EP4764218A1

EP4764218A1 - Rotary compressor and refrigeration cycle device

Info

Publication number: EP4764218A1
Application number: EP25852800.9A
Authority: EP
Inventors: Akichika MORISHIMA; Takazo Sotojima; Junki Yamaguchi; Kento GONOME
Original assignee: Daikin Industries Ltd
Current assignee: Daikin Industries Ltd
Priority date: 2024-09-30
Filing date: 2025-09-08
Publication date: 2026-06-24
Also published as: WO2026070338A1; JP7853624B2; JP2026062501A

Abstract

A lower end surface of a roller (36) of a rotary compressor (1) has a horizontal surface (36a) parallel to an upper surface of a rear head (42), and an inclined surface (36b) continuous with an outer peripheral side of the horizontal surface (36a); extending to an outer peripheral surface (36o) of the roller (36); and inclined so as to be positioned higher at a radially outer side.

Description

TECHNICAL FIELD

The present disclosure relates to a rotary compressor and a refrigeration cycle apparatus including the rotary compressor. The rotary compressor is a compressor that eccentrically rotates a roller in a cylinder to compress gas in a cylinder chamber formed in the cylinder. In general, the rotary compressor has a vane for dividing the cylinder chamber. The rotary compressor includes a so-called rolling-piston-type compressor in which a roller eccentrically rotates while a vane separate from the roller contacts the roller; a so-called swing-type compressor in which a vane integrated with a roller swings as the roller eccentrically rotates; a so-called hinge-vane-type compressor in which a roller eccentrically rotates while the tip end of a vane is rotatably fitted in a recess of the outer peripheral surface of the roller; and the like.

BACKGROUND ART

Conventionally, a rotary compressor has been used in which a refrigerant is compressed by a cylindrical roller eccentrically rotating in a cylinder chamber that is comprised of a cylindrical cylinder and end plates which close the upper and lower ends of the cylinder (see, for example, Patent Document 1). In this rotary compressor, the compressed refrigerant is discharged into a casing through a port penetrating the end plate. Therefore, the outer periphery of a compression mechanism having the cylinder, the roller, and the end plates is a high-pressure space.
In the above rotary compressor, an upper corner portion formed between the inner peripheral surface and the upper end surface of the roller and a lower corner portion formed between the inner peripheral surface and the lower end surface of the roller are chamfered, whereby high-pressure lubricant is supplied to the upper chamfered portion and the lower chamfered portion. Further, in the above rotary compressor, the horizontal projection area of the chamfered surface of the lower corner portion is made larger than or equal to the horizontal projection area of the chamfered surface of the upper corner portion so that the vertical component of force exerted on the roller by the high-pressure lubricant is larger at the chamfered portion of the lower corner portion than at the chamfered portion of the upper corner portion. In the above rotary compressor, with the above configuration, during the operation in which a fitting portion between an eccentric portion of a drive shaft and the roller is filled with high-pressure lubricant, the vertically upward force exerted on the roller by the high-pressure lubricant supplied to the upper and lower chamfered portions is made larger than the vertically downward force, whereby the roller is pushed up from bottom to top, then the gaps between the upper and lower end surfaces of the roller and the end plates facing the upper and lower end surfaces are made uniform, and then the leakage loss of the roller end surface is reduced.

CITATION LIST

PATENT DOCUMENT

Patent Document 1: Japanese Unexamined Patent Publication No. 2010-31733

SUMMARY OF THE INVENTION

TECHNICAL PROBLEM

However, in the above rotary compressor, it is conceivable in principle that the force by the high-pressure lubricant does not act only on the upper and lower chamfered portions of the roller but acts on the entire part of the upper end surface and the lower end surface of the roller. Therefore, in the above rotary compressor, it is conceivable that even if the upper corner portion and the lower corner portion are chamfered, the projection areas of the upper end surface and the lower end surface of the roller that includes the chamfered portions are substantially equal to each other, and therefore the vertically upward force enabling the roller to float up is not generated by the high-pressure lubricant. Therefore, in the above rotary compressor, there is a problem that the gaps between the upper and lower end surfaces of the roller and the end plates facing those end surfaces cannot be made uniform and the leakage loss cannot be sufficiently reduced.
An object of the present disclosure is to provide a rotary compressor in which a roller can float up to an appropriate height during the operation.

SOLUTION TO THE PROBLEM

A first aspect of the present disclosure is directed to a rotary compressor including:

a casing (10);
a drive shaft (70) provided in the casing (10) and extending in a top-bottom direction;
a motor (20) provided in the casing (10) and configured to drive the drive shaft (70); and
a compression mechanism (30) provided in the casing (10); coupled to the drive shaft (70); and driven and rotated by the motor (20) to compress a refrigerant,
wherein
the compression mechanism (30) has
- a cylinder (34) having a cylindrical shape,
- an upper end plate (41) and a lower end plate (42) closing an upper opening end surface and a lower opening end surface of the cylinder (34),
- a cylindrical roller (36) attached to the drive shaft (70) and eccentrically rotating in the cylinder (34), and
- a vane (37) dividing a space between the cylinder (34) and the roller (36) into a first chamber (51) on a suction side and a second chamber (52) on a discharge side,
a lower end surface of the roller (36) has a first surface (36a) parallel to an upper surface of the lower end plate (42) and a second surface (36b) continuous with an outer peripheral side of the first surface (36a) and extending to an outer peripheral surface (36o) of the roller (36), and
the second surface (36b) is inclined so as to be positioned higher at a radially outer side.

In the first aspect, an outer peripheral part of the lower end surface of the roller (36) is formed as the second surface (36b) which is inclined so as to be positioned higher at a radially outer side, whereby the wedge-shaped gap (90) is formed on the outer peripheral side between the lower end surface of the roller (36) and the lower end plate (42). With this configuration, during the operation, the high-pressure lubricant equivalent in pressure to the inside of the casing (10) is supplied to each sliding portion of the compression mechanism (30), and part of the lubricant reaches the space between the cylinder (34) and the roller (36). Then this lubricant having reached the space is pushed into the radially inner side (the center of the roller (36)) through the outer peripheral inlet of the wedge-shaped gap (90) by eccentric rotation of the roller (36) in the cylinder (34). The high-pressure lubricant pushed in this manner exerts an upward force on the roller (36), which causes the roller (36) to float upward (the wedge effect). Therefore, according to the first aspect, it is possible to provide the rotary compressor (1) in which the roller (36) can float up to an appropriate height during the operation.
A second aspect of the present disclosure is an embodiment of the first aspect. In the second aspect,

the second surface (36b) satisfies Expression (1) below: $(ΔH \div 2 \times 0.75) \div (0.0055 \times m^{- 0.893}) \leq X \leq (ΔH \div 2 \times 1.25) \div (0.00006 m^{- 1.118})$
where
a sum of a mass of the roller (36) and a mass of the vane (37) is m,
a distance in an axial direction of the drive shaft (70) between an outer end of the second surface (36b) and the first surface (36a) is X, and
a difference between a height of the cylinder (34) and a height of the roller (36) is ΔH.

For various versions of the rotary compressor (1) with various sizes, the inventors of the present application have estimated the range of the height of the wedge-shaped gap (90) (the distance in the axial direction of the drive shaft (70) between the outer end of the second surface (36b) and the first surface (36a)) that enables the roller (36) to float up to an appropriate height during the operation; and then the inventors have derived Expression (1) expressed using the sum of the mass of the roller (36) and the mass of the vane (37) and the difference in height between the cylinder (34) and the roller (36).
In the second aspect, for various versions of the rotary compressor with various sizes, by forming the second surface (36b) so that Expression 1 is satisfied, it is possible to easily form various versions of the rotary compressor with various sizes in which the roller (36) can float up to an appropriate height during the operation.
A third aspect of the present disclosure is an embodiment of the first or second aspect. In the third aspect,

the rotary compressor is a swing-type compressor including a piston (35) in which the roller (36) and the vane (37) are integrated together, and
assuming that an angular position of a center line (M) of the vane (37) with respect to a center of the roller (36) in an circumferential direction of the roller (36) is 0° and that an angular position of the roller (36) in the circumferential direction increases while moving in a rotation direction of the drive shaft (70),
the second surface (36b) is formed in a range at least from an angular position of a side surface of the vane (37) that is on a suction side to an angular position of 180° in the circumferential direction of the roller (36).

