Patent Publication Number: US-8985984-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. 20  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. When the compressor  501  shifts from a suction process to a compression process, a large amount of refrigerant flows back into the suction port  527   a  from the auxiliary compression chamber  527 . This causes a significant 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. 
     It is an object of the present invention to improve 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; 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; 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 of the present invention; 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 
     The rotary compressor of the present invention has a cylinder and a plurality of vanes attached to the cylinder. The plurality of vanes divide the space between the cylinder and a piston, and thereby, a first compression chamber and a second compression chamber are formed within the cylinder. The second compression chamber has a smaller volume than the first compression chamber. The first compression chamber can be used as a main compression chamber. The second compression chamber can be used as a compression chamber for compressing a working fluid injected into the rotary compressor. 
     The working fluid is introduced into the second compression chamber through a second suction port. The second suction port is provided with a suction check valve. With this valve, it is possible to prevent the working fluid drawn into the second compression chamber from flowing back outside the second compression chamber through the second suction port. Therefore, the rotary compressor of the present invention can achieve a high compressor efficiency. 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. 4  is an enlarged cross-sectional view of a suction check valve. 
         FIG. 5A  shows side and plan views of a valve body. 
         FIG. 5B  shows side and plan views of a valve stopper. 
         FIG. 6  is a perspective view of a compression mechanism. 
         FIG. 7  is a schematic diagram showing the operation of the rotary compressor with the rotation angle of a shaft. 
         FIG. 8A  is a PV diagram of a first compression chamber. 
         FIG. 8B  is a PV diagram of a second compression chamber. 
         FIG. 9  is a PV diagram of the second compression chamber showing the compression work that can be reduced by injection. 
         FIG. 10A  is a schematic diagram showing the operation of a rotary compressor provided with no suction check valve. 
         FIG. 10B  is a PV diagram of a second compression chamber shown in  FIG. 10A . 
         FIG. 11  is a schematic diagram showing a modification designed to have an obtuse angle between a first vane and a second vane. 
         FIG. 12A  is a schematic diagram of a modification of the vanes. 
         FIG. 12B  is a schematic diagram of another modification of the vanes. 
         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 refrigeration cycle apparatus according to a second embodiment of the present invention. 
         FIG. 16  is a longitudinal cross-sectional view of a rotary compressor used in the refrigeration cycle apparatus shown in  FIG. 15 . 
         FIG. 17A  is a transverse cross-sectional view of the rotary compressor shown in  FIG. 16 , taken along the line D-D. 
         FIG. 17B  is a transverse cross-sectional view of the rotary compressor shown in  FIG. 16 , taken along the line E-E. 
         FIG. 18  is a schematic diagram showing the relationship between the thickness of a first cylinder and that of a second cylinder. 
         FIG. 19  is a partial configuration diagram showing modified first and second injection paths. 
         FIG. 20  is a configuration diagram of a conventional heat pump type heating apparatus. 
         FIG. 21  is a transverse cross-sectional view of a conventional rolling piston compressor having only one vane. 
     
    
    
     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) having a tip in contact with the outer peripheral surface of the piston  8  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 . 
     The piston  8  and one selected from the first vane  32  and the second vane  33  may be constituted by a single component, i.e., a so-called swing piston. One selected from the first vane  32  and the second vane  33  may be coupled to the piston  8 . 
     A first spring  36  is disposed behind the first vane  32 . A second spring  37  is disposed behind the second vane  37 . The first spring  36  and the second spring  37  press the first vane  32  and the second vane  33 , respectively, toward the center of the shaft  4 . The rear end of the first vane groove  34  and the rear end of the second vane groove  35  are each 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  and the rear surface of the second vane  33 . 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 . 
     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. 
     As shown in  FIG. 4 , 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. 5A , 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. 5B , 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. 5A , 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. 4  and  FIG. 6 , 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. 3  to  FIG. 6  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. 7 . The angles in  FIG. 7  represent the rotation angles of the shaft  4 . The angles shown in  FIG. 7  are merely examples, and each process does not always start or end at the angle shown in  FIG. 7 . 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. 7 , 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. 7 , 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. 7 , 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. 7 , 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. 8A  and  FIG. 8B  show the PV diagrams of the first compression chamber  25  and the second compression chamber  26  respectively. As shown in  FIG. 8A , 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. 8B , 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. 9 . Thereby, the efficiency of the refrigeration cycle apparatus  100  is increased. It should be noted that  FIG. 8B  and  FIG. 9  are PV diagrams obtained by assuming that the dead volume caused by the suction check valve  50  is zero. 
     For information,  FIG. 10A  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. 10A , 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. 10B , 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. 10B  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. 8A ,  FIG. 8B ,  FIG. 9  and  FIG. 10B , the vertical axis (pressure axis) and the horizontal axis (volume axis) are drawn on the same scale.  FIG. 10A  and  FIG. 10B  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. 11 . 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). 
