rolling piston compressor provided with a plurality of vanes (blades) so as to form a first compression chamber and a second compression chamber within a cylinder has been proposed (see Patent Literature 2). 
       FIG. 15  is a configuration diagram of a heat pump type heating apparatus described in FIG. 3 of Patent Literature 2. A heat pump type heating apparatus  500  includes a rolling piston compressor  501 , a condenser  503 , an expansion mechanism  504 , a gas-liquid separator  507 , and an evaporator  509 , and is configured to compress a gas phase refrigerant from the evaporator  509  and an intermediate pressure gas phase refrigerant separated in the gas-liquid separator  507 , respectively, in the compressor  501 . Vanes  525  and  535  attached to a cylinder  522  of the compressor  501  divide the space between the cylinder  522  and a rotor  523  into a main compression chamber  526  and an auxiliary compression chamber  527 . The main compression chamber  526  has a suction port  526   a  and a discharge port  526   b . The auxiliary compression chamber  527  has a suction port  527   a  and a discharge port  527   b . The suction port  526   a  is connected to the evaporator  509 , and the suction port  527   a  is connected to the gas-liquid separator  507 . The discharge port  526   b  and the discharge port  527   b  are merged together and connected to the condenser  503 . 
     CITATION LIST 
     Patent Literature 
     
         
         Patent Literature 1 JP 2006-112753 A 
         Patent Literature 2 JP 03(1991)-53532 B 
       
    
     SUMMARY OF INVENTION 
     Technical Problem 
     The present inventors have studied in detail the heat pump type heating apparatus  500  described in Patent Literature 2 to determine whether it can be practically used. As a result, they have ascertained that the compressor  501  has the following technical problems. 
     First, as shown in  FIG. 16 , in a conventional rolling piston compressor having only one vane, a force to press a vane  540  against a piston  543  is generated mainly due to a difference between a pressure applied to a front surface  541  of the vane  540  and a pressure applied to a rear surface  542  thereof. If the compressor is a high-pressure shell type compressor, a pressure equal to a discharge pressure (high pressure) is applied to the rear surface  542  of the vane  540 . The vane  540  has the front surface  541  having an arc shape in plan view, and is in contact with the piston  543  at the front surface  541 . When only one vane  540  is provided in one cylinder, the right side of the front surface  541  with respect to the point of contact between the vane  540  and the piston  543  is always exposed to a suction pressure (low pressure) from a suction port  544 . The left side of the front surface  541  is exposed to a pressure that varies between the suction pressure (low pressure) and the discharge pressure (high pressure). Even when the left side of the front surface  541  is exposed to the discharge pressure (high pressure), the right side of the front surface  541  is always exposed to the suction pressure (low pressure), and thus a sufficient pressure difference is maintained between the front surface  541  and the rear surface  542 . Therefore, a force great enough to press the vane  540  against the piston  543  is always applied to the vane  540 . 
     On the other hand, in a rolling piston compressor  501  described in Patent Literature 2, two vanes are provided in one cylinder. Pressing forces applied to the two vanes are discussed based on the same logic applied to a rolling piston compressor having only one vane. As shown in  FIG. 15 , one side of the front surface of the vane  525  is always exposed to a suction pressure (low pressure) from the suction port  526   a . The other side of the front surface of the vane  525  is exposed to a pressure in the auxiliary compression chamber  527 . The pressure in the auxiliary compression chamber  527  varies between a pressure (intermediate pressure) of a gas phase refrigerant separated in the gas-liquid separator  507  and a discharge pressure (high pressure). Therefore, if it is assumed that the rolling piston compressor  501  is a high-pressure shell type compressor, a force great enough to press the vane  525  against the piston  523  is applied to the vane  525 . 
     Next, one side of the front surface of the vane  535  is always exposed to a suction pressure from the suction port  527   a , that is, the pressure (intermediate pressure) of the gas phase refrigerant separated in the gas-liquid separator  507 . The other side of the front surface of the vane  535  is exposed to a pressure in the main compression chamber  526 . The pressure in the main compression chamber  526  varies between the suction pressure (low pressure) and the discharge pressure (high pressure). Therefore, the pressing force applied to the vane  535  (minimum pressing force) is less than the pressing force applied to the vane  525  and that applied to the vane  540  of the conventional rolling piston compressor. 
     If the pressing force applied to the vane is small, a malfunction called “vane jumping” may occur. As stated herein, “vane jumping” means a phenomenon in which the tip of the vane loses contact with the piston. Vane jumping may cause a significant decrease in the compressor efficiency. 
     It is an object of the present invention to prevent vane jumping in a rotary compressor that can be applied to the injection technique. 
     Solution to Problem 
     The present invention provides a rotary compressor including: a cylinder; a piston disposed within the cylinder so as to form a space between the piston itself and the cylinder; a shaft to which the piston is fitted; a first vane for dividing the space along a circumferential direction of the piston, the first vane being attached to the cylinder at a first angular position along a rotation direction of the shaft; and a second vane for further dividing the space divided by the first vane along the circumferential direction of the piston so that a first compression chamber and a second compression chamber having a smaller volume than the first compression chamber are formed within the cylinder, the second vane being attached to the cylinder at a second angular position along the rotation direction of the shaft. The piston and the second vane are integrated together or the piston and the second vane are coupled together. 
     In a preferred embodiment, the rotary compressor of the present invention further includes: a first suction port for introducing a working fluid to be compressed in the first compression chamber into the first compression chamber; a first discharge port for discharging the working fluid compressed in the first compression chamber outside the first compression chamber from the first compression chamber; a second suction port for introducing the working fluid to be compressed in the second compression chamber into the second compression chamber; a second discharge port for discharging the working fluid compressed in the second compression chamber outside the second compression chamber from the second compression chamber; and a suction check valve provided in the second suction port. 
     In another aspect, the present invention provides a refrigeration cycle apparatus including: the rotary compressor according to the preferred embodiment; a radiator for cooling the working fluid compressed in the rotary compressor; an expansion mechanism for expanding the working fluid cooled in the radiator; a gas-liquid separator for separating the working fluid expanded in the expansion mechanism into a gas phase working fluid and a liquid phase working fluid; an evaporator for evaporating the liquid phase working fluid separated in the gas-liquid separator; a suction flow path for introducing the working fluid that has flowed out of the evaporator into the first suction port of the rotary compressor; and an injection flow path for introducing the gas phase working fluid separated in the gas-liquid separator into the second suction port of the rotary compressor. 
     Advantageous Effects of Invention 
     In the rotary compressor of the present invention, the piston and the second vane are integrated together, or the piston and the second vane are coupled together. In this case, there is essentially no problem of vane jumping. Therefore, the present invention can provide a rotary compressor with a high compressor efficiency, in which vane jumping could never occur. A refrigeration cycle apparatus using the rotary compressor of the present invention can enjoy the benefit of a high injection effect. 
    
    
     
       BRIEF DESCRIPTION OF DRAWINGS 
         FIG. 1  is a configuration diagram of a refrigeration cycle apparatus according to a first embodiment of the present invention. 
