Patent Publication Number: US-2011067434-A1

Title: Hermetic type compressor and refrigeration cycle apparatus

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
     This is a Continuation Application of PCT Application No. PCT/JP2009/059719, filed May 27, 2009, which was published under PCT Article 21(2) in Japanese. 
     This application is based upon and claims the benefit of priority from prior Japanese Patent Application No. 2008-139682, filed May 28, 2008, the entire contents of which are incorporated herein by reference. 
    
    
     BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     The present invention relates to a hermetic type compressor whose bearing structure is improved and a refrigeration cycle apparatus that includes the hermetic type compressor to form a refrigeration cycle. 
     2. Description of the Related Art 
     Frequently, a rotary hermetic type compressor is used in the refrigeration cycle apparatus. In the rotary hermetic type compressor, a motor portion and a compression mechanism portion that is coupled to the motor portion via a rotating shaft (crankshaft) interposed therebetween are accommodated in a closed vessel. In this kind of compressor, a refrigerant is introduced into a compression chamber formed in a cylinder and compressed, whereby a compressive load acts on the rotating shaft. 
     Accordingly, the rotating shaft generates a flexural deformation, and a rotating shaft portion in a flexure direction and a bearing that journals the rotating shaft come into partial contact with each other unless some sort of measure is taken. Smooth rotation of the rotating shaft is spoiled, which leads to damage of the rotating shaft and bearing. Therefore, for example, Jpn. Pat. Appln. KOKAI Publication No. 2004-124834 proposes a bearing structure in order to properly bear the flexural deformation of the rotating shaft. 
     In the technique proposed in Jpn. Pat. Appln. KOKAI Publication No. 2004-124834, according to the flexural deformation of the rotating shaft due to the compressive load in the cylinder, a groove is provided on a cylinder side of a main bearing to allow the flexural deformation of the main bearing, and a center of an internal diameter on the motor side of the main bearing is eccentrically disposed by a predetermined amount with respect to a center of an internal diameter on the cylinder side in a direction of the flexural deformation of the rotating shaft. 
     BRIEF SUMMARY OF THE INVENTION 
     However, in the groove on the cylinder side of the main bearing in the technique, a diameter of an inner circumferential surface of the main bearing is kept constant over a total length, and a thickness between the inner circumferential surface of the groove and an inner circumference of a bearing hole is also kept constant over the total length. 
     Accordingly, although the partially strong contact between the rotating shaft and the bearing can be avoided in a certain range of the groove by the flexure of the bearing, rigidity of the bearing increases rapidly at an end of the groove, and a contact load is concentrated on the end of the groove. Therefore, local abrasion is generated, and bearing reliability cannot sufficiently be enhanced. 
     In view of the foregoing, an object of the invention is to provide a hermetic type compressor in which, according to the flexural deformation of the rotating shaft due to the compressive load in the cylinder, uneven contact with the rotating shaft is prevented in at least one of the main bearing and sub-bearing, thereby achieving the enhancement of the reliability and a longer operation life. 
     Another object of the invention is to provide a refrigeration cycle apparatus that includes the hermetic type compressor to form the refrigeration cycle, thereby improving refrigeration efficiency. 
     A hermetic type compressor of the present invention comprises, a motor portion and a compression mechanism portion that are coupled to the motor portion with a rotating shaft interposed therebetween are accommodated in a closed vessel, the compression mechanism portion comprises a cylinder that comprises an internal diameter hole; and a main bearing and a sub-bearing in which a bearing hole that journals the rotating shaft is provided and the internal diameter hole of the cylinder is closed to form a compression chamber in the compression mechanism portion, at least one of the main bearing and the sub-bearing have a circular groove that is opened toward the compression chamber side, an inner circumferential surface of the circular groove is tapered such that a diameter increases gradually from the compression chamber side toward an opposite side of the compression chamber side, and a depth of the circular groove is set to at least 40% of a diameter of the bearing hole. 
     A refrigeration cycle apparatus of the present invention comprises, the hermetic type compressor; a condenser; an expansion device; and an evaporator. 
    
    
     
       BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWING 
         FIG. 1  is a refrigeration cycle configuration diagram of a refrigeration cycle apparatus according to a first embodiment of the invention and a longitudinal sectional view of a hermetic type compressor. 
         FIG. 2  is an enlarged longitudinal sectional view of a compression mechanism portion of the hermetic type compressor. 
         FIG. 3  is an enlarged longitudinal sectional view of a compression mechanism portion of a hermetic type compressor according to a second embodiment of the invention. 
         FIG. 4  is a longitudinal sectional view of a main part of a hermetic type compressor according to a third embodiment of the invention. 
         FIG. 5  is a longitudinal sectional view of a main part of a hermetic type compressor according to a fourth embodiment of the invention. 
         FIG. 6  is a characteristic diagram of a circular groove depth effect in the invention. 
         FIG. 7  is a characteristic diagram of a circular groove minimum wall thickness effect in the invention. 
         FIG. 8  is a characteristic diagram of a circular groove minimum seal width effect in the invention. 
         FIG. 9  is a characteristic diagram of a circular groove slope effect in the invention. 
         FIG. 10  is a longitudinal sectional view of a hermetic type compressor according to a modification of the third embodiment of the invention. 
