Patent Publication Number: US-2009220356-A1

Title: Swash plate type variable displacement compressor

Description:
BACKGROUND OF THE INVENTION 
     The present invention relates to a variable displacement compressor for use in an automotive air conditioner, and the like. 
     Generally, a variable displacement compressor (hereinafter referred to as “compressor”) is known as a compressor for use in an automotive air conditioner that is operable to variably control its displacement. This type of compressor uses a displacement control valve for adjusting pressure in a crank chamber to change the inclination angle of a swash plate accommodated in the crank chamber, thereby to adjust the stroke length of pistons and hence to control the displacement of the compressor. 
     Japanese Unexamined Patent Application Publication No. 10-54350 discloses the compressor having a valve disposed in a bleed passage connecting the crank chamber to a suction pressure region of the compressor. The valve includes a valve body a coil spring and a counterweight. The coil spring urges the valve body in the direction that causes the valve body to open a valve hole. When the rotational speed of the rotary shaft reaches a predetermined value, the valve body is moved in the direction that causes the valve body to close the valve hole by centrifugal force acting on the counterweight, which closes the bleed passage and stops the flow of refrigerant gas from the crank chamber into the suction region through the bleed passage. During the compression operation under a large displacement, the valve closes the bleed passage and the pressure in the crank chamber is gradually increased by blow-by gas flowing into the crank chamber. Thus, the displacement of the compressor is decreased so that the compression load is reduced and the contact pressure acting on various sliding surfaces of the compressor is reduced, accordingly. 
     However, according to the reference No. 10-54350, the bleed passage is closed by the valve when the rotational speed of the rotary shaft reaches the predetermined value or more, with the result that the amount of refrigerant gas drawn from the crank chamber into the suction pressure region becomes zero. In this operating state, it takes a long time to increase the displacement of the compressor, and the displacement recovery performance of the compressor is deteriorated, because the bleed passage has been closed thereby to prevent the refrigerant gas from being rapidly drawn from the crank chamber. 
     The present invention, which has been made in light of the above problems, is directed to a swash plate type variable displacement compressor which ensures the performance to recover displacement of the compressor during the operation at a low rotational speed and to reduce power loss during the operation at a high rotational speed. 
     SUMMARY OF THE INVENTION 
     In accordance with an aspect of the present inventions a swash plate type variable displacement compressor includes a housing including a cylinder block having a cylinder bore formed therein, a crank chamber formed in the housing, a rotary shaft extending through the crank chamber, and a swash plate connected to the rotary shaft. The rotary shaft is rotatably supported by the housing. The swash plate is integrally rotatable with the rotary shaft and inclinable relative to the rotary shaft. The compressor further includes a piston received in the cylinder bore to be reciprocally movable, a discharge pressure region for receiving discharge pressure gas, a suction pressure region for receiving, suction pressure gas, a supply passage connecting the crank chamber to the discharge pressure region and first and second bleed passages. The supply passage is provided with a displacement control valve. The pressure in the crank chamber is varied by adjusting the opening of the displacement control valve to change the inclination angle of the swash plate thereby to control the displacement of the compressor. The first bleed passage connecting the crank chamber to the suction pressure region is provided with a valve and the second bleed passage constantly connecting the crank chamber to the suction pressure region is provided with a throttle. The valve operates to close the first bleed passage according to the magnitude of centrifugal force generated by the rotation of the rotary shaft. 
     Other aspects and advantages of the invention will become apparent from the following description, taken in conjunction with the accompanying drawings, illustrating by way of example the principles of the invention. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The features of the present invention that are believed to be novel are set forth with particularity in the appended claims. The invention together with objects and advantages thereof, may best be understood by reference to the following description of the presently preferred embodiments together with the accompanying drawings in which: 
         FIG. 1  is a longitudinal cross-sectional view of a swash plate type variable displacement compressor according to a first preferred embodiment of the present invention; 
         FIG. 2  is an enlarged cross-sectional view of a valve used in the compressor according to the first preferred embodiment of the present invention; 
         FIG. 3  is an enlarged fragmentary cross-sectional view of the compressor according to the first preferred embodiment of the present invention; 
         FIG. 4  is a schematic block diagram illustrating the compressor according to the first preferred embodiment of the present invention; 
         FIG. 5  is a schematic graph showing a relation between the rotational speed of a rotary shaft of the compressor and the total cross-sectional area of throttle opening in bleed passages of the compressor according to the first preferred embodiment of the present invention; 
         FIG. 6  is an enlarged fragmentary cross-sectional view of the compressor showing the valve according to a second preferred embodiment of the present invention; 
         FIG. 7  is an enlarged fragmentary cross-sectional view of the compressor showing the valve according to a third preferred embodiment of the present invention, and 
         FIG. 8  is a schematic view as seen in the direction of the arrow D in  FIG. 7 . 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     The following will describe a swash plate type variable displacement compressor (hereinafter referred to as “compressor”) according to the first preferred embodiment of the present invention with reference to  FIGS. 1 through 5 . Referring to  FIG. 1 , the compressor designated by numeral  10  has a housing  11  forming the outer shell of the compressor  10 . The housing  11  includes a cylinder block  12 , a front housing  13  joined to the front end of the cylinder block  12 , and a rear housing  14  joined to the rear end of the cylinder block  12 . The cylinder block  12  has a plurality of cylinder bores  12 A formed therein. In  FIG. 1 , the left side of the drawing corresponds to the front side of the compressor  10 , and the right side of the drawing corresponds to the rear side of the compressor  10 . The front housing  13 , the cylinder block  12  and the rear housing  14  are fastened together in the longitudinal direction of the compressor  10  by a plurality of bolts  15  (only one bolt being shown) inserted through the front housing  13 , the cylinder block  12  and the rear housing  14 , thus the housing  11  of the compressor  10  is formed thereby. 