         FIG. 2  is a longitudinal cross-sectional view of a rotary compressor used in the refrigeration cycle apparatus shown in  FIG. 1 . 
         FIG. 3  is a transverse cross-sectional view of the rotary compressor shown in  FIG. 2 , taken along the line A-A. 
         FIG. 4A  is a schematic plan view showing a structure for preventing vane jumping. 
         FIG. 4B  is a schematic plan view showing another structure for preventing vane jumping. 
         FIG. 4C  is a schematic plan view showing still another structure for preventing vane jumping. 
         FIG. 4D  is a schematic plan view showing still another structure for preventing vane jumping. 
         FIG. 4E  is a schematic plan view showing still another structure for preventing vane jumping. 
         FIG. 4F  is a schematic plan view showing still another structure for preventing vane jumping. 
         FIG. 5  is an enlarged cross-sectional view of a suction check valve. 
         FIG. 6A  shows side and plan views of a valve body. 
         FIG. 6B  shows side and plan views of a valve stopper. 
         FIG. 7  is a perspective view of a compression mechanism. 
         FIG. 8  is a schematic diagram showing the operation of the rotary compressor with the rotation angle of a shaft. 
         FIG. 9A  is a PV diagram of a first compression chamber. 
         FIG. 9B  is a PV diagram of a second compression chamber. 
         FIG. 10  is a PV diagram of the second compression chamber showing the compression work that can be reduced by injection. 
         FIG. 11A  is a schematic diagram showing the operation of a rotary compressor provided with no suction check valve. 
         FIG. 11B  is a PV diagram of a second compression chamber shown in  FIG. 11A . 
         FIG. 12  is a schematic diagram showing a modification designed to have an obtuse angle between a first vane and a second vane. 
         FIG. 13  is a longitudinal cross-sectional view of a rotary compressor according to a modification. 
         FIG. 14  is a transverse cross-sectional view of the rotary compressor shown in  FIG. 13 , taken along the line B-B. 
         FIG. 15  is a configuration diagram of a conventional heat pump type heating apparatus. 
         FIG. 16  is a transverse cross-sectional view of a conventional rolling piston compressor having only one vane. 
         FIG. 17  is a schematic diagram showing a problem that may occur when a second vane is not coupled with a piston. 
     
    
    
     DESCRIPTION OF EMBODIMENTS 
     Hereinafter, embodiments of the present invention will be described with reference to the accompanying drawings. The present invention is not limited by the embodiments described below. The embodiments and modifications can be combined with one another, without departing from the spirit and scope of the invention. 
     First Embodiment 
       FIG. 1  is a configuration diagram of a refrigeration cycle apparatus according to the present embodiment. A refrigeration cycle apparatus  100  includes a rotary compressor  102 , a first heat exchanger  104 , a first expansion mechanism  106 , a gas-liquid separator  108 , a second expansion mechanism  110 , and a second heat exchanger  112 . These components are connected in a loop in this order by flow paths  10   a  to  10   d  so as to form a refrigerant circuit  10 . The flow paths  10   a  to  10   d  are typically constituted by refrigerant pipes. The refrigerant circuit  10  is filled with a refrigerant, such as hydrofluorocarbon or carbon dioxide, as a working fluid. 
     The refrigeration cycle apparatus  100  further includes an injection flow path  10   j . The injection flow path  10   j  has one end connected to the gas-liquid separator  108  and the other end connected to the rotary compressor  102 , and introduces a gas phase refrigerant separated in the gas-liquid separator  108  directly into the rotary compressor  102 . The injection flow path  10   j  is typically constituted by a refrigerant pipe. A pressure reducing valve may be provided in the injection flow path  10   j . An accumulator may be provided in the injection flow path  10   j.    
     A four-way valve  116 , as a switching mechanism capable of switching the flow direction of the refrigerant, is provided in the refrigerant circuit  10 . When the four-way valve  116  is controlled as indicated by solid lines in  FIG. 1 , the refrigerant compressed in the rotary compressor  102  is supplied to the first heat exchanger  104 . In this case, the first heat exchanger  104  functions as a radiator (condenser) for cooling the refrigerant compressed in the rotary compressor  102 . The second heat exchanger  112  functions as an evaporator for evaporating a liquid phase refrigerant separated in the gas-liquid separator  108 . On the other hand, when the four-way valve  116  is controlled as indicated by dashed lines in  FIG. 1 , the refrigerant compressed in the rotary compressor  102  is supplied to the second heat exchanger  112 . In this case, the first heat exchanger  104  functions as an evaporator and the second heat exchanger  112  functions as a radiator. The four-way valve  116  allows, for example, an air conditioner using the refrigeration cycle apparatus  100  to have both cooling and heating functions. 
     The rotary compressor  102  is a device for compressing the refrigerant to a high temperature and high pressure state. The rotary compressor  102  has a first suction port  19  (main suction port) and a second suction port  20  (injection suction port). The flow path  10   d  is connected to the first suction port  19  so that the refrigerant that has flowed out of the first heat exchanger  104  or the second heat exchanger  112  is introduced into the rotary compressor  102 . The injection path  10   j  is connected to the second suction port  20  so that the gas refrigerant separated in the gas-liquid separator  108  is introduced into the rotary compressor  102 . 
     The first heat exchanger  104  is typically constituted by an air-refrigerant heat exchanger or a water-refrigerant heat exchanger. The second heat exchanger  112  also is typically constituted by an air-refrigerant heat exchanger or a water-refrigerant heat exchanger. When the refrigeration cycle apparatus  100  is used for an air conditioner, both the first heat exchanger  104  and the second heat exchanger  112  are constituted by air-refrigerant heat exchangers. When the refrigeration cycle apparatus  100  is used for a water heater or a hot water heater, the first heat exchanger  104  is constituted by a water-refrigerant heat exchanger, and the second heat exchanger  112  is constituted by an air-refrigerant heat exchanger. 
     The first expansion mechanism  106  and the second expansion mechanism  110  are devices for expanding the refrigerant cooled in the first heat exchanger  104  (or the second heat exchanger  112 ) as a radiator or the liquid phase refrigerant separated in the gas-liquid separator  108 . The first expansion mechanism  106  and the second expansion mechanism  110  are typically constituted by expansion valves. A preferred expansion valve is an opening adjustable valve, such as, for example, an electronic expansion valve. The first expansion mechanism  106  is provided in the flow path  10   b  between the first heat exchanger  104  and the gas-liquid separator  108 . The second expansion mechanism  110  is provided in the flow path  10   c  between the gas-liquid separator  108  and the second heat exchanger  112 . The expansion mechanisms  106  and  110  each may be constituted by a positive displacement expander capable of recovering power from the refrigerant. 
     The gas-liquid separator  108  separates the refrigerant expanded in the first expansion mechanism  106  or the second expansion mechanism  110  into a gas phase refrigerant and a liquid phase refrigerant. The gas-liquid separator  108  is provided with an inlet for the refrigerant expanded in the first expansion mechanism  106  or the second expansion mechanism  110 , an outlet for the liquid phase refrigerant, and an outlet for the gas phase refrigerant. One end of the injection flow path  10   j  is connected to the outlet for the gas phase refrigerant. 