The third aspect is directed to a swing-type compressor including the piston (35) in which the roller (36) and the vane (37) are integrated together. As shown in FIG. 6C, in the swing-type compressor, the amount of leakage of lubricant into the cylinder chamber through the gaps above and below the roller (36) becomes large particularly when the rotation angle of the drive shaft (70) is 0° (360°).
Here, assuming that the second surface (36b) is formed on the roller (36) in the circumferential direction in the angular range of α to α + 180°, the wedge effect described above is obtained effectively when the rotation angle of the drive shaft (70) is α, and the roller (36) floats up effectively.
Therefore, in the third aspect, the second surface (36b) is formed in the range at least from the angular position of the side surface of the vane (37) that is on the suction side to the angular position of 180° in the circumferential direction of the roller (36). According to this configuration, the wedge effect is produced effectively when the rotation angle of the drive shaft (70) is 0°, and the roller (36) can float up reliably when the rotation angle of the drive shaft (70) is 0° at which the amount of leakage of lubricant into the cylinder chamber is significantly large. Therefore, according to the third aspect, it is possible to reduce the amount of leakage of lubricant into the cylinder chamber, and thus it is possible to reduce the leakage loss.
A fourth aspect of the present disclosure is an embodiment of the third aspect. In the fourth aspect,
the second surface (36b) is formed in a range from the angular position of the side surface of the vane (37) that is on the suction side to an angular position of a side surface of the vane (37) that is on a discharge side in the circumferential direction of the roller (36).
In the fourth aspect, the second surface (36b) is formed on an almost entire circumference of the roller (36) in the circumferential direction while the angular range in which the vane (37) is present in the circumferential direction of the roller (36) is excluded. Since the roller (36) can float up whatever the rotation angle of the drive shaft (70) may be, it is possible to reduce the amount of leakage of lubricant into the space (the cylinder chamber) between the cylinder (34) and the roller (36), and therefore it is possible to reduce the leakage loss.
A fifth aspect of the present disclosure is an embodiment of the third aspect. In the fifth aspect,
the second surface (36b) is formed in a range from the angular position of the side surface of the vane (37) that is on the suction side to a predetermined angular position of 180° or more and 270° or less in the circumferential direction of the roller (36).
Here, from the viewpoint of reducing the amount of leakage of lubricant into the cylinder chamber through the gaps above and below the roller (36) by enabling the roller (36) to float up to an appropriate height, the range in which the second surface (36b) is formed (the angular range in which the second surface (36b) is formed in the circumferential direction of the roller (36)) should be as large as possible in one preferred embodiment. On the other hand, if the second surface (36b) is formed by chamfering the outer peripheral corner portion of the lower end portion of the roller (36) and then the wedge-shaped gap (90) is formed between the second surface (36b) and the upper surface of the lower end plate (42), it is possible to reduce the amount of leakage of lubricant into the cylinder chamber through the gaps above and below the roller (36) by enabling the roller (36) to float up, but the refrigerant compressed in the second chamber (52) may leak into the first chamber (51) through the gap (90), which may cause performance degradation. As the range in which the second surface (36b) is formed is made larger, the leakage of a refrigerant through the gap (90) more likely occurs, which may more likely cause performance degradation. Therefore, from the viewpoint of reducing the leakage of a refrigerant from the second chamber (52) to the first chamber (51) through the gap (90), the range in which the second surface (36b) is formed should be as small as possible in one preferred embodiment.
Therefore, in the fifth aspect, the second surface (36b) is formed only in the range from the angular position of the side surface of the vane (37) that is on the suction side to a predetermined angular position (180° or more and 270° or less) in the circumferential direction of the roller (36), and the second surface (36b) is not formed in the range from the predetermined angular position to the angular position of the side surface of the vane (37) that is on the discharge side. With this configuration, according to the fifth aspect, it is possible to reduce the leakage of a refrigerant from the second chamber (52) to the first chamber (51) through the wedge-shaped gap (90) produced by forming the second surface (36b).
A sixth aspect of the present disclosure is a refrigeration cycle apparatus including the rotary compressor of any one of the first to fifth aspects.

BRIEF DESCRIPTION OF THE DRAWINGS

[FIG. 1] FIG. 1 is a piping system diagram of a refrigeration cycle apparatus according to a first embodiment.
[FIG. 2] FIG. 2 is a longitudinal sectional view of a rotary compressor according to the first embodiment.
[FIG. 3] FIG. 3 is a lateral sectional view of a compression mechanism of the rotary compressor.
[FIG. 4] FIG. 4 is a partially enlarged longitudinal sectional view of the rotary compressor.
[FIG. 5] FIG. 5 shows operation of the compression mechanism.
[FIG. 6A] FIG. 6A is a graph showing how the amount of leakage of lubricant into a first chamber through gaps above and below a rotor changes while a drive shaft rotates one time, where the broken line indicates a change in which an inclined surface is not formed on the lower surface of the rotor, and the solid line indicates a change in which the inclined surface is formed on the lower surface of the rotor.
[FIG. 6B] FIG. 6B is a graph showing how the amount of leakage of lubricant into a second chamber through gaps above and below a rotor changes while a drive shaft rotates one time, where the broken line indicates a change in which an inclined surface is not formed on the lower surface of the rotor, and the solid line indicates a change in which the inclined surface is formed on the lower surface of the rotor.
[FIG. 6C] FIG. 6C is a graph showing how the amount of leakage of lubricant into a cylinder chamber through gaps above and below a rotor changes while a drive shaft rotates one time, where the broken line indicates a change in which an inclined surface is not formed on the lower surface of the rotor, and the solid line indicates a change in which the inclined surface is formed on the lower surface of the rotor.
[FIG. 7A] FIG. 7A is a graph showing a result of estimation of the relation between the cutting height and the floating amount for each cutting angle for the rotary compressor with the piston mass of 0.01 kg.
[FIG. 7B] FIG. 7B is a graph showing a result of estimation of the relation among the cutting angle, the cutting height, and the floating amount for the rotary compressor with the piston mass of 0.1 kg.
[FIG. 8A] FIG. 8A is a graph showing the relation between the slope and the mass of the piston where the floating amount at the cutting angle of 10° is expressed as a function of the cutting height.
[FIG. 8B] FIG. 8B is a graph showing the relation between the slope and the mass of the piston where the floating amount at the cutting angle of 45° is expressed as a function of the cutting height.
[FIG. 9] FIG. 9 is a graph showing the relation between the cutting height and the floating amount and the range of an appropriate floating amount at the cutting angle of 10° and 45°.
[FIG. 10] FIG. 10 is a lateral sectional view of a compression mechanism of a rotary compressor according to a second embodiment.

DESCRIPTION OF EMBODIMENTS

<<First Embodiment>>

Embodiments of the present invention will be described in detail below with reference to the drawings. The embodiments below are merely exemplary ones in nature, and are not intended to limit the scope, applications, or use of the present invention. Features of the embodiment, variations, and other examples described below can be combined or partially substituted within the range where the present invention can be embodied. The terms "top" and "bottom" refer to the directions of a rotary compressor (1) in front view (see FIG. 2). Some drawings may be illustrated without hatching for the sake of easy understanding.

(1) Refrigeration Cycle Apparatus

As shown in FIG. 1, the rotary compressor (1) of this example is applied to a refrigeration cycle apparatus (100). In the following, the rotary compressor (1) may be simply referred to as a compressor (1). The refrigeration cycle apparatus (100) is an air conditioner for conditioning air in an indoor space, for example.
The refrigeration cycle apparatus (100) has an outdoor unit (7) disposed outdoors and an indoor unit (8) disposed indoors. The outdoor unit (7) includes the compressor (1), an accumulator (2), a four-way switching valve (3), an outdoor heat exchanger (4), and an expansion valve (5). The indoor unit (8) includes an indoor heat exchanger (6).
The refrigeration cycle apparatus (100) includes a refrigerant circuit (9). The compressor (1), the four-way switching valve (3), the outdoor heat exchanger (4), the expansion valve (5), and the indoor heat exchanger (6) are connected to the refrigerant circuit (9). The refrigerant flows through the refrigerant circuit (9) to create a refrigeration cycle.
The refrigeration cycle apparatus (100) performs the heating operation and the cooling operation by switching the four-way switching valve (3). In the cooling operation, a first refrigeration cycle is performed. Specifically, when a first port (p1) and a third port (p3) of the four-way switching valve (3) communicate with each other and a second port (p2) and a fourth port (p4) communicate with each other (the solid lines in FIG. 1), the indoor heat exchanger (6) functions as an evaporator, and the outdoor heat exchanger (4) functions as a radiator. In the heating operation, a second refrigeration cycle is performed. Specifically, when the first port (p1) and the fourth port (p4) of the four-way switching valve (3) communicate with each other and the second port (p2) and the third port (p3) communicate with each other (the broken lines in FIG. 1), the indoor heat exchanger (6) functions as a radiator, and the outdoor heat exchanger (4) functions as an evaporator.

(2) Rotary Compressor

As shown in FIG. 2, the rotary compressor (1) includes a casing (10), a drive shaft (70), a motor (20), and a compression mechanism (30). The drive shaft (70), the motor (20), and the compression mechanism (30) are housed in the casing (10). The compressor (1) is a so-called high-pressure dome-type compressor where the refrigerant compressed in the compression mechanism (30) is discharged into an internal space (60) of the casing (10) so that the pressure in the internal space (60) becomes high.

(2-1) Casing

The casing (10) is vertically long. Specifically, the casing (10) includes a barrel portion (11) formed in a cylindrical shape and extending in the top-bottom direction, an upper casing head (12) closing the upper end of the barrel portion (11), and a lower casing head (13) closing the lower end of the barrel portion (11). The upper casing head (12) and the lower casing head (13) are relatively thick. A suction pipe (14) is provided in a lower portion of the barrel portion (11), and a discharge pipe (15) passes through an upper portion of the upper casing head (12). A bottom portion (a lower portion) of the casing (10) is an oil reservoir in which lubricant is stored.

