Patent Publication Number: US-8985976-B2

Title: Two-stage rotary expander, expander-integrated compressor, and refrigeration cycle apparatus

Description:
TECHNICAL FIELD 
     The present invention relates to a two-stage rotary expander, an expander-integrated compressor, and a refrigeration cycle apparatus. 
     BACKGROUND ART 
     There have been proposed refrigeration cycle apparatuses in which an expander recovers the expansion energy of a working fluid, and the recovered energy is used for a part of the work of the compressor. As one of such refrigeration cycle apparatuses, a refrigeration cycle apparatus using an expander-integrated compressor is known (see Patent Literature 1). 
       FIG. 28  shows a conventional refrigeration cycle apparatus using an expander-integrated compressor. This refrigeration cycle apparatus includes a compressor (compression mechanism)  201 , a radiator  202 , an expander (expansion mechanism)  203 , and an evaporator  204 . These components are connected to each other by pipes so as to form a main circuit  208 . The compressor  201  and the expander  203  are coupled together by a shaft  207 . A motor  206  for rotationally driving the shaft  207  is disposed between the compressor  201  and the expander  203 . The compressor  201 , the expander  203 , the shaft  207 , and the motor  206  constitute the expander-integrated compressor. 
     This refrigeration cycle apparatus further includes a secondary circuit  209  that is connected to the main circuit  208  so as to be provided in parallel to the expander  203 . The secondary circuit  209  branches from the main circuit  208  between the outlet of the radiator  202  and the inlet of the expander  203 , and merges with the main circuit  208  between the outlet of the expander  203  and the inlet of the evaporator  204 . A working fluid flowing through the main circuit  208  expands in the positive-displacement expander  203 . The working fluid flowing through the secondary circuit  209  expands in an expansion valve  205 . 
     The working fluid is compressed by the compressor  201 . The compressed working fluid is delivered to the radiator  2 , and cooled in the radiator  202 . The working fluid expands in the expander  203  or the expansion valve  205 , and then is heated in the evaporator  204 . The expander  203  recovers the expansion energy of the working fluid, and converts the recovered energy into the rotational energy of the shaft  207 . This rotational energy is used as part of the work for driving the compressor  201 . As a result, the power consumption of the motor  206  is reduced. 
     How the refrigeration cycle apparatus operates when the expansion valve  205  is fully closed will be described. 
     First, the suction volume of the compressor  201 , the suction volume of the expander  203 , the rotational speed of the shaft  207  are denoted as Vcs, Ves, and N, respectively. In this case, the volumetric flow rate of the working fluid at the inlet of the compressor  201  is expressed as (Vcs×N). The volumetric flow rate of the working fluid at the inlet of the expander  203  is expressed as (Ves×N). Since the mass flow rate of the working fluid in the secondary circuit  209  is zero, the mass flow rate thereof in the compressor  201  and that in the expander  203  are equal to each other. If this mass flow rate is denoted as G, the density of the working fluid at the inlet of the compressor  201  is expressed as {G/(Vcs×N)}. The density of the working fluid at the inlet of the expander  203  is expressed as {G/(Ves×N)}. Based on these formulas, the ratio between the density of the working fluid at the inlet of the compressor  201  and that at the inlet of the expander  203  is expressed as {G/(Vcs×N)}/{G/(Ves×N)}. That is, the density ratio (Ves/Vcs) is always constant regardless of the rotational speed of the shaft  207  (constraint of constant density ratio). 
       FIG. 29  shows a Mollier diagram of a CO 2  refrigeration cycle. The compression process in the compressor  201 , the heat radiation process in the radiator  202 , the expansion process in the expander  203 , and the evaporation process in the evaporator  204  correspond to AB, BC, CD, and DA, respectively. The ratio between the density of the working fluid at the inlet of the compressor  201  (Point A) and that at the inlet of the expander  203  (Point C) is (Ves/Vcs). If the density at Point A is ρ 0 , the density ρ c  at Point C is (Vcs/Ves)ρ 0 . When the density ρ 0  of the working fluid at the inlet of the compressor  201  (Point A) is constant, the state of the working fluid at the inlet of the expander  203  (Point C) always changes along the line that satisfies the relationship of ρ c =(Vcs/Ves)ρ 0 . That is, the temperature and pressure of the working fluid at Point C cannot be controlled freely. The refrigeration cycle has an optimum high pressure at which the highest coefficient of performance (COP) is achieved at a certain heat source temperature (for example, an outside air temperature). Therefore, if the temperature and pressure cannot be controlled freely, it is difficult to operate the refrigeration cycle apparatus efficiently. 
     There have been several proposals to avoid the constraint of constant density ratio. For example, in the refrigeration cycle apparatus shown in  FIG. 28 , the constraint of constant density ratio can be avoided by opening the expansion valve  205  to allow a part of the working fluid to flow into the secondary circuit  209 . This method, however, has a problem in that the expansion energy of the working fluid flowing through the secondary circuit  209  cannot be recovered, which reduces the effect of improving the COP. 
     Patent Literature 2 discloses an expander including an auxiliary chamber that can communicate with an expansion chamber. With this expander, the volumetric capacity of the expansion chamber can be increased or decreased by increasing or decreasing the volumetric capacity of the auxiliary chamber. The suction volume of the expander Ves changes with an increase or a decrease in the volumetric capacity of the expansion chamber. Thus, the constraint of constant density ratio can be avoided. Nevertheless, this expander has a problem in that the working fluid remains in the auxiliary chamber. It also has another problem of sealing a piston for increasing or decreasing the volumetric capacity of the auxiliary chamber. 
     CITATION LIST 
     Patent Literature 
     Patent Literature 1 JP 2001-116371 A 
     Patent Literature 2 JP 2006-46257 A 
     SUMMARY OF INVENTION 
     Technical Problem 
     The present invention has been made in view of the above circumstances, and it is an object of the present invention to provide a two-stage rotary expander in which both the avoidance of the constraint of constant density ratio and the efficient power recovery can be achieved. It is another object of the present invention to provide an expander-integrated compressor using this two-stage rotary expander. It is still another object of the present invention to provide a refrigeration cycle apparatus using this expander-integrated compressor. 
     Solution to Problem 
     The present invention provides a two-stage rotary expander including: a first cylinder; a first piston disposed rotatably in the first cylinder; a second cylinder disposed concentrically with the first cylinder; a second piston disposed rotatably in the second cylinder; a shaft on which the first piston and the second piston are mounted; a first vane, disposed slidably in a first vane groove formed in the first cylinder, for partitioning a space between the first cylinder and the first piston into a first suction space and a first discharge space; a second vane, disposed slidably in a second vane groove formed in the second cylinder, for partitioning a space between the second cylinder and the second piston into a second suction space and a second discharge space; an intermediate plate for separating the first cylinder from the second cylinder, the intermediate plate having a through-hole that communicates the first discharge space with the second suction space so as to form one expansion chamber; and a variable vane mechanism for controlling movement of the first vane so that a ratio of a period P 2  to a period P 1  (P 2 /P 1 ) can be adjusted, where P 1  denotes the period during which the first vane is in contact with the first piston in the course of one rotation of the shaft, and P 2  denotes the period during which the first vane is spaced from the first piston in the course of one rotation of the shaft. 
     In another aspect, the present invention provides an expander-integrated compressor including: a compression mechanism for compressing a working fluid; an expansion mechanism for expanding the working fluid; and a shaft that couples the compression mechanism and the compression mechanism. In this expander-integrated compressor, the expansion mechanism is constituted by the above-mentioned two-stage rotary expander of the present invention. 
     In still another aspect, the present invention provides a refrigeration cycle apparatus including: the above-mentioned expander-integrated compressor of the present invention; a radiator for cooling a working fluid that has been compressed in a compression mechanism of the expander-integrated compressor; and an evaporator for evaporating a working fluid that has been expanded in an expansion mechanism of the expander-integrated compressor. 
     Advantageous Effects of Invention 
     The two-stage rotary expander of the present invention includes a variable vane mechanism for controlling the movement of the first vane. By the action of the variable vane mechanism, the first vane is spaced from the first piston during the period P 2 , which is a part of the period of one rotation of the shaft, so that the working fluid in the first suction space can flow directly into the first discharge space. When the ratio (P 2 /P 1 ) changes under the control of the movement of the first vane, the suction volume (volumetric flow rate) of the expansion mechanism also changes. That is, the constraint of constant density ratio can be avoided. In addition, since the power can be recovered from the entire amount of the working fluid, a high power recovery efficiency can be achieved. 
     Here, the minimum value of the period P 2  may be zero. When the period P 2  is zero, the first vane and the first piston are in contact with each other all the time, and thus the suction volume of the two-stage rotary expander is minimized. More specifically, the variable vane mechanism controls the movement of the first vane so that one of the following (a) and (b) is achieved. 
     (a) The variable vane mechanism controls the movement of the first vane so that a first mode and a second mode can be switched to each other. In the first mode, the first vane is always in contact with the first piston, and in the second mode, the period of one rotation of the shaft includes the period P 1  during which the first vane is in contact with the first piston and the period P 2  during which the first vane is spaced from the first piston. 
     (b) The variable vane mechanism controls the movement of the first vane so that the period of one rotation of the shaft includes the period P 1  during which the first vane is in contact with the first piston and the period P 2  during which the first vane is spaced from the first piston, and that the ratio of the period P 2  to the period P 1  (P 2 /P 1 ) can be adjusted. 
     The two-stage rotary expander of the present invention can be used suitably as an expansion mechanism of an expander-integrated compressor in which it is difficult to control the rotational speed of the compression mechanism and the rotational speed of the expansion mechanism independently. In the refrigeration cycle apparatus using such an expander-integrated compressor, power can be recovered efficiently by controlling the variable vane mechanism properly. Accordingly, a high COP can be achieved. 
