Patent Publication Number: US-2004052647-A1

Title: Compressor

Description:
TECHNICAL FIELD  
       [0001] The present invention relates to a compressor that is driven selectively by an external drive source and an electric part.  
       BACKGROUND ART  
       [0002] Japanese Laid-Open Patent Publication No. 2001-140757 discloses a hybrid compressor for a vehicle that is driven by an electric part, which is a motor, while the external drive source, which is an engine of the vehicle, is stopped.  
       [0003] The hybrid compressor includes a drive shaft and a compression mechanism, which is driven by the drive shaft. A rotary body is secured to the drive shaft to rotate integrally with the drive shaft. The rotary body supports a pulley via a bearing such that the pulley rotates relative to the rotary body. The rotary body has a rotor, which forms part of the motor so that the drive shaft is rotated by the motor. A one-way clutch is located on a power transmission path between the pulley and the rotary body. The one-way clutch permits power transmission from the pulley to the rotary body such that rotational force in only one direction is transmitted. Accordingly, although the engine for traveling is stopped, the compression mechanism is driven by the motor. When the motor drives the compression mechanism, the power of the motor is prevented from being transmitted to the engine.  
       [0004] The one-way clutch eliminates the need to use an electromagnetic clutch for selectively permitting and discontinuing power transmission between the pulley and the rotary body, which simplifies the structure of the compressor. However, in the case the one-way clutch is used, if an abnormality, such as a dead lock, occurs in the compression mechanism while the compression mechanism is driven by the engine, an excessive load is applied to the engine.  
       DISCLOSURE OF THE INVENTION  
       [0005] Accordingly, it is an objective of the present invention to provide a compressor that reduces weight, size, and cost, and prevents an excessive load from being applied to an external drive source when the compressor has an abnormality.  
       [0006] To achieve the above objective, the present invention provides a compressor, which includes a compressor main body, an electric part, a first rotating body, a second rotating body, a first cylinder, a second cylinder, a one-way clutch, a first ball bearing, a second ball bearing, and a discontinuing mechanism. The compressor main body includes a housing and a drive shaft, which is supported by the housing. The compressor main body compresses refrigerant in accordance with rotation of the drive shaft. The electric part at least functions as a motor. The first rotating body is rotated by an external drive source. The second rotating body is secured to the drive shaft to rotate integrally with the drive shaft. The second rotating body is operably coupled to the first rotating body and power is directly transmitted from the electric part to the second rotating body. The first cylinder and the second cylinder are located on the first rotating body. The first cylinder and the second cylinder are apart from each other. The one-way clutch is located on a power transmission path between the first cylinder and the second cylinder. The first ball bearing is located between the first cylinder and the second rotating body. The second ball bearing is located between the second cylinder and the housing. The discontinuing mechanism is located on a power transmission path between the external drive source and the drive shaft and discontinues excessive torque transmission from the external drive source to the drive shaft. 
     
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
     [0007]FIG. 1 is a cross-sectional view illustrating a compressor according to a first embodiment of the present invention;  
     [0008]FIG. 2 is a cross-sectional view illustrating the control valve of the compressor shown in FIG. 1;  
     [0009]FIG. 3 is a cross-sectional view for explaining the movement of the operating rod of the control valve shown in FIG. 2;  
     [0010]FIG. 4 is an enlarged cross-sectional view illustrating the power transmission mechanism of the compressor shown in FIG. 1;  
     [0011]FIG. 5( a ) is a front view illustrating a downstream pulley member according to a second embodiment of the present invention;  
     [0012]FIG. 5( b ) is a cross-sectional view illustrating a power transmission mechanism according to a second embodiment;  
     [0013] FIGS.  6 ( a ) and  6 ( b ) are partial cross-sectional views illustrating the one-way clutch according to the first embodiment; and  
     [0014]FIG. 7 is a partial cross-sectional view illustrating a power transmission mechanism according to a modified embodiment. 
    
    
     BEST MODE FOR CARRYING OUT THE INVENTION  
     [0015] A first embodiment of the present invention will now be described with reference to FIGS.  1  to  4 ,  6 ( a ), and  6 ( b ). The left side of FIG. 1 is defined as the front end of the compressor, and the right side of FIG. 1 is defined as the rear end of the compressor.  
     [0016] As shown in FIG. 1, a compressor main body C, which forms part of a vehicular air conditioner, includes a cylinder block  11 , a front housing member  12 , and a rear housing member  14 . The front housing member  12  is secured to the front end of the cylinder block  11 . The rear housing member  14  is attached to the rear end of the cylinder block  11  with a valve plate assembly  13  located in between. The cylinder block  11 , the front housing member  12 , the valve plate assembly  13 , and the rear housing member  14  form a housing assembly of the compressor main body C.  
     [0017] The cylinder block  11  and the front housing member  12  define a control pressure zone, which is a crank chamber  15 . The housing assembly of the compressor main body C rotatably supports a drive shaft  16 , which extends through the crank chamber  15 . The front end of the drive shaft  16  is supported by a radial bearing  12 A, which is secured to the front wall of the front housing member  12 . The rear end of the drive shaft  16  is supported by a radial bearing  11 A, which is secured to the cylinder block  11 .  
     [0018] The front end of the drive shaft  16  projects outside through the front wall of the front housing member  12 . The front end of the drive shaft  16  is operably coupled to an external drive source, which is an engine E of a vehicle in the first embodiment, via a power transmission mechanism PT and a belt  18 . The belt  18  is wound about a first rotating body, which is a pulley  17  in the first embodiment. The pulley  17  forms part of the power transmission mechanism PT. The power transmission mechanism PT and the compressor main body C form the compressor.  
     [0019] An electric part, which is a motor generator MG in the first embodiment, is located between the pulley  17  and the drive shaft  16  in a power transmission path between the engine E of the vehicle and the drive shaft  16 . The motor generator MG is formed by an induction machine and functions as a motor and a generator. In the first embodiment, when the engine E of the vehicle is running, power from the engine E is always transmitted to the drive shaft  16  and the motor generator MG. At this time, the motor generator MG functions as a generator. If air conditioning is necessary when the engine E of the vehicle is stopped, the motor generator MG functions as a motor to rotate the drive shaft  16 .  
     [0020] A lug plate  19  is located in the crank chamber  15  and is secured to the drive shaft  16  to rotate integrally with the drive shaft  16 . A drive plate, which is a swash plate  20  in the first embodiment, is located in the crank chamber  15 . The swash plate  20  is supported on the drive shaft  16 , and slides along and inclines with respect to the drive shaft  16 . The swash plate  20  is coupled to the lug plate  19  via a hinge mechanism  21 . The hinge mechanism  21  causes the swash plate  20  to rotate integrally with the lug plate  19  and the drive shaft  16 , and permits the swash plate  20  to slide in the axial direction of the drive shaft  16  and incline with respect to the drive shaft  16 .  
     [0021] The minimum inclination angle of the swash plate  20  is determined by a ring  22 , which is secured to the drive shaft  16 , and a spring  23 , which is located between the ring  22  and the swash plate  20 . The swash plate  20  is movable between a minimum inclination angle position indicated by a solid line in FIG. 1, and a maximum inclination angle position indicated by a chain double-dashed line in FIG. 1. When the swash plate  20  is located at the minimum inclination angle position, the angle between the swash plate  20  and a surface that is perpendicular to the axis of the drive shaft  16  is closest to zero.  
