JP7853624B2 - Rotary compressor and refrigeration cycle system - Google Patents

Description

This disclosure relates to a rotary compressor and a refrigeration cycle system equipped therewith. A rotary compressor is a compressor that compresses the gas in a cylinder chamber formed within a cylinder by eccentrically rotating a roller within the cylinder. Rotary compressors generally have vanes for partitioning the cylinder chamber. Rotary compressors include so-called rolling piston type, where separate vanes contact the roller while the roller rotates eccentrically; so-called swing type, where vanes integrally formed with the roller oscillate in conjunction with the roller's eccentric rotation; and so-called hinge vane type, where the tip of the vane is rotatably fitted into a recess on the outer surface of the roller while the roller rotates eccentrically.

Conventionally, rotary compressors have been used in which a refrigerant is compressed by the eccentric rotation of a cylindrical roller within a cylinder chamber formed by a cylindrical cylinder and end plates that close the upper and lower ends of the cylinder (see, for example, Patent Document 1). In the above rotary compressor, the compressed refrigerant is discharged into the casing through a port that penetrates the end plates. Therefore, the outer circumference of the compression mechanism having the cylinder, roller, and end plates is a high-pressure space.

In the rotary compressor described above, chamfers are provided on both the upper corner formed between the inner circumferential surface and the upper end surface of the roller, and the lower corner formed between the inner circumferential surface and the lower end surface of the roller, so that high-pressure lubricating oil is supplied to the upper and lower chamfered portions. Furthermore, in the rotary compressor described above, the horizontal projected area of the chamfered surface of the lower corner is made greater than or equal to the horizontal projected area of the chamfered surface of the upper corner, so that the vertical component of the force exerted on the roller by the high-pressure lubricating oil is greater on the chamfered surface of the lower corner than on the chamfered surface of the upper corner. In the rotary compressor described above, with the above configuration, during operation when the eccentric portion of the drive shaft and the fitting portion of the roller are filled with high-pressure lubricating oil, the vertically upward force exerted on the roller by the high-pressure lubricating oil supplied to the upper and lower chamfered portions is greater than the vertically downward force, thereby pushing the roller upward from below, equalizing the gap between the upper and lower end surfaces of the roller and the end plate facing them, and reducing leakage loss at the roller end surface.

Japanese Patent Publication No. 2010-31733

However, in the rotary compressor described above, in principle, the force from the high-pressure lubricating oil acts not only on the chamfered portions of the upper and lower ends of the roller, but on the entire surface of both the upper and lower ends of the roller. Therefore, even if chamfers are provided on the upper and lower corners of the rotary compressor, the projected area of the upper and lower ends of the roller, including the chamfered portions, is approximately equal. As a result, it is believed that the high-pressure lubricating oil does not generate a vertically upward force that would lift the roller. Consequently, the rotary compressor has the problem of not being able to achieve uniformity in the gap between the upper and lower ends of the roller and the opposing end plates, thus failing to adequately reduce leakage losses.

The purpose of this disclosure is to provide a rotary compressor capable of raising the rollers to an appropriate height during operation.

A first aspect of this disclosure is,
Casing (10),
A drive shaft (70) is provided within the casing (10) and extends in the vertical direction,
A motor (20) is provided within the casing (10) and drives the drive shaft (70),
The device comprises a compression mechanism (30) provided within the casing (10), connected to the drive shaft (70), and rotationally driven by the motor (20) to compress the refrigerant.
The above compression mechanism (30) is
A cylindrical cylinder (34) and
The upper end plate (41) and lower end plate (42) close the upper and lower open end faces of the cylinder (34),
A cylindrical roller (36) is attached to the drive shaft (70) and rotates eccentrically within the cylinder (34),
The space between the cylinder (34) and the roller (36) is divided into a first chamber (51) on the intake side and a second chamber (52) on the discharge side by a vane (37).
The lower end surface of the roller (36) has a first surface (36a) parallel to the upper surface of the lower end plate (42), and a second surface (36b) that is continuous with the outer circumference of the first surface (36a) and extends to the outer circumference surface (36o) of the roller (36).
The second surface (36b) described above is a rotary compressor that is inclined so that it is positioned higher as it moves radially outward.