    
    
     
       BRIEF DESCRIPTION OF DRAWINGS 
         FIG. 1  is a configuration diagram showing a refrigeration cycle apparatus according to a first embodiment of the present invention. 
         FIG. 2  is a longitudinal cross-sectional view of an expander-integrated compressor shown in  FIG. 1 . 
         FIG. 3A  is a transverse cross-sectional view of the expander-integrated compressor shown in  FIG. 2 , taken along the line D 1 -D 1 . 
         FIG. 3B  is a transverse cross-sectional view of the expander-integrated compressor shown in  FIG. 2 , taken along the line D 2 -D 2 . 
         FIG. 4A  is a partially enlarged view of  FIG. 3A , showing a variable vane mechanism at the minimum suction volume.  FIG. 4B  is a partially enlarged view of  FIG. 3A , showing the variable vane mechanism at a larger suction volume than in  FIG. 4A . 
         FIG. 5  is a diagram showing the operating principle of an expansion mechanism at the minimum suction volume. 
         FIG. 6  is a diagram showing the operating principle of the expansion mechanism at a larger suction volume than in  FIG. 5 . 
         FIG. 7A  is a graph corresponding to  FIG. 5 , showing the position of the tip of a first vane. 
         FIG. 7B  is a graph corresponding to  FIG. 6 , showing the position of the tip of the first vane. 
         FIG. 8  is a configuration diagram showing a refrigeration cycle apparatus according to a second embodiment of the present invention. 
         FIG. 9  is a configuration diagram showing a refrigeration cycle apparatus according to a third embodiment of the present invention. 
         FIG. 10A  is a partially enlarged view of a variable vane mechanism using an electric actuator. 
         FIG. 10B  is a partially enlarged view of the variable vane mechanism at a larger suction volume than in  FIG. 10A . 
         FIG. 11  is a configuration diagram showing a refrigeration cycle apparatus according to a fourth embodiment of the present invention. 
         FIG. 12  is a longitudinal cross-sectional view of an expander-integrated compressor shown in  FIG. 11 . 
         FIG. 13A  is a transverse cross-sectional view of the expander-integrated compressor shown in  FIG. 12 , taken along the line D 3 -D 3 . 
         FIG. 13B  is a transverse cross-sectional view of the expander-integrated compressor shown in  FIG. 12 , taken along the line D 4 -D 4 . 
         FIG. 14A  is a partially enlarged view of  FIG. 13A , showing a variable vane mechanism at the minimum confined volume. 
         FIG. 14B  is a partially enlarged view of  FIG. 13A , showing the variable vane mechanism at a larger confined volume than in  FIG. 14A . 
         FIG. 15  is a diagram showing the operating principle of an expansion mechanism at the minimum confined volume. 
         FIG. 16  is a diagram showing the operating principle of the expansion mechanism at a larger confined volume than in  FIG. 15 . 
         FIG. 17A  is a graph showing the position of the tip of a first vane with respect to the rotation angle of a shaft. 
         FIG. 17B  is a graph showing the pressure of a working fluid with respect to the rotation angle of the shaft. 
         FIG. 17C  is a graph showing the volumetric capacity of a working chamber with respect to the rotation angle of the shaft. 
         FIG. 18  is a transverse cross-sectional view of a modified variable vane mechanism of the fourth embodiment. 
         FIG. 19  is a configuration diagram showing a refrigeration cycle apparatus according to a fifth embodiment of the present invention. 
         FIG. 20  is a partially enlarged view of a variable vane mechanism using an electromagnetic force to brake the first vane. 
         FIG. 21  is a partially enlarged view of another example of a variable vane mechanism using an electromagnetic force to brake the first vane. 
         FIG. 22  is a partially enlarged view of a variable vane mechanism for applying a load to brake the first vane. 
         FIG. 23  is a partially enlarged view of another example of a variable vane mechanism for applying a load to brake the first vane. 
         FIG. 24  is a diagram showing how to control an electric actuator. 
         FIG. 25  is a timing diagram showing how to control the electric actuator. 
         FIG. 26  is a configuration diagram showing a refrigeration cycle apparatus according to a sixth embodiment of the present invention. 
         FIG. 27  is a graph showing the relationship between power generator efficiency and rotation speed. 
         FIG. 28  is a configuration diagram showing a conventional refrigeration cycle apparatus using an expander-integrated compressor. 
         FIG. 29  is a Mollier diagram of a CO 2  refrigeration cycle. 
     
    
    
     DESCRIPTION OF EMBODIMENTS 
     Hereinafter, some of the embodiments of the present invention will be described with reference to the drawings. 
     (First Embodiment) 
     As shown in  FIG. 1 , a refrigeration cycle apparatus  200 A of the present embodiment includes a compression mechanism  2 , a radiator  101 , an expansion mechanism  3 , an evaporator  102 , and a plurality of pipes  103   a  to  103   d  for connecting these components to each other so as to form a refrigerant circuit. The compression mechanism  2  and the expansion mechanism  3  are coupled together by a shaft  5  so as to constitute an expander-integrated compressor  100 . The basic operation of the refrigeration cycle apparatus  200 A is as described in the background art. 
     The expansion mechanism  3  of the expander-integrated compressor  100  is provided with a variable vane mechanism  60 . The variable vane mechanism  60  has a function of changing the volume (volumetric flow rate) of a working fluid to be drawn into the expansion mechanism  3  during one rotation of the shaft  5 . In other words, it has a function of changing the suction volume of the expansion mechanism  3 . The constraint of constant density ratio can be avoided by changing the volumetric flow rate of the expansion mechanism  3  according to the operation state of the refrigeration cycle apparatus  200 A. 
     In the present embodiment, a method of injecting a high-pressure working fluid into the expansion chamber is employed as a method of changing the volumetric flow rate of the expansion mechanism  3 . That is, the variable vane mechanism  60  can be a mechanism for injecting the working fluid into the expansion chamber. 
     The refrigeration cycle apparatus  200 A further includes a pressure supply circuit  110  for driving the actuator of the variable vane mechanism  60 . It should be noted, however, that in the present embodiment, this pressure supply circuit  110  is not a supply circuit for the working fluid to be injected into the expansion chamber. The pressure supply circuit  110  includes a throttle valve  104 , a pipe  105  and a fine passage  106 . The working fluid, whose pressure is adjusted to a predetermined one by the pressure supply circuit  110 , is supplied to the variable vane mechanism  60 . 
     The pipe  105  has one end connected to a portion (pipe  103   b ) between the radiator  101  and the expansion mechanism  3  in the refrigerant circuit, and the other end connected to the variable vane mechanism  60  of the expansion mechanism  3 . The throttle valve  104  is an opening-adjustable valve (for example, an electric expansion valve), and is provided on the pipe  105 . The portion between the throttle valve  104  and the variable vane mechanism  60  in the pipe  105  and the portion (pipe  103   c ) from the outlet of the expansion mechanism  3  to the inlet of the evaporator  102  in the refrigerant circuit are connected by the fine passage  106 . A specific example of the fine passage  106  is a capillary. 
     As shown in  FIG. 2 , the expander-integrated compressor  100  includes a closed casing  1 , the compression mechanism  2 , the expansion mechanism  3 , a motor  4 , and the shaft  5 . The compression mechanism  2  is disposed in the upper part in the closed casing  1 . The expansion mechanism  3  is disposed in the lower part in the closed casing  1 . The motor  4  is disposed between the compression mechanism  2  and the expansion mechanism  3 . The compression mechanism  2 , the motor  4 , and the expansion mechanism  3  are coupled together by the shaft  5  so as to transmit power therebetween. 
     The compression mechanism  2  is actuated when the motor  4  drives the shaft  5 . The expansion mechanism  3  recovers the power from the expanding working fluid and provides the recovered power to the shaft  5  so as to assist the motor  4  in driving the shaft  5 . Specific examples of the working fluid include refrigerants such as carbon dioxide and hydrofluorocarbon. 
     In the present embodiment, the positions of the compression mechanism  2 , the motor  4 , and the expansion mechanism  3  are determined so that the axial direction of the shaft  5  coincides with the vertical direction. This positional relationship between the compression mechanism  2  and the expansion mechanism  3  in the present embodiment may be reversed. That is, the compression mechanism  2  may be disposed in the lower part in the closed casing  1 , and the expansion mechanism  3  may be disposed in the upper part in the closed casing  1 . 
     The closed casing  1  has an interior space  24  for accommodating the components. The interior space  24  of the closed casing  1  is filled with the working fluid that has been compressed in the compression mechanism  2 . The bottom of the closed casing  1  is used as an oil reservoir  25 . Oil is used to ensure the lubrication and sealing of the sliding parts in the compression mechanism  2  and the expansion mechanism  3 . The amount of oil in the oil reservoir  25  is regulated so that the oil level is maintained below the motor  4 . Therefore, it is possible to prevent the rotor of the motor  4  from agitating the oil and thus prevent a decrease in the motor efficiency and an increase in the amount of oil discharged into the refrigerant circuit. 
     The scroll compression mechanism  2  includes an orbiting scroll  7 , a stationary scroll  8 , an Oldham ring  11 , a bearing member  10 , a muffler  16 , a suction pipe  13 , a discharge pipe  15 , and a reed valve  19 . The bearing member  10  is fixed to the closed casing  1  by a technique, such as welding or shrink fitting, to support the shaft  5 . The stationary scroll  8  is fixed to the bearing member  10  by a fastening member such as a bolt. The orbiting scroll  7  is fitted to the eccentric axis  5   a  of the shaft  5  between the stationary scroll  8  and the bearing member  10 , and is prevented by the Oldham ring  11  from rotating on its own axis. 