         FIG. 11  is a plan view of a discharge valve mechanism mounted on an intermediate partition plate of the modification. 
         FIG. 12  is a sectional view of an intermediate partition plate and a discharge valve mechanism of a first example of the modification. 
         FIG. 13  is a sectional view of an intermediate partition plate and a discharge valve mechanism of a first example of the modification. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     Embodiments of the invention will be described below with reference to the drawings.  FIG. 1  is a longitudinal sectional view of a hermetic type compressor  1  and a refrigeration cycle configuration diagram of a refrigeration cycle apparatus R. 
     In  FIG. 1 , the numeral  1  designates a hermetic type rotary compressor (hereinafter simply referred to as “compressor”), and the compressor  1  is described later. A refrigerant pipe P is connected to an upper end portion of the compressor  1 . A condenser  2 , an expansion valve (expansion device)  3 , an evaporator  4 , and an accumulator  5  are sequentially provided in the refrigerant pipe P. The refrigerant pipe P is also connected to a side portion of the compressor  1  from the accumulator  5 , thereby forming a refrigeration cycle of the refrigeration cycle apparatus R. 
     The compressor  1  will be described next. The compressor  1  comprises a closed vessel  10 . A motor portion  11  is accommodated on an upper portion side in the closed vessel  10 , and a compression mechanism portion  12  is accommodated on a lower portion side. The motor portion  11  and the compression mechanism portion  12  are coupled to each other via a rotating shaft  13  interposed therebetween. 
     A discharge portion la formed by a hole portion is provided in an upper surface portion of the closed vessel  10 , and the refrigerant pipe P communicated with the condenser  2  is connected to the discharge portion  1   a.  A suction portion  1   b  formed by a hole portion is provided in a circumferential wall in a lower portion of the closed vessel  10 , and the refrigerant pipe P communicated with the accumulator  5  is connected to the suction portion  1   b.    
     The motor portion  11  comprises a rotor  15  and a stator  16 . The rotor  15  is fitted in and fixed to a rotating shaft  13 . An inner circumferential surface of the stator  16  faces an outer circumferential surface of the rotor  15  with a narrow gap, and the stator  16  is fitted in and fixed to an inner circumferential wall of the closed vessel  10 . 
     The compression mechanism portion  12  will be described below with reference to  FIGS. 1 and 2 .  FIG. 2  is an enlarged longitudinal sectional view illustrating the compression mechanism portion  12 . 
     The compression mechanism portion  12  comprises a cylinder  20 , a main bearing  21 , and a sub-bearing  22 . The cylinder  20  is fitted in and fixed to the inner circumferential wall of the closed vessel  10 , and an internal diameter hole S is made in an axial center of the cylinder  20 . The main bearing  21  is mounted on an upper surface of the cylinder  20 . The sub-bearing  22  is mounted on a lower surface of the cylinder  20 . The cylinder internal diameter hole S is closed by the main bearing  21  and the sub-bearing  22  to form a space, and the space constitutes a compression chamber (hereinafter referred to as “cylinder chamber”) S. 
     In the rotating shaft  13 , a portion between the motor portion  11  and the upper surface of the cylinder  20  is inserted and journaled in a bearing hole N made in the main bearing  21 . In the rotating shaft  13 , a portion between the lower surface and a lower end of the cylinder  20  is inserted and journaled in a bearing hole N made in the sub-bearing  22 . 
     The main bearing  21  and the sub-bearing  22  comprise flanges  21   a  and  22   a  and cylindrical pivot portions  21   b  and  22   b,  respectively. The flanges  21   a  and  22   a  close the cylinder internal diameter hole S. The cylindrical pivot portions  21   b  and  22   b  are projected along axial center portions of the flanges  21   a  and  22   a  while integrated with the flanges  21   a  and  22   a,  and the cylindrical pivot portions  21   b  and  22   b  comprise the bearing holes N in which the rotating shaft  13  is journaled. Circular grooves K are provided in the main bearing  21  and the sub-bearing  22 , and the circular grooves K are described later. 
     An eccentric portion  13   a  whose center axis is eccentrically located by an eccentric amount e is integrally provided in the rotating shaft  13 . A rolling piston (hereinafter simply referred to as “roller”)  25  is fitted in a circumferential surface of the eccentric portion  13   a.  The roller  25  and the eccentric portion  13   a  are accommodated in the cylinder chamber S, and part of an outer circumferential wall of the roller  25  is designed to come into linear contact with a circumferential wall of the cylinder chamber S along an axis direction. Accordingly, a position at which the outer circumferential wall of the roller  25  comes into contact with the circumferential wall of the cylinder chamber S moves gradually in a circumferential direction by the rotation of the rotating shaft  13 . 
     A blade chamber (not illustrated) is provided in the cylinder  20 . A compression spring is accommodated in the blade chamber, and a blade that receives a back pressure from the compression spring is movably accommodated. A leading end edge of the blade is in contact with part of the outer circumferential wall of the roller  25  along the axis direction, and therefore the blade always divides the cylinder chamber S into two. 
     A discharge hole  26  is made in the main bearing  21 . A position at which the discharge hole  26  is made is located near a region where the blade comes into contact with the roller  25 , and the position constitutes one side portion of the region. A discharge valve mechanism  27  is provided in the discharge hole  26 , and the discharge valve mechanism  27  is covered with a valve cover  28  mounted on the main bearing  21 . A guide hole  28   c  is made in the valve cover  28  so as to be opened into the closed vessel  10 . 