     Meanwhile, the present inventors have also ascertained that a rotary compressor having a plurality of vanes has the following problem. 
     As shown in  FIG. 21 , 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. 20 , 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. Particularly in the case where the suction check valve  50  is provided in the second suction port  20  as in the present embodiment, vane jumping is likely to occur. As a means for preventing the occurrence of vane jumping, the following configurations can be proposed. The occurrence of vane jumping can be prevented by adopting at least one of the following configurations. 
     In a configuration shown in  FIG. 12A , the width W 4  of the second vane  33  is smaller than the width W 3  of the first vane  32 . Instead of or in addition to the adjustment of the width, the weight of the second vane  33  may be adjusted to be smaller than that of the first vane  32 . Even if the size of the first vane  32  is equal to that of the second vane  33 , the weight of the second vane can be reduced by using a lighter material for the second vane  33  than that for the first vane  32 . For example, in the case where the first vane  32  is made of a metal containing iron as a main component (i.e., a component having the largest content in terms of mass percentage), the second vane  33  can be formed of a material containing aluminum as a main component. The “width of the vane” means the dimension of the vane in a direction perpendicular to the axial direction of the shaft  4  and the longitudinal direction of the vane. 
     In a configuration shown in  FIG. 12B , the seal length L 2  of the second vane  33  is shorter than the seal length L 1  of the first vane  32 . In other words, the second vane  33  is shorter than the first vane  32 . The “seal length” means the longitudinal length of the contact surface between the vane and the vane groove when the vane is pushed all the way into the vane groove. As the second spring  37 , a spring having a larger spring constant than the first spring  36  may be used. 
     In each of the above configurations, the inertial force acting on the second vane  33  can be reduced. With the use of a spring having a large spring constant, the spring pressing force can be increased. Therefore, even if the pressing force generated by the difference between the pressure applied to the front surface of the vane and the pressure applied to the rear surface thereof is small, jumping of the second vane  33  can be prevented. 
     (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. 6 . 
     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. 
     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 . 
     Second Embodiment 
       FIG. 15  is a configuration diagram of a refrigeration cycle apparatus according to a second embodiment. A refrigeration cycle apparatus  200  of the present embodiment is different from the refrigeration cycle apparatus  100  of the first embodiment in that injection is performed in two steps. Since the injection is performed in two steps, the refrigeration cycle apparatus  200  is highly effective particularly when it is used for heating or hot water supply. Hereinafter, the components that have been described in the first embodiment are denoted by the same reference numerals, and no further description thereof is given. 
     The refrigeration cycle apparatus  200  includes a rotary compressor  302 , a first heat exchanger  104 , a first expansion mechanism  106 , a first gas-liquid separator  108 , a second expansion mechanism  110 , a second gas-liquid separator  109 , a third expansion mechanism  111 , and a second heat exchanger  112 . These components are connected in a loop in this order by flow paths  10   a  to  10   e  so as to form a refrigerant circuit  10 . A four-way valve  116 , as a switching mechanism capable of switching the flow direction of a refrigerant, is provided in the refrigerant circuit  10 . 
     The first expansion mechanism  106  expands the refrigerant cooled in the first heat exchanger  104  as a radiator. The first gas-liquid separator  108  separates the refrigerant expanded in the first expansion mechanism  106  into a gas phase refrigerant and a liquid phase refrigerant. The second expansion mechanism  110  expands the liquid phase refrigerant separated in the first gas-liquid separator  108 . The second gas-liquid separator  109  separates the refrigerant expanded in the second expansion mechanism  110  into a gas phase refrigerant and a liquid phase refrigerant. The third expansion mechanism  111  expands the liquid phase refrigerant separated in the second gas-liquid separator  109 . After passing through the third expansion mechanism  111 , the refrigerant flows into the second heat exchanger  112  as an evaporator. The function of the four-say valve  116  allows the refrigerant to flow also in the direction opposite to the above direction. 
     The rotary compressor  302  has a first suction port  19 , a second suction port  20 , a third suction port  23 , and a fourth suction port  24 . The suction flow path  10   d  introduces the refrigerant that has flowed out of the first heat exchanger  104  or the second heat exchanger  112  into each of the first suction port  19  and the third suction port  23  of the rotary compressor  302 . 
     The refrigeration cycle apparatus  200  further includes a first injection flow path  10   j  and a second injection flow path  10   k . The first injection flow path  10   j  has one end connected to the first gas-liquid separator  108  and the other end connected to the rotary compressor  302 , and introduces the gas refrigerant separated in the first gas-liquid separator  108  to the rotary compressor  302 . The second injection flow path  10   k  has one end connected to the second gas-liquid separator  109  and the other end connected to the rotary compressor  302 , and introduces the gas refrigerant separated in the second gas-liquid separator  109  to the rotary compressor  302 . 