     The orbiting scroll  7 , with its spiral wrap  7   a  meshing with the wrap  8   a  of the stationary scroll  8 , moves in an orbit as the shaft  5  rotates. A crescent-shaped working chamber  12  formed between the wrap  7   a  and the wrap  8   a  decreases its volumetric capacity as it moves inwardly, and compresses the working fluid drawn through the suction pipe  13 . The compressed working fluid pushes open the reed valve  19  to be discharged into the interior space  16   a  of the muffler  16  through a discharge hole  8   b  formed in the center of the stationary scroll  8 . The working fluid further is discharged into the interior space  24  of the closed casing  1  through a flow path  17  penetrating the stationary scroll  8  and the bearing member  10 . Then, the working fluid is delivered to the radiator  101  through the discharge pipe  15 . 
     The compression mechanism  2  may be constituted by another type of positive displacement compression mechanism (for example, a rotary compression mechanism). 
     The motor  4  includes a stator  21  fixed to the closed casing  1  and a rotor  22  fixed to the shaft  5 . Electric power is supplied from a power source  108  to the motor  4  through a terminal  107  provided above the closed casing  1  (see  FIG. 1 ). 
     The shaft  5  may be made up of a single part, or may be made up of a combination (coupling) of a plurality of parts. If the shaft  5  is made up of a combination of a plurality of parts, the assembly is easy, and in particular, the alignment of the compression mechanism  2  and the expansion mechanism  3  is easy. 
     The expansion mechanism  3  has a structure of a multi-stage rotary expander. Specifically, the expansion mechanism  3  includes a first cylinder  42 , a second cylinder  44  with a greater thickness than the first cylinder  42 , and an intermediate plate  43  for separating the first cylinder  42  from the second cylinder  44 . The first cylinder  42  and the second cylinder  44  are disposed concentrically with each other. As shown in  FIG. 3A  and  FIG. 3B , the expansion mechanism  3  further includes a first piston (first roller)  46 , a first vane  48 , a first spring  50 , a second piston (second roller)  47 , a second vane  49 , and a second spring  51 . The first cylinder  42  has the variable vane mechanism  60  built therein. 
     As shown in  FIG. 3A , the first piston  46  is fitted to the eccentric portion  5   c  of the shaft  5  so as to rotate eccentrically in the first cylinder  42 . The first vane  48  is provided slidably in a first vane groove  42   a  formed in the first cylinder  42 . One end (tip) of the first vane  48  is in contact with the first piston  46 . The first spring  50  is in contact with the other end (rear end) of the first vane  48  and pushes the first vane  48  toward the first piston  46 . 
     As shown in  FIG. 3B , the second piston  47  is fitted to the eccentric portion  5   d  of the shaft  5  so as to rotate eccentrically in the second cylinder  44 . The second vane  49  is provided slidably in a second vane groove  44   a  formed in the second cylinder  44 . One end of the second vane  49  is in contact with the second piston  47 . The second spring  51  is in contact with the other end of the second vane  49  and pushes the second vane  49  toward the second piston  47 . 
     As shown in  FIG. 2 , the expansion mechanism  3  further includes a lower bearing member  41  and an upper bearing member  45 . The upper bearing member  45  is fitted in the closed casing  1  with no space therebetween. The components such as the cylinders and the intermediate plate are fixed to the closed casing  1  by the upper bearing member  45 . The lower bearing member  41  and the intermediate plate  43  close the first cylinder  42  from below and above respectively. The intermediate plate  43  and the upper bearing member  45  close the second cylinder  44  from below and above respectively. As a result, a working chamber is formed in each of the first cylinder  42  and the second cylinder  44 . A suction port  42   p  for drawing the working fluid into the working chamber of the first cylinder  42  is formed in the lower bearing member  41 . A discharge port  45   q  for discharging the working fluid from the working chamber of the second cylinder  44  is formed in the upper bearing member  45 . 
     As shown in  FIG. 3A , a suction-side working chamber  55   a  and a discharge-side working chamber  55   b  are formed in a space inside the first cylinder  42 . The working chamber  55   a  and the working chamber  55   b  are partitioned by the first piston  46  and the first vane  48 . As shown in  FIG. 3B , a suction-side working chamber  56   a  and a discharge-side working chamber  56   b  are formed in a space inside the second cylinder  44 . The working chamber  56   a  and the working chamber  56   b  are partitioned by the second piston  47  and the second vane  49 . Hereinafter, the working chambers  55   a ,  55   b ,  56   a , and  56   b  are also referred to as a first suction space  55   a , a first discharge space  55   b , a second suction space  56   a , and a second discharge  56   b , respectively 
     The total volumetric capacity of the working chamber  56   a  and the working chamber  56   b  in the second cylinder  44  is greater than that of the working chamber  55   a  and the working chamber  55   b  in the first cylinder  42 . The discharge side working chamber  55   b  in the first cylinder  42  and the suction-side working chamber  56   a  in the second cylinder  44  communicate with each other through a through-hole  43   a  formed in the intermediate plate  43 . Thus, the working chamber  55   b  and the working chamber  56   a  function as a single expansion chamber. 
     In the present embodiment, the thickness of the first cylinder  42  and that of the second cylinder  44  are made different from each other to obtain a greater total volumetric capacity of the working chamber  56   a  and the working chamber  56   b  than that of the working chamber  55   a  and the working chamber  55   b . In this regard, it is also possible to adopt a configuration in which the inner diameters of the cylinders or the outer diameters of the pistons are made different from each other. Furthermore, the second piston  47  and the second vane  49  may be integrated as a single unit, called a swinging piston. 
     As shown in  FIG. 2 , the expansion mechanism  3  further includes a suction pipe  52  for drawing the working fluid to be expanded directly from the outside of the closed casing  1 , and a discharge pipe  53  for discharging the expanded working fluid directly to the outside of the closed casing  1 . The suction pipe  52  is inserted directly into the lower bearing member  41  and connected to the suction port  41   p  so that the working fluid can be delivered from the outside of the closed casing  1  to the working chamber  55  of the first cylinder  42 . The discharge pipe  53  is inserted directly into the upper bearing member  43  and connected to the discharge port  45   q  so that the working fluid can be delivered from the working chamber  56  of the second cylinder  44  to the outside of the closed casing  1 . 
     The working fluid to be expanded passes through the suction pipe  52  and the suction port  41   p , and then flows into the working chamber  55   a  of the first cylinder  42 . The working fluid that has flowed into the working chamber  55   a  of the first cylinder  42  moves to the working chamber  55   b  as the shaft  5  rotates, and expands in the expansion chamber formed by the working chamber  55   b , the through-hole  43   a , and the working chamber  56   a , while rotating the shaft  5 . The working fluid thus expanded is delivered to the outside of the closed casing  1  through the working chamber  56   b , the discharge port  45   q , and the discharge pipe  53 . 
       FIG. 4A  shows an enlarged view of the variable vane mechanism at the minimum suction volume.  FIG. 4B  shows an enlarged view of the variable vane mechanism at a larger suction volume than in  FIG. 4A . In the present description, a period during which the tip of the first vane  48  is in contact with the first piston  46  in the course of one rotation of the shaft  5  is denoted as P 1 , and a period during which the tip of the first vane  48  is spaced from the first piston  46  in the course of one rotation of the shaft  5  is denoted as P 2 . During the period P 2 , the working fluid can flow from the first suction space  55   a  into the first discharge space  55   b . The variable vane mechanism  60  controls the movement of the first vane  48  so that the ratio of the period P 2  to the period P 1 (P/P 1 ) can be adjusted. The length of the period P 1  and the length of the period P 2  each can be represented by an angle (in degrees). When the ratio (P 2 /P 1 ) changes, the suction volume (volumetric flow rate) of the expansion mechanism  3  also changes. That is, the constraint of constant density ratio can be avoided. The power recovery efficiency can be optimized by adjusting the ratio (P 2 /P 1 ) according to the heat source temperature (for example, an outside air temperature). 
     In the present embodiment, the suction volume of the expansion mechanism  3  is minimum when the period P 2  is 0, that is, when the first vane  48  and the first piston  46  are always in contact with each other. In this regard, the minimum value of the period P 2  may be greater than zero. 
     As shown in  FIG. 4A  and  FIG. 4B , the variable vane mechanism  60  includes a stopper  61  and an actuator  62 . The stopper  61  serves to limit the range of movement of the first vane  48 . The actuator  62  serves to move the stopper  61  in the direction from a position for increasing the range of the movement of the first vane  48  to a position for reducing the range of the movement, or in the opposite direction. This mechanism is advantageous in that the actuator  62  moves the stopper  61  so that the length of the stroke of the first vane  48  can be changed mechanically. Furthermore, this mechanism rarely requires a high precision control technique because the stopper  61  does not need to be moved according to the rotation angle of the shaft  5 , and therefore is highly reliable. 
     Specifically, the actuator  62  is composed of a main body  65 , a pressure chamber  67  in which the main body  65  is placed, and a passage  69  for supplying a fluid to the pressure chamber  67 . The main body  65  includes a portion working with the stopper  61 , and determines, based on the pressure of the fluid, the position of the stopper  61  with respect to the longitudinal direction of the first vane groove  42   a . Thus, in the present embodiment, a fluid pressure actuator is used as the actuator  62 . The working fluid in the refrigeration cycle apparatus  200 A is used as the fluid to be supplied to the pressure chamber  67 . The use of the working fluid as a power source allows some leakage of the working fluid from the pressure chamber  67  to the first vane groove  42   a . Therefore, tight sealing is not required. 