     The front housing  13  has a crank chamber  16  formed therein, whose rear end is closed by the cylinder block  12 . A rotary shaft  17  extends through the center of the crank chamber  16  and is rotatably supported by the front housing  13  and the cylinder block  12  through radial bearings  18 ,  19 , respectively. A shaft seal mechanism  20  is disposed in slide contact with the circumferential surface of the rotary shaft  17  at a position forward of the radial bearing  18  supporting the front part of the rotary shaft  17 . The seal mechanism  20  has a lip seal member to prevent refrigerant gas in the crank chamber  16  from leaking out through the clearance between the front housing  13  and the rotary shaft  17 . The rotary shaft  17  is connected at the front end thereof to an external drive source (not shown) through a power transmission mechanism (not shown either) so as to be rotated by the external drive source. 
     A lug plate  21  is fixedly mounted on the rotary shaft  17  in the crank chamber  16  so as to rotate integrally therewith. A swash plate  23  as a part of displacement changing mechanism  22  of the compressor  10  is provided on the rotary shaft  17  behind the lug plate  21  and supported in such a way that it is slidable in the axial direction of the rotary shaft  17  and inclinable relative to the axis of the rotary shaft  17 . A hinge mechanism  24  is interposed between the swash plate  23  and the lug plate  21 , through which the swash plate  23  and the lug plate  21  are connected such that the swash plate  23  is integrally rotatable with the lug plate  21  and the rotary shaft  17 , while inclinable relative to the rotary shaft  17 . 
     A coil spring  25  is disposed on the rotary shaft  17  between the lug plate  21  and the swash plate  23 . A sleeve  26  is slidably disposed on the rotary shaft  17  and urged rearward by the pressing force of the coil spring  25 . The swash plate  23  is urged by the coil spring  25  through the sleeve  26  rearward or in the direction that decreases the inclination angle of the swash plate  23 . The inclination angle of the swash plate  23  means an angle between the swash plate  23  and an imaginary plane that is perpendicular to the axis of the rotary shaft  17 . The swash plate  23  has a restricting portion  23 A projecting from the front end thereof and abutable with the lug plate  21 ′, thereby restricting the maximum inclination angle of the swash plate  23 . The rotary shaft  17  has a snap ring  27  fitted thereon behind the swash plate  23 . The rear end of the swash plate  23  is abutable with the snap ring  27 , thereby restricting the minimum inclination angle of the swash plate  23 . Referring to  FIG. 1 , the swash plate  23  indicated by the solid line represents the position at the maximum inclination angle thereof, and the swash plate  23  indicated by the double-dashed line represents the position at the minimum inclination angle thereof. 
     Each cylinder bore  12 A of the cylinder block  12  receives therein a reciprocally movable single-headed piston  29 . The piston  29  engages at the neck portion thereof with the outer periphery of the swash plate  23 ′ through a pair of shoes  30 . As the swash plate  23  is rotated with the rotary shaft  17 , each piston  29  is reciprocated in its associated cylinder bore  12 A through the pair of shoes  30 . 
     As shown in  FIG. 1 , the front end of the rear housing  14  is joined to the rear end of the cylinder block  12  through a valve plate  32 . The rear housing  14  has a suction chamber  38  which serves as a suction pressure region formed at a center region thereof. The suction chamber  38  is in communication with a compression chamber  31  defined by the cylinder bore  12 A through a suction port  36  formed through the valve plate  32 . The rear housing  14  also has a discharge chamber  39  which serves as a discharge pressure region formed at a circumferential region thereof. The discharge chamber  39  and the suction chamber  38  are separated by a partition wall  14 A. The valve plate  32  defining the compression chamber  31  with the piston  29  in the cylinder bore  12 A has a discharge port  37  formed therethrough in communication with the discharge chamber  39 . The suction port  36  and the discharge port  37  for each cylinder bore  12 A are provided with a suction valve  33  and a discharge valve  34 , respectively. 