     [0022] Cylinder bores  24  (only one shown) are formed in the cylinder block  11 . The cylinder bores  24  extend along the drive shaft  16  and are arranged about the axis of the drive shaft  16  at equal angular intervals. A single headed piston  25  is accommodated in each cylinder bore  24  to reciprocate in the cylinder bore  24 . Both ends of each cylinder bore  24  are closed by the valve plate assembly  13  and the corresponding piston  25 . Each cylinder bore  24  defines a compression chamber the volume of which varies in accordance with the reciprocation of the corresponding piston  25 . Each piston  25  is coupled to the peripheral portion of the swash plate  20  by a pair of shoes  26 . Therefore, when the swash plate  20  rotates with the drive shaft  16 , the shoes  26  convert the rotation of the swash plate  20  into reciprocation of the pistons  25 .  
     [0023] The cylinder block  11 , the drive shaft  16 , the lug plate  19 , the swash plate  20 , the hinge mechanism  21 , the pistons  25 , and the shoes  26  form a compression mechanism of a variable displacement piston type compressor.  
     [0024] The rear housing member  14  defines a suction pressure zone, which is a suction chamber  27 , and a discharge pressure zone, which is a discharge chamber  28 . The openings of the suction chamber  27  and the discharge chamber  28  that face the valve plate assembly  13  are closed by the valve plate assembly  13 . The valve plate assembly  13  has suction ports  29 , suction valve flaps  30 , discharge ports  31 , and discharge valve flaps  32 . Each set of the suction port  29 , the suction valve flap  30 , the discharge port  31 , and the discharge valve flap  32  corresponds to one of the cylinder bores  24 . When each piston  25  moves from the top dead center position to the bottom dead center position, refrigerant gas in the suction chamber  27  is drawn into the corresponding cylinder bore  24  via the corresponding suction port  29  and suction valve flap  30 . When each piston  25  moves from the bottom dead center position to the top dead center position, refrigerant gas in the corresponding cylinder bore  24  is compressed to a predetermined pressure and is discharged to the discharge chamber  28  via the corresponding discharge port  31  and discharge valve flap  32 .  
     [0025] A crank chamber pressure control mechanism controls the internal pressure of the crank chamber  15 , or a crank pressure Pc, to control the inclination angle of the swash plate  20 . The crank chamber pressure control mechanism is formed by a bleed passage  33 , a supply passage ( 34 ,  96 ,  98 ), and a control valve  35 . The bleed passage  33  is formed in the housing assembly of the compressor main body C to connect the suction chamber  27  to the crank chamber  15 . The supply passage ( 34 ,  96 ,  98 ) connects the crank chamber  15  to the discharge chamber  28 . The control valve  35  is located in the supply passage ( 34 ,  96 ,  98 ). The supply passage ( 34 ,  96 ,  98 ) includes a pipe  96 , a second pressure introduction passage  98 , and a communication passage  34 . The pipe  96  is connected to an outlet of the discharge chamber  28 . The second pressure introduction passage  98  extends from a second pressure monitoring point P 2  located in the pipe  96  to the control valve  35 . The communication passage  34  extends from the control valve  35  to the crank chamber  15 .  
     [0026] Refrigerant in the crank chamber  15  is released into the suction chamber  27  through the bleed passage  33 . On the other hands, the control valve  35  adjusts the flow rate of refrigerant supplied to the crank chamber  15  from the discharge chamber  28  through the supply passage ( 34 ,  96 ,  98 ). In accordance with a change in the crank pressure Pc, the difference between the crank pressure Pc and the pressure in each cylinder bore  24  is changed, which alters the inclination angle of the swash plate  20 . As a result, the stroke of each piston  25 , that is, the discharge displacement (displacement of the compressor main body C) per one rotation of the drive shaft  16 , is controlled. In the compressor main body C of the first embodiment, the discharge displacement per one rotation of the drive shaft  16  approaches zero when the swash plate  20  is located at the minimum inclination angle position.  
     [0027] The suction chamber  27  and the discharge chamber  28  are connected to each other by an external refrigerant circuit  90 . The compressor main body C and the external refrigerant circuit  90  form the refrigerant circuit of the vehicular air-conditioner. The external refrigerant circuit  90  includes a condenser  91 , a decompression device, which is a temperature type expansion valve  92  in this embodiment, and an evaporator  93 . The opening degree of the expansion valve  92  is feedback controlled based on the temperature of the refrigerant detected by a heat sensitive tube  94 , which is located at the outlet or downstream of the evaporator  93 , or the pressure at the outlet of the evaporator  93 . The expansion valve  92  supplies appropriate amount of liquid refrigerant to the evaporator  93  in accordance with the heat load applied to the refrigerant circuit to adjust the flow rate of refrigerant in the external refrigerant circuit  90 .  
     [0028] The external refrigerant circuit  90  includes a pipe  95 , or a low pressure passage, which connects the outlet of the evaporator  93  to the suction chamber  27  of the compressor main body C. The low pressure passage and the suction chamber  27  form a low pressure zone. The external refrigerant circuit  90  also includes the pipe  96 , or a high pressure passage, which connects the discharge chamber  28  of the compressor main body C to the inlet of the condenser  91 . The high pressure passage and the discharge chamber  28  form a high pressure zone. The compressor main body C draws in and compresses refrigerant gas that is introduced into the suction chamber  27  from the low pressure passage and discharges the compressed gas to the discharge chamber  28 , which is connected to the high pressure passage.  
     [0029] As the flow rate of refrigerant (refrigerant flow rate Q) that flows through the refrigerant circuit increases, the pressure loss per unit length of the circuit or the pipe increases. That is, the pressure loss between the pressure monitoring points P 1  and P 2  located in the refrigerant circuit has a positive correlation with the refrigerant flow rate Q in the refrigerant circuit. A primary pressure ΔPX, which is the pressure loss between two pressure monitoring points P 1  and P 2 , that is, the difference between the pressure PdH at the first pressure monitoring point P 1  and the pressure PdL at the second pressure monitoring point P 2 , reflects the refrigerant flow rate Q in the refrigerant circuit.  
     [0030] In the first embodiment, the first pressure monitoring point P 1  for monitoring high pressure in the upstream side is located in the discharge chamber  28 , which corresponds to the most upstream section of the pipe  96 . The second pressure monitoring point P 2  for monitoring low pressure in the downstream side is located in the pipe  96  apart from the first pressure monitoring point P 1  by a predetermined distance. The pressure PdH at the first pressure monitoring point P 1  is introduced to the control valve  35  through a first pressure introduction passage  97  (shown in FIG. 2 only) and the pressure PdL at the second pressure monitoring point P 2  is introduced to the control valve  35  through the second pressure introduction passage  98 .  
     [0031] Pressure difference increasing means for clarifying or increasing the primary pressure ΔPX, or a fixed restrictor  99 , is located in the pipe  96  between the pressure monitoring points P 1  and P 2 . Since the fixed restrictor  99  is located between the pressure monitoring points P 1  and P 2 , the pressure monitoring points P 1  and P 2  need not be separated from each other by a large amount. Therefore, the second pressure monitoring point P 2  can be located close to the compressor main body C, which shortens the second pressure introduction passage  98  between the second pressure monitoring point P 2  and the control valve  35 . Although the pressure PdL at the second pressure monitoring point P 2  is decreased compared to the pressure PdH at the first pressure monitoring point P 1  by the operation of the fixed restrictor  99 , the pressure PdL is sufficiently higher than the crank pressure Pc.  