     Other devices such as an accumulator and an internal heat exchanger may be provided in the refrigerant circuit  10 . 
       FIG. 2  is a longitudinal cross-sectional view of the rotary compressor  102  used in the refrigeration cycle apparatus  100  shown in  FIG. 1 .  FIG. 3  is a transverse cross-sectional view of the rotary compressor  102  shown in  FIG. 2 , taken along the line A-A. The rotary compressor  102  includes a closed casing  1 , a motor  2 , a compression mechanism  3 , and a shaft  4 . The compression mechanism  3  is disposed in the lower part of the closed casing  1 . The motor  2  is disposed above the compression mechanism  3  in the closed casing  1 . The compression mechanism  3  and the motor  2  are coupled by the shaft  4 . A terminal  21  for supplying electric power to the motor  2  is provided on the top of the closed casing  1 . An oil reservoir  22  for holding lubricating oil is formed in the bottom of the closed casing  1 . 
     The motor  2  is constituted by a stator  17  and a rotor  18 . The stator  17  is fixed to the inner wall of the closed casing  1 . The rotor  18  is fixed to the shaft  4  and rotates together with the shaft  4 . 
     A discharge pipe  11  is provided in the top wall of the closed casing  1 . The discharge pipe  11  penetrates the top wall of the closed casing  1  and opens into an internal space  13  of the closed casing  1 . The discharge pipe  11  serves as a discharge flow path for discharging the refrigerant compressed in the compression mechanism  3  outside the closed casing  1 . That is, the discharge pipe  11  constitutes a part of the flow path  10   a  shown in  FIG. 1 . During the operation of the rotary compressor  102 , the internal space  13  of the closed casing  1  is filled with the compressed refrigerant. That is, the rotary compressor  102  is a high-pressure shell type compressor. In the high-pressure shell type rotary compressor  102 , since the motor  2  can be cooled by the refrigerant, an increase in the motor efficiency can be expected. When the refrigerant is heated by the motor  2 , the heating capability of the refrigeration cycle apparatus  100  also is increased. 
     The compression mechanism  3  is driven by the motor  2  to compress the refrigerant. As shown in  FIG. 2  and  FIG. 3 , the compression mechanism  3  has a cylinder  5 , a main bearing  6 , an auxiliary bearing  7 , a piston  8 , a muffler  9 , a first vane  32 , a second vane  33 , a first discharge valve  43 , a second discharge valve  44 , and a suction check valve  50 . In the present embodiment, only the second suction port  20  of the first and second suction ports  19  and  20  is provided with the suction check valve  50 . 
     The shaft  4  has an eccentric portion  4   a  projecting outwardly in a radial direction. The piston  8  is disposed within the cylinder  5 . Within the cylinder  5 , the piston  8  is fitted to the eccentric portion  4   a  of the shaft  4 . A first vane groove  34  and a second vane groove  35  are formed in the cylinder  5 . The first vane groove  34  is formed at a first angular position along the rotation direction of the shaft  4 . The second vane groove  35  is formed at a second angular position along the rotation direction of the shaft  4 . 
     A first vane  32  (blade) having a tip in contact with the outer peripheral surface of the piston  8  is slidably fitted in the first vane groove  34 . The first vane  32  divides the space between the cylinder  5  and the piston  8  along the circumferential direction of the piston  8 . A second vane  33  (blade) is slidably fitted in the second vane groove  35 . The second vane  33  further divides the space between the cylinder  5  and the piston  8  along the circumferential direction of the piston  8 . Thereby, a first compression chamber  25  (main compression chamber) and a second compression chamber  26  (injection compression chamber) having a smaller volume than the first compression chamber  25  are formed within the cylinder  5 . 
     A first spring  36  pressing the first vane  32  toward the center of the shaft  4  is disposed behind the first vane  32 . The rear end of the first vane groove  34  is in communication with the internal space  13  of the closed casing  1 . Therefore, the pressure in the internal space  13  of the closed casing  1  is applied to the rear surface of the first vane  32 . The second vane  33  is coupled to the piston  8 . Therefore, no spring is disposed behind the second vane  33 . However, a spring may be disposed behind the second vane  33 . The second vane groove  35  also is in communication with the internal space  13  of the closed casing  1 . Lubricating oil stored in the oil reservoir  22  is supplied to the first vane groove  34  and the second vane groove  35 . 
     In the present description, the position of the first vane  32  and the first vane groove  34  is defined as a position of “0 degrees (a first angle)” along the rotation direction of the shaft  4 . In other words, the rotation angle of the shaft  4  at the moment when the first vane  32  is pushed all the way into the first vane groove  34  by the piston  8  is defined as “0 degrees”. The rotation angle of the shaft  4  at the moment when the second vane  33  is pushed all the way into the second vane groove  35  by the piston  8  corresponds to “a second angle”. In the present embodiment, the angle θ (degrees) from the first angular position where the first vane  32  is disposed to the second angular position where the second vane  33  is disposed is, for example, in the range of 270 to 350 degrees in the rotation direction of the shaft  4 . In other words, the angle (360-θ) between the first vane  32  and the second vane  33  is in the range of 10 to 90 degrees. When the angle θ is 270 degrees or more, the amount of refrigerant flowing back into the first suction pipe  14  from the first compression chamber  25  through the first suction port  19  is small enough for the suction process of the first compression chamber  25 . Therefore, there is no need to provide a check valve in the first suction port  19 . 
     In the present embodiment, the piston  8  is provided with a recessed portion  8   s , and the second vane  33  is provided with a projecting portion  33   t . The projecting portion  33   t  of the second vane  33  is fitted in the recessed portion  8   s  of the piston  8  so that the piston  8  and the second vane  33  are coupled together. Since the piston  8  and the second vane  33  are coupled together, the second vane  33  always follows the movement of the piston  8 . Therefore, there is substantially no problem of vane jumping of the second vane  33 . 
     As shown in  FIG. 4A , the second vane  33  includes a sliding portion  33   a  fitted in the second vane groove  35  and the projecting portion  33   t  located at the tip of the sliding portion  33   a . The projecting portion  33   t  has a circular shape in plan view. The recessed portion  8   s  of the piston  8  in which the projecting portion  33   t  is fitted also has a circular shape in plan view. The projecting portion  33   t  and the recessed portion  8   s  can rotate relatively to each other while maintaining the coupling of the second vane  33  and the piston  8 . When the piston  8  rotates, the second vane  33  slides in the second vane groove  35 . In addition, the projecting portion  33   t  of the second vane  33  rotates in the recessed portion  8   s  of the piston  8 . 
     The width 1  of the projecting portion  33   t  of the second vane  33  is smaller than the width W 2  of the sliding portion  33   a  in the width direction of the second vane  33 . Since such a configuration facilitates the final polishing of the sliding portion  33   a , the production cost of the second vane  33  can be reduced. The “width of the vane” means the dimension of the vane in the direction perpendicular to the axial direction of the shaft  4  and to the longitudinal direction of the vane. 
     The structure capable of preventing vane jumping is not limited to the structure shown in  FIG. 4A . Some specific examples are described below. 