     In the cylinder  20 , the hole portion constituting the suction portion  1   b  is provided across the region where the blade comes into contact with the roller  25  from the discharge hole  26 . The suction portion  1   b  is communicated with the closed vessel  10  while radially piercing the cylinder  20 , and the suction portion  1   b  is connected to the refrigerant pipe P communicated with the accumulator  5 . 
     The circular grooves K, provided in the main bearing  21  and the sub-bearing  22 , will be described in detail. 
     The circular groove K provided in the main bearing  21  and the circular groove K provided in the sub-bearing  22  have the same structure, shape, and dimensions. At this point, only the circular groove K of the main bearing  21  is described. In the circular groove K of the sub-bearing  22 , the same component is designated by the same numeral, and the description is not repeated. 
     The circular groove K is provided from an intersection portion of the flange  21   a  and cylindrical pivot portion  21   b  constituting the main bearing  21  to the cylindrical pivot portion  21   b.  The circular groove K comprises an opening end Kd that is opened to the cylinder chamber S, and the circular groove K is formed deeper from the opening end Kd toward the side of the motor portion  11  that is the opposite side of the cylinder chamber S. 
     The opening end Kd of the circular groove K is concentric with the bearing hole N made in the main bearing  21 , and the opening end Kd is formed into a ring shape having a predetermined width. In the circular groove K, a distance between an outer circumferential surface Km and a circumferential surface of the bearing hole N is kept constant from the opening end Kd in a depth direction, while a distance between an inner circumferential surface Kq and the circumferential surface of the bearing hole N increases gradually. 
     In other words, while the diameter is evenly formed along the axis direction in the outer circumferential surface Km of the circular groove K, the inner circumferential surface Kq is tapered such that the diameter increases gradually along the axis direction. Therefore, the wall thickness from the circumferential surface of the bearing hole N to the inner circumferential surface Kq of the circular groove K becomes minimum (thinnest) at the opening end Kd of the circular groove K and increases gradually from the opening end Kd in the depth direction. 
     On the assumption that the inner circumferential surface Kq of the circular groove K is tapered such that the diameter increases gradually from the opening end Kd that is the side of the cylinder chamber S toward the opposite side of the cylinder chamber S, a depth L of the circular groove K is set to at least 40% of a diameter D of the bearing hole N for the later-described reason, where L is a depth of the circular groove K and D is a diameter (that is also a shaft diameter of the rotating shaft  13 ) of the bearing hole N. 
     In the circular groove K, because the inner circumferential surface Kq is tapered, a wall thickness b that is a distance between the inner circumferential surface Kq and the circumferential surface of the bearing hole N becomes minimum at the opening end Kd facing the cylinder chamber S. For the later-described reason, the wall thickness b between the inner circumferential surface Kq of the circular groove K and the circumferential surface of the bearing hole N is set so as to satisfy a relationship of an equation (1): 
       0.09×diameter  D  of bearing hole N≧minimum wall thickness  b≧ 0.04×diameter  D  of bearing hole  N    (1)
 
     Assuming that e is an eccentric amount of the eccentric portion  13   a  of the rotating shaft  13  and r is an outer circumferential radius of the roller  25 , the outer circumferential radius g of the circular groove K is set so as to satisfy relationships of equations (2) and (3) for the later-described reason: 
       0.5 mm≦[outer circumferential radius  r  (mm) of roller 25−eccentric amount  e  (mm) of eccentric portion 13 a ]−outer circumferential radius  g  (mm) of circular groove  K    (2)
 
       outer circumferential radius  g  (mm) of circular groove  K &gt;diameter  D  (mm) of bearing hole  N/ 2+minimum wall thickness  b  (mm)   (3)
 
     The action of the compressor  1  and freezing action of the refrigeration cycle apparatus R will be described below. 
     When a current is passed through the motor portion  11  constituting the compressor  1 , the rotor  15  is rotated by a rotating magnetic field generated by the stator  16 , thereby rotating the rotating shaft  13  integrated with the rotor  15 . A driving torque acts on the rotating shaft  13  from the motor portion  11 , and the eccentric portion  13   a  provided in the rotating shaft  13  is eccentrically rotated while integrated with the roller  25  in the cylinder chamber S. 
     Therefore, a negative pressure is partially generated in the cylinder chamber S, and the refrigerant is introduced from the accumulator  5  through the refrigerant pipe P. The refrigerant is introduced into the space region partitioned by the circumferential surface of the roller  25 , the circumferential surface of the cylinder chamber S, and the blade, and a volume of the space region is reduced in association with the eccentric rotation of the roller  25 , thereby compressing the refrigerant. 
     When the space region is minimized, the refrigerant is raised to a high temperature while attaining a predetermined high-pressure state. The discharge valve mechanism  27  is opened by the compressed gas refrigerant, the refrigerant is introduced to the closed vessel  10  through a valve cover  28 , and the closed vessel  10  is filled with the gas refrigerant. The high-temperature, high-pressure gas refrigerant with which the closed vessel  10  is filled is discharged from the discharge portion  1   a  to the refrigerant pipe P. 