(2-2) Drive Shaft

The drive shaft (70) is disposed so as to extend in the top-bottom direction in the casing (10). An upper portion of the drive shaft (70) is coupled to a rotor (22) (described later) of the motor (20). A lower portion of the drive shaft (70) has an upper main shaft portion (71), an eccentric portion (72), and a lower main shaft portion (73) in sequence from top to bottom. The upper main shaft portion (71) and the lower main shaft portion (73) share the same axis. The eccentric portion (72) has a larger diameter than the upper main shaft portion (71) and the lower main shaft portion (73), and is eccentric to the axis of the upper main shaft portion (71) and the lower main shaft portion (73). The eccentric portion (72) is coupled with a piston (35) (described later) of the compression mechanism (30). The drive shaft (70) is driven by the motor (20) to eccentrically rotate the piston (35).
An oil supply passage (80) that guides the lubricant to each sliding portion of the compression mechanism (30) is formed in a lower portion of the drive shaft (70). The oil supply passage (80) has one main passage (81) extending in the top-bottom direction and a plurality of branch passages (82) extending in the radial direction from the main passage (81) to the outer peripheral surface of the drive shaft (70). In the first embodiment, one branch passage (82) is provided at every one of a position in the middle of the upper main shaft portion (71) in the top-bottom direction as well as above the compression mechanism (30) and a position in the middle of the eccentric portion (72) in the top-bottom direction. The lower end of the drive shaft (70) is provided with an oil tube (not shown) that draws the lubricant stored in the bottom portion of the casing (10) to the oil supply passage (80).

(2-3) Motor

The motor (20) is housed in the casing (10). The motor (20) drives the compression mechanism (30). The motor (20) is disposed above the compression mechanism (30) in the casing (10). The motor (20) has a stator (21) formed in a tubular shape along the inner peripheral surface of the barrel portion (11), and a rotor (22) disposed inside the stator (21). The stator (21) is fixed to the inner surface of the barrel portion (11) by welding. A method of fixing the stator (21) to the barrel portion (11) is not limited to welding and may be shrink fitting or press fitting.

(2-4) Compression Mechanism

As shown in FIGS. 2 to 4, the compression mechanism (30) is housed in the casing (10). The compression mechanism (30) compresses a sucked refrigerant and discharges a compressed refrigerant to the internal space (60) of the casing (10). The compression mechanism (30) includes a cylinder (34), a front head (41), a rear head (42), and the piston (35). The compression mechanism (30) is fixed to the inner surface of the barrel portion (11) by welding.

(2-4-1) Cylinder

The cylinder (34) is a thick disk-shaped member. As shown in FIG. 3, the cylinder (34) has a cylinder bore (31), a vane housing hole (32), and a suction port (55).
The cylinder bore (31) is a circular hole penetrating the cylinder (34) in the thickness direction. The cylinder bore (31) is formed in a center portion of the cylinder (34). The cylinder bore (31) houses a roller (36) of the piston (35) in particular.
A cylinder chamber (S) is formed in the cylinder (34). Specifically, the cylinder chamber (S) is formed between the inner peripheral surface of the cylinder (34) defining the outer edge of the cylinder bore (31) and the piston (35).
The vane housing hole (32) is a hole extending from the inner peripheral surface of the cylinder (34) (i.e., the outer edge of the cylinder bore (31)) toward the outside of the cylinder (34) in the radial direction. The vane housing hole (32) penetrates the cylinder (34) in the thickness direction. The vane housing hole (32) houses a vane (37) of the piston (35).

(2-4-2) Front Head

As shown in FIGS. 2 and 4, the front head (41) is an upper end plate that closes an upper end portion (an upper opening end surface) of the cylinder (34) in the axial direction. Specifically, the front head (41) closes the end portion of the cylinder (34) that is close to the motor (20) in the axial direction (the upper opening end surface of the cylinder (34) in FIG. 2). The front head (41) is one example of an upper end plate (41) of the present disclosure. The front head (41) includes a first body portion (41a) and an upper bearing portion (41b). The first body portion (41a) and the upper bearing portion (41b) are integrated together.
The first body portion (41a) is formed in a substantially circular thick plate shape so that the outer peripheral portion is thicker in the top-bottom direction than the inner peripheral portion. The lower surface of the first body portion (41a) is in close contact with the upper end surface of the cylinder (34). The upper bearing portion (41b) is formed in a cylindrical shape extending from the first body portion (41a) toward the motor (20) (or toward the upper side in FIG. 2). The upper bearing portion (41b) is disposed in a center portion of the first body portion (41a). The upper bearing portion (41b) rotatably supports the upper main shaft portion (71) of the drive shaft (70).
The front head (41) has a discharge port (24). The discharge port (24) penetrates the first body portion (41a) of the front head (41) in the top-bottom direction and is opened in the upper surface. The discharge port (24) communicates with a second chamber (52) (described later) of the cylinder chamber (S). The discharge port (24) is provided with a reed valve (not shown). The reed valve is configured so that if the pressure of a refrigerant in the second chamber (52) is more than or equal to a predetermined value, part of the reed valve that closes the opening end surface of the discharge port (24) is released from the opening end surface to discharge the refrigerant.
The front head (41) is provided with a discharge muffler (25) that is formed in a bowl shape to cover the first body portion (41a) from above and surround the upper bearing portion (41b). The discharge muffler (25) is provided so as to cover the discharge port (24) and the reed valve (not shown). The discharge muffler (25) has an outflow port (not shown) through which the refrigerant discharged into a muffler space (26) in the discharge muffler (25) flows out. The refrigerant discharged from the discharge port (24) to the muffler space (26) in the discharge muffler (25) flows out to the internal space (60) of the casing (10) through the outflow port.

(2-4-3) Rear Head

The rear head (42) is a lower end plate that closes a lower end portion (a lower opening end surface) of the cylinder (34) in the axial direction. Specifically, the rear head (42) closes the end portion of the cylinder (34) that is opposite to the motor (20) in the axial direction (the lower opening end surface of the cylinder (34) in FIG. 2). The rear head (42) includes a second body portion (42a) and a lower bearing portion (42b).
The second body portion (42a) is formed in a substantially circular thick plate shape. The upper surface of the second body portion (42a) is in close contact with the lower end surface of the cylinder (34). The lower bearing portion (42b) is formed in a cylindrical shape extending from the second body portion (42a) toward the opposite side to the cylinder (34) (or toward the lower side in FIG. 2). The lower bearing portion (42b) is disposed in a center portion of the second body portion (42a). The lower bearing portion (42b) rotatably supports the lower main shaft portion (73) of the drive shaft (70).

(2-4-4) Piston

As shown in FIG. 3, the piston (35) is housed in the cylinder (34) (the cylinder bore (31)) and eccentrically rotates in the cylinder (34). The piston (35) slides on both the front head (41) and the rear head (42). The piston (35) has the roller (36) and the vane (37).
The roller (36) is formed in an annular shape. Specifically, the roller (36) is formed in a cylindrical shape that is slightly thick. The eccentric portion (72) of the drive shaft (70) is slidably inserted into the roller (36). The roller (36) revolves along the inner peripheral surface of the cylinder (34) when the drive shaft (70) rotates. The height hp of the roller (36) is lower than the height hc of the cylinder (34) so that the roller (36) can slide in the cylinder (34).
As shown in FIG. 4, a lower end portion of the roller (36) has an inclined surface (36b) formed by chamfering the outer peripheral corner portion, which will be described in detail later. That is, the lower end surface of the roller (36) has the horizontal surface (the first surface) (36a) that is parallel to the upper surface of the rear head (42) and the inclined surface (the second surface) (36b) that is continuous with the outer peripheral side of the horizontal surface (36a) and extends to the outer peripheral surface (36o) of the roller (36). On the other hand, the upper end portion of the roller (36) does not have an inclined surface, and the upper end surface of the roller (36) has only a horizontal surface parallel to the upper surface of the front head (41).
As shown in FIG. 3, the vane (37) is integrated with the roller (36). The vane (37) protrudes radially outward from the outer peripheral surface of the roller (36). The vane (37) is fitted in the vane housing hole (32). The vane (37) is sandwiched between a pair of swing bushes (54a, 54b) provided in a bush groove (53) extending radially outward from the inner peripheral surface of the cylinder (34). The vane (37) restricts the roller (36) from rotating when the roller (36) revolves. The vane (37) divides the cylinder chamber (S) into a first chamber (51) on the suction side and the second chamber (52) on the discharge side.