     In an example shown in  FIG. 4B , the piston  8  is provided with a projecting portion  8   t , and the second vane  33  is provided with a recessed portion  33   s . The projecting portion  8   t  of the piston  8  is fitted in the recessed portion  33   s  of the second vane  33  so that the piston  8  and the second vane  33  are coupled together. That is, there is no particular limitation on the structure for coupling the vane to the piston. 
     Next, in an example shown in  FIG. 4C , the piston  8  and the first vane  32  are constituted by an integrally formed swing piston  56 . That is, the first vane  32  is integrated with the piston  8 . A bush  57  (first bush) is disposed in the first vane groove  34  (bush groove). The bush  57  is composed of two members each having an approximately semicircular column shape. The outer peripheral surface of the semicircular columnar member includes a flat surface and a circular arc surface. The flat surface of the semicircular columnar member faces the side surface of the first vane  32 , and the circular arc surface thereof faces the circular arc surface of the first vane groove  34 . That is, the bush  57  slidably holds the first vane  32 , and the bush  57  itself can slide relative to the cylinder  5 . As the piston  8  rotates, the first vane  32  moves back and force in the first vane groove  34  while changing its posture little by little. As just described, the first vane  32  is swingably disposed in the first vane groove  34  of the cylinder  5  by means of the bush  57 . The bush  57  also can rotate (swing) in the first vane groove  34 . 
     On the other hand, the second vane  33  is coupled to the piston  8 . Specifically, as described with reference to  FIG. 4A , the projecting portion  33   t  of the second vane  33  is fitted in the recessed portion  8   s  of the piston  8 . A bush  58  (second bush) holding the second vane  33  is provided at the second angular position so that the second vane  33  can swing as the piston  8  rotates. The movement of the bush  58  disposed in the second vane groove  35  is the same as that of the bush  57  disposed in the first vane groove  34 . The projecting portion  33   t  of the second vane  33  and the recessed portion  8   s  of the piston  8  can rotate relatively to each other while maintaining the coupling of the second vane  33  and the piston  8 . The second vane  33  moves in the same way as the first vane  32 , except that the former is coupled to the piston  8  while the latter is integrated with the piston  8 . 
     With a configuration shown in  FIG. 4C , not only the second vane  33  but also the first vane  32  can be prevented from jumping. Since the first vane  32  and the second vane  33  swing in the vane groove  34  and the vane groove  35  respectively, the piston  8  can rotate smoothly. As described with reference to  FIG. 4B , the projecting portion  8   t  of the piston  8  may be fitted in the recessed portion  33   s  of the second vane  33 . 
     Next, in an example shown in  FIG. 4D , the same structure as the structure described with reference to  FIG. 4A  is employed for the second vane  33 . In addition to this structure, the piston  8  is further provided with an other recessed portion  8   c , and the first vane  32  is provided with a projecting portion  32   t . The projecting portion  32   t  of the first vane  32  is fitted in the other recessed portion  8   c  of the piston  8 . A bush  57  (first bush) holding the first vane  32  is provided at the first angular position so that the first vane  32  can swing as the piston  8  rotates. More specifically, the bush  57  is disposed in the first vane groove  34 . 
     In the fitting structure, there is no limitation on the positional relationship between the projecting portion and the recessed portion. That is, as described with reference to  FIG. 4B , the piston  8  may be provided with a projecting portion and the second vane  33  may be provided with a recessed portion. Furthermore, the piston  8  may be provided with an other projecting portion and the first vane  32  may be provided with a recessed portion. In this case, the other projecting portion of the piston  8  can be fitted in the recessed portion of the first vane  32 . 
     Instead of the first vane  32 , the second vane  33  may be configured to swing. Both the first vane  32  and the second vane  33  may be configured to swing. That is, a first bush  57  holding the first vane  32  may be provided at the first angular position and/or a second bush  58  (see  FIG. 4C ) holding the second vane  33  may be provided at the second angular position so that at least one selected from the first vane  32  and the second vane  33  can swing as the piston  8  rotates. 
     Next, in an example shown in  FIG. 4E , the piston  8  and the second vane  33  are constituted by an integrally formed swing piston  59 . The structure of the first vane  32  is not particularly limited. In the example shown in  FIG. 4E , the first vane  32  has the same structure as a vane used in a typical rolling piston compressor. That is, the first vane  32  is not coupled to the piston  8 , nor is it integrated with the piston  8 . 
     Also in an example shown in  FIG. 4F , the piston  8  and the second vane  33  are constituted by a swing piston  59 . In addition, the swing piston  59  is provided with a recessed portion  8   c , and the first vane  32  is provided with a projecting portion  32   t . The projecting portion  32   t  of the first vane  32  is fitted in the recessed portion  8   c  of the swing piston  59  so that the swing piston  59  and the first vane  32  are coupled together. A bush  57  holding the first vane  32  is provided at the first angular position so that the first vane  32  can swing as the piston  8  rotates. In the example shown in  FIG. 4F , the swing piston  59  may be provided with a projecting portion and the first vane  32  may be provided with a recessed portion. In this case, the projecting portion of the swing piston  59  can be fitted in the recessed portion of the first vane  32 . 
     With the structures described with reference to  FIG. 4A  to  FIG. 4F , it is possible to reliably prevent the second vane  33  from separating from the piston  8 . Furthermore, in the structures described with reference to  FIG. 4A  to  FIG. 4F , the axial rotation of the piston  8  is inhibited. The “axial rotation of the piston  8 ” means that the piston  8  can rotate freely with respect to the eccentric portion  4   a  of the shaft  4 , the first vane  32 , and the second vane  33 . When the axial rotation of the piston  8  is inhibited, a specific part of the piston  8  always faces the second compression chamber  26  and the other part thereof always faces the first compression chamber  25 . The temperature of the refrigerant compressed in the second compression chamber  26  is slightly lower (for example, by about 10° C.) than that of the refrigerant compressed in the first compression chamber  25 . Therefore, during the operation of the rotary compressor  102 , the temperature of the specific part of the piston  8  is slightly lower than that of the other part thereof. If the temperature of the specific part is lower than that of the other part, the refrigerant drawn into the second compression chamber  26  is less likely to receive heat from the piston  8 . Since the refrigerant drawn into the second compression chamber  26  is less likely to receive heat from the piston  8 , a decrease in the volumetric efficiency of the second compression chamber  26  caused by the expansion of the refrigerant drawn thereinto can be suppressed. 
     Referring back to  FIG. 2  and  FIG. 3 , the other components are described. 
     As shown in  FIG. 2 , the main bearing  6  and the auxiliary bearing  7  are disposed on and beneath the cylinder  5  to close the cylinder  5 . The muffler  9  is provided on the main bearing  6  and covers the first discharge valve  43  and the second discharge valve  44 . A discharge port  9   a  for discharging the compressed refrigerant to the internal space  13  of the closed casing  1  is formed in the muffler  9 . The shaft  4  penetrates the central portion of the muffler  9  and is rotatably supported by the main bearing  6  and the auxiliary bearing  7 . 