     The condenser  2  performs heat exchange of the gas refrigerant for outside air or water, and the gas refrigerant is condensed and liquefied into a liquid refrigerant. The liquid refrigerant is introduced to the expansion valve  3  to perform adiabatic expansion, the liquid refrigerant is introduced to the evaporator  4  to perform the heat exchange for air around a region where the evaporator  4  is disposed, and the liquid refrigerant is evaporated. 
     Evaporative latent heat is seized from the surrounding region in association with the evaporation of the refrigerant. That is, the freezing action acts on the surrounding region. The refrigerant evaporated in the evaporator  4  is introduced to the accumulator  5  to perform gas-liquid separation. The refrigerant is sucked into the cylinder chamber S of the compressor  1 , the refrigerant is compressed again to change into the high-temperature, high-pressure gas refrigerant, and the refrigeration cycle is repeated. 
     Thus, a suction stroke in which the refrigerant to which the gas-liquid separation is performed is sucked from the accumulator  5 , a compression stroke in which the sucked refrigerant is compressed, and a discharge stroke in which the compressed refrigerant is discharged are continuously performed in the cylinder chamber S constituting the compression mechanism portion  12 . 
     Particularly, in the compression stroke, the compressive load is applied to the rotating shaft  13  by the compressed high-pressure gas refrigerant, whereby the flexural deformation of the rotating shaft  13  is generated, from a macroscopic point of view. Specifically, the flexural deformation of the rotating shaft  13  is generated in an opposite direction to the compressive load direction during the compression action. 
     However, because the main bearing  21  and the sub-bearing  22  comprise the circular grooves K set to the above-described conditions, the uneven contact of the rotating shaft  13  with the main bearing  21  and sub-bearing  22  is not generated, irrespective of the flexural deformation of the rotating shaft  13 , and the smooth rotation is secured. 
     More specifically, the bearing hole N that is the inner surface of the main bearing  21  is deformed so as to follow the rotating shaft  13  in which the flexural deformation is generated by receiving the load, and an area where the evenness of the gap between the rotating shaft  13  and the main bearing  21  is retained is expanded. Accordingly, the ability to form an oil film of lubrication oil between the rotating shaft  13  and the main bearing  21  is improved, and the oil film is securely formed even if the rotating shaft  13  is rotated at low speed. 
     There are circumstances in which the formation of the oil film can hardly be maintained, such conditions being when the number of rotations of the rotating shaft  13  is decreased, viscosity of the lubrication oil is reduced, or the compressive load is increased. That is, the contact between the rotating shaft  13  and the main bearing  21  makes a transition to a mixed lubrication state in which not only the rotating shaft  13  and the main bearing  21  come into contact with each other while the oil film is interposed therebetween, but also metallic materials come into solid-state contact with each other due to the surface roughness of the rotating shaft  13  and main bearing  21  to support the load. 
     Even if the solid-state contact cannot be avoided, the surface of the bearing hole N of the main bearing  21  is continuously deformed to prevent the generation of a locally high contact force. The generation of seizing or local bearing abrasion can be prevented to provide the high-reliability main bearing  21 . Because the sub-bearing  22  comprises the circular groove K having completely the same structure, a similar effect can be obtained in the sub-bearing  22 . 
     The circular groove K of the embodiment will be described in comparison with a flexible-structure groove described in Jpn. Pat. Appln. KOKAI Publication No. 2004-124834. From the viewpoint of the formation of the oil film, desirably a gap is evenly formed along the axis line direction between the main bearing  21  that journals the rotating shaft  13  and the rotating shaft  13  in which the flexural deformation is generated by receiving the compressive load in the cylinder chamber S. 
     The flexural deformation of the rotating shaft  13  is maximized on the side of the cylinder chamber S in which the compressive load is applied to the rotating shaft  13  and decreases gradually with distance from the side of the cylinder chamber S. As described above, when the circular groove K is formed in the main bearing  21 , rigidity of the internal diameter of the main bearing  21  is low on the side of the cylinder chamber S in which the rotating shaft  13  has the large flexural deformation, and the rigidity increases gradually with distance from the side of the cylinder chamber S. 
     Therefore, the inner surface of the main bearing  21  is deformed by following the deformation of the rotating shaft  13 , and the deformable circular groove K is formed deeper than the flexible-structure groove, so that the circular groove K is greatly deformed in a wide area to follow the rotating shaft  13 . Additionally, the rigidity of the internal diameter of the main bearing  21  increases gradually with distance from the side of the cylinder chamber. S, so that a fluctuation in load applied to the main bearing  21  in the axis direction can be reduced. 
     On the other hand, in the flexible-structure groove, because the wall thickness between the groove inner surface and the circumferential surface of the bearing hole is kept constant over the total length of the groove, the rigidity of the circumferential surface of the bearing hole is kept constant. Therefore, the rigidity is small in the groove portion, the rigidity increases rapidly in the portion in which the groove is terminated, and the fluctuation in load applied to the bearing also increases. Accordingly, the oil film is easily broken in the portion in which the groove is terminated. This cannot be solved even if the groove depth is simply increased. 
     In the embodiment, the circular groove K is provided, and the depth of the groove K and the wall thickness between the groove K and the bearing hole N are increase to enhance the strength. The rigidity of the internal diameter of the main bearing  21  increases with distance from the side of the cylinder chamber S, the oil film is evenly formed in the whole of the main bearing  21 , and the fluid lubrication state can be maintained in the wide operating area. 