     When the piston  29  moves toward the bottom dead center from the top dead center thereof, refrigerant gas in the suction chamber  38  is drawn into the compression chamber  31  through the suction port  36  and the suction valve  33 . Refrigerant gas drawn into the compression chamber  31  is compressed to a predetermined pressure by the motion of the piston  29  from the bottom dead center to the top dead center thereof, and discharged into the discharge chamber  39  through the discharge port  37  and the discharge valve  34 . 
     A supply passage  42  is formed in the cylinder block  12  and the rear housing  14  to connect the discharge chamber  39  to the crank chamber  16 . An electromagnetic displacement control valve  35  is disposed in the supply passage  42 . The displacement control valve  35  is in communication with the suction chamber  38  through a pressure sensing passage  61 . The opening of the displacement control valve  35  is adjustable according to the detected pressure in the suction chamber  38  or in response to any external command signals. Adjusting the opening of the displacement control valve  35  varies the flow rate of high-pressure refrigerant gas introduced from the discharge chamber  39  into the crank chamber  16 . The pressure differential between the crank chamber  16  and the compression chamber  31  across the piston  29  is varied, thereby changing the inclination angle of the swash plate  23 . Accordingly, the stroke length of the piston  29  is varied thereby to control the displacement of the compressor  10 . 
     The center of the cylinder block  12  has a shaft hole  43  therethrough, and a recess  44  located behind and in communication with the shaft hole  43 . The rear end of the rotary shaft  17  is inserted into and supported by the shaft hole  43  through the radial bearing  19 . The compressor  10  of the first preferred embodiment includes a first bleed passage  48  and a second bleed passage  58 . A passage hole  45  forming a part of the first bleed passage  48  extends in the rotary shaft  17  along its center axis. The front end portion of the passage hole  45  is opened to the crank chamber  16  at a position adjacent to the radial bearing  18  and the shaft seal mechanism  20 . The passage hole  45  is closed at the rear end by a plug  60 . A valve  50  is mounted on the rotary shaft  17  at the rear end portion thereof in the recess  44 . The valve  50  will be described in detail later. 
     A thrust bearing  46  and a support spring  47  are interposed between the rear end of the rotary shaft  17  and the valve plate  32 . The recess  44  is in communication with the suction chamber  38  through a communication hole  49  formed at the center of the valve plate  32 . The communication hole  49  serves as a throttle for restricting flow rate of refrigerant gas drawn from the crank chamber  16  into the suction chamber  38 . The aforementioned first bleed passage  48  includes the passage hole  45 , the recess  44 , the valve  50  and the communication hole  49  so as to connect the crank chamber  16  to the suction chamber  38 . 
     The valve  50  is provided for opening or closing the first bleed passage  48 . As shown in  FIG. 2 , the rotary shaft  17  has plane seating surfaces  51 ,  52  formed by cutting off the top and bottom of the circumferential surface of the rear end portion of the rotary shaft  17 , respectively. A valve hole  53  is formed in the radial direction of the rotary shaft  17 , or the compressor  10  so as to provide fluid communication between the seating surfaces  51 ′,  52  and also to be in communication with the passage hole  45 . The valve hole  53  is larger in diameter on the side opened to the seating surface  51  than the opposite side opened to the seating surface  52 . A valve body  54  is movably mounted on the rotary shaft  17  so as to open or close the valve hole  53 . The valve body  54  is disposed on the side of the seating surface  51 , and a counterweight  55  connected to the valve body  54  through a connecting portion  56  is disposed on the side of the seating surface  52 . A coil spring  57  serving as an urging member is provided-between the seating surface  51  and the valve body  54  for urging the valve body  54  toward its opened position. 
     A centrifugal force acting on the counterweight  55  is increased with an increase in rotational speed of the rotary shaft  17 , with the result that the counterweight  55  is moved away from the axis of the rotary shaft  17 . Accordingly, the valve body  54  is moved toward the axis of the rotary shaft  17  against the urging force of the coil spring  57  and brought into contact with the seating surface  51 , thereby to close the valve hole  53 . On the other hand, the centrifugal force acting on the counterweight  55  is decreased with a decrease in rotational speed of the rotary shaft  17 , with the result that the urging force of the coil spring  57  becomes greater than the centrifugal force acting on the counterweight  55 . Accordingly, the valve body  54  is moved away from the axis of the rotary shaft  17  by the urging force of the coil spring  57 , thereby to open the valve hole  53 .  FIGS. 1 and 2  show the valve  50  in its opened position during compressor operation at a high rotational speed of the rotary shaft  17 , and  FIG. 3  shows the valve  50  in its closed position during compressor operation at a low rotational speed of the rotary shaft  17 . 