     [0032] As shown in FIG. 2, the control valve  35  includes an inlet valve portion  101 , which forms the upper half of the control valve  35 , and a solenoid portion  102 , which forms the lower half of the control valve  35 . The inlet valve portion  101  adjusts the opening degree, or the restricting degree, of the supply passage, which connects the second pressure monitoring point P 2  to the crank chamber  15 . The solenoid portion  102  is an electromagnetic actuator for urging the operating rod  103  located inside the control valve  35  based on the external current supply control. The operating rod  103  includes a coupling portion  105 , which is the distal end, a valve body  106 , which is located at the substantial center, and a guide rod portion  107 , which is the proximal end. The valve body  106  forms part of the guide rod portion  107 . Assuming that the diameter of the coupling portion  105  is represented by d 1 , the diameter of the guide rod portion  107  is represented by d 2 , d 1  is smaller than d 2 . Assuming that the circle ratio is n, the cross-sectional area SB of the coupling portion  105  is represented by π(d 1 /2) 2 , and the cross-sectional area SD of the guide rod portion  107  is represented by π(d 2 /2) 2 .  
     [0033] A valve housing  108  of the control valve  35  includes a cap  109 , an upper half main body  110 , which forms the main outline of the inlet valve portion  101 , and a lower half main body  111 , which forms the main outline of the solenoid portion  102 . A valve chamber  112  and a communication passage  113  are defined in the upper half main body  110  of the valve housing  108 . A pressure sensing chamber  114  is defined between the upper half main body  110  and the cap  109 , which is inserted in the upper portion of the upper half main body  110 . The operating rod  103  is located inside the valve chamber  112 , the communication passage  113 , and a pressure sensing chamber  114  and is movable in the axial direction (vertical direction in FIG. 2). The valve chamber  112  and the communication passage  113  are selectively communicated in accordance with the position of the operating rod  103 . The communication passage  113  and part of the pressure sensing chamber  114  (second pressure chamber  115 , which will be described later) are always communicated.  
     [0034] The bottom wall of the valve chamber  112  is formed by the upper end of a fixed iron core  116 , which will be described later. A port  117  is formed in the circumferential wall of the valve housing  108 , which surrounds the valve chamber  112 . The port  117  connects the valve chamber  112  to the crank chamber  15  via the communication passage  34 , which is the downstream section of the supply passage. A port  118  is formed in the circumferential wall of the valve housing  108 , which surrounds the pressure sensing chamber  114  (the second pressure chamber  115 ). The port  118  connects the communication passage  113  to the second pressure monitoring point P 2  via the pressure sensing chamber  114  (the second pressure chamber  115 ) and the second pressure introduction passage  98 , which is the upstream section of the supply passage. The port  117 , the valve chamber  112 , the communication passage  113 , the pressure sensing chamber  114  (the second pressure chamber  115 ), and the port  118  form a passage inside the control valve, and the passage inside the control valve forms part of the supply passage.  
     [0035] The valve body  106  of the operating rod  103  is located inside the valve chamber  112 . The inner diameter d 3  of the communication passage  113  is greater than the diameter d 1  of the coupling portion  105  of the operating rod  103  and is smaller than the diameter d 2  of the guide rod portion  107 . That is, the cross-sectional area (opening area) SC of the communication passage  113  is π(d 3 /2) 2 , and the opening area SC is greater than the cross-sectional area SB of the coupling portion  105  and is smaller than the cross-sectional area SD of the guide rod portion  107 . Therefore, a step located between the valve chamber  112  and the communication passage  113  function as a valve seat  119  and the communication passage  113  function as a valve hole. When the operating rod  103  moves from the lowermost position shown in FIG. 2 to the uppermost position at which the valve body  106  contacts the valve seat  119 , the communication passage  113  is disconnected. The valve body  106  of the operating rod  103  moves in the axial direction to adjust the opening degree of the supply passage.  
     [0036] A first pressure sensing member, which is a movable wall  120  in the first embodiment, is located in the pressure sensing chamber  114  and is movable in the axial direction. The movable wall  120  is cup-shaped. The bottom wall of the movable wall  120  divides the pressure sensing chamber  114  into a high pressure chamber, which is a first pressure chamber  121  in the first embodiment, and a low pressure chamber, which is a second pressure chamber  115  in the first embodiment. The movable wall  120  serves as a partition between the first pressure chamber  121  and the second pressure chamber  115 . The movable wall  120  does not permit fluid to directly move between the first pressure chamber  121  and the second pressure chamber  115 . The cross-sectional area SA of the movable wall  120  is greater than the opening area SC of the communication passage  113 .  
     [0037] The first pressure chamber  121  is always communicated with the first pressure monitoring point P 1 , which is the discharge chamber  28 , via a port  122 , which is formed in the cap  109  and the first pressure introduction passage  97 . On the other hand, the second pressure chamber  115  is always communicated with the second pressure monitoring point P 2  via the port  118 , which is part of the supply passage, and the second pressure introduction passage  98 . That is, the pressure PdH at the first pressure monitoring point P 1  is introduced into the first pressure chamber  121 , and the pressure PdL at the second pressure monitoring point P 2  is introduced into the second pressure chamber  115 . Therefore, the upper and lower surfaces of the movable wall  120  serve as pressure receiving surfaces that are exposed to the pressures PdH and PdL, respectively. The movable wall  120  is displaced in accordance with the difference (the primary ΔPX) between the pressure PdH and the pressure PdL.  
     [0038] The distal end of the coupling portion  105  of the operating rod  103  enters the second pressure chamber  115 . The distal end of the coupling portion  105  is attached to the movable wall  120 . The first pressure chamber  121  accommodates a return spring  123 . The return spring  123  urges the movable wall  120  from the first pressure chamber  121  toward the second pressure chamber  115 .  
     [0039] The solenoid portion  102  includes a cup-shaped cylinder  124 . A fixed iron core  116  is fitted in the upper portion of the cylinder  124  so that a solenoid chamber  125  is defined in the cylinder  124 . A plunger, which is a movable iron core  126  in the first embodiment, is accommodated in the solenoid chamber  125  and is movable in the axial direction. A guide hole  127  is formed at the center of the fixed iron core  116  and extends in the axial direction. A guide rod portion  107  of the operating rod  103  is located in the guide hole  127  and is movable in the axial direction. A little space (not shown) is formed between the inner circumferential surface of the guide hole  127  and the guide rod portion  107 . The valve chamber  112  and the solenoid chamber  125  are connected to each other via the space. That is, the solenoid chamber  125  is exposed to the crank pressure Pc that is the same as the crank pressure Pc in the valve chamber  112 .  