     Even if the contact between the rotating shaft  13  and the main bearing  21  makes the transition from the fluid lubrication state to the mixed lubrication state in which the lubrication state including the solid-state contact state is maintained, because the circular groove K is deep and flexible, the solid-state contact is generated in the depth range of the circular groove K in which the elastic deformation can be generated, and the main bearing  21  is elastically deformed to prevent the uneven contact with the rotating shaft  13 . Therefore, seizing and the like are not generated. 
     As described above, there is the setting condition that the inner circumferential surface Kq of the circular groove K is tapered. The setting condition is fixed on the following basis. First, the basis on which the depth L of the circular groove K is set to at least 40% of the diameter D of the bearing hole N will be described, on the assumption that the inner circumferential surface Kq of the circular groove K is tapered such that the diameter increases gradually from the surface facing the cylinder chamber S toward the opposite side of the cylinder chamber S. 
     In the bearing hole N of the main bearing  21 , the portion in which the circumferential surface of the rotating shaft  13  is particularly effectively journaled is a portion from an end portion of the bearing hole N to a length corresponding to the diameter of the bearing hole N. The depth L of the circular groove K is formed equal to or more than 40% of the diameter D of the bearing hole N. 
     Therefore, the inner surface (bearing hole N) of the main bearing  21  is deformed so as to follow the deformation of the rotating shaft  13 , which desirably affects the formation of the oil film between the rotating shaft  13  and the main bearing  21  and the contact of the rotating shaft  13  with the main bearing  21  due to the deformation of the rotating shaft  13 . 
     This can be described with reference to  FIG. 6 .  FIG. 6  is a characteristic diagram illustrating a groove depth effect. In  FIG. 6 , a horizontal axis indicates the depth of the circular groove K, and a vertical axis indicates the oil film thickness of the lubricant oil formed between the rotating shaft  13  and the main bearing  21  and the contact force between the rotating shaft  13  and the main bearing  21 . In  FIG. 6 , a solid-line indicates the contact force and a broken-line indicates the oil film thickness. Where the depth of the circular groove K is indicated by a ratio to the shaft diameter (diameter) D of the rotating shaft  13  (bearing hole N). 
     When the depth of the circular groove K in which the inner circumferential surface Kg is tapered becomes zero, the contact force between the rotating shaft  13  and the main bearing  21  becomes maximum (100), and the oil film is hardly formed. The oil film is formed in the thinnest state at a point where the contact force is weakened to some extent. The contact force decreases rapidly with increasing depth of the circular groove K, and the oil film thickness is thickened in inverse proportion to the decreasing contact force. 
     Particularly, when the depth of the circular groove K exceeds 0.4 (40% of the shaft diameter ratio), a degree to which the contact force decreases changes from the rapidly decreasing state to the gradually decreasing state, the oil film thickness exceeds a necessary oil film thickness ( 1 ), and the oil film thickness is maintained at 1 or more. 
     In the fluid lubrication “state in which only the oil film of the lubrication oil is interposed between the rotating shaft  13  and the main bearing  21 , the oil film thickness is thickened by increasing the groove depth, a tilt of the rotating shaft  13  increases to keep the oil film thickness substantially constant when the depth of the circular groove K becomes at least 40% of the shaft diameter ratio of the rotating shaft  13 . 
     On the other hand, the contact load of the rotating shaft  13  and the main bearing  21  in the mixed lubrication state exhibits a characteristic in which the contact load can be reduced with increasing depth of the circular groove K. However, when the depth of the circular groove K becomes at least 40% of the shaft diameter ratio of the rotating shaft  13 , the tilt of the rotating shaft  13  increases, and a decreasing ratio of the contact load becomes small. 
     In the circular groove K whose inner circumferential surface Kq is tapered, the wall thickness b that is the distance between the inner circumferential surface Kq and the bearing hole N becomes minimum (thinnest) at the opening end Kd facing the cylinder chamber S. 
     The minimum wall thickness b between the inner circumferential surface Kg of the circular groove K and the circumferential surface of the bearing hole N is set so as to satisfy the relationship of the equation (1): 
       0.09×diameter  D  of bearing hole  N ≧minimum wall thickness  b≧ 0.04×diameter  D  of bearing hole  N    (1)
 
     This can be described with reference to  FIG. 7 .  FIG. 7  is a characteristic diagram illustrating a circular groove minimum wall thickness effect. In  FIG. 7 , the horizontal axis indicates the minimum wall thickness (shaft diameter ratio) b of the circular groove K, and the vertical axis indicates the contact force. In  FIG. 7 , the solid-line change indicates the contact force, and a maximum allowable contact force is set to 0.5. 
     When the minimum wall thickness b of the circular groove K decreases excessively, a lack of rigidity is generated in the main bearing  21 , and the deformation becomes large. At this point, even if the oil film thickness can be secured in the fluid lubrication state, the contact load increases in the mixed lubrication state. 
     On the other hand, when the minimum wall thickness b of the circular groove K increases excessively, the rigidity increases excessively to hardly generate the deformation, and the contact load also increases in the mixed lubrication state. Therefore, the proper value of the minimum wall thickness to the contact load is set as illustrated in  FIG. 7  and the equation (1). 