     As shown in  FIG. 2  and  FIG. 3 , in the present embodiment, the first suction port  19  and the second suction port  20  are formed in the cylinder  5 . The first suction port  19  introduces the refrigerant to be compressed in the first compression chamber  25  into the first compression chamber  25 . The second suction port  20  introduces the refrigerant to be compressed in the second compression chamber  26  into the second compression chamber  26 . The first suction port  19  and the second suction port  20  may each be formed in the main bearing  6  or the auxiliary bearing  7 . 
     In the present embodiment, the second suction port  20  has a smaller opening area than the first suction port  19 . The smaller the opening area of the second suction port  20  is, the smaller the sizes of the parts of the suction check valve  50  are. This is important in suppressing an increase in dead volume caused by the suction check valve  50  and in providing a design margin. When the opening area of the first suction port  19  is S 1  and the opening area of the second suction port  20  is S 2 , the opening areas S 1  and S 2  satisfy, for example, 1.1≦(S 1 /S 2 )≦30. The “dead volume” refers to the volume that does not serve as a working chamber. Generally, a large dead volume is not preferable for a positive displacement fluid machine. 
     As shown in  FIG. 3 , the first suction pipe  14  (main suction pipe) and the second suction pipe  16  (injection suction pipe) are connected to the compression mechanism  3 . The first suction pipe  14  is fitted in the cylinder  5  through the barrel portion of the closed casing  1  so as to supply the refrigerant to the first suction port  19 . The first suction pipe  14  constitutes a part of the flow path  10   d  shown in  FIG. 1 . The second suction pipe  16  is fitted in the cylinder  5  through the barrel portion of the closed casing  1  so as to supply the refrigerant to the second suction port  20 . The second suction pipe  16  constitutes a part of the injection flow path  10   j  shown in  FIG. 1 . 
     The compression mechanism  3  further is provided with a first discharge port  40  (main discharge port) and a second discharge port  41  (injection discharge port). The first discharge port  40  and the second discharge port  41  are each formed in the main bearing  6  in a manner as to penetrate the main bearing  6  in the axial direction of the shaft  4 . The first discharge port  40  discharges the refrigerant compressed in the first compression chamber  25  outside the first compression chamber  25  (into the internal space of the muffler  9  in the present embodiment) from the first compression chamber  25 . The second discharge port  41  discharges the refrigerant compressed in the second compression chamber  26  outside the second compression chamber  26  (into the internal space of the muffler  9  in the present embodiment) from the second compression chamber  26 . The first discharge port  40  and the second discharge port  41  are provided with a first discharge valve  43  and a second discharge valve  44  respectively. When the pressure in the first compression chamber  25  exceeds the pressure in the internal space  13  of the closed casing  1  (high pressure of the refrigeration cycle), the first discharge valve  43  opens. When the pressure in the second compression chamber  26  exceeds the pressure in the internal space  13  of the closed casing  1 , the second discharge valve  44  opens. 
     The muffler  9  serves as a discharge flow path connecting the internal space  13  of the closed casing  1  and each of the first discharge port  40  and the second discharge port  41 . The refrigerant discharged outside the first compression chamber  25  through the first discharge port  40  and the refrigerant discharged outside the second compression chamber  26  through the second discharge port  41  are merged together in the muffler  9 . The merged refrigerant flows into the discharge pipe  11  through the internal space  13  of the closed casing  1 . The motor  2  is disposed in the closed casing  1  to be located in the flow path of the refrigerant from the muffler  9  to the discharge pipe  11 . With such a configuration, efficient cooling of the motor  2  by the refrigerant and efficient heating of the refrigerant by the heat of the motor  2  can be achieved. 
     In the present embodiment, the second discharge port  41  has a smaller opening area than the first discharge port  40 . The smaller the opening area of the second discharge port  41  is, the more the dead volume caused by the second discharge port  41  can be reduced. When the opening area of the first discharge port  40  is S 3  and the opening area of the second discharge port  41  is S 4 , the opening areas S 3  and S 4  satisfy, for example, 1.1≦(S 3 /S 4 )≦15. 
     The opening area S 2  of the second suction port  20  may be equal to the opening area S 1  of the first suction port  19  in some cases. Furthermore, the opening area S 4  of the second discharge port  41  may be equal to the opening area S 3  of the first discharge port  40  in some cases. The size of each of the suction ports and the discharge ports should be determined appropriately in view of the flow rate of the refrigerant at that port. More specifically, the size should be determined in view of the balance between the dead volume and the pressure loss. 
     For the reason described below, the rotary compressor  102  of the present embodiment includes not only the discharge valves  43  and  44  but also a suction check valve  50  provided in the second suction port  20 . In the compressor  501  described in Patent Literature 2, when it shifts from a suction process to a compression process, a large amount of refrigerant may flow back into the suction port  527   a  from the auxiliary compression chamber  527 . This causes a decrease in compressor efficiency. Therefore, even if the compressor  501  described in Patent Literature 2 is used to construct a refrigeration cycle apparatus, an increase in the COP of the refrigeration cycle cannot be expected. The suction check valve  50  can solve this problem. 
     As shown in  FIG. 5 , the suction check valve  50  includes a valve body  51  and a valve stopper  52 . A shallow groove  5   g  having a strip shape in plan view is formed on the top surface  5   p  of the cylinder  5 , and the valve body  51  and the valve stopper  52  are fitted in the groove  5   g . The groove  5   g  extends outwardly in a radial direction of the cylinder  5  and is in communication with the second compression chamber  26 . The second suction port  20  opens into the bottom of the groove  5   g . Specifically, the second suction port  20  is constituted by a closed-end hole formed in the cylinder  5 , and the other end of the hole opens into the bottom of the groove  5   g . In the cylinder  5 , a suction flow path  5   f  extending from the outer peripheral surface of the cylinder  5  to the center thereof is formed so as to supply the refrigerant to the second suction port  20 . The suction pipe  16  is connected to the suction flow path  5   f.    
     As shown in  FIG. 6A , the valve body  51  has a back surface  51   q  for closing the second suction port  20  and a front surface  51   p  to be exposed to the atmosphere in the second compression chamber  26  when the second suction port  20  is closed. The range of movement of the valve body  51  of the suction check valve  50  is determined in the second compression chamber  26 . The valve body  51  has a thin plate shape as a whole. Typically, the valve body  51  is constituted by a thin metal plate (reed valve). 
     As shown in  FIG. 6B , the valve stopper  52  has a supporting surface  52   q  for limiting the amount of displacement of the valve body  51  in the thickness direction thereof when the second suction port  20  is opened. The supporting surface  52   q  forms a slightly curved surface so that the thickness of the valve stopper  52  decreases as it approaches the second compression chamber  26 . That is, the valve stopper  52  has a shoetree-like shape as a whole. The front end surface  52   t  of the valve stopper  52  has a shape of a circular arc having the same radius of curvature as the inner radius of the cylinder  5 . 