In the first embodiment, a portion of the outer circumference of the lower end surface of the roller (36) is made into a second surface (36b) that is inclined upward as it moves radially outward, thereby forming a wedge-shaped gap (90) on the outer circumference between the lower end surface of the roller (36) and the lower end plate (42). With this configuration, during operation, high-pressure lubricating oil equivalent to the pressure inside the casing (10) is supplied to each sliding part of the compression mechanism (30), and some of it reaches the space between the cylinder (34) and the roller (36). As the roller (36) rotates eccentrically within the cylinder (34), the lubricating oil that reaches this space is pushed from the outer circumference entrance to the back (towards the center of the roller (36)) in the wedge-shaped gap (90). The high-pressure lubricating oil pushed in in this way acts an upward force on the roller (36), causing the roller (36) to lift up (wedge effect). Therefore, according to the first embodiment, a rotary compressor (1) can be provided that can raise the roller (36) to an appropriate height during operation.

A second aspect of this disclosure is, in the first aspect,
The second surface (36b) above is,
The sum of the weight of the roller (36) and the weight of the vane (37) is m.
Let X be the distance between the outer edge of the second surface (36b) and the first surface (36a) along the axial direction of the drive shaft (70).
When the difference between the height of the cylinder (34) and the height of the roller (36) is ΔH,
It satisfies the following formula (1).
(ΔH÷2×0.75)÷(0.0055×m ^-0.893 )≦X≦(ΔH÷2×1.25)÷(0.00006m ^-1.118 )...(1)

The inventors of this application calculated the range of wedge-shaped gap heights (the distance along the axial direction of the drive shaft (70) between the outer edge of the second surface (36b) and the first surface (36a)) that would allow the roller (36) to float to an appropriate height during operation for various rotary compressors of different sizes. They then derived the above formula (1), which is expressed using the sum of the weights of the roller (36) and the vane (37), and the height difference between the cylinder (34) and the roller (36).

In the second embodiment, by forming the second surface (36b) of rotary compressors of various sizes to satisfy the above formula (1), rotary compressors of various sizes that can levitate the rollers (36) to an appropriate height during operation can be easily constructed.

A third aspect of this disclosure is a first or second aspect,
The rotary compressor described above is a swing-type compressor equipped with a piston (35) in which the roller (36) and the vane (37) are integrally formed.
When 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 is set to 0°, and the angular position of the roller (36) in the circumferential direction increases as it moves in the rotational direction of the drive shaft (70),
The second surface (36b) described above is a rotary compressor formed over a range of at least 180° from the angular position of the suction-side surface of the vane (37) in the circumferential direction of the roller (36).

The third embodiment is a swing-type compressor equipped with a piston (35) in which the roller (36) and vane (37) are integrally formed. In a swing-type compressor, as shown in Figure 6C, the amount of lubricating oil leaking from the gaps above and below the roller (36) into the cylinder chamber is particularly large when the rotation angle of the drive shaft (70) is 0° (360°).

Incidentally, if a second surface (36b) is formed in the circumferential direction of the roller (36) within an angular range from α to α + 180°, the aforementioned wedge effect is effectively obtained when the rotation angle of the drive shaft (70) is α, and the roller (36) effectively levitates.

Therefore, in the third embodiment, the second surface (36b) is formed over a range of at least 180° from the angular position of the intake-side surface of the vane (37) in the circumferential direction of the roller (36). According to the above configuration, a wedge effect is effectively generated when the rotation angle of the drive shaft (70) is 0°, and the roller (36) can be reliably levitated when the rotation angle of the drive shaft (70), which is particularly high in terms of lubricating oil leakage into the cylinder chamber, is 0°. Therefore, according to the third embodiment, the amount of lubricating oil leakage into the cylinder chamber can be reduced, and leakage losses can be reduced.