     The refrigeration cycle apparatus  200  of the present embodiment is different from the refrigerant cycle apparatus  100  of the first embodiment in that the former has the second gas-liquid separator  109  and the second injection flow path  10   k  in addition to the first gas-liquid separator  108  and the first injection flow path  10   j . Furthermore, the rotary compressor  302  used in the refrigeration cycle apparatus  200  of the second embodiment is configured to perform injection in two steps. 
     As shown in  FIG. 16 ,  FIG. 17A , and  FIG. 17B , the rotary compressor  302  includes the compression mechanism  3  described in the first embodiment and a second compression mechanism  90  having the same structure as the compression mechanism  3 . The second compression mechanism  90  is disposed concentrically with the first compression mechanism  3  so that they share the shaft  4 . The compression mechanism  3 , the cylinder  5 , the piston  8 , the eccentric portion  4   a , and the suction check valve  50  of the rotary compressor  102  described in the first embodiment are defined as a first compression mechanism  3 , a first cylinder  5 , a first piston  8 , a first eccentric portion  4   a , and a first suction check valve  50 , respectively. 
     As shown in  FIG. 16  and  FIG. 17B , the second compression mechanism  90  has a second cylinder  75 , a second piston  78 , a third vane  92 , a fourth vane  93 , a third suction port  23 , a third discharge port  45 , a third discharge valve  47 , a fourth suction port  24 , a fourth discharge port  46 , a fourth discharge valve  48 , and a second suction check valve  56 . The second cylinder  75  is disposed concentrically with the first cylinder  5 . The second piston  78  is disposed within the second cylinder  75  so as to form a second space between the second piston itself and the second cylinder  75 . The shaft  4  has a second eccentric portion  4   b , and the second piston  78  is fitted to the second eccentric portion  4   b . The third vane  92  is attached to the second cylinder  75  at a third angular position along the rotation direction of the shaft  4 , and divides the second space along the circumferential direction of the second piston  78 . The fourth vane  93  is attached to the second cylinder  75  at a fourth angular position along the rotation direction of the shaft  4 , and further divides the second space divided by the third vane  92  so that a third compression chamber  27  and a fourth compression chamber  28  having a smaller volume than the third compression chamber  27  are formed within the second cylinder  75 . The third suction port  23  introduces a working fluid to be compressed in the third compression chamber  27  into the third compression chamber  27 . The third discharge port  45  discharges the working fluid compressed in the third compression chamber  27  outside the third compression chamber  27  from the third compression chamber  27 . The fourth suction port  24  introduces the working fluid to be compressed in the fourth compression chamber  28  into the fourth compression chamber  28 . The fourth discharge port  46  discharges the working fluid compressed in the fourth compression chamber  28  outside the fourth compression chamber  28  from the fourth compression chamber  28 . The second suction check valve  56  is provided in the fourth suction port  24 . As described above, the second compression mechanism  90  has essentially the same structure as the first compression mechanism  3 . 
     That is, the first cylinder  5 , the first piston  8 , the first vane  32 , the second vane  33 , the first suction port  19 , the first discharge port  40 , the first discharge valve  43 , the second suction port  20 , the second discharge port  41 , the second discharge valve  44 , and the first suction check valve  50  of the first compression mechanism  3  correspond to the second cylinder  75 , the second piston  78 , the third vane  92 , the fourth vane  93 , the third suction port  23 , the third discharge port  45 , the third discharge valve  47 , the fourth suction port  24 , the fourth discharge port  46 , the fourth discharge valve  48 , and the second suction check valve  57  of the second compression mechanism  90 , respectively. The first vane groove  34 , the first spring  36 , the second vane groove  35 , and the second spring  37  of the first compression mechanism  3  correspond to the third vane groove  94 , the third spring  96 , the fourth vane groove  95 , and the fourth spring  97  of the second compression mechanism  90 , respectively. Furthermore, the first compression chamber  25  and the second compression chamber  26  of the first compression mechanism  3  correspond to the third compression chamber  27  and the fourth compression chamber  28  of the second compression mechanism  90 , respectively. The first angular position and the second angular position correspond to the third angular position and the fourth angular position, respectively. Furthermore, the first suction pipe  14  and the second suction pipe  16  of the rotary compressor  102  correspond to the third suction pipe  84  and the fourth suction pipe  86  of the rotary compressor  302 , respectively. All the structures and descriptions of the first compression mechanism  3  can be applied to those of the second compression mechanism  90  correspondingly. 
     In the rotary compressor  302 , the angular difference between a direction in which the first eccentric portion  4   a  projects and a 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  78  is 180 degrees in the rotation direction of the shaft  4 . The effects obtained in this configuration are the same as those described for the rotary compressor  202  shown in  FIG. 13 . 