     The main body  65  includes a slider  63  disposed slidably in the pressure chamber  67  to partition the pressure chamber  67  into sections, and a spring  64  provided in one section  67   b  of the pressure chamber  67  partitioned by the slider  63 . The stopper  61  is integrated with the slider  63 . The passage  69  is connected to the other section  67   a  of the pressure chamber  67  partitioned by the slider  63 . Like the first vane groove  42   a , the pressure chamber  67  and the passage  69  are spaces formed in the first cylinder  42 . The pipe  105  of the pressure supply circuit  110 , which has been described with reference to  FIG. 1 , is connected to the passage  69 . The position of the stopper  61  with respect to the longitudinal direction of the first vane groove  42   a  is determined based on the force applied to the slider  63  by the working fluid that has been supplied to the pressure chamber  67   a  through the pipe  105  and the passage  69  and the force applied to the slider  63  by the spring  64 . The stopper  61  can move, together with the slider  63 , in the direction parallel to the longitudinal direction of the first vane groove  42   a . In such a configuration, the position of the stopper  61  can be changed freely and continuously by adjusting the pressure in the pressure chamber  67   a . This means that the power recovery efficiency can be optimized easily. 
     Furthermore, it is possible to adopt not only the mechanism for changing the position of the stopper  61  continuously but also the mechanism for changing the position of the stopper  61  stepwise. In some cases, the position of the stopper  61  may only need to be changed from one position with a larger ratio (P 2 /P 1 ) to the other position with a smaller ratio (P 2 /P 1 ), or from the other position to the one position. 
     The pressure chamber  67  and the passage  69  may be formed in the bearing member  41  of the expansion mechanism  3  (see  FIG. 2 ). That is, the variable vane mechanism  60  may be built in the bearing member  41 . The stopper  61  and the slider  63  may be constituted by separate components. In this case, the slider  63  and the stopper  61  may be coupled together by direct fitting, or they may be coupled together by another member. 
     The first vane  48  has a recessed portion  48   k  (notched groove) for laterally receiving the stopper  61 . The pressure chamber  67  of the fluid pressure actuator  62  is formed adjacent to the first vane groove  42   a  in the first cylinder  42 . A groove  68  for allowing the stopper  61  to pass through is formed between the first vane groove  42   a  and the pressure chamber  67 . One end of the stopper  61  is fixed to the slider  63  and the other end thereof is inserted into the recessed portion  48   k  so that the stopper  61  extends from the pressure chamber  67  to the first vane groove  42   a  through the groove  68 . In such a configuration, the range of the movement of the first vane  48  can be limited easily by fitting the stopper  61  in the recessed portion  48   k  of the first vane  48 . 
     The relationship of Lc&gt;Ws+Tmax is satisfied when the length of the recessed portion  48   k  with respect to the longitudinal direction of the first vane groove  42   a  is Lc, the width of the stopper  61  with respect to this longitudinal direction is Ws, and the maximum length of the stroke of the first vane  48  is Tmax. When this relationship is satisfied, the period P 2  of 0 can be selected, that is, the interference between the first vane  48  and the stopper  61  can be avoided, and as a result, a wide range of adjustment of the suction volume can be achieved. 
     In the operation mode (first mode) shown in  FIG. 4A , the pressure chamber  67   a  is filled with the high-pressure working fluid, and thus the slider  63  and the stopper  61  are pressed downward. When the stopper  61  is in this position, the stopper  61  and the first vane  48  do not interfere with each other, and thus the range of the movement of the first vane  48  is not limited. The first vane  48  can move freely within the maximum stroke Tmax, so that the contact state between the first vane  48  and the first piston  46  is always maintained. 
     On the other hand, in the operation mode (second mode) shown in  FIG. 4B , the pressure chamber  67   a  is filled with the low-pressure or intermediate-pressure working fluid, and thus the slider  63  and the stopper  61  move to a position above the position shown in  FIG. 4A . Specifically, the slider  63  and the stopper  61  move to the position where the force applied to the slider  63  by the working fluid filled in the pressure chamber  67   a  and the force applied to the slider  63  by the spring  64  (elastic force) are balanced with each other. When the stopper  61  is in this position, the stopper  61  and the first vane  48  interfere with each other, and thus the range of the movement of the first vane  48  is limited. As a result, the first vane  48  cannot move to the lowest point. During the period P 2  in which the movement of the first vane  48  is restricted by the stopper  61 , the first vane  48  is spaced from the first piston  46 . During this period, the high-pressure working fluid filled in the working chamber  55   a  (first suction space) flows directly into the working chamber  55   b  (first discharge space) filled with the intermediate-pressure working fluid. 
     When the pressure in the pressure chamber  67   a  is changed, the position of the stopper  61  changes, and the period P 2  (injection period) changes accordingly. The lower the pressure in the pressure chamber  67   a  is, the higher the stopper  61  is positioned. Therefore, the range of the movement of the first vane is reduced accordingly. Then, the period P 1  in which the first vane  48  is in contact with the first piston  46  becomes progressively shorter while the period P 2  becomes progressively longer, and as a result, the working fluid in the working chamber  55   a  flows more into the working chamber  55   b . In this way, the amount of the working fluid injected into the expansion chamber can be adjusted by adjusting the pressure in the pressure chamber  67   a . In other words, the suction volume of the expansion mechanism  3  can be adjusted freely. 
     The pressure in the pressure chamber  67   a  can be adjusted by the throttle valve  104  of the pressure adjustment circuit  110 . That is, the position of the stopper  61  can be controlled by adjusting the opening of the throttle valve  104 . When the opening of the throttle valve  104  is increased, the pressure in the pressure chamber  67   a  increases, and the stopper  61  moves downward. As a result, the injection amount decreases to a smaller value or to zero. When the opening of the throttle valve  104  is reduced, the pressure in the pressure chamber  67   a  decreases, and the stopper  61  moves upward. As a result, the injection amount increases. 
     As described with reference to  FIG. 1 , the fine passage  106  bridges the pipe  105  and the pipe  103   c  between the throttle valve  104  and the variable vane mechanism  60 . Therefore, the pressure in the pressure chamber  67   a  of the variable vane mechanism  60  can be changed between the high pressure and the low pressure of the refrigeration cycle by adjusting the opening of the throttle valve  104 . The amount of the working fluid flowing through the fine passage  106  is so small that it has little effect on the power recovery efficiency. 
     Next, the operating principle of the expansion mechanism  3  at the minimum suction volume is described with reference to  FIG. 5 . 
     As shown in Step A 1  in  FIG. 5 , when the first piston  46  rotates in a counterclockwise direction and the suction port  41   p  is opened, the drawing of the working fluid into the first suction space  55   a  (suction process) starts. Next, as shown in Step B 1  and Step C 1  in  FIG. 5 , as the first piston  46  rotates, the working fluid is further drawn into the first suction space  55   a . As shown in Step D 1  in  FIG. 5 , when the first piston  46  further rotates and the suction port  41   p  is closed, the drawing of the working fluid into the first suction space  55   a  is completed. 
     When the suction process is completed, the first suction space  55   a  is shifted to the first discharge space  55   b . As described with reference to  FIG. 3A  and  FIG. 3B , the first discharge space  55   b  and the second suction space  56   a  communicate with each other through the through-hole  43   a . As shown in Steps A 1  to C 1  in  FIG. 5 , the working fluid filled in the first discharge space  55   b  moves to the second suction space  56   a  of the second cylinder  44  through the through-hole  43   a  as the first piston  46  rotates. The increase in the volumetric capacity of the second suction space  56   a  with the rotation of the shaft  5  is greater than the decrease in the volumetric capacity of the first discharge space  55   b . Therefore, the working fluid expands in the first discharge space  55   b , the through-hole  43   a , and the second suction space  56   a  (expansion process). When the first piston  46  closes the through-hole  43   a  completely, the movement of the working fluid into the second suction space  56   a  and the expansion thereof are completed. 
     When the expansion process is completed, the second suction space  56   a  is shifted to the second discharge space  56   b , as described with reference to  FIG. 3B . The discharge of the working fluid filled in the second discharge space  56   b  to the outside through the discharge port  45   q  (discharge process) starts. When the second piston  47  further rotates and the discharge port  45   q  is closed, the discharge of the working fluid in the second discharge space  56   b  to the outside is completed. By repeating the above processes, the working fluid expands and the expansion energy is recovered. 
     Next, the operating principle of the expansion mechanism  3  at a larger suction volume than in  FIG. 5  will be described with reference to  FIG. 6 . 
     As shown in Step A 2  in  FIG. 6 , when the first piston  46  rotates in a counterclockwise direction and the suction port  41   p  is opened, the drawing of the working fluid into the first suction space  55   a  (suction process) starts. Next, as shown in Step B 2  in  FIG. 6 , when the first piston  46  further rotates, the first vane  48  and the stopper  61  interfere with each other, and thus the first vane  48  is prevented from moving (downward). As a result, the first vane  48  is detached from the first piston  46 , and a flow path from the first suction space  55   a  to the first discharge space  55   b  is formed. Thus, the high-pressure working fluid in the first discharge space  55   a  flows into the first discharge space  55   b . The high-pressure working fluid also flows into the second suction space  56   a  that communicates with the first discharge space  55   b . That is, the first vane  48  is detached from the first piston  46  in the course of the expansion of the working fluid in the expansion chamber, so that the working fluid to be expanded is injected into the expansion chamber. 