     Referring back to  FIG. 1 , the second bleed passage  58  connecting the crank chamber  16  to the suction chamber  38  is formed in the cylinder block  12 . The second bleed passage  58  has a throttle hole  59  formed in the valve plate  32  which functions as a fixed throttle for throttling the flow rate of the refrigerant gas. The crank chamber  16  is in constant communication with the suction chamber  38  through the second bleed passage  58 . 
     Referring to  FIG. 4  showing a schematic block diagram illustrating the compressor  10  according to the first preferred embodiment, the discharge chamber  39  is in communication with the crank chamber  16  through the supply passage  42  in which the displacement control valve  35  is disposed. The crank chamber  16  is in communication with the suction chamber  38  through the first bleed passage  48  and the second bleed passage  58 . The first bleed passage  48  is provided with the valve  50  operable to open or close according to the magnitude of the centrifugal force and the second bleed passage  58  is provided with the throttle hole  59  serving as a fixed throttle. 
       FIG. 5  is a schematic graph showing a relation between rotational speed N of the rotary shaft  17  of the compressor  10  and total cross-sectional area AS of the throttle opening which is the sum of the cross-sectional areas of the throttle openings formed in the first and second bleed passages  48 ,  58  according to the first preferred embodiment. In the graph of  FIG. 5 , the cross-sectional areas of the communication hole  49  provided in the first bleed passage  48  and the throttle hole  59  provided in the second bleed passage  58  are designated by reference symbols AA, AB, respectively. During the operation of the compressor  10  at a low rotational speed, the valve  50  is in its opened position. In this state, the relation among total cross-sectional area AS 1  of the throttle opening, the cross-sectional area AA of the communication hole  49  and the cross-sectional area AB of the throttle hole  59  is expressed by AS 1 =AA+AB. On the other hand, during the operation of the compressor  10  at a high rotational speed, the valve  50  is in its closed position, that is, the first bleed passage  48  is closed and only the second bleed passage  58  is opened. When the rotational speed of the rotary shaft  17  is at or higher than NC1, the relation among total cross-sectional area AS 2  of the throttle opening, the cross-sectional area AA of the communication hole  49  and the cross-sectional area AB of the throttle hole  59  is expressed by AS 2 =AB. The flow rate of refrigerant, gas drawn from the crank chamber  16  into the suction chamber  38  through the first and second bleed passages  48 ,  58  is proportional to the total cross-sectional area AS of the throttle opening. Therefore, the flow rate of refrigerant gas during the operation at a low rotational speed that is expressed by AS 1  (=AA+AB) is larger than that during the operation at a high rotational speed that is expressed by AS 2  (=AB). The cross-sectional areas AA and AB are previously set at any values suitable to ensure both of the displacement recovery and power loss reduction during the operation of the compressor  10 . The diameter of the fully opened valve hole  53  is set such that the cross-sectional area of such valve hole  53  is larger than the cross-sectional area AA of the communication hole  49 . 
     The following will describe the operation of the compressor constructed as described above. As the rotary shaft  17  is rotated by the external drive source such as a vehicle engine, the swash plate  23  is rotated with the rotary shaft  17  through the lug plate  21  and the hinge mechanism  24 . Accordingly, the rotational movement of the swash plate  23  is converted into reciprocating movement of the piston  29  by way of the shoes  30 . The piston  29  is reciprocated in the cylinder bore  12 A, thereby causing refrigerant gas to be drawn from the suction chamber  38  into the compression chamber  31  through the suction port  36  and the suction valve  33 . Then the refrigerant gas is compressed in the compression chamber  31  to a predetermined pressure and discharged into the discharge chamber  39  through the discharge port  37  and the discharge valve  34 . Most of the high-pressure refrigerant gas discharged into the discharge chamber  39  is delivered to the external refrigeration circuit (not shown), while a part of the high-pressure refrigerant gas in the discharge chamber  39  is drawn into the crank chamber  16  through the supply passage  42  for varying the inclination of the swash plate  23 . 
     The opening of the displacement control valve  35  provided in the supply passage  42  is adjusted to control the relation between the flow rate of refrigerant gas introduced from the discharge chamber  39  into the crank chamber  16  and the flow rate of refrigerant gas flowing out from the crank chamber  16  into the suction chamber  38  through the first and second bleed passages  48 ,  58 . A crank chamber pressure PC in the crank chamber  16  is determined by this relation of the refrigerant gas. As the opening of the displacement control valve  35  is adjusted to change the crank chamber pressure PC in the crank chamber  16 , the pressure differential between the crank chamber  16  and the compression chambers  31  through the piston  29  varies thereby to change the inclination angle of the swash plate  23 . Thus; the stroke length of the piston  29  is changed and the displacement of the compressor  10  is changed accordingly. 