(3) Operation

In the compressor (1), when the motor (20) is started to rotate the rotor (22), the drive shaft (70) rotates, and the eccentric portion (72) rotates eccentrically in the cylinder (34) (the cylinder chamber (S)), as shown in FIG. 5. Then, as the eccentric portion (72) rotates eccentrically, the roller (36) of the piston (35), which is restricted from rotating, revolves along the inner peripheral surface of the cylinder (34).
The position of a contact point (CP) between the roller (36) and the cylinder (34) will be described. Assuming that as shown in FIG. 3, the angular position of the center line (M) of the vane (37) with respect to the center of the roller (36) in the circumferential direction of the roller (36) is 0° and that the angular position of the roller (36) in the circumferential direction increases while moving in the rotation direction of the drive shaft (70), the roller (36) comes into contact with the cylinder (34) at the angular position that is equal to the rotation angle of the drive shaft (70) as shown in FIG. 5. That is, the angular position of the contact point (CP) between the roller (36) and the cylinder (34) in the circumferential direction of the roller (36) is equal to the rotation angle of the drive shaft (70).
Specifically, when the rotation angle of the drive shaft (70) is 0° or 360° (the state (A) in FIG. 5), the roller (36) is in contact with the cylinder (34) at the base portion of the vane (37) which is located at the angular position of 0°. When the rotation angle of the drive shaft (70) is 45° (the state (B) in FIG. 5), the roller (36) is in contact with the cylinder (34) at the angular position of 45°. When the rotation angle of the drive shaft (70) is 90° (the state (C) in FIG. 5), the roller (36) is in contact with the cylinder (34) at the angular position of 90°. When the rotation angle of the drive shaft (70) is 135° (the state (D) in FIG. 5), the roller (36) is in contact with the cylinder (34) at the angular position of 135°. When the rotation angle of the drive shaft (70) is 180° (the state (E) in FIG. 5), the roller (36) is in contact with the cylinder (34) at the angular position of 180°. When the rotation angle of the drive shaft (70) is 225° (the state (F) in FIG. 5), the roller (36) is in contact with the cylinder (34) at the angular position of 225°. When the rotation angle of the drive shaft (70) is 270° (the state (G) in FIG. 5), the roller (36) is in contact with the cylinder (34) at the angular position of 270°. When the rotation angle of the drive shaft (70) is 315° (the state (H) in FIG. 5), the roller (36) is in contact with the cylinder (34) at the angular position of 315°.
A suction phase of sucking a refrigerant into the cylinder chamber (S) will be described. When the drive shaft (70) slightly rotates from the state in which the rotation angle is 0° (the state (A) in FIG. 5) and exceeds a suction start angle, the suction port (55) and the first chamber (51) communicate with each other to start suction of a refrigerant into the first chamber (51). The suction start angle is the rotation angle of the drive shaft (70) at which the roller (36) is in contact with the cylinder (34) at a first position (P1), which is an inner peripheral end of the suction port (55) and which the angle is minimum (or at which the contact point (CP) between the roller (36) and the cylinder (34) is located at the first position (P1)).
The refrigerant is sucked from the suction pipe (14) through the suction port (55). Then, as the rotation angle of the drive shaft (70) increases, the volume of the first chamber (51) gradually increases and the amount of refrigerant sucked into the first chamber (51) increases (the states (B) to (H) in FIG. 5). When the rotation angle of the drive shaft (70) exceeds 360° and further exceeds a suction end angle, the first chamber (51) turns into the second chamber (52). The first chamber (51) communicates with the suction port (55) until the rotation angle of the drive shaft (70) reaches the suction end angle. Therefore, the suction phase continues until the rotation angle of the drive shaft (70) reaches the suction end angle. The suction end angle is the rotation angle of the drive shaft (70) at which the contact point (CP) between the roller (36) and the cylinder (34) is located at a second position (P2) , which is the inner peripheral end of the suction port (55) and which the angle is maximum.
A compression phase of compressing a refrigerant in the cylinder chamber (S) will be described. When the rotation angle of the drive shaft (70) exceeds the suction end angle, the suction phase shifts to the compression phase. The second chamber (52) is shut off from the suction port (55), and compression of a refrigerant in the second chamber (52) is started. As the rotation angle of the drive shaft (70) further increases, the volume of the second chamber (52) decreases and the pressure in the second chamber (52) increases. When the pressure in the second chamber (52) exceeds a predetermined pressure, the compression phase ends and shifts to a discharge phase.
The discharge phase of discharging a refrigerant from the cylinder chamber (S) will be described. When the pressure in the second chamber (52) exceeds a predetermined pressure, the reed valve (not shown) opens, and the second chamber (52) and the discharge port (24) communicate with each other. When the reed valve opens, the refrigerant in the second chamber (52) is discharged into the muffler space (26) in the discharge muffler (25) through the discharge port (24). The refrigerant discharged into the muffler space (26) flows into the internal space (60) in the casing (10) through the outflow port (not shown) of the discharge muffler (25), and then is discharged to the outside of the compressor (1) through the discharge pipe (15). This discharge phase of discharging the refrigerant continues until the rotation angle of the drive shaft (70) reaches 360°.
As described above, the compressor (1) continuously performs operation of compressing a refrigerant by repeating the suction phase, the compression phase, and the discharge phase in the cylinder chamber (S).

(4) Gap between Roller and Cylinder

As described above, the roller (36) is formed with the height hp lower than the height hc of the cylinder (34) so that the roller (36) can slide in the cylinder (34). Therefore, there are gaps between the upper end surface of the roller (36) and the lower end surface of the front head (41) and between the lower end surface of the roller (36) and the upper end surface of the rear head (42). The gaps formed above and below the roller (36) are supplied and sealed with the high-pressure lubricant stored in the bottom portion of the casing (10) (the lubricant in the oil reservoir) through the oil supply passage (80). Specifically, as the drive shaft (70) rotates, the lubricant drawn to the main passage (81) through the oil tube (not shown) provided at the lower end portion of the drive shaft (70) flows out of the branch passage (82) opening in the middle of the eccentric portion (72) of the drive shaft (70) in the top-bottom direction and then is supplied to the gaps.
In consideration of the slidability of the roller (36), it is necessary for the gaps above and below the roller (36) to be large to some extent, but if the gaps are too large, the high-pressure lubricant supplied to the gaps is likely to leak into the cylinder chamber (S) (the first chamber (51) and the second chamber (52)) between the roller (36) and the cylinder (34). A large amount of refrigerant is dissolved in the high-pressure lubricant. If such high-pressure lubricant in which a large amount of refrigerant is dissolved leaks into the cylinder chamber (S) in which the pressure is relatively low, a large amount of refrigerant is separated from the leaked high-pressure lubricant. That is, a large amount of high-pressure refrigerant flows into the cylinder chamber (S) in which the pressure is relatively low (or in other words, a so-called refrigerant leakage occurs), and the efficiency of the compressor significantly decreases. In order to achieve both ensuring the slidability and reducing the leakage loss, while the gaps above and below the roller (36) are narrowed to the minimum necessary, the roller (36) is configured to float up to an appropriate height in one preferred embodiment when the compressor (1) is operating so that the gaps above and below the roller (36) are equal in height.
Therefore, in the rotary compressor (1) of the present disclosure, the inclined surface (36b) is formed on the lower end portion of the roller (36) so that the roller (36) floats up to an appropriate height (the height at which the gaps above and below the roller (36) are substantially equal in height) during the operation. The inclined surface (36b) will be described in detail below.

(5) Inclined Surface

As shown in the enlarged view of FIG. 4, the lower end portion of the roller (36) has the inclined surface (36b) formed by chamfering the outer peripheral corner portion. The inclined surface (36b) is inclined so as to be positioned higher at a radially outer side. Part of the lower end surface of the roller (36) that is other than the inclined surface (36b) is formed as the horizontal surface (36a) parallel to the upper surface of the rear head (42).
As shown in FIG. 3, the inclined surface (36b) is formed on the entire circumference of the roller (36) while the angular range in which the vane (37) is present in the circumferential direction of the roller (36) is excluded so as not to cross the vane (37), or more specifically, on the area from the angular position of the surface of the vane (37) that is on the suction side in the circumferential direction of the roller (36) (the line that connects the root-side end portion of the surface that is on the suction side and the center of the roller (36) in FIG. 3) to the angular position of the side surface of the vane (37) that is on the discharge side in the circumferential direction of the roller (36) (the line that connects the root-side end portion of the surface that is on the discharge side and the center of the roller (36) in FIG. 3).
As shown in the enlarged view of FIG. 4, the inclined surface (36b) is continuous with the outer peripheral side of the horizontal surface (36a) and extends to the outer peripheral surface (36o) of the roller (36). The cutting angle φ of the inclined surface (36b) (the angle of the inclined surface (36b) with respect to the horizontal surface (36a)) is an angle of 10° or more and 45° or less. The cutting height X of the inclined surface (36b) (the distance in the axial direction of the drive shaft (70) between the outer end of the inclined surface (36b) and the horizontal surface (36a)) is a height of 0.01 mm or more and 5 mm or less. The cutting height X of the inclined surface (36b) is less than or equal to half of the thickness t (the width in the radial direction) of the roller (36) in one preferred embodiment.