     As shown in Step C 2  in  FIG. 6 , when the first piston  46  further rotates and the first vane  48  and the first piston  46  again come into contact with each other, the first suction space  55   a  and the first discharge space  55   b  are separated again by the first vane  48 . Thus, the flow of the working fluid from the first suction space  55   a  to the first discharge space  55   b  is inhibited. As shown in Step D 2  in  FIG. 6 , when the first piston  46  further rotates and the suction port  41   p  is closed, the drawing of the working fluid into the first suction space  55   a  is completed. When the suction process is completed, the first suction space  55   a  is shifted to the first discharge space  55   b . The first discharge space  55   b  and the second suction space  56   a  communicate with each other through the through-hole  43   a , and the expansion process starts. The operations in Steps A 2  to D 2  in  FIG. 6  are repeated in this manner. 
       FIG. 7A  is a graph corresponding to  FIG. 5 , showing the position of the tip of the first vane. The vertical axis represents the position of the tip of the first vane  48 . The position of the tip of the first vane  48  corresponds to the distance from the rotational axis of the shaft  5  to the tip of the first vane  48 . The horizontal axis represents the rotation angle of the shaft  5  with respect to the position of the shaft at the moment when the first piston  46  occupies the top dead center. Specifically, the rotation angles t 0 , t 1 , t 2 , and t 3  are 0 degree, 180 degrees, 360 degrees, and 540 degrees, respectively. The “top dead center” means the position of the piston in a state in which the vane is pressed into the vane groove most inwardly. The “bottom dead center” means the position of the piston 180-degree opposite to the “top dead center”. 
     At the angles t 0  and t 2  at which the first piston  46  is in the top dead center, the tip of the first vane  48  is in the upper limit position  30   a  farthest from the rotational axis of the shaft  5 . At the angles t 1  and t 3  at which the first piston  46  is in the bottom dead center, the tip of the first vane  48  is in the lower limit position  30   b  nearest to the rotational axis of the shaft  5 . The tip of the first vane  48  undergoes simple harmonic motion in synchronism with the rotation of the shaft  5 . 
       FIG. 7B  is a graph corresponding to  FIG. 6 , showing the position of the tip of the first vane. At the angles t 0  and t 2 , the tip of the first vane  48  is in the upper limit position  30   a , as in  FIG. 5 . When the stopper  61  prevents the first vane  48  from moving downward at the angle T 1 , the tip of the first vane  48  occupies the position  30   c  between the upper limit position  30   a  and the lower limit position  30   b . When the first vane  48  and the first piston  46  again come into contact with each other at the angle T 2 , the tip of the first vane  48  begins to be displaced to the upper limit position  30   a . During the period P 2  (the period T 2 -T 1  and the period T 4 -T 3 ) in which the tip of the first vane  48  stays in the position  30   c , the working fluid is injected into the expansion chamber. The injection amount of the working fluid increases or decreases according to the length of the period P 2 . In other words, it increases or decreases according to the ratio of the period P 2  to the period P 1  (P 2 /P 1 ). The length of the period P 2  varies depending on the pressure in the pressure chamber  67   a  of the variable vane mechanism  60 . 
     The range of the ratio (P 2 /P 1 ) is not particularly limited. For example, P 2  is in the range of 0 to 180 (degrees) (0≦P 2 ≦180) and P 2 /P 1  is in the range of 0 to 1 (0≦P 2 /P 1 ≦1). That is, the position of the stopper  61  may be adjusted so that the period P 2  falls within the period in which the rotation angle of the shaft  5  is in the range of 90 to 270 degrees, if the rotation angle of the shaft  5  at the moment when the first piston  46  occupies the top dead center is defined as 0 degree. 
     As described above, with the expansion mechanism  3  provided with the variable vane mechanism  60 , the working fluid can be injected into the expansion chamber at the same time as it is drawn into the first suction space  55   a . Therefore, the volume of the working fluid to be drawn into the expansion mechanism  3  during one rotation of the shaft can be changed. Furthermore, the injection amount can be changed by adjusting the opening of the throttle valve  104 . 
     (Second Embodiment) 
       FIG. 8  shows a refrigeration cycle apparatus according to the second embodiment of the present invention. A refrigeration cycle apparatus  200 B of the present embodiment includes, instead of the pressure supply circuit  110 , a pipe  112  connecting the pipe  103   c  and the variable vane mechanism  60 . This refrigeration cycle apparatus  200 B is different from that of the first embodiment in that the discharge pressure of the expansion mechanism  3  is supplied to the pressure chamber  76   a  of the variable vane mechanism  60 . In the following embodiments, the same components are designated by the same reference numerals, and no further description is given. 
     In the refrigeration cycle apparatus  200 B, the position of the stopper  61  changes according to the discharge pressure of the expansion mechanism  3 , and thus the ratio (P 2 /P 1 ) changes. The lower the discharge pressure of the expansion mechanism  3  is, the higher the stopper  61  is positioned. As a result, the period P 2  in which the first piston  46  and the first vane  48  are spaced from each other is increased, and thus the injection amount increases. Conversely, the higher the discharge pressure of the expansion mechanism  3  is, the lower the stopper  61  is positioned. As a result, the period P 2  in which the first piston  46  and the first vane  48  are spaced from each other is reduced, and thus the injection amount decreases. In this way, the position of the stopper  61  changes automatically according to the discharge pressure of the expansion mechanism  3 , and thus the injection amount increases or decreases automatically. Therefore, efficient operation can be achieved without adjustment of the opening of the valve, or the like. 
     (Third Embodiment) 
     The actuator of the variable vane mechanism is not limited to a fluid pressure actuator.  FIG. 9  is a configuration diagram showing a refrigeration cycle apparatus using an electric actuator as an actuator of the variable vane mechanism. This refrigeration cycle apparatus  200 C has an expander-integrated compressor  100 C. The expansion mechanism  3  in the expander-integrated compressor  100 C is provided with a variable vane mechanism  60 C including an electric actuator. The electric actuator of the variable vane mechanism  60 C is connected to an external controller  70 . The operation of the electric actuator can be controlled by the external controller. The refrigeration cycle apparatus  200 C has an advantage in that the pressure supply circuit  110  described with reference to  FIG. 1  can be omitted. Furthermore, since the positioning accuracy of the stopper can be increased easily by the electric actuator, the injection amount can be optimized more easily. 
     As shown in  FIG. 10A  and  FIG. 10B , in the variable vane mechanism  60 C, a rotary motor  74  is used as an actuator for moving the stopper  610 . The rotary motor  74  and the stopper  610  are coupled together so that the position of the stopper  610  with respect to the longitudinal direction of the first vane groove  42   a  changes when the rotary motor  74  is driven. 
     Specifically, a slide bar  75  with a male-threaded outer peripheral surface is attached to the rotary motor  74 . A groove  76  that communicates with the first vane groove  42   a  through the groove  68  is formed in the first cylinder  42 . A female thread is cut on the inner peripheral surface of the groove  76 . The slide bar  75  is disposed rotatably in the groove  76  in such a manner that the male and female threads are engaged with each other. The stopper  610  is constituted by a component having a T-shaped transverse cross-section. One end of the stopper  610  is inserted into the recessed portion  48   k  of the first vane  48 , and the other end of the stopper  610  is accommodated in the groove  76 . In the groove  76 , the tip of the slide bar  75  is fitted rotatably to the other end of the stopper  610 . When the rotary motor  74  is driven, the slide bar  75  rotates and moves forward or backward in the groove  76 . Along with the movement of the slide bar  75 , the stopper  610  moves in the direction parallel to the longitudinal direction of the first vane groove  42   a . The function and movement of the stopper  610  are basically the same as those of the stopper  61  described in the first embodiment. 
     As shown in  FIG. 10A , when the rotary motor  74  is rotated in the normal direction to press the slide bar  75  and the stopper  610  downward, the stopper  610  and the first vane  48  do not interfere with each other. Therefore, the range of the movement of the first vane  48  is not limited. The first vane  48  can move freely within the maximum stroke Tmax, so that the contact state between the first vane  48  and the first piston  46  is always maintained. 
     On the other hand, as shown in  FIG. 10B , when the rotary motor  74  is rotated in the reverse direction to press the slide bar  75  and the stopper  610  upward, the stopper  610  and the first vane  48  interfere with each other. Therefore, the range of the movement of the first vane  48  is limited, so that the first vane  48  cannot move to the lowest point. During the period P 2  in which the movement of the first vane  48  is restricted by the stopper  610 , the first vane  48  is spaced from the first piston  46 . During this period, the high-pressure working fluid filled in the first suction space  55   a  flows directly into the first discharge space  55   b  (expansion chamber) filled with the intermediate-pressure working fluid. 
     The stopper  610  can be moved by controlling the driving of the rotary motor  74  by the external controller  70  ( FIG. 9 ). When the stopper  610  is moved, the period P 2  in which the first vane  48  is spaced from the first piston  46  changes, and thus the injection amount changes. Since the stopper  610  can be locked securely, the injection amount can be maintained at a constant value easily. 
     A linear motor may be used instead of the rotary motor  74 . A solenoid may be used as an electric actuator. Furthermore, the rotary motor  74  may be a servomotor or a stepping motor. With any of these motors, the position of the stopper  610  with respect to the longitudinal direction of the first vane groove  42   a  can be controlled precisely. Alternatively, a simple positioning element may be used to detect the positions of the slide bar  75  and the stopper  610  and control the driving of the rotary motor  74  based on the detection results. For example, one or a plurality of limit switches may be provided along the longitudinal direction of the slide bar  75 , so that the driving of the rotary motor  74  can be controlled based on the detection signals of the limit switches. 
     Furthermore, the injection amount can be controlled based on the discharge pressure of the expansion mechanism  4  or the evaporation temperature of the working fluid in the evaporator  102 . The injection amount may be controlled based on at least one temperature selected from the group consisting of the discharge temperature of the compression mechanism  2 , the suction temperature of the compression mechanism  2 , and the suction temperature of the expansion mechanism  3 . This also applies to the other embodiments. 