     The valve body  51  is disposed in the groove  5   g  so as to open and close the second suction port  20 . The valve stopper  52  is disposed in the groove  5   g  so that the supporting surface  52   q  is exposed to the atmosphere in the second compression chamber  26  when the valve body  51  closes the second suction port  20 . The valve body  51  and the valve stopper  52  are fixed to the cylinder  5  by a fastening member  54  such as a bolt. The rear end of the valve body  51  cannot move between the valve stopper  52  and the groove  5   g , but the front end of the valve body  51  is not fixed and can swing. In a plan view of the valve stopper  52  and the second suction port  20 , the second suction port  20  and the supporting surface  52   q  of the valve stopper  52  lie on top of each other. 
     The total thickness of the valve body  51  and the valve stopper  52  near the rear end of the valve stopper  52  is almost equal to the depth of the groove  5   g . When the valve body  51  and the valve stopper  52  are fitted into the groove  5   g , the level of the top surface  52   p  of the valve stopper  52  coincides with that of the cylinder  5  in the thickness direction of the cylinder  5 . 
     As shown in  FIG. 6A , the valve body  51  has a widened portion  55  for opening and closing the second suction port  20 . The maximum width W 1  of the widened portion  55  is greater than the width W 2  of the front end of the valve stopper  52 , in other words, greater than the width of the groove  5   g  at a position where it faces the cylinder  5 . With the widened portion  55 , an increase in the dead volume can be suppressed while the seal width for closing the second suction port  20  is secured. 
     As shown in  FIG. 5  and  FIG. 7 , the depth of the groove  5   g  is, for example, smaller than a half of the thickness of the cylinder  5 . The valve stopper  52  occupies a large part of the groove  5   g . Only a small part of the groove  5   g  remains as the range of movement of the valve body  51 . 
     The suction check valve  50  operates in the following manner as the shaft  5  rotates. When the pressure in the second compression chamber  26  falls below the pressure in the suction flow path  5   f  and the second suction pipe  16 , the valve body  51  is displaced to conform to the shape of the supporting surface  52   q  of the valve stopper  52 . In other words, the valve body  51  is pushed up. Thereby, the second suction port  20  is brought into communication with the second compression chamber  26 , so that the refrigerant is supplied to the second compression chamber  26  through the second suction port  20 . On the other hand, when the pressure in the second compression chamber  26  exceeds the pressure in the suction flow path  5   f  and the second suction pipe  16 , the valve body  51  returns to its original flat shape. Thereby, the second suction port  20  is closed. Therefore, it is possible to prevent the refrigerant drawn into the second compression chamber  26  from flowing back to the suction flow path  5   f  and the second suction pipe  16  through the second suction port  20 . 
     With the structural features of the suction check valve  50  of the present embodiment described above, it is possible to suppress an increase in dead volume caused by the presence of a check valve in the suction port. That is, the suction check valve  50  contributes to a high compressor efficiency. Accordingly, the refrigeration cycle apparatus  100  using the rotary compressor  102  of the present embodiment has a high COP. 
     The second suction port  20  may be formed in the main bearing  6  or the auxiliary bearing  7 . In this case, the suction check valve  50  having the structure described with reference to  FIG. 5 , etc. can be provided in the main bearing  6  or the auxiliary bearing  7 . A member (closing member) for closing the cylinder  5  may be provided between the main bearing  6  (or the auxiliary bearing  7 ) and the cylinder  5 . The suction check valve  50  may be provided in that member. 
     Next, the operation of the rotary compressor  102  is described in time series with reference to  FIG. 8 . The angles in  FIG. 8  represent the rotation angles of the shaft  4 . The angles shown in  FIG. 8  are merely examples, and each process does not always start or end at the angle shown in  FIG. 8 . A suction process of drawing the refrigerant into the first compression chamber  25  starts when the shaft  4  has a rotation angle of 0 degrees and takes place until the shaft  4  has a rotation angle of approximately 360 degrees. The refrigerant drawn into the first compression chamber  25  is compressed as the shaft  4  rotates. The compression process continues until the pressure in the first compression chamber  25  exceeds the pressure in the internal space  13  of the closed casing  1 . In  FIG. 8 , the compression process starts when the shaft  4  has a rotation angle of 360 degrees and takes place until the shaft  4  has a rotation angle of 540 degrees. A process of discharging the compressed refrigerant outside the first compression chamber  25  takes place until the point of contact between the cylinder  5  and the piston  8  passes the first discharge port  40 . In  FIG. 8 , the discharge process starts when the shaft  4  has a rotation angle of 540 degrees and takes place until the shaft  4  has a rotation angle of (630+α) degrees. “α” denotes an angle between the angular position of 270 degrees and the second angular position where the second vane  33  is disposed. 
     On the other hand, a suction process of drawing the refrigerant into the second compression chamber  26  starts when the shaft  4  has a rotation angle of (270+α) degrees and takes place until the shaft  4  has a rotation angle of (495+α/2) degrees. (495+α/2) is a rotation angle of the shaft  4  at which the second compression chamber  26  has a maximum volume. The refrigerant drawn into the second compression chamber  26  is compressed as the shaft  4  rotates. The compression process continues until the pressure in the second compression chamber  26  exceeds the pressure in the internal space  13  of the closed casing  1 . In  FIG. 8 , the compression process starts when the shaft  4  has a rotation angle of (495+α/2) degrees and takes place until the shaft  4  has a rotation angle of 630 degrees. A process of discharging the compressed refrigerant outside the second compression chamber  26  takes place until the point of contact between the cylinder  5  and the piston  8  passes the second discharge port  41 . In  FIG. 8 , the discharge process starts when the shaft  4  has a rotation angle of 630 degrees and takes place until the shaft  4  has a rotation angle of 720 degrees. 
       FIG. 9A  and  FIG. 9B  show the PV diagrams of the first compression chamber  25  and the second compression chamber  26  respectively. As shown in  FIG. 9A , the suction process in the first compression chamber  25  is represented by a change from Point A to Point B. The volume of the first compression chamber  25  becomes maximum at Point B. However, since the first compression chamber  25  is not provided with a check valve, a small amount of refrigerant flows back into the first suction port  19  from the first compression chamber  25  between Point B and Point C. Therefore, the actual suction volume (confined volume) of the first compression chamber  25  is identified as the volume at Point C. The compression process is represented by a change from Point C to Point D. The discharge process is represented by a change from Point D to Point E. 
     As shown in  FIG. 9B , the suction process in the second compression chamber  26  is represented by a change from Point F to Point G. The backflow amount of the refrigerant from the second compression chamber  26  into the second suction port  20  is nearly zero owing to the function of the suction check valve  50 . Therefore, the maximum volume of the second compression chamber  26  is equal to the actual suction volume. The compression process is represented by a change from Point G to Point H. The discharge process is represented by a change from Point H to Point I. Since the second compression chamber  26  draws and compresses a gaseous refrigerant having an intermediate pressure, the compression work corresponding to the area of a shaded region can be reduced, as shown in  FIG. 10 . Thereby, the efficiency of the refrigeration cycle apparatus  100  is increased. It should be noted that  FIG. 9B  and  FIG. 10  are PV diagrams obtained by assuming that the dead volume caused by the suction check valve  50  is zero. 