     [0040] The proximal end of the operating rod  103  is accommodated in the solenoid chamber  125 . That is, the lower end of the guide rod portion  107  is located inside the solenoid chamber  125  and is fitted and calked to a hole made through the center of the movable iron core  126 . Therefore, the movable iron core  126  and the operating rod  103  integrally move in the vertical direction. The solenoid chamber  125  accommodates a shock absorbing spring  128 , which urges the movable iron core  126  toward the fixed iron core  116 . In other words, the shock absorbing spring  128  urges the movable iron core  126  and the operating rod  103  upward. The force of the shock absorbing spring  128  is smaller than the force of the return spring  123 . The return spring  123  functions as restore means for returning the movable iron core  126  and the operating rod  103  to the lowermost position when the solenoid portion  102  is de-excited.  
     [0041] A coil  129  is wound about the fixed iron core  116  and the movable iron core  126 . A drive signal is sent to the coil  129  from a drive circuit  131  based on a command from a controller  130 . The coil  129  generates an electromagnetic force the magnitude of which corresponds to the value of current supplied to the coil  129 . The electromagnetic force attracts the movable iron core  126  toward the fixed iron core  116 , which urges the operating rod  103  upward. The current supplied to the coil  129  is controlled by adjusting the applied voltage to the coil  129 . The applied voltage is generally controlled by means for changing the voltage or means that utilizes pulse-width modulation. The pulse-width modulation is a method for adjusting the average voltage by applying a pulse voltage having a constant cycle and changing the on time of the pulse. The applied voltage is represented by the pulse voltage multiplied by the on time of the pulse divided by the pulse cycle. The on time of the pulse divided by the pulse cycle is referred to as a duty ratio. The voltage control that makes use of the pulse width modulation is sometimes referred to as a duty control. When the pulse width modulation is employed, the current intermittently varies, which reduces the hysteresis of the electromagnet. It is also common to measure the value of current that flows through the coil  129 , and feedback control the applied voltage based on the measured current value. In the first embodiment, the duty control is employed. Due to the structure of the control valve  35 , a smaller duty ratio increases the opening degree of the control valve  35 . A greater duty ratio decreases the opening degree of the control valve  35 .  
     [0042] The opening degree of the control valve shown in FIG. 2 is determined by the axial position of the operating rod  103 , which includes the valve body  106 . The operating conditions and the characteristics of the control valve  35  will become apparent by considering, in a comprehensive manner, the forces that act on each part of the operating rod  103 .  
     [0043] As shown in FIG. 3, a downward force f 1  of the return spring  123  and a downward force based on the primary pressure ΔPX, which is the difference between the pressure PdH and the pressure PdL applied to the movable wall  120 , act on the coupling portion  105  of the operating rod  103 . The pressure receiving area of the upper surface of the movable wall  120  is represented by SA but the pressure receiving area of the lower surface of the movable wall  120  is represented by (SA-SB). Assume that downward direction is defined as the positive direction. The sum ΣF 1  of the forces that act on the coupling portion  105  is expressed by the following equation I. 
     Σ F   1 = PdH×SA−PdL ( SA - SB )+ f   1   (Equation I) 
     [0044] On the other hand, an upward force f 2  of the shock-absorbing spring  128  and an upward electromagnetic force F, which is generated by the solenoid portion  102 , act on the guide rod portion  107  of the operating rod  103 . The pressures applied to all the exposed surfaces of the valve body  106 , the guide rod portion  107 , and the movable iron core  126  are simplified as follows. First, the upper end surface  132  of the valve body  106  is divided into the inner circumferential section and the outer circumferential section by an imaginary cylinder (shown by two broken lines in FIG. 3), which extends from the inner circumferential surface of the communication passage  113 . The pressure PdL acts in a downward direction on the inner circumferential section (area: SC-SB). The crank pressure Pc acts in a downward direction on the outer circumferential section (area: SD-SC). Taking the pressure balance between the upper and lower surfaces of the movable iron core  126  into account, the crank pressure Pc in the solenoid chamber  125  urges the lower end surface  133  of the guide rod portion  107  upward by the area corresponding to the cross-sectional area SD of the guide rod portion  107 . Assume that the upward direction is defined as the positive direction. The sum ΣF 2  of the forces that act on the valve body  106  and the guide rod portion  107  is expressed by the following equation II.  
                     ∑   F2     =            F   +   f2   -     PdL        (     SC   -   SB     )       -     Pc        (     SD   -   SC     )       +     Pc   ×   SD                   =            F   +   f2   +     Pc   ×   SD     -     PdL        (     SC   -   SB     )                       (     Equation                 II     )                       
 
     [0045] In the process of calculating the equation II, —Pc×SD was canceled by +Pc×SD, and the term Pc×SC remained. This means that the effective pressure receiving area of the guide rod portion  107 , which includes the valve body  106 , related to the crank pressure Pc can be expressed as SD-(SD-SC)=SC when considering on the assumption that the crank pressure Pc intensively acts on only the lower end surface  133  of the guide rod portion  107  when the crank pressure Pc acts on the upper and lower end surfaces  132 ,  133  of the guide rod portion  107 . As far as the crank pressure Pc is concerned, the effective pressure receiving area of the guide rod portion  107  is equal to the opening area SC of the communication passage  113  regardless of the cross-sectional area SD of the guide rod portion  107 . In this specification, when pressures of the same kind act on both ends of a member such as a rod, the pressure receiving area having an effect that can be assumed that the pressure acts intensively on one end only is called the “effective pressure receiving area” 
     [0046] Since the operating rod  103  is an integrated member formed by connecting the coupling portion  105  to the guide rod portion  107 , its axial position is determined by the dynamic balance of ΣF 1 =ΣF 2 . After the equation ΣF 1 =ΣF 2  is sorted, the following equation III is obtained. 
       F−f   1 + f   2 =( PdH - PdL ) SA +( PdL - Pc ) SC   (Equation III) 
     [0047] In the Equation III, f 1 , f 2 , SA, SC are parameters that are defined in the steps of mechanical design. The electromagnetic force F is a variable parameter that changes in accordance with the power supplied to the coil  129 . The pressure PdH, PdL and the crank pressure Pc are variable parameters that change in accordance with the driving condition of the compressor. As apparent from equation III, the control valve  35  automatically controls the opening degree such that gas pressure load obtained by multiplying the primary pressure ΔPX, or PdH-PdL, and a secondary pressure ΔPY, or PdL-Pc, by the corresponding pressure receiving area and the total load of the force f 1  and f 2  of the electromagnetic force F and the spring  123 ,  128  are balanced. The operating rod  103  is a second pressure sensing member, which is displaced in accordance with the pressure difference between the pressure PdL and the crank pressure Pc.  
     [0048] In the control valve  35  according to the first embodiment having the above mentioned operating characteristics, the opening degree is determined in the following manner under each circumstance. When no current is supplied to the coil  129 , or when the duty ratio is zero percent, the force of the return spring  123  (more specifically, the force of f 1 -f 2 ) becomes dominant and positions the operating rod  103  at the lowermost position shown in FIG. 2. The valve body  106  is spaced from the valve seat  119  by the greatest distance, which fully opens the inlet valve portion  101 . When a current of the minimum duty ratio within a variable range of the duty ratio is supplied to the coil  129 , the upward electromagnetic force F is at least greater than the downward force f 2  of the return spring  123 . The sum of the upward electromagnetic force F generated by the solenoid portion  102  and the upward force f 2  of the shock-absorbing spring  128  acts against the sum of the downward force f 1  of the return spring  123  and the downward force based on the secondary pressure ΔPY and the primary pressure ΔPX. As a result, the position of the valve body  106  relative to the valve seat  119  is determined such that equation III is satisfied, which determines the opening degree of the control valve  35 . Accordingly, the flow rate of gas to the crank chamber  15  through the supply passage is determined. Then, the crank pressure Pc is adjusted in accordance with the relationship between the flow rate of gas through the supply passage and the flow rate of gas flowing out from the crank chamber  15  through the bleed passage  33 .  