     (Fourth Embodiment) 
     As shown in  FIG. 11 , the basic configuration of a refrigeration cycle apparatus  400 A of the present embodiment is the same as that of the first embodiment described with reference to  FIG. 1 . The refrigeration cycle apparatus  400 A includes an expander-integrated compressor  300  having a variable vane mechanism  130 . In the present embodiment, a method of changing the confined volume of the expansion chamber is employed as a method of changing the volumetric flow rate of the expansion mechanism  3 . The confined volume means the volumetric capacity of the expansion chamber at the time when the working fluid begins to expand. That is, the variable vane mechanism  130  can be a volume-changeable mechanism for changing the volumetric capacity of the expansion chamber at the start of the expansion. 
     The refrigeration cycle apparatus  400 A further includes a pressure supply circuit  110  for adjusting the opening of a valve in the variable vane mechanism  130 . The configuration of the pressure supply circuit  110  is as described with reference to  FIG. 1 . 
     As shown in  FIG. 12 ,  FIG. 13A , and  FIG. 13B , the configuration of the expander-integrated compressor  300  is basically the same as that of the expander-integrated compressor  100  described with reference to  FIG. 2 , except that the variable vane mechanism  130  provided in the expansion mechanism  3 . 
       FIG. 14A  shows an enlarged view of the variable vane mechanism when it is controlled to have the minimum confined volume.  FIG. 14B  shows an enlarged view of the variable vane mechanism when it is controlled to have a larger confined volume than in  FIG. 14A . Also in the present embodiment, the period in which the tip of the first vane  48  is in contact with the first piston  46  in the course of one rotation of the shaft  5  is denoted as P 1 , and the period in which the tip of the first vane  48  is spaced from the first piston  46  in the course of one rotation of the shaft  5  is denoted as P 2 . During the period P 2 , the working fluid can flow from the first suction space  55   a  into the first discharge space  55   b . The variable vane mechanism  130  controls the movement of the first vane  48  so that the ratio of the period P 2  to the period P 1  (P 2 /P 1 ) can be adjusted. 
     In the present embodiment, the point in time when the first piston  46  reaches the top dead center is defined as the starting point of the period P 2 . Therefore, the confined volume of the expansion chamber formed by the first discharge space  55   b , the through-hole  43   a , and the second suction space  56   a  changes according to the ratio (P 2 /P 1 ). When the confined volume of the expansion chamber changes, the suction volume (volumetric flow rate) of the expansion mechanism  3  also changes. As a result, the constraint of constant density ratio can be avoided. The power recovery efficiency can be optimized by adjusting the ratio (P 2 /P 1 ) according to the heat source temperature (for example, an outside air temperature). 
     Also in the present embodiment, the confined volume is minimum when the period P 2  is 0, that is, when the first vane  48  and the first piston  46  are always in contact with each other. The minimum value of the period P 2  may be greater than zero, of course. 
     As shown in  FIG. 14A  and  FIG. 14B , the variable vane mechanism  130  includes an oil chamber  142 , a first oil passage  144 , a second oil passage  146 , a first valve  148 , a second valve  149 , and a pressure supply passage  147 . The oil chamber  142  communicates with the first vane groove  42   a  so that the oil can be supplied to the first vane groove  42   a  and the oil can be received from the first vane groove  42   a . In the present embodiment, a part of the first vane groove  42   a  is used as the oil chamber  142 . 
     In the present embodiment, the expansion mechanism  3  is placed in the lower part in the closed casing  1 , and the space around the expansion mechanism  3  is filled with oil. The first oil passage  144  opens directly into the oil reservoir  25 . Therefore, no oil pump is needed to pump the oil into the first oil passage  144 . 
     Through the first oil passage  144 , the oil in the oil reservoir  25  is supplied to the oil chamber  142  and the oil in the oil chamber  142  is discharged to the oil reservoir  25 . The first valve  148  is an opening-adjustable valve provided in the first oil passage  144  so that the flow resistance (the inflow resistance and the outflow resistance) of the first oil passage  144  can be increased or decreased. If the flow resistance of the first oil passage  144  is increased or decreased, the flow rate of the oil flowing into the oil chamber  142  can be adjusted, and thus the movement of the first vane  48  can be controlled. This mechanism rarely requires a high precision control technique because the opening of the first valve  148  does not need to be adjusted according to the rotation angle of the shaft  5 , and therefore is highly reliable. 
     Specifically, the first valve  148  has a valve body  151 , a spring  152 , and a pressure chamber  153 . The valve body  151  and the spring  152  are placed in the pressure chamber  153 . The spring  152  is placed behind the valve body  151  so that an elastic force is applied to the rear end surface of the valve body  151 . The pressure supply passage  147  is connected to the portion of the pressure chamber  153  where the spring  152  is placed so that the pressure of the control fluid can be applied to the rear end surface of the valve body  151 . The pressure of the control fluid and the elastic force of the spring  152  are applied to the rear end surface of the valve body  151 . The position of the valve body  151  is determined according to the pressure of the control fluid supplied to the pressure chamber  153 . 
     On the side of the head of the valve body  151 , the range of the movement of the valve body  151  overlaps the first oil passage  144 . As shown in  FIG. 14A , when the valve body  151  occupies the most backward position, the cross-sectional area of the first oil passage  144  is maximum. As shown in  FIG. 14B , when the valve body  151  occupies the most forward position, the cross-sectional area of the first oil passage  144  is minimum. The minimum cross-sectional area of the first oil passage  144  is, for example, about half the maximum cross-sectional area of the first oil passage  144 . Thus, the first valve  148  is structured as a flow rate control valve. 
     As a control fluid to be supplied to the pressure chamber  153  of the first valve  148 , the working fluid in the refrigeration cycle apparatus  400  A is used. The use of the working fluid as a power source allows some leakage of the working fluid from the pressure chamber  153  to the first oil passage  144 . Therefore, tight sealing is not required. 
     As shown in  FIG. 12  and  FIG. 13A , in the present embodiment, the first vane groove  42   a  is closed by the bearing member  42  and the intermediate plate  43 . Therefore, the oil is supplied to the oil chamber  142  only through the first oil passage  144 . As an oil passage for discharging the oil in the oil chamber  142  to the oil reservoir  25 , the second oil passage  146  is provided. The second oil passage  146  communicates the oil chamber  142  with the oil reservoir  25  by a route different from the first oil passage  144 . The second oil passage  146  is provided with the second valve  149 . 
     The second valve  149  has a valve body  155 , a spring  156 , and an accommodation space  157 . The valve body  155  can occupy the positions for closing and opening the second oil passage  146 . The spring  156  is disposed in the accommodation space  157 . The accommodation space  157  may communicate with the oil reservoir  25  so that the valve body  155  can move smoothly. When the oil in the oil chamber  142  is discharged to the oil reservoir  25 , the valve body  155  is pushed by the oil and opens the second oil passage  146 . Conversely, when the oil in the oil reservoir  25  is supplied to the oil chamber  142 , the valve body  155  is subjected to an elastic force from the spring  156  and closes the second oil passage  146 . In this way, the direction of the flow of the oil in the second oil passage  146  is limited substantially only to the direction from the oil chamber  142  to the oil reservoir  25  by the second valve  149 . That is, the second valve  149  is structured as a direction control valve. The phrase “is limited substantially to . . . ” is not intended to exclude completely an unavoidable slight flow. 
     Even if the second oil passage  146  and the second valve  149  are omitted, the ratio (P 2 /P 1 ) can be adjusted, and therefore the variable vane mechanism  130  can work properly. When the oil in the oil chamber  142  is discharged to the oil reservoir  25 , the first vane  48  is strongly pressed by the first piston  46 . Therefore, even if the outflow resistance of the first oil passage  144  is high to some extent, the oil is discharged without any problem. However, such a high outflow resistance increases pressure loss. Furthermore, the valve body  151  of the first valve  148  flutters from side to side, which makes it difficult to set an intended confined volume. 
     In contrast, when the second oil passage  146  is provided, the oil in the oil chamber  142  is discharged to the oil reservoir  25  through both the first oil passage  144  and the second oil passage  146 . In particular, since the oil is discharged relatively freely to the oil reservoir  25  through the second oil passage  146 , an increase in the power recovery efficiency can be expected. Furthermore, since the second valve  149  as a direction control valve is provided in the second oil passage  146 , it is possible to prevent the oil in the oil reservoir  25  from being supplied to the oil chamber  142  through the second oil passage  146 . As a result, the rate of oil supply to the oil chamber  142  can be controlled precisely, and thus the confined volume can be adjusted more easily. 
     The oil chamber may be formed outside the first vane groove  42   a  on the condition that the oil can flow freely therebetween. For example, the oil chamber may be formed immediately behind the first vane groove  42   a . Furthermore, the first valve  148  may be provided at the end portion of the first oil passage  144 . The second valve  149  may be provided at the end portion of the second oil passage  146 . 
     In the operation mode (first mode) shown in  FIG. 14A , the pressure chamber  153  is filled with the low-pressure working fluid, and thus the first valve  148  is fully opened. When the first valve  148  is fully opened, the flow resistance of the first oil passage  144  is low. Therefore, the oil in the oil reservoir  25  can be supplied to the oil chamber  142  smoothly. As a result, a load enough to maintain the contact between the first vane  48  and the first piston  46  is applied continuously to the rear end surface of the first vane  48 . The first vane  48  can follow the movement of the first piston  46 , and thus the contact state between the first vane  48  and the first piston  46  is always maintained. 