(6) Wedge Effect

Since the inclined surface (36b) is formed, whereby a wedge-shaped gap (90) (a triangular shape when viewed in the longitudinal section) is formed on the outer peripheral side between the lower end surface of the roller (36) and the upper end surface of the rear head (42). With this configuration, during the operation, the high-pressure lubricant is supplied to each sliding portion of the compression mechanism (30), and part of the lubricant reaches the cylinder chamber (S) between the cylinder (34) and the roller (36). Then this lubricant having reached the cylinder chamber (S) is pushed into the radially inner side (the center of the roller (36)) through the outer peripheral inlet of the wedge-shaped gap (90) by eccentric rotation of the roller (36). The high-pressure lubricant pushed into the radially inner side in this manner exerts an upward force on the roller (36), which causes the roller (36) to float upward. That is, the high-pressure lubricant pushed into the radially inner side of the wedge-shaped gap (90) acts as a wedge, thereby causing the roller (36) to float upward (the wedge effect).
Since the inclined surface (36b) is formed on the outer peripheral side of the roller (36), the lubricant present between the cylinder (34) that does not rotate and the roller (36) that eccentrically rotates flows into the wedge-shaped gap (90), and the relative speed between the lubricant flowing into the gap (90) and the inclined surface (36b) as a surface on which the lubricant acts becomes higher than when the inclined surface is formed on the inner peripheral side of the roller (36). Therefore, the upward force which acts on the roller (36) due to the wedge effect becomes larger, which enables the roller (36) to float upward enough.
The rotary compressor (1) of the first embodiment is a swing-type compressor (1) including the piston (35) in which the roller (36) and the vane (37) are integrated together. In the swing-type compressor (1), as the drive shaft (70) rotates, the roller (36) eccentrically rotates (revolves) in the cylinder (34) but does not rotate on its own axis. Therefore, as described above, assuming that the angular position of the center line (M) of the vane (37) with respect to the center of the roller (36) in the circumferential direction of the roller (36) is 0° and that the angular position of the roller (36) in the circumferential direction increases while moving in the rotation direction of the drive shaft (70), the roller (36) comes into contact with the cylinder (34) at the angular position that is equal to the rotation angle of the drive shaft (70). That is, the angular position of the roller (36) in the circumferential direction at the contact point (CP) between the roller (36) and the cylinder (34) is equal to the rotation angle of the drive shaft (70).
As described above, if the gaps formed above and below the roller (36) are too large, the high-pressure lubricant supplied to the gaps is likely to leak into the cylinder chamber (S) during the operation. Leakage of the lubricant into the cylinder chamber (S) causes three types of leakage loss of refrigerant (the cylinder roller (CR) leakage loss).
The first leakage loss is the loss caused by leakage through the gaps close to the suction chamber, where the high-pressure lubricant leaks into the first chamber (51) through the gaps above and below the roller (36). Specifically, a refrigerant separated from the lubricant having leaked into the first chamber (51) that is in a low pressure state flows into the first chamber (51), whereby the suction volume decreases and the efficiency of the compressor decreases.
The second leakage loss is the loss caused by leakage through the gaps close to the compression chamber, where the high-pressure lubricant leaks into the second chamber (52) in which the pressure is relatively low through the gaps above and below the roller (36). Specifically, a refrigerant separated from the lubricant having leaked into the second chamber (52) in which the pressure is relatively low flows into the second chamber (52), whereby the compression load increases and the efficiency of the compressor decreases.
The third leakage loss is the loss caused by roller trajectory leakage, where the roller (36) revolves along the inner peripheral surface of the cylinder (34) so that the lubricant having leaked into the second chamber (52) flows into the first chamber (51) (in other words, the location in the cylinder chamber (S) where the refrigerant has leaked shifts from the second chamber (52) to the first chamber (51)). Due to the third loss caused by roller trajectory leakage, similarly to the first loss caused by leakage through the gaps close to the suction chamber, the suction volume decreases and the efficiency of the compressor decreases.
FIGS. 6A to 6C are the graphs showing how the amount of leakage of lubricant changes while the drive shaft (70) rotates one time. The broken line in FIG. 6A indicates a change in the amount of leakage of lubricant into the first chamber (51) (the sum of the amount of leakage through the gaps close to the suction chamber and the amount of leakage due to the roller trajectory leakage). The broken line in FIG. 6B indicates a change in the amount of leakage of lubricant into the second chamber (52) (the amount of leakage through the gaps close to the compression chamber). The broken line in FIG. 6C shows the total sum of the amount of leakage indicated by the broken lines in FIGS. 6A and 6B, that is, a change in the amount of leakage of lubricant into the cylinder chamber (S).
The amount of leakage through the gaps close to the suction chamber, which leads to the first leakage loss, is determined by the difference between the pressure in the internal space (60) of the casing (10) (the pressure of lubricant) and the internal pressure of the first chamber (51) and the arc length of part of the outer peripheral surface of the roller (36) that faces the first chamber (51), and the amount of leakage increases as the volume of the first chamber (51) increases (or as the rotation angle increases). The amount of leakage due to the roller trajectory leakage, which leads to the third leakage loss, also increases as the volume of the first chamber (51) increases (or as the rotation angle increases). Therefore, the amount of leakage of lubricant into the first chamber (51) indicated by the broken line in FIG. 6A (the sum of the amount of leakage through the gaps close to the suction chamber and the amount of leakage due to the roller trajectory leakage) increases as the volume of the first chamber (51) increases (or as the rotation angle increases). Specifically, the amount of leakage is smallest when the rotation angle of the drive shaft (70) is around 0°; then the amount of leakage increases as the rotation angle increases; and then the amount of leakage is largest when the rotation angle of the drive shaft (70) is 360°.
On the other hand, the amount of leakage through the gaps close to the compression chamber, which is indicated by the broken line in FIG 6B and leads to the second leakage loss, is determined by the difference between the pressure in the internal space (60) of the casing (10) (the pressure of lubricant) and the internal pressure of the second chamber (52) and the arc length of part of the outer peripheral surface of the roller (36) that faces the second chamber (52); and the amount of leakage increases as the volume of the second chamber (52) decreases (or as the rotation angle increases). Specifically, the amount of leakage is largest when the rotation angle of the drive shaft (70) is around 0°; then the amount of leakage decreases as the rotation angle increases; and then the amount of leakage becomes zero when the rotation angle of the drive shaft (70) is around 180° at which the discharge phase starts.
Therefore, when the rotation angle of the drive shaft (70) shifts from 0° to 180°, the amount of leakage of lubricant into the cylinder chamber (S) indicated by the broken line in FIG. 6C is significantly affected by the leakage through the gaps close to compression chamber indicated by the broken line in FIG. 6B; the amount of leakage is largest when the rotation angle of the drive shaft (70) is around 0°; then the amount of leakage decreases as the rotation angle increases; and then the amount of leakage is smallest when the rotation angle of the drive shaft (70) is around 180° at which the discharge phase starts. On the other hand, when the rotation angle of the drive shaft (70) shifts from around 180° at which the discharge phase starts to 360°, the leakage through the gaps close to compression chamber indicated by the broken line in FIG. 6B is zero, and therefore the amount of leakage of lubricant into the first chamber (51) indicated by the broken line in FIG. 6A is reflected in the amount of leakage as it is. Specifically, the amount of leakage is smallest when the rotation angle of the drive shaft (70) is around 180°; then the amount of leakage increases as the rotation angle increases; and then the amount of leakage is largest when the rotation angle of the drive shaft (70) is 360°. In this way, in the swing-type compressor (the rotary compressor) (1), the amount of leakage of lubricant into the cylinder chamber (S) through the gaps above and below the roller (36) increases and decreases while the drive shaft (70) rotates one time, and the amount of leakage becomes large particularly when the rotation angle of the drive shaft (70) is 0° (360°).
As indicated by the broken line in FIG. 6C, the amount of leakage of lubricant into the cylinder chamber (S) increases and decreases while the drive shaft (70) rotates one time, but the lubricant keeps leaking into the cylinder chamber (S) to cause the leakage loss while the rotation angle of the drive shaft (70) shifts from 0° to 360°. Therefore, in the first embodiment, the inclined surface (36b) is formed on an almost entire circumference of the roller (36) in the circumferential direction (while the angular range in which the vane (37) is present in the circumferential direction of the roller (36) is excluded) so that whatever the rotation angle of the drive shaft (70) may be, the roller (36) floats up to an appropriate height (the height at which the gaps above and below the roller (36) are substantially equal in height).
Here, assuming that the inclined surface (36b) is formed in the angular range of α to α + 180° in the circumferential direction of the roller (36), the wedge effect described above is obtained effectively when the rotation angle of the drive shaft (70) is α, and the roller (36) floats up effectively. In the first embodiment, the inclined surface (36b) is formed on an almost entire circumference of the roller (36) in the circumferential direction, and therefore the roller (36) floats up whatever the rotation angle of the drive shaft (70) may be.
In FIGS. 6A to 6C, the graphs with the solid lines show how the amount of leakage of lubricant changes while the drive shaft (70) rotates one time, in the rotary compressor (1) of the first embodiment in which the inclined surface (36b) is formed on an almost entire circumference of the roller (36) in the circumferential direction. The solid line in FIG. 6A indicates a change in the amount of leakage of lubricant into the first chamber (51) (the sum of the amount of leakage through the gaps close to the suction chamber and the amount of leakage due to the roller trajectory leakage). The solid line in FIG. 6B indicates a change in the amount of leakage of lubricant into the second chamber (52) (the amount of leakage through the gaps close to the compression chamber). The solid line in FIG. 6C shows the total sum of the amount of leakage indicated by the solid lines in FIGS. 6A and 6B, that is, a change in the amount of leakage of lubricant into the cylinder chamber (S).
As described above, the inclined surface (36b) is formed on an almost entire circumference of the roller (36) in the circumferential direction, whereby as shown in FIG. 6A, the amount of leakage of lubricant into the first chamber (51) (the sum of the amount of leakage through the gaps close to the suction chamber and the amount of leakage due to the roller trajectory leakage) changes from the graph of the broken line to the graph of the solid line and decreases significantly. Further, as shown in FIG. 6B, the amount of leakage of lubricant into the second chamber (52) (the amount of leakage of lubricant through the gaps close to the compression chamber) also changes from the graph of the broken line to the graph of the solid line and decreases significantly. Therefore, as shown in FIG. 6C, the amount of leakage of lubricant into the cylinder chamber (S) also changes from the graph of the broken line to the graph of the solid line and decreases significantly.