     [0049] As shown in FIGS. 1 and 4, the pulley  17  includes an upstream rotating body, which is an upstream pulley member  17 A, and a downstream rotating body, which is a downstream pulley member  17 B. The downstream pulley member  17 B is formed by a first cylinder, which is a first inner cylinder  17 C, and a first disk-like portion  17 D, which is integrally formed with the front end of the first inner cylinder  17 C and extends radially outward. The upstream pulley member  17 A is formed by an outer cylinder  17 E, about which the belt  18  is wound, a second cylinder, which is a second inner cylinder  17 F, and a second disk-like portion  17 G, which is integrally formed with the outer cylinder  17 E and the second inner cylinder  17 F to couple the outer cylinder  17 E and the second inner cylinder  17 F with each other.  
     [0050] Breakable members, which are columnar power transmission pins  17 H (only two are shown) are secured to the peripheral portion of the first disk-like portion  17 D at equal angular intervals about the axis of the first disk-like portion  17 D. The power transmission pins  17 H are fitted in through holes formed in the peripheral portion. The power transmission pins  17 H project rearward and extends substantially parallel to the axial direction of the drive shaft  16 . The power transmission pins  17 H form a discontinuing mechanism on a power transmission path between the engine E and the drive shaft  16  for preventing excessive power transmission. In the first embodiment, the power transmission pins  17 H are formed by sintered metal. The sintered metal has the fatigue ratio σ W /σ B  of approximately 0.5. σ W  represents the fatigue strength and σ B  represents the tensile strength.  
     [0051] An annular elastic member, which is a rubber damper  17 J is attached to the front end of the outer cylinder  17 E of the upstream pulley member  17 A to extend along the circumferential direction of the upstream pulley member  17 A. Through holes  17 K are formed in the rubber damper  17 J at positions corresponding to the power transmission pins  17 H. Each power transmission pin  17 H is fitted in one of the through holes  17 K. Therefore, in the pulley  17  according to the first embodiment, the power that is transmitted to the upstream pulley member  17 A via the belt  18  is transmitted to the downstream pulley member  17 B via the rubber damper  17 J and the power transmission pins  17 H. That is, the rubber damper  17 J and the power transmission pins  17 H are located on the power transmission path between the upstream pulley member  17 A and the downstream pulley member  17 B.  
     [0052] In the first embodiment, the upstream pulley member  17 A, the downstream pulley member  17 B, the power transmission pins  17 H, and the rubber damper  17 J form the pulley  17 . The first inner cylinder  17 C and the second inner cylinder  17 F are substantially coaxial and are apart from each other in the axial direction. The pulley  17  has an internal space surrounded by the upstream pulley member  17 A and the downstream pulley member  17 B.  
     [0053] A second rotating body, which is a hub  40  in the first embodiment, is secured to the front end of the drive shaft  16 . The hub  40  has a cylindrical support portion  40 A and a disk-like portion  40 B, which is integrally formed with the support portion  40 A and extends radially outward. The support portion  40 A has a female screw portion, which is screwed to a male screw portion formed at the front end of the drive shaft  16 . The hub  40  has a cylindrical portion  40 C, which is integrally formed with the disk-like portion  40 B and extends forward from the periphery of the disk-like portion  40 B, and a substantially disk-like flange  40 D, which is integrally formed with the cylindrical portion  40 C and extends radially outward from the front end of the cylindrical portion  40 C.  
     [0054] The support portion  40 A is located radially inward of the inner cylinders  17 C,  17 F, that is, closer to the axis of the drive shaft  16 . The disk-like portion  40 B is located between the inner cylinders  17 C,  17 F in the axial direction of the drive shaft  16 . The cylindrical portion  40 C is located radially outward of the first inner cylinder  17 C.  
     [0055] A one-way clutch unit  50  is located between the cylindrical portion  40 C and the first inner cylinder  17 C. The one-way clutch unit  50  includes a one-way clutch  51  and a first ball bearing, which is a bearing  52  in the first embodiment. The bearing  52  is located at the rear of the one-way clutch  51 .  
     [0056] The one-way clutch unit  50  has an outer ring  53 , which is secured to the inner circumferential surface of the cylindrical portion  40 C, and an inner ring  54 , which is secured to the outer circumferential surface of the first inner cylinder  17 C and is arranged such that the inner ring  54  is surrounded by the outer ring  53 . The bearing  52  includes rolling elements, which are balls  55 . The balls  55  are arranged in a line in the circumferential direction between the outer ring  53  and the inner ring  54 . The balls  55  permit the outer ring  53  to rotate relative to the inner ring  54 .  
     [0057] As shown in FIGS.  6 ( a ) and  6 ( b ), the one-way clutch  51  includes accommodating recesses  56 , which are formed in the inner circumferential surface of the outer ring  53 . The accommodating recesses  56  are arranged at equal angular intervals about the axis of the drive shaft  16 . A power transmission surface  57  is formed on one end (the leading end in the clockwise direction in FIG. 6( a )) of each accommodating recess  56 . Each accommodating recess  56  accommodates a roller  58  the axis of which is parallel to the axis of the drive shaft  16 . Each roller  58  is movable between a position in which the roller  58  engages with the corresponding power transmission surface  57  (see FIG. 6( a )), and a position apart from the engaging position (see FIG. 6( b )). A spring seat  59  is located at the end of each accommodating recess  56  opposite to the power transmission surface  57 . A spring  60  is located between each spring seat  59  and the corresponding roller  58  to urge the roller  58  toward the engaging position.  
     [0058] As shown in FIG. 6( a ), when the inner ring  54  is rotated in the direction shown by an arrow (clockwise direction) by the power transmitted from the engine E of the vehicle via the pulley  17 , each roller  58  moves to the engaging position by the force of the corresponding spring  60 . Then, each roller  58  is engaged between the power transmission surface  57  and the outer circumferential surface of the inner ring  54 . The outer ring  53  is thus rotated in the same direction as the inner ring  54 . Therefore, when the engine E of the vehicle is running, power of the engine E of the vehicle is transmitted to the drive shaft  16  via the pulley  17 , one-way clutch  51 , and the hub  40  so that the drive shaft  16  is always rotated.  
     [0059] When the engine E of the vehicle is stopped, that is, when the pulley  17  is stopped, if the outer ring  53  is rotated in the direction shown by an arrow (clockwise direction) as shown in FIG. 6( b ), each roller  58  separates from the engaging position against the force of the corresponding spring  60 . Therefore, the outer ring  53  runs idle with the inner ring  54 .  
     [0060] As shown in FIGS. 1 and 4, a support cylinder  12 B projects from the front wall of the front housing member  12  of the compressor main body C to surround the front end of the drive shaft  16 . A support  62 A of a stator fixing member  62  is fitted to the support cylinder  12 B. A second ball bearing, which is a bearing  63  in the first embodiment, is located between the support  62 A and the second inner cylinder  17 F of the upstream pulley member  17 A. That is, the pulley  17  is supported by the one-way clutch unit  50  and the bearing  63 , which are located axially separate from each other.  