A fourth aspect of this disclosure is, in the third aspect,
The second surface (36b) described above is formed over a range from the angular position of the suction-side surface of the vane (37) to the angular position of the discharge-side surface of the vane (37) in the circumferential direction of the roller (36).

In the fourth embodiment, the second surface (36b) is formed around almost the entire circumference of the roller (36), except for an angular range where the circumferential vanes (37) of the roller (36) are located. Since the roller (36) can be levitated regardless of the rotation angle of the drive shaft (70), the amount of lubricating oil leaking into the space between the cylinder (34) and the roller (36) (cylinder chamber) can be reduced, thereby reducing leakage losses.

A fifth aspect of this disclosure is, in the fourth aspect,
The second surface (36b) is formed over a range from an angular position on the suction side of the vane (37) in the circumferential direction of the roller (36) to a predetermined angular position of 180° to 270°.

Incidentally, from the viewpoint of reducing the amount of lubricating oil leaking into the cylinder chamber from the gaps formed above and below the roller (36) by raising the roller (36) to an appropriate height, it is preferable to make the formation range of the second surface (36b) (the angular range in the circumferential direction of the roller (36) from which the second surface (36b) is formed) as wide as possible. On the other hand, if the outer peripheral corner of the lower end of the roller (36) is chamfered to form the second surface (36b), and a wedge-shaped gap (90) is formed between the second surface (36b) and the upper surface of the lower end plate (42), the amount of lubricating oil leaking into the cylinder chamber from the gaps above and below the roller (36) can be reduced by raising the roller (36), but there is a risk that the refrigerant compressed in the second chamber (52) will leak into the first chamber (51) through the gap (90), leading to a decrease in performance. The wider the formation range of the second surface (36b), the more likely refrigerant leakage through the gap (90) will occur, and the higher the risk of a decrease in performance. Therefore, from the viewpoint of reducing refrigerant leakage from the second chamber (52) to the first chamber (51) through the gap (90), it is preferable to make the formation area of the second surface (36b) as narrow as possible.

Therefore, in the fifth embodiment, the second surface (36b) is formed only in the range from the angular position of the suction-side side of the vane (37) in the circumferential direction of the roller (36) to a predetermined angular position (180° to 270°), and the second surface (36b) is not formed in the range from the predetermined angular position in the circumferential direction of the roller (36) to the angular position of the discharge-side side of the vane (37). With this configuration, according to the fifth embodiment, refrigerant leakage from the second chamber (52) to the first chamber (51) through the wedge-shaped gap (90) caused by the formation of the second surface (36b) can be reduced.

A sixth aspect of this disclosure is a refrigeration cycle system comprising a rotary compressor according to any one of the first to fifth aspects.