     Assuming that e is the eccentric amount of the eccentric portion  13   a  that is provided integral with the rotating shaft  13  and r is the outer circumferential radius of the roller  25 , the outer circumferential radius g of the circular groove K is set so as to satisfy the relationships of the equations (2) and (3): 
       0.5 mm≦[outer circumferential radius  r  (mm) of roller 25−eccentric amount  e  (mm) of eccentric portion 13 a ]−outer circumferential radius  g  (mm) of circular groove  K    (2)
 
       outer circumferential radius  g  (mm) of circular groove  K &gt;diameter  D  (mm) of bearing roller  N/ 2+minimum wall thickness  b  (mm)   (3)
 
     When the opening end Kd of the circular groove K is communicated with the cylinder chamber S, the refrigerant introduced to the cylinder chamber S remains partially in the circular groove K, and the circular groove K becomes a dead volume. Therefore, in order to prevent the dead volume of the circular groove K, a minimum seal width is formed to exert a seal function between an external diameter of the roller  25  and an external diameter of circular groove K. 
     Particularly, the equation (2) can be described with reference to  FIG. 8 .  FIG. 8  illustrates a minimum seal width effect. In  FIG. 8 , the horizontal axis indicates a minimum seal width (mm), and the vertical axis indicates a performance ratio. 
     The performance ratio is 0.2 when the minimum seal width becomes 0, and the performance ratio does not change even if the minimum seal width increases to about 0.3 mm. The performance ratio increases when the minimum seal width exceeds about 0.3 mm, and the performance ratio increases rapidly when the minimum seal width exceeds 0.4 mm. 
     The performance ratio becomes a peak when the minimum seal width is about 0.5 mm, and the performance ratio is substantially kept constant even if the minimum seal width increases from about 0.5 mm. 
     In the equation (2), [outer circumferential radius r (mm) of roller  25 −eccentric amount e (mm) of eccentric portion  13   a ]−outer circumferential radius g (mm) of circular groove K is the minimum seal width. As can be seen from  FIG. 8 , the minimum seal width of 0.5 mm or more is required. 
     As described above, the inner circumferential surface Kq of the circular groove K is tapered, and setting of a slope angle θ becomes one of necessary conditions. That is, the contact force between the rotating shaft  13  and the main bearing  21  varies depending on the slope angle θ. The circular groove K is formed such that the slope of the inner circumferential surface Kq increases (the slope angle θ decreases) as much as possible, thereby exerting the large contact load reducing effect. 
       FIG. 9  is a characteristic diagram illustrating a groove slope effect. In  FIG. 9 , the horizontal axis indicates the slope of the inner circumferential surface Kg of the circular groove K, and the vertical axis indicates the contact force between the rotating shaft  13  and the main bearing  21 . 
     The contact force is maximized (1 or more) when the slope of the circular groove K is close to zero (0). With increasing groove slope, the contact force decreases, and therefore the oil film thickness increases as described above. 
     Further, as illustrated in  FIG. 2 , there is another setting condition that the main bearing  21  comprises the flange  21   a  whose wall thickness H is set to the depth L of the circular groove K or less. 
     Therefore, the rigidity of the coupling portion between the cylindrical pivot portion  21   b  and the flange  21   a  that supports the whole of the main bearing  21  is reduced to deform the whole of the main bearing  21 , whereby a property of following the rotating shaft  13  is enhanced to improve the effect of the circular groove K. 
       FIG. 3  is an enlarged longitudinal sectional view of a compression mechanism portion  12  according to a second embodiment of the invention. Because a basic configuration of a compression mechanism portion  12  is identical to that of  FIG. 2 , the same component is designated by the same numeral (only main part), and the description of the same component is not repeated. In the second embodiment, a diameter D 1  of a portion (bearing hole Na) that is journaled in a main bearing  21  of a rotating shaft  13  differs from a diameter D 2  of a portion (bearing hole Nb) that is journaled in a sub-bearing  22 . Actually, the diameter D 1  of the portion journaled in the main bearing  21  of the rotating shaft  13  is formed larger than the diameter D 2  of the portion journaled in the sub-bearing  22  (D 1 &gt;D 2 ). 
     Because the diameter D 1  is formed larger than the diameter D 2 , it is necessary to secure a seal width of a circular groove K with respect to a cylinder chamber S in an end face of a roller  25 . Therefore, an inner circumferential surface Kq of the circular groove K is hardly tapered, and a groove Ka having an even width in the depth direction is provided. 
     That is, the tapered inner circumferential surface Kq of the circular groove K is provided only in a rotating shaft portion that is journaled in the sub-bearing  22  having a small diameter, and the seal width of the end face of the roller  25  is secured with respect to the cylinder chamber S. 
     Because the length in the axis direction of the cylindrical pivot portion  22   b  is shorter than that of the main bearing  21 , the flexural deformation becomes large, and the load also becomes large. Therefore, the circular groove K whose inner circumferential surface Kq is tapered is extremely advantageously provided. 
     In the circular groove K whose inner circumferential surface Kq is tapered, the dimensions and configuration are similar to those of the first embodiment, and the effect similar to that of the first embodiment is obtained. However, the overlapping description is not repeated. 
       FIG. 4  is a longitudinal sectional view illustrating a hermetic type compressor  1 A according to a third embodiment of the invention with part of the hermetic type compressor  1 A omitted. 