     [0061] The stator fixing member  62  has a cylindrical mounting portion  62 B for mounting a stator  61 , which forms part of the motor generator MG, and a substantially disk-like coupling portion  62 C, which couples the mounting portion  62 B to the support  62 A. The coupling portion  62 C is located between the inner cylinders  17 C,  17 F in the axial direction of the drive shaft  16  and is located at the rear of the disk-like portion  40 B. The mounting portion  62 B is located radially outward of the cylindrical portion  40 C and the second inner cylinder  17 F.  
     [0062] The stator  61  is mounted on the outer circumferential surface of the mounting portion  62 B. The stator  61  includes a fixed iron core and a coil, which is wound about the fixed iron core. A rotor  64 , which forms part of the motor generator MG, is secured to the outer circumferential portion of the flange  40 D. The rotor  64  is arranged about the stator  61 . The rotor  64  includes a rotor core and a rotor conductor, which is secured to the rotor core. The motor generator MG is located in the internal space of the pulley  17 .  
     [0063] The coil of the stator  61  is connected to a battery (not shown) via a motor drive circuit (not shown), which includes an inverter, converter, and the like. The motor drive circuit controls whether to store the power generated by the coil in the battery or to supply power from the battery to the coil in accordance with a command from a motor control unit, which is not shown.  
     [0064] When the battery needs to be charged while the engine E is running, the motor control unit controls the motor drive circuit such that the motor generator MG functions as an induction generator and the motor generator MG generates electric power. When the rotor  64  is rotated with the hub  40  by the power transmission from the engine E of the vehicle, electricity is generated at the coil and the electricity is stored in the battery via the motor drive circuit.  
     [0065] When the battery need not be charged while the engine E of the vehicle is running, the motor control unit controls the motor drive circuit such that the motor generator MG does not generate electricity. This is achieved when the motor control unit commands the motor drive circuit not to supply exciting current to the motor generator MG, which is formed by an induction machine. In this state, magnetic force does not act between the stator  61  and the rotor  64 . Therefore, although the rotor  64  is rotated by the power from the engine E of the vehicle, energy loss, such as heat generation due to iron loss of the stator  61  and the rotor  64 , is not caused. Although the rotor  64  is rotated by power from the engine E of the vehicle, torque fluctuation of the drive shaft  16  based on the magnetic force is not caused.  
     [0066] If it is determined, based on the external information, that air conditioning (cooling) is necessary while the engine E of the vehicle is stopped, the motor control unit controls the motor drive circuit such that the motor generator MG functions as an induction motor. That is, rotational force is generated at the rotor  64  by the power supplied from the motor drive circuit to the coil. The rotational force is transmitted to the drive shaft  16  via the hub  40 . Accordingly, the vehicle passenger compartment can be air conditioned while the engine E of the vehicle is stopped.  
     [0067] When the motor generator MG functions as a motor and rotates the hub  40 , the one-way clutch  51  operates to stop the power transmission between the hub  40  and the pulley  17 . Thus, the power of the motor generator MG is prevented from being transmitted to the engine E of the vehicle.  
     [0068] In the first embodiment, drive force transmitted to the upstream pulley member  17 A from the engine E of the vehicle is transmitted to the downstream pulley member  17 B via the rubber damper  17 J and the power transmission pins  17 H. Since the rubber damper  17 J is located on the power transmission path between the upstream pulley member  17 A and the downstream pulley member  17 B, misalignment between the axis of the bearing  52  and the axis of the bearing  63  is absorbed. That is, the deformation of the rubber damper  17 J reduces stress generated on the bearings  12 A,  52 , and  63  due to the misalignment. The rubber damper  17 J prevents the rotation vibration of the drive shaft  16  caused by the compression reaction force in the compression mechanism, or the torque fluctuation, from being transmitted from the downstream pulley member  17 B to the upstream pulley member  17 A.  
     [0069] Due to the operation of the one-way clutch  51  that permits transmission of power only in one rotational direction, the rotation vibration that acts in the other direction is not transmitted from the hub  40  to the pulley  17 .  
     [0070] In the first embodiment, when the magnitude of the transmission torque between the upstream pulley member  17 A and the downstream pulley member  17 B is not large enough to affect the engine E of the vehicle, that is, during the normal power transmission state, power transmission from the engine E of the vehicle to the drive shaft  16  is continued. However, if an abnormality (such as a deadlock) occurs in the compressor main body C and the transmission torque exceeds an acceptable value, the power transmission pins  17 H are broken by the excessive load, which stops power transmission from the upstream pulley member  17 A to the downstream pulley member  17 B. Accordingly, the engine E of the vehicle is prevented from being adversely affected by the excessive transmission torque.  
     [0071] The first embodiment provides the following advantages.  
     [0072] (1) The one-way clutch  51  is located on the power transmission path between the first inner cylinder  17 C of the pulley  17  and the cylindrical portion  40 C of the hub  40 . Therefore, for example, as compared to a case where the electromagnetic clutch is located on the power transmission path between the pulley  17  and the hub  40 , the size and the weight of the mechanism for selectively discontinuing power transmission between the drive shaft  16  and the pulley  17  are reduced. This facilitates reducing the size of the power transmission mechanism PT and reducing the size and weight of the compressor that includes the power transmission mechanism PT. Since a controller for selectively connecting and disconnecting the electromagnetic clutch is unnecessary, the structures of the power transmission mechanism PT and the compressor are simplified. This reduces the cost of the power transmission mechanism PT and the compressor.  
     [0073] (2) The motor generator MG is located in the internal space of the pulley  17  surrounded by the upstream pulley member  17 A and the downstream pulley member  17 B. Therefore, the size of the power transmission mechanism PT is reduced by effectively using the internal space.  
     [0074] (3) The first inner cylinder  17 C and the second inner cylinder  17 F are apart from each other in substantially the axial direction of the pulley  17 . Therefore, for example, as compared to a case where the first inner cylinder  17 C and the second inner cylinder  17 F are located at the same axial position, a space for accommodating the motor generator MG is easily obtained.  
     [0075] (4) The first inner cylinder  17 C and the second inner cylinder  17 F, which are apart from each other in the axial direction, are supported by the bearings  52  and  63 , respectively. Therefore, when an external force is applied to the pulley  17 , the pulley  17  is prevented from being inclined with respect to the axis of the drive shaft  16 . Thus, partial wear of each part of the pulley  17  and bad engagement of the one-way clutch  51  caused by the inclination of the pulley  17  are suppressed.  
     [0076] (5) The discontinuing mechanism is located on the power transmission path between the engine E of the vehicle and the drive shaft  16 . Therefore, for example, although an abnormality, such as a deadlock, occurs in the compressor main body C, an excessive load is prevented form being applied to the engine E of the vehicle.  
     [0077] (6) The power transmission pins  17 H located on the power transmission path between the upstream pulley member  17 A and the downstream pulley member  17 B break when the transmission torque is excessive to stop the power transmission. That is, in the first embodiment, the power transmission between the upstream rotating body and the downstream rotating body can be disconnected by breaking the breakable members, which are the power transmission pins  17 H.  