     For information,  FIG. 11A  is a schematic diagram showing the operation of a rotary compressor without a suction check valve. The angle between two vanes is 90 degrees. A compression chamber  536  and a suction port  537  correspond to the second compression chamber  26  and the second suction port  20 , respectively, of the present embodiment. In the state shown in the left side of  FIG. 11A , the compression chamber  536  has a maximum volume. However, during the rotation of the shaft  534  from the state shown in the left side to the state shown in the right side, a refrigerant flows from the compression chamber  536  back into the suction port  537  (backflow process). 
     In fact, as shown in  FIG. 11B , when the maximum volume is represented as a volume at Point J, the volume at the moment when the compression actually starts (actual suction volume) is represented as a volume at Point G. That is, a considerable percentage of the refrigerant (corresponding to a volume obtained by subtracting the volume at Point G from the volume at Point J) is pushed out of the compression chamber  536  in the backflow process. Therefore, a very large loss occurs. A shaded region in  FIG. 11B  represents the sum of a loss that occurs when the compression chamber  536  draws the refrigerant from Point F to Point J and a loss that occurs due to the backflow of the refrigerant when the volume of the compression chamber  536  decreases from Point J to Point G (the sum is an unnecessary compression work). Furthermore, there is a concern that the backflow of the refrigerant causes pulsation, which may increase noise and vibration. The rotary compressor  102  of the present embodiment can solve these problems. 
     In each of  FIG. 9A ,  FIG. 9B ,  FIG. 10  and  FIG. 11B , the vertical axis (pressure axis) and the horizontal axis (volume axis) are drawn on the same scale.  FIG. 11A  and  FIG. 11B  are diagrams for explaining the problems that may occur without a suction check valve, and are not the prior art of the present invention. 
     Next, the positional relationship between the first vane  32  and the second vane  33  is described. The positional relationship between them is also closely related to the timing of opening and closing the suction check valve  50 . The open/close timing of the suction check valve  50  also depends on the type of the refrigerant, the intended use of the refrigeration cycle apparatus  100 , etc. 
     According to the present embodiment, the angle θ between the first angular position (0 degrees) where the first vane  32  is disposed and the second angular position where the second vane  33  is disposed is set to 270 degrees or more in the rotation direction of the shaft  4 . The angle θ should be set appropriately depending on the flow rate of the refrigerant to be compressed in the first compression chamber  25  and the flow rate of the refrigerant to be compressed in the second compression chamber  26 . 
     However, the amount of the refrigerant flowing from the first compression chamber  25  back into the first suction port  19  increases as the angle θ decreases. An appropriate range of angles θ is, for example, 270≦θ≦350. 
     Of course, the optimum angle θ varies depending on the intended use of the refrigeration cycle apparatus  100 . It is conceivable to set the angle θ to less than 270 degrees, as shown in  FIG. 12 . The amount of the refrigerant flowing from the first compression chamber  25  back into the first suction port  19  increases as the angle θ decreases. In order to prevent the refrigerant from flowing from the first compression chamber  25  back into the first suction port  19 , a suction check valve can be provided also in the first suction port  19 . 
     The above findings indicate that the suction check valve  50  prevents the refrigerant drawn into the second compression chamber  26  from flowing back outside the second compression chamber  26  through the second suction port  20  during the period defined as (i), (ii) or (iii): (i) during a period from a point of time when the second compression chamber  26  reaches a maximum volume to a point of time when the second compression chamber  26  reaches a minimum volume (almost equal to 0); (ii) during a period from the point of time when the second compression chamber  26  reaches the maximum volume to a point of time when the compressed refrigerant begins to be discharged outside the second compression chamber  26  through the second discharge port  41 ; and (iii) during a period from the point of time when the second compression chamber  26  reaches the maximum volume to a point of time when the point of contact between the cylinder  5  and the piston  8  passes the second suction port  20  as the shaft  4  rotates. When the angle θ is relatively large, the suction check valve  50  prevents the backflow during the period (i). When the angle θ is relatively small, the suction check valve  50  prevents the backflow during the period (ii) or (iii). 
     The suction check valve  50  contributes significantly to an increase in compressor efficiency. However, from the viewpoint of preventing vane jumping, the suction check valve  50  has an adverse effect. First, the case where a suction check valve is not provided is considered with reference to  FIG. 15 . In the case where a suction check valve is not provided, one side of the front surface of the vane  535  is exposed to a discharge pressure (high pressure) in the compression chamber  526  at the moment when the piston  523  pushes the vane  535  into the vane groove in the state shown in  FIG. 15 . The other side of the front surface of the vane  535  is exposed to a suction pressure (intermediate pressure) in the suction port  527   a . Therefore, if it is assumed that the rolling piston compressor  501  is a high-pressure shell type compressor, a certain pressing force is always applied to the vane  535  based on the difference between the pressure applied to the front surface and the pressure applied to the rear surface. 
     Next, the case where a suction check valve is provided in the second suction port but the second vane is not coupled to the piston is considered with reference to  FIG. 17 . One side of the front surface of the second vane  552  is exposed to a discharge pressure (high pressure) in the first compression chamber  554  at the moment when the piston  558  pushes the second vane  552  in the state shown in  FIG. 17 . The other side of the front surface of the second vane  552  is exposed to a pressure in the second compression chamber  556 . In the state shown in  FIG. 17 , the pressure in the second compression chamber  556  is equal or close to the discharge pressure (high pressure), although it cannot be definitely determined because it depends also on design conditions such as the angle θ. That is, in the state shown in  FIG. 17 , a pressing force applied to the second vane  552  based on the difference between the pressure applied to the front surface and the pressure applied to the rear surface is almost zero, and only a pressing force of the spring  553  is applied to the second vane  552 . If the piston  558  passes the top dead center of the second vane  552  in this state, the second vane  552  cannot follow the movement of the piston  558  because an outward inertial force is applied to the second vane  552 . As a result, vane jumping may occur. 
     As described above, the suction check valve  50  is closely related to the problem of vane jumping. Therefore, in the case where the suction check valve  50  is provided to prevent the backflow of the refrigerant, it is desirable to actively adopt the structures described with reference to  FIG. 4A  to  FIG. 4F  in order to prevent vane jumping. A combination of the suction check valve  50  and the structure for preventing vane jumping can provide the rotary compressor  102  with a very high compressor efficiency.  FIG. 17  is a diagram for explaining the problems that may occur when the second vane is not coupled to the piston, and is not the prior art of the present invention. 
     Modification 
       FIG. 13  is a longitudinal cross-sectional view of a rotary compressor according to a modification. A rotary compressor  202  has a structure in which components such as a cylinder is added to the rotary compressor  102  shown in  FIG. 2 . In the present modification, the compression mechanism  3 , the cylinder  5 , the piston  8  and the eccentric portion  4   a  shown in  FIG. 2  are defined as a first compression mechanism  3 , a first cylinder  5 , a first piston  8 , and a first eccentric portion  4   a , respectively. The detailed structure of the first compression mechanism  3  is as described with reference to  FIG. 2  to  FIG. 7 . 