     When the cooling load is large due to high temperature in the vehicle compartment, a suction pressure PS in the suction chamber  38  is high and there is substantially no pressure differential between the pressures in the compression chambers  31  and the crank chamber pressure PC in the crank chamber  16  through the piston  29 . (or PS≈PC). In this case, the displacement control valve  35  is controlled to be closed so that the supply passage  42  prevents high-pressure refrigerant gas in the discharge chamber  39  from flowing into the crank chamber  16 . Since the crank chamber pressure PC in the crank chamber  16  is substantially the same as the suction pressure PS, refrigerant gas does not flow from the crank chamber  16  through the first and second bleed passages  48 ,  58  into the suction chamber  38 . Thus, as Indicated by the solid line in  FIG. 1 , the swash plate  23  is moved to its maximum inclination angle position to increase the stroke of the piston  29 , thereby to increase the displacement of the compressor  10 . During the maximum displacement operation of the compressor  10 , the refrigerant gas does not circulate through the supply passage  42 , the first and second bleed passages  48 ,  58 , with the result that the compressor  10  is efficiently operated. 
     When the cooling load is decreased due to a decrease of the temperature in the vehicle compartment, the suction pressure PS in the suction chamber  38  is also decreased. In this case, the displacement control valve  35  is controlled to be opened in accordance with the decrease in the suction pressure PS. Accordingly high-pressure refrigerant gas in the discharge chamber  39  is introduced into the crank chamber  16  through the supply passage  42 . As a result, the crank chamber pressure PC in the crank chamber  16  is increased and the pressure differential between the crank chamber  16  and the compression chambers  31  through the piston  29  increases. The inclination angle of the swash plate  23  becomes small in accordance with the increase of the pressure differential, thereby decreasing the displacement of the compressor  10 . 
     During the variable displacement operation of the compressor  10 , in particular, when the rotational speed of the rotary shaft  17  is low, the centrifugal force generated by the rotation of the rotary shaft  17  is small. In this case, the valve body  54  of the valve  50  provided in the first bleed passage  48  is positioned so as to open the valve hole  53 , as shown in  FIG. 2 . The second bleed passage  58  has the throttle hole  59  for constant communication between the crank chamber  16  and the suction chamber  38 . That is, the first bleed passage  48  provided with the valve  50  and the second bleed passage  58  provided with the throttle hole  59  are opened. Thus, the refrigerant gas is drawn from the crank chamber  16  into the suction chamber  38  rapidly and, therefore, the displacement of the compressor  10  is controlled appropriately in accordance with a change in the cooling load. 
     As the rotational speed of the rotary shaft  17  is increased, the centrifugal force generated by the rotation of the rotary shaft  17  is increased. That is, the centrifugal force acting on the counterweight  55  of the valve  50  is also increased. As shown in  FIG. 3 , the valve body  54  is moved toward the axis of the rotary shaft  17  by the centrifugal force acting against the urging force of the coil spring  57  so as to be brought into contact with the seating surface  51 , thereby to close the valve hole  53 . The first bleed passage  48  provided with the valve  50  is closed and only the second bleed passage  58  provided with the throttle hole  59  is opened. Thus, the flow rate of refrigerant gas drawn from the crank chamber  16  into the suction chamber  38  is decreased. The decrease of the flow rate of refrigerant gas circulating within the compressor means the increase the flow rate of refrigerant gas in the external refrigeration circuit, thus reducing the power loss of the compressor  10 . 
     As the cooling load is decreased to be nearly zero due to further decrease of the temperature in the vehicle compartment, the suction pressure PS in the suction chamber  38  is further decreased accordingly and the displacement control valve  35  becomes fully opened. In this case, a large amount of high-pressure refrigerant gas is introduced from the discharge chamber  39  into the crank chamber  16  through the supply passage  42 , thereby to increase the crank chamber pressure PC in the crank chamber  16 . As a result, the pressure differential between the crank chamber  16  and the compression chamber  31  through the piston  29  is increased. As indicated by the double-dashed line in  FIG. 1 , the swash plate  23  is moved to its minimum inclination angle position to decrease the stroke length of the piston  29 , thereby to change the displacement of the compressor  10  to the minimum. During the minimum displacement operation (or OFF operation), the displacement of the compressor  10  is not zero. When the compressor  10  is operated at a high rotational speed during the minimum displacement operation, the flow rate of refrigerant gas circulating within the compressor  10  is further decreased thereby to decrease the level of the minimum displacement. Thus, the power loss during the minimum displacement operation is reduced. 
     The following will describe the recovery of the compressor  10  from the minimum displacement state. The increase of the displacement of the compressor  10  from the OFF operation is dependent on the rate of refrigerant gas from the crank chamber  16  into the suction chamber  38 . When the rotary shaft  17  is rotated at a low speed, the first bleed passage  48  provided with the valve  50  and the second bleed passage  58  provided with the throttle hole  59  are both opened. Therefore, the refrigerant gas is drawn from the crank chamber  16  into the suction chamber  38  rapidly and the crank chamber pressure PC in the crank chamber  16  is decreased accordingly rapidly. Thus, the recovery of the compressor  10  from the minimum displacement state is improved. 