     [0078] (7) The power transmission pins  17 H are formed by sintered metal. Since the sintered metal has a relatively low ductility, the magnitude of the transmission torque required for breaking the power transmission pins  17 H is easily set. The fatigue ratio σ W /σ B  of the sintered metal is relatively easily maintained at a high value. Therefore, during the normal power transmission state, the durability against repeated stress that acts on the power transmission pins  17 H is maintained relatively high, and the balance between the durability and the transmission torque amount for breaking the power transmission pins  17 H is easily optimized. Therefore, the power transmission pins  17 H reliably transmit power during the normal power transmission state showing satisfactory durability and reliably blocks power when the transmission torque becomes excessive.  
     [0079] (8) The breakable members are simple pins  17 H. Therefore, the structures of the breakable members and the through holes  17 K are simplified, which facilitates the manufacture and reduces the cost of the power transmission mechanism PT.  
     [0080] (9) The elastic member, which is the rubber damper  17 J, is located on the power transmission path between the upstream pulley member  17 A and the downstream pulley member  17 B. Therefore, deformation of the rubber damper  17 J reduces the stress generated on the bearings  12 A,  52 , and  63  caused by the misalignment between the axis of the bearing  52  and the axis of the bearing  63  due to the manufacturing tolerance and the like. Therefore, the durability of the compressor is improved.  
     [0081] (10) The rubber damper  17 J, which serves as a buffer member, reduces the torque fluctuation that is transmitted from the downstream pulley member  17 B to the upstream pulley member  17 A.  
     [0082] (11) The bearing  52  is formed by balls  55 , which are arranged in a line in the circumferential direction between the outer ring  53  and the inner ring  54 . Therefore, for example, as compared to a structure in which balls  55  are arranged in the axial direction of the bearing  52 , the axial length of the bearing  52  is reduced.  
     [0083] (12) The compressor main body C reduces the discharge displacement per one rotation of the drive shaft  16  to substantially zero. Since the discharge displacement of the compressor main body C during rotation of the drive shaft  16  is reduced to substantially zero, unnecessary load is hardly applied to the engine E of the vehicle when air conditioning is not required.  
     [0084] (13) According to the control valve  35  of the first embodiment, the discharge displacement (flow rate of refrigerant) of the compressor main body C per unit time, which greatly affects the load torque of the compressor main body C, is controlled directly from the outside. Also, for example, the flow rate of refrigerant is controlled to be less than or equal to a predetermined amount in an accurate and responsive manner without using a refrigerant flow rate sensor or the like.  
     [0085] A second embodiment of the present invention will now be described with reference to FIGS.  5 ( a ) and  5 ( b ). A compressor of the second embodiment has the same structure as the first embodiment except that mainly the downstream pulley member and the structure of the coupling portion between the downstream pulley member and the upstream pulley member are modified. Therefore, like or the same reference numerals are given to those components that are like or the same as the corresponding components of the first embodiment, and detailed explanations are omitted.  
     [0086] As shown in FIGS.  5 ( a ) and  5 ( b ), a downstream rotating body, which is a downstream pulley member  70  in the second embodiment, has a first inner cylinder  70 A, which is fitted in the inner ring  54  of the one-way clutch unit  50 . The downstream pulley member  70  also has an outer ring  70 C, which is formed integrally with the first inner cylinder  70 A via breakable members, which are spokes  70 B in the second embodiment. The spokes  70 B (four in the second embodiment) extend from the first inner cylinder  70 A radialy toward the outer ring  70 C. The spokes  70 B couple the first inner cylinder  70 A to the outer ring  70 C to permit power transmission.  
     [0087] In the second embodiment, the downstream pulley member  70 , which includes the integrally formed first inner cylinder  70 A, the spokes  70 B, and the outer ring  70 C, is made by sintered metal. The sintered metal is set such that the fatigue ratio σ W /σ B  is maintained approximately at 0.5.  
     [0088] The annular elastic member, which is the rubber damper  71 , is located between the rear end of the outer ring  70 C and the front end of the outer cylinder  17 E of the upstream pulley member  17 A. The rubber damper  71  is fixed to the outer ring  70 C and the outer cylinder  17 E.  
     [0089] In the second embodiment, the drive force transmitted from the engine E of the vehicle to the upstream pulley member  17 A is transmitted to the hub  40  via the rubber damper  71 , the outer ring  70 C, and the spokes  70 B. That is, the rubber damper  71  and the spokes  70 B are located on the power transmission path between the upstream pulley member  17 A and the hub  40 . The rubber damper  71  absorbs misalignment between the axis of the bearing  52  and the axis of the bearing  63 . The rubber damper  71  prevents the torque fluctuation from being transmitted from the downstream pulley member  70  to the upstream pulley member  17 A.  
     [0090] In the second embodiment, when the transmission torque between the outer ring  70 C of the downstream pulley member  70  and the first inner cylinder  70 A is not large enough to affect the engine E of the vehicle, that is, during normal power transmission state, the power transmission from the engine E to the drive shaft  16  is continued. However, if an abnormality (such as a deadlock) occurs in the compressor main body C and the transmission torque exceeds the acceptable value, the spokes  70 B are broken by the excessive load, which stops the power transmission from the upstream pulley member  17 A to the hub  40 . Accordingly, the engine E of the vehicle is prevented from being adversely affected by the excessive transmission torque.  
     [0091] The second embodiment provides the following advantage in addition to the advantages (1) to (7) and (9) to (13) of the first embodiment.  
     [0092] (14) Since the power transmission can be stopped by breaking the downstream pulley member  70 , a breakable member formed by a member other than the downstream pulley member  70  need not be provided. Therefore, a procedure for mounting the breakable member formed by another member to the downstream pulley member  70  is omitted.  
     [0093] It should be apparent to those skilled in the art that the present invention may be embodied in many other specific forms without departing from the spirit or scope of the invention. Particularly, it should be understood that the invention may be embodied in the following forms.  
     [0094] In the first embodiment, the pulley  17  has the upstream pulley member  17 A and the downstream pulley member  17 B, and the breakable members, which are the power transmission pins  17 H, are located on the power transmission path between the upstream pulley member  17 A and the downstream pulley member  17 B. On the contrary, the hub  40  may have an upstream rotating body and a downstream rotating body, and a breakable member may be located on a power transmission between the rotating bodies.  
     [0095] In the second embodiment, the downstream pulley member  70  is formed of sintered metal. The power transmission is stopped when the downstream pulley member  70  is broken. However, the hub  40  may be made of sintered metal, and the power transmission may be stopped when the hub  40  is broken.  
     [0096] The compressor according to the first and second embodiments includes the discontinuing mechanism on the pulley  17  ( 70 ). On the contrary, the discontinuing mechanism may be located on a power transmission path between the pulley  17  ( 70 ) and the hub  40 , such as a power transmission path between the pulley  17  and the one-way clutch unit  50 , or between the one-way clutch unit  50  and the hub  40 . The discontinuing mechanism may be located on the one-way clutch unit  50 . The discontinuing mechanism may be located on a power transmission path between the hub  40  and the drive shaft  16 .  
     [0097] In the above embodiments, the fatigue ratio σ W /σ B  of the sintered metal that form the breakable member is maintained approximately at 0.5. However, the fatigue ratio may be varied as long as the breakable member can be broken when the excessive transmission torque is applied to the breakable member.  