     The first injection flow path  10   j  introduces the gas phase refrigerant separated in the first gas-liquid separator  108  into the second suction port  20  of the rotary compressor  302 . The second injection flow path  10   k  introduces the gas phase refrigerant separated in the second gas-liquid separator  109  into the fourth suction port  24  of the rotary compressor  302 . Since both the first compression mechanism  3  and the second compression mechanism  90  can compress the refrigerant having an intermediate pressure, a further increase in the efficiency of the rotary compressor  302  can be expected. 
     (Modification) 
     The first compression chamber  25  may have a volume different from that of the third compression chamber  27 . The second compression chamber  26  may have a volume different from that of the fourth compression chamber  28 . For example, in the modification shown in  FIG. 18 , the thickness H 2  of the second cylinder  75  is greater than the thickness H 1  of the first cylinder  5 . Therefore, the fourth compression chamber  28  (second injection compression chamber) has a larger volume than the second compression chamber  26  (first injection compression chamber). In this case, the refrigerant can be supplied to the second compression chamber  26  from a high pressure side injection flow path (for example, the first injection flow path  10   j ), while the refrigerant can be supplied to the fourth compression chamber  28  from a low pressure side injection flow path (for example, the second injection flow path  10   k ). This means that a relatively low pressure refrigerant is compressed in the fourth compression chamber  28  having a relatively large volume, while a relatively high pressure refrigerant is compressed in the second compression chamber  26  having a relatively small volume. This allows the second compression chamber  26  and the fourth compression chamber  28  to draw just enough gaseous refrigerant generated in the first gas-liquid separator  108  and the second gas-liquid separator  109 , respectively. The injection of just enough gaseous refrigerant into the rotary compressor  302  enables highly efficient operation of the refrigeration cycle apparatus  200 . 
     The ratio of the volume of the fourth compression chamber  28  to the volume of the second compression chamber  26  cannot be definitely determined because it depends on the type of the refrigerant, the intended use of the refrigeration cycle apparatus  100 , etc. As an example, the compression mechanisms  3  and  90  can be designed to satisfy 1.1≦(V 2 /V 1 )≦30, where V 1  is the volume of the second compression chamber  26 , and V 2  is the volume of the fourth compression chamber  28 . The volume of the compression chamber can be adjusted by changing various design values such as the height of the cylinder, the inner diameter of the cylinder, the outer diameter of the piston, and the amount of projection of the eccentric portion of the shaft. The volume of the compression chamber can also be adjusted by changing the positional relationship between the two vanes, of course. When the volume of the second compression chamber  26  and the volume of the fourth compression chamber  28  are adjusted to satisfy the above relationship by allowing at least one design value selected from the height of the cylinder, the inner diameter of the cylinder, the outer diameter of the piston, and the amount of projection of the eccentric portion of the shaft to differ between the first compression mechanism  3  and the second compression mechanism  90 , the volumes of the compression chambers can be optimized without changing the positions of the vanes. 
     In the refrigeration cycle apparatus  200  shown in  FIG. 15 , the flow direction of the refrigerant is switched by controlling the four-way valve  116 . Therefore, as shown in  FIG. 19 , a flow path switching portion  122  can be provided so as to introduce the refrigerant in the first injection flow path  10   j  into one selected from the second suction port  20  and the fourth suction port  24  of the rotary compressor  302  and to introduce the refrigerant in the second injection flow path  10   k  into the other of the second suction port  20  and the fourth suction port  24  of the rotary compressor  302 . 
     The flow path switching portion  122  has a first three-way valve  118 , a second three-way valve  119 , a first bypass flow path  120 , and a second bypass flow path  121 . The first three-way valve  118  is provided in the first injection flow path  10   j . The second three-way valve  119  is provided in the second injection flow path  10   k . The first bypass flow path  120  connects one outlet of the first three-way valve  118  and the second injection flow path  10   k . The second bypass flow path  121  connects one outlet of the second three-way valve  119  and the first injection flow path  10   j . When the three-way valves  118  and  119  are controlled as indicated by solid lines, the refrigerant in the first injection flow path  10   j  is introduced into the second suction port  20  and the refrigerant in the second injection flow path  10   k  is introduced into the fourth suction port  24 . When the three-way valves  118  and  119  are controlled as indicated by dashed lines, the refrigerant in the first injection flow path  10   j  is introduced into the fourth suction port  24  and the refrigerant in the second injection flow path  10   k  is introduced into the second suction port  20 . This control allows appropriate pressure refrigerants to be supplied to the second compression chamber  26  and the fourth compression chamber  28  respectively, even if the flow directions of the refrigerants are changed. 
     INDUSTRIAL APPLICABILITY 
     The refrigeration cycle apparatus of the present invention can be used for water heaters, hot water heating apparatuses, air conditioners, etc.