(7) First Method of Forming Inclined Surface

For various versions of the rotary compressors (1) with various sizes, the inventors of the present application have studied a method of easily forming the inclined surface (36b) that enables the roller (36) to float up to an appropriate height during the operation and have devised the following first method of forming the inclined surface.
The first method of forming the inclined surface is to derive Expression (1) below and then form (design) the inclined surface (36b) by using Expression (1), where Expression (1) shows the range of the cutting height X of the inclined surface (36b) (the distance in the axial direction of the drive shaft (70) between the outer end of the inclined surface (36b) and the horizontal surface (36a)) that enables the roller (36) to float up to an appropriate height during the operation by using the mass m of the piston (35) (the sum of the mass of the roller (36) and the mass of the vane (37)) and the difference ΔH between the height of the cylinder (34) and the height of the roller (36). $(ΔH \div 2 \times 0.75) \div (0.0055 \times m^{- 0.893}) \leq X \leq (ΔH \div 2 \times 1.25) \div (0.00006 m^{- 1.118})$
If the inclined surface (36b) is formed while the cutting angle φ of the inclined surface (36b) is an angle of 10° or more and 45° or less and the cutting height X of the inclined surface (36b) satisfies Expression (1), it is possible to easily form the rotary compressor (1) in which the roller (36) can float up to an appropriate height during the operation.

First, for the following five types of rotary compressor (1) with different sizes (the mass m of the piston (35) is different), the relation between the cutting height X of the inclined surface (36b) and the floating amount Y of the piston (35) has been estimated where the cutting angle φ of the inclined surface (36b) is set to 10°, 20°, 30°, and 45°. FIG. 7A shows the estimation result of type A where the piston (35) is lightest, and FIG. 7B shows the estimation result of type E where the piston (35) is heaviest. All the five types of rotary compressor (1) have the configuration described above and are different only in size (and thus in mass) from each other:

A: the rotary compressor where the mass m of the piston (35) is 0.01 kg;
B: the rotary compressor where the mass m of the piston (35) is 0.025 kg;
C: the rotary compressor where the mass m of the piston (35) is 0.05 kg;
D: the rotary compressor where the mass m of the piston (35) is 0.075 kg; and
E: the rotary compressor where the mass m of the piston (35) is 0.1 kg.

The estimation starts with calculating the upward force py which acts on the roller (36) due to the wedge effect when the rotary compressor (1) is operating in a state in which the floating amount of the piston (35) is zero (Step 1). The upward force py which acts on the roller (36) due to the wedge effect is calculated using Expressions (2) to (5) below. Expressions (2) to (5) below contain x, η, r, a, δ, θ, h, X, and φ, each defined as follows:

x: the distance (mm) from the outer peripheral surface of the roller (36) in the radial direction;
η: the viscosity of lubricant (mPa·s);
r: the eccentricity (mm) of the eccentric portion (72);
a: the outer peripheral radius (mm) of the roller (36);
δ: the angle (rad) of the inclined surface (36b) where the moving direction of the piston (35) is defined as 0° $(- π / 2 < δ < π / 2);$
θ: the rotation angle (rad) of the drive shaft (70);
h: the height (mm) of the inclined surface (36b) with respect to the horizontal surface (36a) at a position of the distance x from the outer peripheral surface of the roller (36) in the radial direction (the distance in the axial direction of the drive shaft (70));
X: the height (mm) of the inclined surface (36b) with respect to the horizontal surface (36a) at the outer peripheral surface of the roller (36) in the radial direction (the distance in the axial direction of the drive shaft (70)); and
φ: the cutting angle (°) of the inclined surface (36b). $\frac{d}{dx} (\frac{h^{3}}{η} \cdot \frac{dp}{dx}) = 6 U \frac{dh}{dx}$ $U = (r + a \cos δ) \cdot \cos δ \cdot dδ \cdot dθ$ $h = X - x \tan ϕ$ $p_{y} = p \cdot cosθ$

Then, the floating amount y of the piston (35) due to the calculated upward force py is determined (Step 2). Next, in a state in which the piston (35) floats up by the floating amount y, the upward force py which acts on the roller (36) due to the wedge effect is calculated (Step 3), and it is determined whether this upward force py and the mass m of the piston (35) are balanced (Step 4).
In Step 4, if the upward force py and the mass m of the piston (35) are balanced, the floating amount y of the piston (35) due to the upward force py is determined, and this determined floating amount y is defined as the floating amount Y of the piston (35).
On the other hand, in Step 4, if the upward force py and the mass m of the piston (35) are not balanced, the process returns to Step 2, and the floating amount y of the piston (35) due to the upward force py is determined again (Step 2). Then, in a state in which the piston (35) floats up by the floating amount y, the upward force py which acts on the roller (36) due to the wedge effect is calculated again (Step 3), and it is determined whether this upward force py and the mass m of the piston (35) are balanced (Step 4).
By repeating the above steps, the floating amount Y of the piston (35) is determined. Then, the relational expressions in which the floating amount Y determined by the above estimation at each cutting angle φ is expressed as a function of the cutting height X have been determined as follows. FIG. 7A shows the result of type A where the piston (35) is lightest, and
FIG. 7E shows the result of type E where the piston (35) is heaviest.

(7-1) type A: m = 0.01 $φ = 10 : Y = 0.3313 X$ $φ = 20 : Y = 0.0938 X$ $φ = 30 : Y = 0.0366 X$ $φ = 45 : Y = 0.0103 X$
(7-2) type B: m = 0.025 $φ = 10 : Y = 0.1532 X$ $φ = 20 : Y = 0.0395 X$ $φ = 30 : Y = 0.0149 X$ $φ = 45 : Y = 0.0041 X$
(7-3) type C: m = 0.05 $φ = 10 : Y = 0.0817 X$ $φ = 20 : Y = 0.0075 X$ $φ = 30 : Y = 0.0202 X$ $φ = 45 : Y = 0.002 X$
(7-4) type D: m = 0.075 $φ = 10 : Y = 0.0559 X$ $φ = 20 : Y = 0.0135 X$ $φ = 30 : Y = 0.0049 X$ $φ = 45 : Y = 0.0009 X$
(7-5) type E: m = 0.1 $φ = 10 : Y = 0.0424 X$ $φ = 20 : Y = 0.0101 X$ $φ = 30 : Y = 0.0037 X$ $φ = 45 : Y = 0.0009 X$

The estimation results of types A to E show that the smaller the mass m of the piston (35) and the smaller the cutting angle φ of the inclined surface (36b), the larger the slope of the line segment (the slope of the function) that indicates the relation between the floating amount Y and the cutting height X of the inclined surface (36b); and in contrast, the larger the mass m of the piston (35) and the larger the cutting angle φ of the inclined surface (36b), the smaller the slope of the line segment (the slope of the function) that indicates the relation between the floating amount Y and the cutting height X of the inclined surface (36b).
Accordingly, next, regarding types A to E, among the functions A10 to E45 determined at the cutting angles φ (10°, 20°, 30°, and 45°), the functions A10, B10, C10, D10, and E10 at the cutting angle φ of 10° where the slope I is maximum and the functions A45, B45, C45, D45, and E45 at the cutting angle φ of 45° where the slope I is minimum have been each employed to calculate the approximate expression that indicates the relation between the slope I and the mass m of the piston (35).
Specifically, as shown in FIG. 8A, the approximate expression which shows the relation between the slope I of each of the functions A10, B10, C10, D10, and E10 at the cutting angle φ of 10° and the mass m of the piston (35) and in which the slope I of the functions A10, B10, C10, D10, and E10 at the cutting angle φ of 10° is expressed as a function of the mass m of the piston (35) has been found to be I = 0.0055m^-0.893.
Specifically, as shown in FIG. 8B, the approximate expression which shows the relation at the cutting angle φ of 45° between the slope I of each of the functions A45, B45, C45, D45, and E45 and the mass m of the piston (35) and in which the slope I of the functions A45, B45, C45, D45, and E45 at the cutting angle φ of 45° is expressed as a function of the mass m of the piston (35) has been found to be I = 0.00006m^-1.118.
Next, the range of the cutting height X in which the floating amount Y can have an appropriate value has been determined by using the approximate expression I = 0.0055m^-0.893 of the slope I of the functions A10, B10, C10, D10, and E10 at the cutting angle φ of 10° and the approximate expression I = 0.00006m^-1.118 of the slope I of the functions A45, B45, C45, D45, and E45 at the cutting angle φ of 45°.
Specifically, the appropriate range of the floating amount Y has been defined as ΔH ÷ 2 × 0.75 ≤ Y ≤ ΔH ÷ 2 × 1.25; and the range of the cutting height X (Expression (1), or X1 ≤ X ≤ X2 in FIG. 9) that can be included in the hatched range surrounded as shown in FIG. 9 by a line segment L1 indicating the relation between the floating amount Y and the cutting height X at the cutting angle φ of 10° (Y = 0.0055 × m^-0.893·X), a line segment L2 indicating the relation between the floating amount Y and the cutting height X at the cutting angle φ of 45° (Y = 0.00006m^-1.118·X), a line segment L3 indicating the lower limit value of the appropriate range of the floating amount Y (Y = ΔH ÷ 2 × 0.75), and a line segment L4 indicating the upper limit value of the appropriate range of the floating amount Y (Y = ΔH ÷ 2 × 1.25) has been derived, and then this derived range has been defined as the range of the cutting height X in which the floating amount Y can have an appropriate value.

