HERMETIC TYPE COMPRESSOR AND REFRIGERATION CYCLE APPARATUS

Abstract

As a hermetic type compressor, 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. 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 40% of a diameter of the bearing hole.

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 1a. A suction portion 1b 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 1b.

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 21a and 22a and cylindrical pivot portions 21b and 22b, respectively. The flanges 21a and 22a close the cylinder internal diameter hole S. The cylindrical pivot portions 21b and 22b are projected along axial center portions of the flanges 21a and 22a while integrated with the flanges 21a and 22a, and the cylindrical pivot portions 21b and 22b 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 13a 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 13a. The roller 25 and the eccentric portion 13a 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 28c 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 1b is provided across the region where the blade comes into contact with the roller 25 from the discharge hole 26. The suction portion 1b is communicated with the closed vessel 10 while radially piercing the cylinder 20, and the suction portion 1b 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 21a and cylindrical pivot portion 21b constituting the main bearing 21 to the cylindrical pivot portion 21b. 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 13a 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 13a]−outer circumferential radius g (mm) of circular groove K   (2)

outer circumferential radius g (mm) of circular groove K>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 13a 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 1a 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 13a 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 13a]−outer circumferential radius g (mm) of circular groove K   (2)

outer circumferential radius g (mm) of circular groove K>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 13a]−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 21a 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 21b and the flange 21a 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 D1 of a portion (bearing hole Na) that is journaled in a main bearing 21 of a rotating shaft 13 differs from a diameter D2 of a portion (bearing hole Nb) that is journaled in a sub-bearing 22. Actually, the diameter D1 of the portion journaled in the main bearing 21 of the rotating shaft 13 is formed larger than the diameter D2 of the portion journaled in the sub-bearing 22 (D1>D2).

Because the diameter D1 is formed larger than the diameter D2, 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 22b 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 1A according to a third embodiment of the invention with part of the hermetic type compressor 1A omitted.

Basically, the configuration in which a motor portion 11 and a compression mechanism portion 12A 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 12A is a two-cylinder type compressor 1A that comprises two cylinders 20A and 20B that are provided above and below an intermediate partition plate 30. Each of the cylinders 20A and 20B comprises an internal diameter hole Sa. The internal diameter hole Sa of the cylinder 20A 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 20B on the lower side is closed by a sub-bearing 22 and the intermediate partition plate 30. Eccentric portions 13a and 13b and a roller 25 are accommodated in the first cylinder chamber Sa and the second cylinder chamber Sb, respectively. The eccentric portions 13a and 13b are provided while integrated with the rotating shaft 13, and the eccentric portions 13a and 13b have a phase difference of 180°. The roller 25 is fitted in the eccentric portions 13a and 13b.

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 1A 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 1B according to a fourth embodiment of the invention with part of the hermetic type compressor 1A omitted.

Basically, the hermetic type compressor 1B of the fourth embodiment comprises a compression mechanism portion 12B having a configuration similar to that of the two-cylinder type compression mechanism portion 12A of the third embodiment (see FIG. 4).

In the fourth embodiment, a diameter D1 of a portion journaled in a main bearing 21 of a rotating shaft 13 differs from a diameter D2 of a portion journaled in a sub-bearing 22. The diameter D1 of the portion journaled in the main bearing 21 of the rotating shaft 13 is formed larger than the diameter D2 of the portion journaled in the sub-bearing 22 (D1>D2).

Accordingly, in the compression mechanism portion 12B, similarly to the compression mechanism portion 12 of the second embodiment (see FIG. 3), because the diameter D1 is formed larger than the diameter D2, 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 22b 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 1A 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 1A of the modification of the third embodiment comprises the two-cylinder type compression mechanism portion 12A 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 27A for the first cylinder chamber Sa and a discharge valve mechanism 27A for the second cylinder chamber Sb are provided in an intermediate partition plate 30A that is interposed between two cylinders 20A and 20B.

Because the intermediate partition plate 30A comprises the two discharge valve mechanisms 27A, the intermediate partition plate 30A is divided into two in a thickness direction. As described later, the two discharge valve mechanisms 27A of the intermediate partition plate 30A are mounted while overlapping each other when viewed from above.

FIG. 11 is a plan view of the intermediate partition plate 30A when viewed from a side of a surface in which the discharge valve mechanisms 27A 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 30A is guided to the outside from a communication hole 32 through a groove 31 provided in each of the intermediate partition plates 30A.

FIG. 12 is a longitudinal sectional view of a region where the discharge valve mechanisms 27A are provided in the intermediate partition plates 30A divided into two.

The discharge valve mechanism 27A comprises a discharge valve 33 and a discharge valve guard 34a. 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 34a is formed by a thick plate piece having rigidity, and the discharge valve guard 34a 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 34a 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 34a has a specific thickness because the discharge valve guard 34a has the necessary rigidity. One end of the discharge valve guard 34a mounted on the intermediate partition plate 30A is formed into a flat shape, and the discharge valve guard 34a 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 34a is formed at a certain level from a flat surface formed in the mounting portion.

When the intermediate partition plate 30A directly comprises the discharge valve mechanism 27A, the wall thickness of the intermediate partition plate 30A increases considerably, and the compression mechanism portion 12A is lengthened in the axis direction, which leads to enlargement of the compressor 1A.

When the intermediate partition plate 30A 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 13a 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 34a, a flat portion mounted on the intermediate partition plate 30A 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 34a receives the force of the discharge valve 33, strength is required for the discharge valve guard 34a, and the discharge valve guard 34a 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 34a can be reduced to decrease the wall thickness of the intermediate partition plate 30A. As the height of the compression mechanism portion 12A is reduced, the distance between the eccentric portions 13a 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 27A for the first cylinder chamber Sa and the discharge valve mechanism 27A for the second cylinder chamber Sb may be provided only in the intermediate partition plate 30A.

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 34a.

That is, at the leading ends Z of the discharge valve guards 34a, 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 34a can further be shortened to minimize the thickness of the intermediate partition plate 30A, 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.

Claims

1. A hermetic type compressor in which 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,

wherein 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 has 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 L of the circular groove is set to at least 40% of a diameter D of the bearing hole.

2. The hermetic type compressor according to claim 1, wherein, in the main bearing or sub-bearing that comprises the circular groove, a minimum wall thickness b between an inner circumferential surface of the circular groove and a circumferential surface of the bearing hole is set so as to satisfy a relationship of an equation (1):

0.09×diameter D of bearing hole≧minimum wall thickness b≧0.04×diameter D of bearing hole   (1).

3. The hermetic type compressor according to claim 1, wherein a compression chamber of the compression mechanism portion accommodates an eccentric portion that is eccentrically provided while integrated with the rotating shaft and a rolling piston that is fitted in the eccentric portion to rotate eccentrically in the compression chamber in association with rotation of the rotating shaft, and

assuming that e is an eccentric amount of the eccentric portion and r is an outer circumferential radius of the rolling piston, an outer circumferential radius g of the circular groove satisfies relationships of equations (2) and (3): 0.5 mm≦[outer circumferential radius r (mm) of rolling piston−eccentric amount e (mm) of eccentric portion]−outer circumferential radius g (mm) of circular groove   (2) outer circumferential radius g (mm) of circular groove>diameter D (mm) of bearing hole/2+minimum wall thickness b (mm)   (3).

4. The hermetic type compressor according to claim 1, wherein the main bearing and the sub-bearing have a flange whose wall thickness is set to a depth L of the circular groove or less.

5. A refrigeration cycle apparatus comprising:

the hermetic type compressor according to claim 1;

a condenser;

an expansion device; and

an evaporator.