     [0098] In the above embodiments, the breakable member is made by sintered metal. However, the breakable member may be formed by low-carbon steel. The fatigue ratio σ W /σ B  of low-carbon steel is relatively easy to maintain at a relatively high value (approximately 0.5). Therefore, during the normal power transmission state, the durability against the repeated stress that acts on the breakable member is maintained relatively high, and the balance between the durability and the transmission torque amount for breaking the power transmission pins  17 H is easily optimized.  
     [0099] In the above embodiments, the breakable member is made of metal. However, the breakable member may be made of resin or ceramics as long as the breakable member can be broken at a predetermined transmission torque when an excessive transmission torque is applied to the breakable member.  
     [0100] In the first embodiment, the power transmission pins  17 H may be formed integrally with the downstream pulley member  17 B. When the downstream pulley member  17 B, to which the power transmission pins  17 H are integrally formed, is made by breakable material such as sintered metal, the power transmission can be stopped by breaking of the portion corresponding to the power transmission pins  17 H.  
     [0101] In the first embodiment, the power transmission pins  17 H are fixed to the downstream pulley member  17 B and coupled to the upstream pulley member  17 A via the rubber damper  17 J. On the contrary, the power transmission pins  17 H may be fixed to the upstream pulley member  17 A and coupled to the downstream pulley member  17 B via the rubber damper  17 J.  
     [0102] In the first embodiment, the power transmission pins  17 H may be attached to the downstream pulley member  17 B via a tubular elastic member such as a rubber damper. That is, the power transmission pins  17 H may be coupled to both the upstream pulley member  17 A and the downstream pulley member  17 B via the elastic member.  
     [0103] In the first embodiment, all the power transmission pins  17 H are coupled to one rubber damper  17 J. However, each power transmission pin  17 H may be coupled to separate rubber damper. For example, the power transmission pins  17 H may be formed as shown in FIG. 7. A tubular elastic member having a circular cross-section, which is a rubber damper  80 , is fitted to the rear end of each power transmission pin  17 H. Each rubber damper  80  is accommodated in a damper accommodating recess  81  formed in the front end of the outer cylinder  17 E of the upstream pulley member  17 A. In this case, as compared to the first embodiment, the amount of rubber material used for forming the rubber damper is reduced.  
     [0104] The cross-sections of the power transmission pins  17 H and the hole of the rubber damper  17 J ( 71 ,  80 ) need not be circular. The cross-section of the outline of the rubber damper  80  shown in FIG. 7, in particular, need not be circular.  
     [0105] In the first and second embodiments, the breakable member and the elastic member are separate members. However, the elastic member may also serve as the breakable member. For example, in the first embodiment, the upstream pulley member  17 A and the downstream pulley member  17 B may be coupled to each other via only the rubber damper  17 J. In this case, if an excessive transmission power is applied to the rubber damper  17 J, the power transmission is stopped when the rubber damper  17 J is pulled apart.  
     [0106] In the above embodiments, the elastic member is located on the pulley  17  but may be located on the hub  40 . For example, the elastic member may be located on a power transmission path between the one-way clutch unit  50  and the drive shaft  16 . The elastic member may also be located on a power transmission path between the pulley  17  and the one-way clutch unit  50 , or between the one-way clutch unit  50  and the hub  40 . The elastic member may be located on the power transmission path between the hub  40  and the drive shaft  16 .  
     [0107] The elastic member may be formed of elastic material other than rubber such as elastomer.  
     [0108] The elastic member for absorbing misalignment between the axis of the bearing  52  and the axis of the bearing  63  may be omitted.  
     [0109] In the above embodiments, the discontinuing mechanism is formed by a breakable member. However, the discontinuing mechanism need not be a breakable member. For example, the discontinuing mechanism may be formed by a coupling member that operably couples the upstream rotating body to the downstream rotating body and selectively engages with and disengages from one of the rotating bodies.  
     [0110] The structure of the one-way clutch  51  may be modified as long as the power transmission from the pulley  17  to the drive shaft  16  is permitted and the power transmission from the motor generator MG to the pulley  17  is prevented.  
     [0111] A motor generator that utilizes a permanent magnet may be employed instead of the motor generator MG that is formed by an induction machine. The motor generator that uses a permanent magnet easily obtains a great output as compared to the motor generator that is formed by an induction machine.  
     [0112] An electric part that functions only as a motor may be employed instead of the electric part that functions as a motor and a generator.  
     [0113] The motor generator MG need not be located inside the internal space of the pulley  17  but may be located outside the pulley  17 .  
     [0114] The first inner cylinder  17 C ( 70 A) and the second inner cylinder  17 F may be located at substantially the same position in the axial direction.  
     [0115] The bearing  52  may have multiple lines of balls  55 .  
     [0116] In the above embodiments, the position of the valve body  106  is adjusted by applying the pressure in the pressure sensing chamber  114  to the movable wall  120  accommodated in the pressure sensing chamber  114 . On the contrary, the pressure in the pressure sensing chamber  114  may be applied to a pressure sensing member such as a bellows or a diaphragm located inside the pressure sensing chamber  114  to adjust the axial position of the valve body  106 .  
     [0117] In the above embodiments, the control valve  35  is designed such that the position of the valve body  106  is automatically changed to vary the displacement to cancel the fluctuation of the pressure difference between two pressure monitoring points P 1 , P 2  located in the refrigerant circuit. On the contrary, for example, the position of the valve body  106  may be changed in accordance with the pressure of one pressure monitoring point located in the refrigerant circuit. Also, for example, the position of the valve body  106  may be changed in accordance with a command from the outside.  
     [0118] In the above embodiments, the control valve  35  is designed such that the criteria of the positioning operation of the valve body  106  can be changed by an external control. On the contrary, for example, the control valve  35  may be designed to perform only the automatic positioning operation of the valve body  106  without being controlled from the outside.  
     [0119] Instead of providing the control valve  35  in the supply passage, a control valve may be located in the bleed passage  33 . In this case, the control valve in the bleed passage adjusts the flow rate of refrigerant from the crank chamber  15  to the suction chamber  27  to control the crank pressure Pc. That is, the control valve may be located at any place as long as the control valve is located in at least one of the supply passage and the bleed passage connected to the crank chamber  15 . The discharge chamber  28  and the suction chamber  27  are pressure zones, which are exposed to pressure that is different from the crank pressure Pc. The supply passage and the bleed passage are pressure control passages, which connect the pressure zone to the control pressure zone, which is the crank chamber  15 .  
     [0120] The single-sided compressor main body C, which causes single-headed pistons to perform the compression operation, may be changed to a both-sided compressor main body, which causes double-headed pistons to perform the compression operation in cylinder bores formed on both sides of a crank chamber.  
     [0121] The present invention may be embodied in a compressor that has a drive plate that is rotatably supported by the drive shaft to wobble with respect to the drive shaft. For example, the present invention may be embodied in a wobble plate type compressor.  
     [0122] The discharge displacement per one rotation of the drive shaft  16  may be greater than zero at the minimum displacement of the compressor main body C.  
     [0123] The compressor main body C may be changed to a fixed displacement compressor main body in which the stroke of the pistons  25  is constant.  
     [0124] The present invention may be embodied in rotary compressors such as a scroll compressor.  
     [0125] The first rotating body, which is the pulley  17 , may be changed to, for example, a sprocket or a gear.