     Basically, the configuration in which a motor portion  11  and a compression mechanism portion  12 A that is coupled to the motor portion  11  with a rotating shaft  13  interposed therebetween are accommodated in a closed vessel  10  is similar to that of the first embodiment. 
     The compression mechanism portion  12 A is a two-cylinder type compressor  1 A that comprises two cylinders  20 A and  20 B that are provided above and below an intermediate partition plate  30 . Each of the cylinders  20 A and  20 B comprises an internal diameter hole Sa. The internal diameter hole Sa of the cylinder  20 A on the upper side is closed by a main bearing  21  and the intermediate partition plate  30  to form the first cylinder chamber Sa. 
     The internal diameter hole Sb of the cylinder  20 B on the lower side is closed by a sub-bearing  22  and the intermediate partition plate  30 . Eccentric portions  13   a  and  13   b  and a roller  25  are accommodated in the first cylinder chamber Sa and the second cylinder chamber Sb, respectively. The eccentric portions  13   a  and  13   b  are provided while integrated with the rotating shaft  13 , and the eccentric portions  13   a  and  13   b  have a phase difference of 180°. The roller  25  is fitted in the eccentric portions  13   a  and  13   b.    
     A diameter of a portion journaled in the main bearing  21  of the rotating shaft  13  is equal to a diameter of a portion journaled in the sub-bearing  22 . In other words, diameters of bearing holes N made in the main bearing  21  and sub-bearing  22  are equal to each other. 
     Circular grooves K opened to the cylinder chambers Sa and Sb are provided in the main bearing  21  and the sub-bearing  22 . An inner peripheral surface of the circular groove K is tapered such that a diameter of the inner peripheral surface increases gradually from the surface facing each of the cylinder chambers Sa and Sb toward the opposite side of the cylinder chamber. The depth of the circular groove K is set to at least 40% of the diameter of the bearing hole. 
     Because all the above-described setting conditions are included in the hermetic type compressor  1 A of the third embodiments, similar effects are obtained in both the main bearing  21  and the sub-bearing  22 . 
       FIG. 5  is a longitudinal sectional view illustrating a hermetic type compressor  1 B according to a fourth embodiment of the invention with part of the hermetic type compressor  1 A omitted. 
     Basically, the hermetic type compressor  1 B of the fourth embodiment comprises a compression mechanism portion  12 B having a configuration similar to that of the two-cylinder type compression mechanism portion  12 A of the third embodiment (see  FIG. 4 ). 
     In the fourth embodiment, a diameter D 1  of a portion journaled in a main bearing  21  of a rotating shaft  13  differs from a diameter D 2  of a portion journaled in a sub-bearing  22 . The diameter D 1  of the portion journaled in the main bearing  21  of the rotating shaft  13  is formed larger than the diameter D 2  of the portion journaled in the sub-bearing  22  (D 1 &gt;D 2 ). 
     Accordingly, in the compression mechanism portion  12 B, similarly to the compression mechanism portion  12  of the second embodiment (see  FIG. 3 ), because the diameter D 1  is formed larger than the diameter D 2 , it is necessary to secure a seal width of a circular groove K with respect to a cylinder chamber S in an end face of a roller  25 . Therefore, an inner circumferential surface Kg of the circular groove K is hardly tapered, and a groove Ka having an even width in the depth direction is provided. 
     The tapered inner circumferential surface Kq of the circular groove K is provided only in a portion of a rotating shaft  13  that is journaled in the sub-bearing  22  having a small diameter, and the seal width of the end face of the roller  25  is secured with respect to the cylinder chamber S. 
     Because the length in the axis direction of the cylindrical pivot portion  22   b  is shorter than that of the main bearing  21 , the flexural deformation becomes large, and the load also becomes large. Therefore, the circular groove K whose inner circumferential surface Kq is tapered is extremely advantageously provided. 
       FIG. 10  is a longitudinal sectional view of the hermetic type compressor  1 A according to a modification of the third embodiment of the invention, and the refrigeration cycle is omitted in  FIG. 10 . 
     Basically, the hermetic type compressor  1 A of the modification of the third embodiment comprises the two-cylinder type compression mechanism portion  12 A of the third embodiment (see  FIG. 4 ), diameters of bearing holes N made in a main bearing  21  and a sub-bearing  22  are equal to each other, and the main bearing  21  and the sub-bearing  22  comprise circular grooves K. 
     In the modification of the third embodiment, a discharge valve mechanism  27  for a first cylinder chamber Sa is provided in the main bearing  21 , a discharge valve mechanism  27  for a second cylinder chamber Sb is provided in the sub-bearing  22 , and a discharge valve mechanism  27 A for the first cylinder chamber Sa and a discharge valve mechanism  27 A for the second cylinder chamber Sb are provided in an intermediate partition plate  30 A that is interposed between two cylinders  20 A and  20 B. 
     Because the intermediate partition plate  30 A comprises the two discharge valve mechanisms  27 A, the intermediate partition plate  30 A is divided into two in a thickness direction. As described later, the two discharge valve mechanisms  27 A of the intermediate partition plate  30 A are mounted while overlapping each other when viewed from above. 
       FIG. 11  is a plan view of the intermediate partition plate  30 A when viewed from a side of a surface in which the discharge valve mechanisms  27 A overlap each other. 