     On the other hand, in the operation mode (second mode) shown in  FIG. 14B , the pressure chamber  153  is filled with the high-pressure or intermediate-pressure working fluid, and thus the opening of the first valve  148  is reduced. Specifically, the valve body  151  moves to the position where the force applied to the valve body  151  by the working fluid filled in the pressure chamber  153  and by the spring  152  and the force applied to the valve body  151  by the oil in the first oil passage  144  are balanced with each other. Then, the cross-sectional area of the first oil passage  144  becomes smaller than that in the first mode ( FIG. 14A ). When the cross-sectional area of the first oil passage  144  becomes smaller, a rapid flow of the oil into the oil chamber  142  can be prevented. Then, the flow of the oil into the oil chamber  142  cannot catch up with the downward moving speed of the first vane  48 , and the first vane  48  is spaced from the first piston  46  during the passage of a predetermined period P 2  from the moment when the first piston  46  occupies the top dead center. During this period, the high-pressure working fluid a continues to flow from the first suction space  55   a  into the first discharge space  55   b . At the moment when the first vane  48  again comes into contact with the first piston  46  after the passage of the period P 2 , the expansion chamber is formed by the first discharge space  55   b , the through-hole  43   a , and the second suction space  56   a , and thus the working fluid begins to expand. 
     When the pressure in the pressure chamber  153  is changed, the position of the valve body  151  changes, and thus the flow rate of the oil flowing into the oil chamber  142  changes. The length of the period P 2  changes accordingly. The higher the pressure in the pressure chamber  153  is, the smaller the opening of the first valve  148  becomes, in other words, the smaller the cross-sectional area of the first oil passage  144  becomes, which makes the flow of the oil into the oil chamber less easily. Then, the period P 1  in which the first vane  48  is in contact with the first piston  46  becomes progressively shorter while the period P 2  becomes progressively longer, and the confined volume of the expansion chamber increases. In this way, the confined volume can be adjusted by adjusting the pressure in the pressure chamber  153 . In other words, the suction volume of the expansion mechanism  3  can be adjusted freely. 
     Since the pipe  105  in the pressure adjustment circuit  110  is connected to the pressure supply passage  147  of the variable vane mechanism  130 , the pressure in the pressure chamber  153  can be adjusted by the throttle valve  104  in the pressure adjustment circuit  110 . That is, the opening of the first valve  148  can be controlled by adjusting the opening of the throttle valve  104 . When the opening of the throttle valve  104  is increased, the pressure in the pressure chamber  153  increases, and the opening of the first valve  148  decreases. As a result, the confined volume increases. When the opening of the throttle valve  104  is reduced, the pressure in the pressure chamber  153  decreases, and the opening of the first valve  148  increases. As a result, the confined volume decreases. 
     The pressure in the pressure chamber  153  can be changed between the high pressure and the low pressure of the refrigeration cycle by adjusting the opening of the throttle valve  104 , as in the first embodiment. 
     Next, the operating principle of the expansion mechanism  3  will be described. As shown in Steps A 3  to D 3  in  FIG. 15 , when the confined volume is minimum, the expansion mechanism  3  operates on the same principle as that described in the first embodiment with reference to  FIG. 5 . 
     Next, the operating principle of the expansion mechanism  3  at a larger confined volume than in  FIG. 15  will be described with reference to  FIG. 16 . 
     First, Step A 1  in  FIG. 16  shows a state in which the first piston  46  rotates 360 degrees and the first suction space  55   a  is filled with a high-pressure working fluid. Next, as shown in Step B 4  in  FIG. 16 , when the first piston  46  rotates in a counterclockwise direction, it is spaced from the first vane  48 . This is because the movement of the first vane  48  is restricted by the variable vane mechanism  130  from the moment when the first piston  46  occupies the top dead center. When the first piston  46  is spaced from the first vane  48 , a flow path is formed from the first suction space  55   a  to the first discharge space  55   b , and thus the high-pressure working fluid flows directly from the first discharge space  55   a  into the first discharge space  55   b . The high-pressure working fluid also flows into the second suction space  56   a  that communicates with the first discharge space  55   b . That is, the working fluid does not expand during the period P 2  in which the first piston  46  is spaced from the first vane  48 , and the suction process continues. 
     Next, as shown in Step C 4  in  FIG. 16 , when the first piston  46  further rotates and comes close to the bottom dead center, the first vane  48  catches up with the first piston  46  and again comes into contact with the first piston  46 . The first suction space  55   a  and the first discharge space  55   b  are separated from each other by the first vane  48 , and the flow of the working fluid from the first suction space  55   a  to the first discharge space  55   b  is interrupted. The working fluid begins to expand from the point in time when the first vane  48  and the first piston  46  again come into contact with each other. 
     As shown in Step D 4  in  FIG. 16 , when the first piston  46  further rotates, the volumetric capacity of the first discharge space  55   b  decreases gradually, and the working fluid moves to the second suction space  56   a  while expanding. The operations in Steps A 4  to D 4  in  FIG. 6  are repeated in this manner. 
       FIG. 17A ,  FIG. 17B , and  FIG. 17C  are graphs showing the position of the tip of the first vane, the pressure of the working fluid drawn into the expansion mechanism, and the volumetric capacity of the working chamber, respectively. In each of these graphs, the horizontal axis represents the rotation angle of the shaft  5  obtained when the angle at the moment the first piston  46  occupies the top dead center is defined as a reference angle (of 0 degree). 
     The position of the tip of the first vane  48  shown in the vertical axis in  FIG. 17A  corresponds to the distance from the rotational axis of the shaft  5  to the tip of the first vane  48 . The solid line shows the position of the tip of the first vane  48  in the first mode. The dashed line shows the position of the tip of the first vane  48  in the second mode. In the second mode, the first vane  48  is detached from the first piston  46  at angles of 0 degree and 360 degrees (top dead center), and again comes into contact with the first piston  46  at angles of θ 1  and θ 2  slightly less than the angles of 180 degrees and 540 degrees (bottom dead center). 
     Also in  FIG. 17B , the solid line corresponds to the first mode, and the dashed line corresponds to the second mode, respectively. In the first mode (solid line), the working fluid begins to be drawn into the expansion mechanism at the reference angle, and expands when the rotation angle is in the range of 360 to 720 degrees. On the other hand, in the second mode (dashed line), the working fluid expands when the rotation angle is in the range of the angle θ 2 , which is larger than 360 degrees, to 720 degrees. 
     The volumetric capacity of the working chamber shown in the vertical axis in  FIG. 17C  corresponds to the volumetric capacity of the first suction space  55   a  in the range of 0 to 360 degrees, and to the total volumetric capacity of the first discharge space  55   b  and the second suction space  56   a  in the range of 360 to 720 degrees. In the first mode, the suction process is completed at 360 degrees, and the expansion process is performed in the range of 360 to 720 degrees. On the other hand, in the second mode, the expansion process is performed in the range of the angle θ 2 , which is larger than 360 degrees, to 720 degrees. The total volumetric capacity (confined volume) V 2  of the first discharge space  55   b  and the second suction space  56   a  at the start of the expansion process in the second mode is larger than the total volumetric capacity (confined volume) V 1  in the first mode. 
     The difference in the suction volume ΔV between the first mode and the second mode is expressed as (V 2 -V 1 ) per cycle including the suction process, the expansion process, and the discharge process. This volume difference ΔV increases or decreases according to the length of the period P 2  (in other words, the ratio (P 2 /P 1 )). The length of the period P 2  varies depending on the pressure in the pressure chamber  153  of the variable vane mechanism  130 . The range of the ratio (P 2 /P 1 ) is not particularly limited. For example, the ratio is 0≦(P 2 /P 1 )≦1. This means that the period P 2  falls within the period in which the rotation angle of the shaft  5  is in the range of 0 to 180 degrees, if the rotation angle at the moment when the first piston  46  occupies the top dead center is defined as 0 degree. In the present embodiment, the moment when the first piston  46  occupies the top dead center is the starting point of the period P 2 . 
     As described above, with the expansion mechanism  3  including the variable vane mechanism  130 , the confined volume of the expansion chamber can be changed. Therefore, the volume of the working fluid to be drawn into the expansion mechanism  3  during one rotation of the shaft can be changed. 
     (Modification of Fourth Embodiment) 
       FIG. 18  is a transverse cross-sectional view of a modification of the fourth embodiment. According to this modification, the variable vane mechanism  130  further includes an acceleration port  159  for assisting the first vane  48  in moving downward (in the direction approaching the rotation axis of the shaft  5 ) in the second mode. One end of the acceleration port  159  opens into the first vane groove  42   a  at a predetermined position along the longitudinal direction of the first vane groove  42   a . The other end of the acceleration port  159  opens into the oil reservoir  25 . When the rear end surface of the first vane  48  passes the position of the one end of the acceleration port  159  in the process where the first vane  48  is pushed out of the first vane groove  42   a  by the load applied by the oil and the first spring  50 , the oil in the oil reservoir  25  can flow into the first vane groove  42   a  through the acceleration port  159 . 
     That is, with this acceleration port  159 , even in the case where the cross-sectional area of the first oil passage  144  (see  FIG. 14A ) is set small, when the first vane  48  projects from the first vane groove  42   a  to some extent, the resistance of the oil flowing into the portion (oil chamber  142 ) behind the first vane groove  42   a  drops sharply. Then, the first vane  48  is pushed strongly toward the first piston  46 , and again comes into contact with the first piston  46  immediately. 
     For example, in the case where the resistance of the oil flowing into the portion (oil chamber  142 ) behind the first vane groove  42   a  is very high, the first vane  48  could be kept away from the first piston  46  even if the first piston  46  reaches the bottom dead center. To put it more simply, the period P 2  could continue even after the rotation angle exceeds 180 degrees. In contrast, when the acceleration port  159  is provided, it is possible to ensure that the first vane  48  and the first piston  46  again come into contact with each other before the first piston  46  reaches the bottom dead center. As a result, a sufficiently high ratio of expansion can be obtained, and thus an increase in the power recovery efficiency can be expected. 