(8) Second Method of Forming Inclined Surface

Further, for various versions of the rotary compressors (1) with various sizes, the inventors of the present application have studied a different method from the first method of easily forming the inclined surface (36b) that enables the roller (36) to float up to an appropriate height during the operation and have devised the following second method of forming the inclined surface.
The second method of forming the inclined surface is to derive the relational expression below and then design the inclined surface (36b) by using the relational expression, where the relational expression shows the relation among the mass m of the piston (35), the cutting angle φ, and the cutting height X where the floating amount Y of the roller (36) can be included in the appropriate range (ΔH ÷ 2 × 0.75 ≤ Y ≤ ΔH ÷ 2 × 1.25) during the operation.
First, for three types of rotary compressor (1) where the mass m of the piston (35) is 0.01 kg, 0.05 kg, and 0.1 kg (type A, type C, and type E of the first method of forming the inclined surface), the appropriate range of the floating amount Y (ΔH ÷ 2 × 0.75 ≤ Y ≤ ΔH ÷ 2 × 1.25) has been determined as follows. Note that m = 0.01 kg has resulted in ΔH = 0.015; m = 0.05 kg has resulted in ΔH = 0.022; and m = 0.1 kg has resulted in ΔH = 0.028.

(8-1) type A: m = 0.01 $0.00563 \leq Y \leq 0.009375$
(8-2) type C: m = 0.05 $0.00825 \leq Y \leq 0.0138$
(8-3) type E: m = 0.1 $0.0105 \leq Y \leq 0.0175$

Next, the ranges of the cutting height X and the cutting angle φ that can be included in the appropriate range of the floating amount Y of (8-1) in the relational expression in which the floating amount Y for each cutting angle φ of type A in FIG. 7A is expressed as a function of the cutting height X, the ranges of the cutting height X and the cutting angle φ that can be included in the appropriate range of the floating amount Y of (8-2) in the relational expression in which the floating amount Y for each cutting angle φ of type C is expressed as a function of the cutting height X, and the ranges of the cutting height X and the cutting angle φ that can be included in the appropriate range of the floating amount Y of (8-3) in the relational expression in which the floating amount Y for each cutting angle φ of type E in FIG. 7B is expressed as a function of the cutting height X have been each determined as follows.

(8-4) m = 0.01 $10 ° \leq φ \leq 20 ° and 0.017 \leq X \leq 0.10$ $20 ° \leq φ \leq 30 ° and 0.060 \leq X \leq 0.26$ $30 ° < φ \leq 45 ° and 0.15 \leq X \leq 0.91$
(8-5) m = 0.05 $10 ° < φ \leq 20 ° and 0.10 \leq X \leq 0.68$ $20 ° < φ \leq 30 ° and 0.41 \leq X \leq 1.84$ $30 ° < φ \leq 45 ° and 1.10 \leq X \leq 5.00$
(8-6) m = 0.1 $10 ° < φ \leq 20 ° and 0.25 \leq X \leq 1.73$ $20 ° < φ \leq 30 ° and 1.04 \leq X \leq 4.70$ $30 ° < φ \leq 45 ° and 2.84 \leq X \leq 5.00$

From the above results, the relational expressions shown in (8-7) to (8-9) below have been derived as the relation among the mass m of the piston (35), the cutting angle φ, and the cutting height X where the floating amount Y can be included in an appropriate range. If the inclined surface (36b) is formed while the relational expressions shown in (8-7) to (8-9) are satisfied, it is possible to form the rotary compressor (1) in which the roller (36) can float up to an appropriate height during the operation.

(8-7) m ≤ 0.03 $10 ° \leq φ \leq 45 ° and 0.017 \leq X \leq 0.91$
(8-8) 0.03 < m ≤ 0.06 $10 ° \leq φ \leq 45 ° and 0.10 \leq X \leq 5.00$
(8-9) m > 0.06 $10 ° \leq φ \leq 45 ° and 0.25 \leq X \leq 5.00$

From the above results, it is possible to derive more detailed relational expressions as shown in (8-10) to (8-12) below than the relational expressions shown in (8-7) to (8-9). If the inclined surface (36b) is formed while the relational expressions shown in (8-10) to (8-12) are satisfied, it is possible to form the rotary compressor (1) in which the roller (36) can float up to a more appropriate height during the operation.

(8-10) m ≤ 0.03 $10 ° \leq φ \leq 20 ° and 0.017 \leq X \leq 0.10$ $20 ° < φ \leq 30 ° and 0.060 \leq X \leq 0.26$ $30 ° < φ \leq 45 ° and 0.15 \leq X \leq 0.9$
(8-11) 0.03 < m ≤ 0.06 $10 ° < φ \leq 20 ° and 0.10 \leq X \leq 0.68$ $20 ° < φ \leq 30 ° and 0.41 \leq X \leq 1.84$ $30 ° < φ \leq 45 ° and 1.1 \leq X \leq 5.0$
(8-12) m > 0.06 $10 ° < φ \leq 20 ° and 0.25 \leq X \leq 1.73$ $20 ° < φ \leq 30 ° and 1.04 \leq X \leq 4.7$ $30 ° < φ \leq 45 ° and 2.84 \leq X \leq 5.0$

-Advantages of First Embodiment-

In the rotary compressor (1) of the first embodiment, an outer peripheral part of the lower end surface of the roller (36) is formed as the inclined surface (the second surface) (36b) which is inclined so as to be positioned higher at a radially outer side. With this configuration, the wedge-shaped gap (90) is formed on the outer peripheral side between the lower end surface of the roller (36) and the rear head (the lower end plate) (42). With this gap (90), during the operation, the high-pressure lubricant equivalent in pressure to the inside of the casing (10) is supplied to each sliding portion of the compression mechanism (30), and part of the lubricant reaches the space (the cylinder chamber (S)) between the cylinder (34) and the roller (36). Then this lubricant having reached the cylinder chamber (S) is pushed into the radially inner side (the center of the roller (36)) through the outer peripheral inlet of the wedge-shaped gap (90) by eccentric rotation of the roller (36) in the cylinder (34). The high-pressure lubricant pushed in this manner exerts an upward force on the roller (36), which causes the roller (36) to float upward (the wedge effect). Therefore, according to the first embodiment, it is possible to provide the rotary compressor (1) in which the roller (36) can float up to an appropriate height during the operation. Further, if the roller (36) can float up to an appropriate height during the operation, it is possible to reduce the amount of leakage of lubricant into the cylinder chamber (S) through the gaps above and below the roller (36), and therefore it is possible to reduce the leakage loss.
Further, in the rotary compressor (1) of the first embodiment, since the inclined surface (36b) is formed on the outer peripheral side of the roller (36), the lubricant present between the cylinder (34) that does not rotate and the roller (36) that eccentrically rotates flows into the wedge-shaped gap (90), and the relative speed between the lubricant flowing into the gap (90) and a surface (the inclined surface) on which the lubricant acts becomes higher than when the inclined surface is formed on the inner peripheral side of the roller (36). Therefore, the upward force which acts on the roller (36) due to the wedge effect becomes larger, which enables the roller (36) to float upward enough. Therefore, according to the first embodiment, it is possible to provide the rotary compressor (1) in which the roller (36) can float up to an appropriate height during the operation.
Further, in the first embodiment, for various versions of the rotary compressor (1) with various sizes, the range of the height (the cutting height) X of the wedge-shaped gap (90) that enables the roller (36) to float up to an appropriate height during the operation is estimated; then, Expression (1) expressed using the mass m of the piston (35) (the sum of the mass of the roller (36) and the mass of the vane (37)) and the difference ΔH in height between the cylinder (34) and the roller (36) is derived; and then the inclined surface (36b) is formed so as to satisfy Expression (1). By forming (designing) the inclined surface (36b) in this manner, it is possible to easily form various versions of the rotary compressor (1) with various sizes in which the roller (36) can float up to an appropriate height during the operation.
In the first embodiment, the rotary compressor (1) is a swing-type compressor including the piston (35) in which the roller (36) and the vane (37) are integrated together. As shown in FIG. 6C, in the swing-type compressor, the amount of leakage of lubricant into the cylinder chamber (S) through the gaps above and below the roller (36) becomes large particularly when the rotation angle of the drive shaft (70) is 0° (360°) while the drive shaft (70) rotates one time.
Here, as described above, assuming that the inclined surface (36b) is formed in the angular range of α to α + 180° in the circumferential direction of the roller (36), the wedge effect described above is obtained effectively when the rotation angle of the drive shaft (70) is α, and the roller (36) floats up effectively.
Therefore, in the first embodiment, the inclined surface (36b) is formed in the range at least from the angular position of the side surface of the vane (37) that is on the suction side to the angular position of 180° in the circumferential direction of the roller (36). According to this configuration, the wedge effect is produced effectively when the rotation angle of the drive shaft (70) is 0°, and the roller (36) can float up reliably when the rotation angle of the drive shaft (70) is 0° at which the amount of leakage of lubricant into the cylinder chamber (S) is significantly large. Therefore, according to the first embodiment, it is possible to reduce the amount of leakage of lubricant into the cylinder chamber (S), and thus it is possible to reduce the leakage loss.
Further, in the first embodiment, the inclined surface (36b) is formed on an almost entire circumference of the roller (36) in the circumferential direction while the angular range in which the vane (37) is present in the circumferential direction of the roller (36) is excluded. Since the roller (36) can float up whatever the rotation angle of the drive shaft (70) may be, it is possible to reduce the amount of leakage of lubricant into the cylinder chamber (S), and therefore it is possible to reduce the leakage loss.