     When the rotary shaft  17  is rotated at a high speed, the first bleed passage  48  provided with the valve  50  is closed and only the second bleed passage  58  provided with the throttle hole  59  is opened. Accordingly, the flow rate of refrigerant gas drawn from the crank chamber  16  into the suction chamber  38  is decreased. During the high-speed operation of the compressor  10 , however, an inertial force acting on the piston  29 ′ and the swash plate  23  is increased so as to principally affect the motion of the piston  29  and the swash plate  23  to change in the direction that increases the compression displacement. Thus, the desired compression displacement is achieved rapidly from the minimum displacement state despite the decrease of the flow rate of refrigerant gas drawn from the crank chamber  16  into the suction chamber  38 . 
     The swash plate type variable displacement compressors  10  according to the first preferred embodiment of the present invention offers the following advantageous effects. 
     (1) When the rotational speed of the rotary shaft  17  is low and, therefore, the centrifugal force generated by the rotation of the rotary shaft  17  is small, the valve body  54  of the valve  50  provided in the first bleed passage  48  is positioned so as to open the valve hole  53 . The second bleed passage  58  has the throttle hole  59  providing constant communication between the crank chamber  16  and the suction chamber  38 . That is, the first bleed passage  48  provided with the valve  50  and the second bleed passage  58  provided with the throttle hole  59  are both opened. Thus, the refrigerant gas is drawn from the crank chamber  16  into the suction chamber  38  rapidly and the crank chamber pressure PC is decreased accordingly rapidly, thereby improving the recovery of the compressor  10  from the minimum displacement state.
 
(2) During the variable displacement operation of the compressor  10 , the centrifugal force generated by the rotation of the rotary shaft  17  and acting on the counterweight  55  of the valve  50  is increased with an increase in the rotational speed of the rotary shaft  17 . The valve body  54  is moved toward the axis of the rotary shaft  17  by the centrifugal force acting against the urging force of the coil spring  57  so as to be in contact with the seating surface  51 , with the result that the valve hole  53  is closed. Since the first bleed passage  48  provided with the valve  50  is closed and only the second bleed passage  58  provided with the throttle hole  59  is opened, the flow rate of the refrigerant gas drawn from the crank chamber  16  into the suction chamber  38  is decreased. Such decreased flow rate of refrigerant gas within the compressor  10  contributes to increasing the flow rate of refrigerant gas in the external refrigeration circuit, thereby to reduce the power loss and improve the operating efficiency of the compressor  10 .
 
(3) During the operation of the compressor  10  at a high rotational speed, the first bleed passage  48  provided with the valve  50  is closed and only the second bleed passage  58  provided with the throttle hole  59  is opened. Thus, the refrigerant gas drawn from the crank chamber  16  into the suction chamber  38  is decreased. In particular, during the OFF operation, the minimum displacement of the compressor is further decreased thereby to reduce the power loss. When the compressor  10  is operated at a high rotational speed, the inertial force acting on the piston  29  and the swash plate  23  is increased so as to affect the increase of the compression displacement. Thus, the decrease of the performance of the compressor  10  to recover the displacement of the compressor  10  from the minimum displacement state is prevented despite the decrease of the flow rate of refrigerant gas drawn from the crank chamber  16  into the suction chamber  38 .
 
(4) The valve  50  provided in the first bleed passage  48  formed in the rotary shaft  17  is operable to be opened or closed by utilizing the centrifugal force generated by the rotation of the rotary shaft  17 . Further, the throttle hole  59  is easily provided in the second bleed passage  58  formed in the cylinder block  12  separately from the first bleed passage  48  to ensure a constant flow of refrigerant gas therethrough.
 
(5) The passage hole  45  formed in the rotary shaft  17  along its axis is opened at one end thereof to the crank chamber  16  and has at the other end thereof the valve  50 , which allows the valve  50  to be disposed effectively in the cylinder block  12 .
 
(6) The valve  550  includes the valve body  54 , the coil spring  57  and the counterweight  55 . The coil spring  57  urges the valve body  54  toward its opened position. The centrifugal force generated by the rotation of the rotary shaft  17  and acting on the counterweight  55  causes the valve body  54  to be moved against the urging force of the coil spring  57  and to close the valve hole  53 . When the rotational speed of the rotary shaft  17  is increased, the counterweight  55  is moved away from the axis of the rotary shaft  17  by the increasing centrifugal force acting on the counterweight  55  against the urging force of the coil spring  57 . As a result, the valve body  54  is moved toward its closed position. On the other hand, when the rotational speed of the rotary shaft  17  is decreased, the urging force of the coil spring  57  becomes greater than the centrifugal force acting on the counterweight  55 , so that the valve body  54  is moved to and held at its opened position. The valve  50  is simple in structure as described above and the first bleed-passage  48  is opened or closed reliably in accordance with the rotational speed of the rotary shaft  17 .