     As illustrated by a solid-line arrow of  FIG. 11 , a gas refrigerant that is discharged from a discharge holes  26  made in each of the divided intermediate partition plates  30 A is guided to the outside from a communication hole  32  through a groove  31  provided in each of the intermediate partition plates  30 A. 
       FIG. 12  is a longitudinal sectional view of a region where the discharge valve mechanisms  27 A are provided in the intermediate partition plates  30 A divided into two. 
     The discharge valve mechanism  27 A comprises a discharge valve  33  and a discharge valve guard  34   a.  One end of the discharge valve  33  is supported while separated from a discharge hole  26 . The discharge valve  33  is formed by a thin spring plate, and the other end of the discharge valve  33  is in close contact with the discharge hole  26  so as to close the discharge hole  26 . The discharge valve guard  34   a  is formed by a thick plate piece having rigidity, and the discharge valve guard  34   a  is gently bent from a support portion at one end toward the discharge hole  26  at the other end. 
     A pressure at each of cylinder chambers Sa and Sb increases by the compression action of the refrigerant, the discharge valve  33  is pressed when the pressure reaches a predetermined value, and the discharge valve  33  is elastically deformed to open the discharge hole  26 . Accordingly, the high-pressure gas refrigerant compressed by each of the cylinder chambers Sa and Sb is discharged from the discharge hole  26 . The discharge valve guard  34   a  receives the elastically-deformed discharge valve  33  to regulate further deformation, thereby preventing metal fatigue of the discharge valve  33  as much as possible. 
     The discharge valve guard  34   a  has a specific thickness because the discharge valve guard  34   a  has the necessary rigidity. One end of the discharge valve guard  34   a  mounted on the intermediate partition plate  30 A is formed into a flat shape, and the discharge valve guard  34   a  is bent into a predetermined curved shape from the flat-shape leading end to the other end facing the discharge hole  26 . Therefore, the leading end of the discharge valve guard  34   a  is formed at a certain level from a flat surface formed in the mounting portion. 
     When the intermediate partition plate  30 A directly comprises the discharge valve mechanism  27 A, the wall thickness of the intermediate partition plate  30 A increases considerably, and the compression mechanism portion  12 A is lengthened in the axis direction, which leads to enlargement of the compressor  1 A. 
     When the intermediate partition plate  30 A is thickened, an interval between the first cylinder chamber Sa and the second cylinder chamber Sb is lengthened, and the distance between the eccentric portions  13   a  of the rotating shafts  13  that are accommodated in the first cylinder chamber Sa and the second cylinder chamber Sb. This leads to the degradation of the rigidity of the rotating shaft  13  to cause the increase of flexural deformation, amplification of wobbling, and the degradation of the reliability. 
     Therefore, as illustrated in a first example of  FIG. 12 , in the discharge valve guard  34   a,  a flat portion mounted on the intermediate partition plate  30 A has the same wall thickness, and a bent portion U facing the discharge hole  26  is tapered such that a wall thickness decreases gradually toward the leading end and such that the wall thickness in section becomes the thinnest in the leading end portion. 
     Because the discharge valve guard  34   a  receives the force of the discharge valve  33 , strength is required for the discharge valve guard  34   a,  and the discharge valve guard  34   a  is formed with a predetermined thickness. However, a stress is not applied to the leading end of the bent portion U too much, and no problem occurs even if a section of the leading end of the bent portion U is thinned into the tapered shape. 
     Therefore, a height of the discharge valve guard  34   a  can be reduced to decrease the wall thickness of the intermediate partition plate  30 A. As the height of the compression mechanism portion  12 A is reduced, the distance between the eccentric portions  13   a  of the rotating shafts  13  can be shortened to reduce the flexural deformation or wobbling of the rotating shaft  13 , thereby improving the reliability. 
     Meanwhile, the discharge valve mechanisms  27  of the main bearing  21  and sub-bearing  22  are removed, the discharge valve mechanism  27 A for the first cylinder chamber Sa and the discharge valve mechanism  27 A for the second cylinder chamber Sb may be provided only in the intermediate partition plate  30 A. 
     Alternatively, as illustrated in a second example of  FIG. 13 , only a leading end Z of the bent portion is processed, although the plate thickness is evenly formed from the mounting portion to the bent portion without changing a configuration of the discharge valve guard  34   a.    
     That is, at the leading ends Z of the discharge valve guards  34   a,  surfaces facing each other that are the surfaces that do not collide with the discharge valves  33  are cut into flat shapes so as to be parallel. Therefore, the distance between the mounting portions of the two discharge valve guards  34   a  can further be shortened to minimize the thickness of the intermediate partition plate  30 A, so that the above-described effect is obtained. 
     The invention is not limited to the embodiments, but various modifications can be made at an implementation stage without departing from the scope of the invention. Various inventions can be made by appropriately combining a plurality of constituents disclosed in the embodiments. 
     According to the invention, according to the flexural deformation of the rotating shaft due to the compressive load in the cylinder, the uneven contact with the rotating shaft is prevented in at least one of the main bearing and sub-bearing, thereby achieving the enhancement of the reliability and the longer operation life. Additionally, the hermetic type compressor is provided to form the refrigeration cycle, thereby improving refrigeration efficiency.