<<Second Embodiment>>

A second embodiment is a modified version of the first embodiment in which the configuration of the inclined surface (36b) of the rotary compressor (1) is modified.
Specifically, as shown in FIG. 10, in the second embodiment, the inclined surface (36b) is formed in the range from the angular position of the side surface of the vane (37) that is on the suction side to the angular position of 270° in the circumferential direction of the roller (36), unlike the first embodiment where the inclined surface (36b) is formed on an almost entire circumference of the roller (36) in the circumferential direction (while the angular range in which the vane (37) is present in the circumferential direction of the roller (36) is excluded). Therefore, in the second embodiment, the inclined surface (36b) is not formed in the range from the angular position of 270° to the angular position of the side surface of the vane (37) that is on the discharge side in the circumferential direction of the roller (36).
Here, as described above, from the viewpoint of reducing the amount of leakage of lubricant into the cylinder chamber (S) through the gaps above and below the roller (36) by enabling the roller (36) to float up to an appropriate height, the range in which the inclined surface (36b) is formed (the angular range in which the inclined surface (36b) is formed in the circumferential direction of the roller (36)) should be as large as possible in one preferred embodiment (an almost entire circumference similarly to the first embodiment). On the other hand, if the inclined surface (36b) is formed by chamfering the outer peripheral corner portion of the lower end portion of the roller (36) and then the wedge-shaped gap (90) is formed between the inclined surface (36b) and the upper surface of the rear head (42), it is possible to reduce the amount of leakage of lubricant into the cylinder chamber (S) through the gaps above and below the roller (36) by enabling the roller (36) to float up, but the refrigerant compressed in the second chamber (52) may leak into the first chamber (51) through the gap (90), which may cause performance degradation. As the range in which the inclined surface (36b) is formed is made larger, the leakage of a refrigerant through the gap (90) more likely occurs, which may more likely cause performance degradation. Therefore, from the viewpoint of reducing the leakage of a refrigerant from the second chamber (52) to the first chamber (51) through the gap (90), the range in which the inclined surface (36b) is formed should be as small as possible in one preferred embodiment.
Therefore, in the second embodiment, as described above, the inclined surface (36b) is formed only in the range from the angular position of the side surface of the vane (37) that is on the suction side to the angular position of 270° in the circumferential direction of the roller (36). Since the inclined surface (36b) formed in this manner enables the roller (36) to float up while the rotation angle of the drive shaft (70) is between 0° and 180° in which the amount of leakage of lubricant into the cylinder chamber (S) (mainly to the second chamber (52)) is large, it is possible to reduce the amount of leakage of lubricant into the cylinder chamber (S), and therefore it is possible to reduce the leakage loss. On the other hand, the inclined surface (36b) is not formed in the range from the angular position of 270° to the angular position of the side surface of the vane (37) that is on the discharge side in the circumferential direction of the roller (36), whereby it is possible to reduce the leakage of a refrigerant from the second chamber (52) to the first chamber (51) through the wedge-shaped gap (90) produced by forming the inclined surface (36b).

<<Other Embodiments>>

In the first and second embodiments, as one example of the rotary compressor (1) of the present disclosure, the swing-type compressor including the piston (35) in which the roller (36) and the vane (37) are integrated together has been described, but the rotary compressor (1) of the present disclosure is not limited to a swing-type compressor. The rotary compressor (1) of the present disclosure may be a so-called rolling-piston-type rotary compressor in which a roller eccentrically rotates while a vane separate from the roller contacts the roller, or a so-called hinge-vane-type rotary compressor in which a roller eccentrically rotates while the tip end of a vane is rotatably fitted in a recessed portion of the outer peripheral surface of the roller.
In the first and second embodiments, as one example of the rotary compressor (1) of the present disclosure, the so-called single-cylinder (one-cylinder) rotary compressor including one cylinder (34) has been described, but the rotary compressor (1) of the present disclosure is not limited to a single-cylinder compressor. The rotary compressor (1) of the present disclosure may be a so-called multi-cylinder rotary compressor (1) including a plurality of cylinders (34). If the rotary compressor (1) is a multi-cylinder compressor, the lower end surface of each of the rollers (36) eccentrically rotating in the cylinders (34) may have the inclined surface (the second surface) (36b). According to this configuration, the rollers (36) float up to an appropriate height in the cylinders (34) during the operation.
In the second embodiment, the inclined surface (36b) is formed only in the range from the angular position of the side surface of the vane (37) that is on the suction side to the angular position of 270° in the circumferential direction of the roller (36), and is not formed in the range from the angular position of 270° to the angular position of the side surface of the vane (37) that is on the discharge side. If the inclined surface (36b) is formed only in part of the angular range of the roller (36) in the circumferential direction as described above, the beginning of the inclined surface (36b) is provided in one preferred embodiment at the angular position of the side surface of the vane (37) that is on the suction side in the circumferential direction of the roller (36) similarly to the second embodiment, but the end of the inclined surface (36b) may be provided at a predetermined angular position of 180° or more and 270° or less in the circumferential direction of the roller (36) and is not limited to the position of 270°. If the beginning of the inclined surface (36b) is provided at the angular position of the side surface of the vane (37) that is on the suction side in the circumferential direction of the roller (36) and the end of the inclined surface (36b) is provided at a predetermined angular position of 180° or more and 270° or less in the circumferential direction of the roller (36), the same advantages as those of the second embodiment can be obtained.
While the embodiment and variations have been described above, it will be understood that various changes in form and details can be made without departing from the spirit and scope of the claims. The above embodiment and variations may be appropriately combined or replaced as long as the functions of the target of the present disclosure are not impaired. The expressions of "first", "second", . . . described above are used to distinguish the terms to which these expressions are given, and do not limit the number and order of the terms.

INDUSTRIAL APPLICABILITY

As described above, the present disclosure is useful for a rotary compressor and a refrigeration cycle apparatus.

DESCRIPTION OF REFERENCE CHARACTERS

1: Rotary Compressor
9: Refrigerant Circuit
10: Casing
20: Motor
30: Compression Mechanism
34: Cylinder
35: Piston
36: Roller
36a: Horizontal Surface (First Surface)
36b: Inclined Surface (Second Surface)
36o: Outer Peripheral Surface
37: Vane
41: Front Head (Upper End Plate)
42: Rear Head (Lower End Plate)
51: First Chamber
52: Second Chamber
70: Drive Shaft
100: Refrigeration Cycle Apparatus

Claims

A rotary compressor comprising:
a casing (10);

a drive shaft (70) provided in the casing (10) and extending in a top-bottom direction;

a motor (20) provided in the casing (10) and configured to drive the drive shaft (70);
and

a compression mechanism (30) provided in the casing (10); coupled to the drive shaft (70); and driven and rotated by the motor (20) to compress a refrigerant,

wherein

the compression mechanism (30) has
a cylinder (34) having a cylindrical shape,

an upper end plate (41) and a lower end plate (42) closing an upper opening end surface and a lower opening end surface of the cylinder (34),

a cylindrical roller (36) attached to the drive shaft (70) and configured to eccentrically rotate in the cylinder (34), and

a vane (37) dividing a space between the cylinder (34) and the roller (36) into a first chamber (51) on a suction side and a second chamber (52) on a discharge side,

a lower end surface of the roller (36) has a first surface (36a) parallel to an upper surface of the lower end plate (42) and a second surface (36b) continuous with an outer peripheral side of the first surface (36a) and extending to an outer peripheral surface (36o) of the roller (36), and

the second surface (36b) is inclined so as to be positioned higher at a radially outer side.
The rotary compressor of claim 1, wherein
the second surface (36b) satisfies Expression (1) below: $(ΔH \div 2 \times 0.75) \div (0.0055 \times m^{- 0.893}) \leq X \leq (ΔH \div 2 \times 1.25) \div (0.00006 \times m^{- 1.118})$

where

a sum of a mass of the roller (36) and a mass of the vane (37) is m,

a distance in an axial direction of the drive shaft (70) between an outer end of the second surface (36b) and the first surface (36a) is X, and

a difference between a height of the cylinder (34) and a height of the roller (36) is ΔH.
The rotary compressor of claim 1 or 2, wherein
the rotary compressor is a swing-type compressor including a piston (35) in which the roller (36) and the vane (37) are integrated together, and

assuming that an angular position of a center line (M) of the vane (37) with respect to a center of the roller (36) in a circumferential direction of the roller (36) is 0° and that an angular position of the roller (36) in the circumferential direction increases while moving in a rotation direction of the drive shaft (70),

the second surface (36b) is formed in a range at least from an angular position of a side surface of the vane (37) that is on a suction side to an angular position of 180° in the circumferential direction of the roller (36).
The rotary compressor of claim 3, wherein
the second surface (36b) is formed in a range from the angular position of the side surface of the vane (37) that is on the suction side to an angular position of a side surface of the vane (37) that is on a discharge side in the circumferential direction of the roller (36).
The rotary compressor of claim 3, wherein
the second surface (36b) is formed in a range from the angular position of the side surface of the vane (37) that is on the suction side to a predetermined angular position of 180° or more and 270° or less in the circumferential direction of the roller (36).
A refrigeration cycle apparatus comprising: the rotary compressor of any one of claims 1 to 5.