     As shown in  FIG. 13  and  FIG. 14 , the rotary compressor  202  includes a second compression mechanism  30  in addition to the first compression mechanism  3 . The second compression mechanism  30  has a second cylinder  65 , an intermediate plate  66 , a second piston  68 , an auxiliary bearing  67 , a muffler  70 , a third vane  72 , a third suction port  69 , and a third discharge port  73 . The second cylinder  65  is disposed concentrically with the first cylinder  5 , and separated from the first cylinder  5  by the intermediate plate  66 . 
     The shaft  4  has a second eccentric portion  4   b  projecting outwardly in a radial direction. The second piston  68  is disposed within the second cylinder  65 . Within the second cylinder  65 , the second piston  68  is fitted to the second eccentric portion  4   b  of the shaft  4 . The intermediate plate  66  is disposed between the first cylinder  5  and the second cylinder  65 . A vane groove  74  is formed in the second cylinder  65 . A third vane  72  (blade) having a tip in contact with the outer peripheral surface of the second piston  68  is slidably fitted in the vane groove  74 . The third vane  72  divides the space between the second cylinder  65  and the second piston  68  along the circumferential direction of the second piston  68 . Thereby, a third compression chamber  71  is formed within the second cylinder  65 . The second piston  68  and the third vane  72  may be constituted by a single component, i.e., a so-called swing piston. The third vane  72  may be coupled to the second piston  68 . A third spring  76  pressing the third vane  72  toward the center of the shaft  4  is disposed behind the third vane  72 . 
     A third suction port  69  introduces the refrigerant to be compressed in the third compression chamber  71  into the third compression chamber  71 . A third suction pipe  64  is connected to the third suction port  69 . The third discharge port  73  penetrates the auxiliary bearing  67  and opens into the internal space of the muffler  70 . The refrigerant compressed in the third compression chamber  71  is discharged outside the third compression chamber  71 , specifically, to the internal space of the muffler  70 , from the third compression chamber  71  through the third discharge port  73 . The refrigerant is introduced from the internal space of the muffler  70  into the internal space  13  of the closed casing  1  through the flow path  63  passing through the main bearing  6 , the first cylinder  5 , the intermediate plate  66 , the second cylinder  65  and the auxiliary bearing  67  in the axial direction of the shaft  4 . The flow path  63  may open into the internal space  13  of the closed casing  1 , or into the internal space of the muffler  9 . 
     As described above, the second compression mechanism  30  has the same structure as a compression mechanism of a typical rolling piston compressor having only one vane. 
     The second piston  68  and the third vane  72  may be integrated together. Alternatively, the second piston  68  and the third vane  72  may be coupled together. That is, the structures described with reference to  FIG. 4A  to  FIG. 4F  can be applied to the second piston  68  and the third vane  72 . The problem of vane jumping is less likely to occur for the third vane  72 . However, it can be expected that the shared use of the components between the first compression mechanism  3  and the second compression mechanism  30  can lead to a cost reduction effect. 
     In the rotary compressor  202 , the height, inner diameter and outer diameter of the second cylinder  65  are equal to the height, inner diameter and outer diameter of the first cylinder  5 , respectively. The outer diameter of the first piston  8  is equal to that of the second piston  68 . Since only the third compression chamber  71  is formed within the second cylinder  65 , the first compression chamber  25  has a smaller volume than the third compression chamber  71 . This means that the shared use of the components between the first compression mechanism  3  and the second compression mechanism  30  can lead to a cost reduction and increased ease of assembling. 
     In the present modification, the first compression mechanism  3  and the second compression mechanism  30  are disposed on the upper side and the lower side of the axial direction of the shaft  4 , respectively. The refrigerant compressed in the first compression mechanism  3  is introduced into the internal space of the muffler  9  through the discharge ports  40  and  41  provided in the main bearing  6 . The first compression mechanism  3  has two discharge ports  40  and  41 . Therefore, it is desirable to reduce the distance between the discharge ports  40  and  41  and the internal space  13  of the closed casing  1  as much as possible so as to reduce the pressure loss of the refrigerant in the discharge ports  40  and  41  as much as possible. From this viewpoint, it is preferable to dispose the first compression mechanism  3  on the upper side of the axial direction. 
     However, from another viewpoint, the first compression mechanism  3  may be disposed on the lower side of the axial direction. The reason for this is as follows. The nearer the motor  2  is, the higher the temperature in the closed casing  1  is. This means that the main bearing  6  has a higher temperature than the auxiliary bearing  67  and the muffler  70  during the operation of the rotary compressor  202 . Therefore, when the first compression mechanism  3  is disposed on the upper side and the second compression mechanism  30  is disposed on the lower side, the refrigerant to be introduced into the second compression chamber  26  is likely to be heated. Then, the mass flow rate of the refrigerant to be compressed in the second compression chamber  26  decreases, which also reduces the injection effect. In order to obtain a higher injection effect, the second compression mechanism  30  may be disposed on the upper side and the first compression mechanism  3  having the second compression chamber  26  may be disposed on the lower side. 
     As shown in  FIG. 13 , the angular difference between the direction in which the first eccentric portion  4   a  projects and the direction in which the second eccentric portion  4   b  projects is 180 degrees in the rotation direction of the shaft  4 . In other words, the phase difference between the first piston  8  and the second piston  68  is 180 degrees in the rotation direction of the shaft  4 . In still other words, the timing of the top dead center of the first piston  8  is shifted from the timing of the top dead center of the second piston  68  by 180 degrees. With such a configuration, the vibration generated by the rotation of the first piston  8  can be cancelled by the rotation of the second piston  68 . Furthermore, the compression process in the first compression chamber  25  and the compression process in the third compression chamber  71  are performed almost alternately, and the discharge process in the first compression chamber  25  and the discharge process in the third compression chamber  71  are performed almost alternately. Therefore, the torque variation of the shaft  4  can be reduced, which is advantageous in reducing the motor loss and mechanical loss. The vibration and noise of the rotary compressor  202  also can be reduced. The “timing of the top dead center of the piston” means the timing when the vane is pushed all the way into the vane groove by the piston. 
     When the rotary compressor  202  is used in the refrigeration cycle apparatus  100  shown in  FIG. 1 , the following configuration can be adopted. The refrigeration cycle apparatus  100  has the suction flow path  10   d  for introducing the refrigerant that has flowed out of the first heat exchanger  104  or the second heat exchanger  112  as an evaporator into the first suction port  19  of the rotary compressor  202 . As shown in  FIG. 13 , the suction flow path  10   d  includes a branch portion  14  extending toward the first suction port  19  and a branch portion  64  extending toward the third suction port  69  so that the refrigerant that has flowed out of the first heat exchanger  104  or the second heat exchanger  112  is introduced into both the first suction port  19  and the third suction port  69  of the rotary compressor  202 . In the present embodiment, the first suction pipe  14  constitutes the branch portion  14  and the third suction pipe  64  constitutes the branch portion  64 . With such a configuration, the refrigerant can be introduced smoothly into the first compression chamber  25  and the third compression chamber  71 . The suction flow path  10   d  may branch in the closed casing  1 . 
     INDUSTRIAL APPLICABILITY 
     The refrigeration cycle apparatus of the present invention can be used for water heaters, hot water heating apparatuses, air conditioners, etc.