     (Fifth Embodiment) 
       FIG. 19  is a configuration diagram of a refrigeration cycle apparatus in which a variable vane mechanism for controlling the movement of the first vane by an electrical method is used. This refrigeration cycle apparatus  400 B has an expander-integrated compressor  300 B. The expansion mechanism  3  in the expander-integrated compressor  300 B is provided with a variable vane mechanism  130 B, ( 130 C,  130 D, or  130 E) connected to an external controller  170 . The operation of the variable vane mechanism  130 B is controlled by the external controller  170 . The refrigeration cycle apparatus  400 B has an advantage in that the pressure supply circuit  110  shown in  FIG. 11  can be omitted. In addition, since the variable vane mechanism  130 B controls the movement of the first vane  48  by an electrical method, the confined volume can be optimized easily. 
     The variable vane mechanisms  130 B to  130 E for controlling the movement of the first vane  48  by an electrical method will be described below. In the present embodiment, the rear portion of the first vane groove  42   a  (where the first spring  50  is placed) opens into the oil reservoir  25 , and the oil in the oil reservoir  25  can flow freely into the rear portion of the first vane groove  42   a.    
     The variable vane mechanism  130 B shown in  FIG. 20  is constituted by an electromagnet having a coil  174  and an iron core  172 . The coil  174  applies an electromagnetic force to the first vane  48  to prevent the first vane  48  from following the movement of the first piston  46 . That is, when the coil  174  is energized, the iron core  172  serves as a magnet to attract the first vane  48 . Thereby, the first vane  48  can be prevented from following the movement of the first piston  46 . Typically, the first vane  48  is made of an iron-based metal such as cast iron or carbon steel, and the iron-based metal can be attracted by a magnet. Therefore, the electromagnet can restrict the movement of the first vane  48 . 
     The coil  174  is placed behind the first vane groove  42   a . The iron core  172  penetrates the coil  174 , and the tip of the iron core  172  projects into the first vane groove  42   a . The length of the iron core  172  with respect to the longitudinal direction of the first vane groove  42   a  is determined so that the first vane  48  comes into contact with the iron core  172  when the first vane  48  is pressed most deeply into the first vane groove  42   a . The timing of energizing the coil  172  can be controlled by the external controller  170  (see  FIG. 19 ). The supply of electric current to the coil  172  is started immediately before the first piston  46  reaches the top dead center. The length of the period P 2  in which the first vane  48  is spaced from the first piston  46 , in other words, the confined volume of the expansion mechanism, can be adjusted by controlling the timing of starting and stopping the supply of electric current. 
     The variable vane mechanism  130 C shown in  FIG. 21  is constituted by a coil  176  disposed around the first vane  48 . When the coil  176  is energized, the first vane  48  is subjected to a force that draws it into the coil  176 . That is, the first vane  48  itself acts as a plunger of a solenoid. As in the example shown in  FIG. 20 , the timing of energizing the coil  176  can be controlled by the external controller  170 , and thereby, the confined volume of the expansion mechanism  3  can be adjusted. Since the coil  176  is disposed around the first vane  48 , such a problem as a shortage of space is less likely to occur. 
     In the fourth embodiment, the movement of the first vane  48  merely slows down near the top dead center, but in the examples shown in  FIG. 20  and  FIG. 21 , the first vane  48  can be locked (or the movement thereof can be stopped temporarily) near the top dead center. When the first vane  48  is locked momentarily, the inflow cross-sectional area (width of the space between the first piston  46  and the first vane  48 ) increases, and thus pressure loss can be reduced. 
     The variable vane mechanism  130 D shown in  FIG. 22  is constituted by an electric actuator for applying a load to the first vane  48  to increase the sliding friction between the first vane groove  42   a  and the first vane  48 . Specifically, the variable vane mechanism  130 D is constituted by a solenoid having a coil  181  and a plunger  185 . 
     A groove  183  extending at an approximately right angle to the longitudinal direction of the first vane groove  42   a  is formed in the first cylinder  42 . The plunger  185  is disposed in this groove  183 . The coil  181  is disposed around the plunger  185 . The head of the plunger  185  faces the side surface of the first vane  48 . When the plunger  185  is retracted to the position where it does not interfere with the first vane  48 , the movement of the first vane  48  is not hindered by the variable vane mechanism  130 D (in the first mode). On the other hand, when the plunger  185  is pushed out of the groove  183  to energize the coil  181 , the head of the plunger  185  hits the first vane  48  at a right angle. Thereby, the side surface of the first vane  48  is subjected to a load in the direction toward the inner wall of the first vane groove  42   a , and thus the first vane  48  becomes difficult to move along the longitudinal direction of the first vane groove  42   a.    
     The variable vane mechanism  130 E shown in  FIG. 23  is the same as the variable vane mechanism  130 D described with reference to  FIG. 22  in that it is constituted by an electric actuator for applying a load laterally to the first vane  48 . Specifically, the variable vane mechanism  130 E is constituted by a piezoelectric actuator having a piezoelectric element  186  and a plunger  184  connected to the piezoelectric element  186 . 
     A groove  182  is formed in the first cylinder  42  so as to communicate with a midpoint of the first vane groove  42   a  with respect to its longitudinal direction. The plunger  184  and the piezoelectric element  186  are disposed in the groove  182  so that the head of the plunger  184  faces the first vane  48 . The rear end of the plunger  184  is fixed to the piezoelectric element  186 . The piezoelectric element  186  and the plunger  184  are coupled together so that the displacement of the piezoelectric element  186  is transmitted to the plunger  184 . The action of the plunger  184  is the same as described with reference to  FIG. 22 , except that the piezoelectric element is used instead of the coil. 
     In the examples shown in  FIG. 22  and  FIG. 23 , the variable vane mechanisms  130 D and  130 E are built in the first cylinder  42 . The variable vane mechanisms  130 D and  130 E may, however, be built in the bearing member  41  or the intermediate plate  43 . They may be provided across the bearing member  41 , the first cylinder  42 , and the intermediate plate  43 . 
     Electric current is supplied to each of the variable vane mechanisms shown in  FIGS. 20 to 23  at an appropriate timing. Specifically, the supply of electric current to the coil or the piezoelectric element is controlled based on the rotation angle of the shaft  5 . In order to detect the rotation angle of the shaft  5 , a rotor  191  that rotates with the shaft  5  and a position sensor  193  that can detect the passing of the rotor  191  may be provided, as shown in  FIG. 24 . For example, the rotor  191  is placed 180-degree opposite to the eccentric direction of the eccentric portion  5   c  of the shaft  5  (or to coincide with the eccentric direction). Furthermore, the position sensor  193  is placed at a position corresponding to the bottom dead center of the first piston  46 . 
     With the above configuration, as shown in  FIG. 25 , a sensor signal is transmitted from the position sensor  193  to the external controller  170  when the first piston  46  reaches the top dead center (or the bottom dead center). The external controller  170  can supply electric current to the coil or the piezoelectric element accurately upon receiving the sensor signal from the position sensor  193 . Electric current may be supplied shortly before the first piston  46  reaches the top dead center (=0 degree). This ensures that the movement of the first vane  48  is stopped or retarded. The period of the electric current supply Δθ may be controlled so that a desired confined volume can be obtained. 
     The sensor for detecting the rotation angle (reference position) of the shaft  5  may be provided at a position other than the expansion mechanism  3 . For example, it may be provided in the compression mechanism  2 . 
     Sixth Embodiment 
     The present invention can be applied also to a two-stage rotary expander as a single unit.  FIG. 26  shows a power recovery type refrigeration cycle apparatus  400 C using such a two-stage rotary expander. The refrigeration cycle apparatus  400 C includes a compressor  123 , a radiator  101 , an expander  120 , and an evaporator  102 . As the expander  120 , a two-stage rotary expander having a structure in which the compression mechanism  2  is removed from each of the expander-integrated compressors described above can be used. The expansion energy of the working fluid is converted into electrical energy by a power generator  121  of the expander  120 , and the obtained electrical energy is supplied to a motor  124  of the compressor  123 . 
     The rotational speed of the compressor  123  can be controlled by the motor  124 , and the rotational speed of the expander  120  can be controlled by the power generator  121 . Therefore, this refrigeration cycle apparatus  400 C is essentially free from the constraint of constant density ratio. However, if the two-stage rotary expander provided with the variable vane mechanism is employed, the following advantageous effect can be obtained. 
       FIG. 27  shows the efficiency curve of a typical power generator. The power generator is designed to achieve the highest power generation efficiency at a predetermined rated rotational speed Nr (for example, 60 Hz). Therefore, the power generation efficiency decreases as the difference between the actual rotational speed and the rated rotational speed increases. That is, it is desirable that the rotational speed of the power generator be as close to the rated rotational speed Nr as possible even if it can be controlled by an inverter. However, the amount and density of a working fluid flowing through the refrigeration cycle apparatus vary, and therefore it is difficult to maintain the rotational speed of the power generator close to the rated rotational speed Nr if a conventional expander is used. In contrast, if the two-stage rotary expander provided with the variable vane mechanism is used, the density ratio can be changed while the rated rotational speed Nr is maintained. Therefore, more efficient power recovery can be expected. 
     INDUSTRIAL APPLICABILITY 
     The present invention is suitably applicable to refrigeration cycle apparatuses used for air conditioners and water heaters. The applications of the present invention are not limited to these, and the present invention can be applied to a wide variety of other apparatuses such as a Rankine cycle apparatus.