 
     The following will describe a swash plate type variable displacement compressor according to the second preferred embodiment of the present invention with reference to  FIG. 6 . The compressor of the second preferred embodiment differs from that of the first preferred embodiment in that the rotary shaft  17  is equipped with the function of the second bleed passage  58  of the first embodiment. That is, the second bleed passage of the second embodiment shares a part of the first bleed passage. The rest of the structure of the compressor according to the second preferred embodiment is substantially the same as that of the first preferred embodiment. For the sake of convenience of explanation, therefore, like or same parts or elements will be referred to by the same reference numerals as those that have been used in the first preferred embodiment, and the description thereof will be omitted. 
     As shown in  FIG. 6 , the rotary shaft  17  of the compressor  10  according to the second preferred embodiment has a throttle hole  70  radially bored therethrough at a position adjacent to the valve  50  for providing fluid communication between the passage hole  45  in the rotary shaft  17  and the recess  44  in the cylinder-block  12 . The throttle hole  70  functions as a fixed throttle. The diameter D 1  of the throttle hole  70  is formed smaller than the diameter D 2  of the communication hole  49  (or D 1 &lt;D 2 ). As the previously described first preferred embodiment, the relation among the cross-sectional area AA of the communication hole  49 , the cross-sectional area AB of the throttle hole  59  and the total cross-sectional area AS 1  during the operation at a low rotational speed is expressed by AS 1 =AA+AB, while the total cross-sectional area AS 2  during the operation at a high rotational speed is expressed by AS 2 =AB. Meanwhile, in the second preferred embodiment, the diameter D 1  of the throttle hole  70  and the diameter D 2  of the communication hole  49  are set such that D 1 =AB=AS 2  and D 2 =AA+AB=AS 1  respectively. 
     When the rotational speed of the rotary shaft  17  is low, the centrifugal force generated by the rotation of the rotary shaft  17  is small, so that the valve body  54  of the valve  50  provided in the first bleed passage  48  is positioned so-as to open the valve hole  53 . The rotary shaft  17  has the throttle hole  70  for providing constant communication between the crank chamber  16  and the suction chamber  38 . In this case, the flow rate of the refrigerant gas drawn from the crank chamber  16  into the suction chamber  38  depends on the diameter D 2  of the communication hole  49 . As described above, the diameter D 1  of the throttle hole  70  is smaller than the diameter D 2  of the communication hole  49  (or D 1 &lt;D 2 ). Therefore, the refrigerant gas in the crank chamber  16  is drawn rapidly into the suction chamber  38  trough the recess  44  and the crank chamber pressure PC in the crank chamber  16  is decreased accordingly rapidly, with the result that the performance of the compressor  10  to recover the displacement of the compressor  10  from the minimum displacement state is improved. 
     During high-speed operation of the compressor  10 , the centrifugal force generated by the rotation of the rotary shaft  17  acting on the counterweight  55  is increased and the valve body  54  is moved-toward the axis of the rotary shaft  17  against the urging force of the coil spring  57  until it is brought into contact with the seating surface  51 , thereby to close the valve hole  53 . Thus, the first bleed passage  48  provided with the valve  50  is closed and only the throttle hole  70  is opened. Accordingly, the flow rate of refrigerant gas drawn from the crank chamber  16  into the suction chamber  38 , which depends on the diameter D 1  of the throttle hole  70 , is decreased. However, during the operation of the compressor  10  at a high rotational speed, the inertial force acting on the piston  29  and the swash plate  23  is increased so as to principally affect the motion of the piston  29  and the swash plate  23  to change in the direction that increases the compression displacement. Thus, the desired compression displacement is achieved rapidly from the minimum displacement state despite the decrease of the flow rate of refrigerant gas drawn from the crank chamber  16  into the suction chamber  38 . In addition, the decreased flow rate of refrigerant gas circulating within the compressor  10  during its variable displacement operation means the increase of the flow rate of refrigerant gas in the external refrigeration circuit, thereby to reduce the power loss. During the OFF operation, the level of the minimum displacement of the compressor is further decreased. Thus, the power loss during the minimum displacement operation is also reduced.  FIG. 6  shows the valve  50  in its closed position. 
     In the compressor  10  of the second preferred embodiment, the first bleed passage  48  and the throttle hole  70  are both provided in the rotary shaft  17 . This structure contributes to reduction in production time and cost as compared with a structure wherein the first bleed passage  48  and the throttle hole  70  are provided separately. 
     The following will describe a swash plate type variable displacement compressor according to the third preferred embodiment of the present invention with reference to  FIGS. 7 and 8 . The compressor  10  of the third preferred embodiment differs from that of the second preferred embodiment in that a groove  80  corresponding to the throttle hole  70  of the second preferred embodiment is formed in the valve hole  53  of the valve  50 . The rest of the structure of the compressor  10  according to the third preferred embodiment is substantially the same as that of the second preferred embodiment. For the sake of convenience of explanation, therefore, like or same parts or elements will be referred to by the same reference-numerals as those which have been used in the first and second preferred embodiments, and the description thereof will be omitted. 
     As shown in  FIG. 7 , in the compressor  10  according to the third preferred embodiment, the groove  80  having a certain depth is formed at the opening of the valve hole  53  of the valve  50  on the side of the seating surface  51 . When the valve body  54  is in contact with the seating surface  51  by the centrifugal force, the valve body  54  and the groove  80  cooperate to form a groove slit  81 . The passage hole  45  is communicated with the recess  44  through the groove slit  81  which functions as a throttle. The cross-sectional area of the groove slit  81  when the valve body  54  is in contact with the seating surface  51  is set smaller than that (D 2 ) of the communication hole  49  and set substantially the same as that (D 1 ) of the throttle hole  70  of the second preferred embodiment. 
     When the rotational speed of the rotary shaft  17  is low, the centrifugal force generated by the rotation of the rotary shaft  17  is small and the valve body  54  of the valve  50  provided in the first bleed passage  48  is positioned to open the valve hole  53 . In this case, since the groove slit  81  is not formed, the flow rate of the refrigerant gas depends on the diameter D 2  of the communication hole  49 . Refrigerant gas is drawn from the crank chamber  16  into the suction chamber  38  rapidly and the crank chamber pressure PC in the crank chamber  16  is decreased accordingly rapidly. Thus, the recovery of the compressor  10  from the minimum displacement state is improved. 
     As the rotational speed of the rotary shaft  17  is increased, the centrifugal force generated by the rotation of the rotary shaft  17  is increased and the centrifugal force acting on the counterweight  55  of the valve  50  is also increased. The valve body  54  is then moved toward the axis of the rotary shaft  17  by the centrifugal force against the urging force of the coil spring  57  until it is brought into contact with the seating surface  51  thereby to close the valve hole  53 . In this case, only the groove slit  81  whose diameter is smaller than that of the communication hole  49 , is opened, so that the flow rate of the refrigerant gas drawn from the crank chamber  16  into the suction chamber  38  is decreased. However, during the operation of the compressor  10  at a high rotational speed, the inertial force acting on the piston  29  and the swash plate  23  is increased so as to principally affect the motion of the piston  29  and the swash plate  23  to change in the direction that increases the compression displacement. Thus, the desired compression displacement is achieved rapidly from the minimum displacement state despite the decrease of the flow rate of refrigerant gas drawn from the crank chamber  16  into the suction chamber  38 . In addition, the decreased flow rate of refrigerant gas circulating within the compressor  10  during its variable displacement operation means the increase of the flow rate of refrigerant gas in the external refrigeration circuit, thereby reducing the power loss. Further, during the OFF operation, the compression displacement is decreased further than the minimum displacement. Thus, the power loss during the minimum displacement operation is also reduced.  FIG. 7  shows the valve  50  in its closed position. 
     According to the third preferred embodiment, the groove  80  is merely formed at the opening of the valve hole  53  of the valve  50 . This contributes to simplified structure and further reduction in the production time and cost of the compressor  10 . 
     The present invention is not limited to the first through third preferred embodiments, but it may be variously modified within the scope of the invention. For example, the above embodiments may be modified as exemplified below. 
     In the second and third preferred embodiments, the throttle hole  70  or the groove slit  81  is provided at one end of the rotary shaft  17  as a throttle. Alternatively, a fixed throttle may be formed through the plug  60  closing rear end of the passage-hole  45  so that the passage hole  45  and the recess  44  are in constant communication with each other. 
     In the third preferred embodiment, the groove  80  is provided on the side of the seating surface  51  of the valve hole  53 . Alternatively, the groove  80  may be provided on the surface of the valve body  54 . Further, the valve body  54  may have an elongated hole formed therein as a throttle providing fluid communication between the passage hole  45  and the recess  44 . 
     In the compressor  10  according to the first through third preferred embodiments any kind of refrigerant may be used, including preferably fluorocarbon gas or carbon dioxide. Although the compressor  10  according to the foregoing embodiments have been described as a compressor for compressing refrigerant gas, the present invention does not limit the refrigerant only to gaseous refrigerant. 
     Therefore, the present examples and embodiments are to be considered as illustrative and not restrictive, and the invention is not to be limited to the details given herein but may be modified within the scope of the appended claims.