Figure 1 is a piping diagram of a refrigeration cycle system according to Embodiment 1. Figure 2 is a longitudinal cross-sectional view of a rotary compressor according to Embodiment 1. Figure 3 is a cross-sectional view of the compression mechanism of a rotary compressor. Figure 4 is an enlarged longitudinal cross-sectional view of a part of a rotary compressor. Figure 5 shows the operation of the compression mechanism. Figure 6A is a graph showing the change in the amount of lubricating oil leaking from the gap between the rotor and the first chamber during one rotation of the drive shaft. The dashed line represents the case where an inclined surface is not formed on the lower surface of the rotor, and the solid line represents the case where an inclined surface is formed. Figure 6B is a graph showing the change in the amount of lubricating oil leaking from the gap between the rotor and the second chamber during one rotation of the drive shaft. The dashed line represents the case where an inclined surface is not formed on the lower surface of the rotor, and the solid line represents the case where an inclined surface is formed. Figure 6C is a graph showing the change in the amount of lubricating oil leaking from the gap between the rotor and the cylinder chamber during one rotation of the drive shaft. The dashed line represents the case where no inclined surface is formed on the lower surface of the rotor, and the solid line represents the case where an inclined surface is formed. Figure 7A is a graph showing the estimated relationship between the cutting height and the amount of floating material for each cutting angle in a rotary compressor with a piston mass of 0.01 kg. Figure 7B is a graph showing the estimated relationship between the cutting angle, cutting height, and buoyancy amount for a rotary compressor with a piston mass of 0.1 kg. Figure 8A is a graph showing the relationship between the slope and the weight of the piston when the amount of levitation is expressed as a function of the cutting height for a cutting angle of 10°. Figure 8B is a graph showing the relationship between the slope and the weight of the piston when the amount of levitation is expressed as a function of the cutting height at a cutting angle of 45°. Figure 9 is a graph showing the relationship between the cutting height and the amount of material to be buoyed, as well as the appropriate range of material to be buoyed, when the cutting angles are 10° and 45°. Figure 10 is a cross-sectional view of the compression mechanism of a rotary compressor according to Embodiment 2.

Embodiment 1
Embodiments of the present invention will be described below with reference to the drawings. The following embodiments are essentially preferred examples and are not intended to limit the scope of the present invention, its applications, or its uses. Furthermore, the embodiments, modifications, and other examples described below can be combined or partially replaced to the extent that the present invention is implementable. "Up" and "down" refer to the direction when the rotary compressor (1) is viewed from the front (see Figure 2). In some cases, hatching is omitted in the figures to facilitate understanding of the explanation.

(1) Refrigeration cycle device As shown in Figure 1, the rotary compressor (1) in this example is applied to the refrigeration cycle device (100). Hereafter, the rotary compressor (1) may be simply referred to as the compressor (1). The refrigeration cycle device (100) is, for example, an air conditioning system that provides air conditioning for a room.

The refrigeration cycle system (100) comprises an outdoor unit (7) located outside the room and an indoor unit (8) located inside the room. The outdoor unit (7) houses a 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) houses an indoor heat exchanger (6).

The refrigeration cycle unit (100) includes a refrigerant circuit (9). A compressor (1), a four-way switching valve (3), an outdoor heat exchanger (4), an expansion valve (5), and an indoor heat exchanger (6) are connected to the refrigerant circuit (9). The refrigeration cycle is performed by the flow of refrigerant through the refrigerant circuit (9).

The refrigeration cycle system (100) performs heating and cooling operations by switching the four-way directional control valve (3). In cooling operation, the first refrigeration cycle is performed. Specifically, when the first port (p1) and the third port (p3) of the four-way directional control valve (3) are connected, and the second port (p2) and the fourth port (p4) are connected (solid line in Figure 1), the indoor heat exchanger (6) functions as an evaporator, and the outdoor heat exchanger (4) functions as a radiator. In heating operation, the second refrigeration cycle is performed. Specifically, when the first port (p1) and the fourth port (p4) of the four-way directional control valve (3) are connected, and the second port (p2) and the third port (p3) are connected (dashed line in Figure 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 Figure 2, the rotary compressor (1) comprises a casing (10), a drive shaft (70), a motor (20), and a compression mechanism (30). The drive shaft (70), motor (20), and compression mechanism (30) are housed within the casing (10). The rotary compressor (1) is configured as a so-called high-pressure dome type, in which the refrigerant compressed in the compression mechanism (30) is discharged into the internal space (60) of the casing (10), and the internal space (60) becomes high-pressure.

(2-1) Casing The casing (10) is formed in an elongated shape. Specifically, the casing (10) comprises a cylindrical body (11) extending in the vertical direction, an upper end plate (12) that closes the upper end of the body (11), and a lower end plate (13) that closes the lower end of the body (11). The upper end plate (12) and the lower end plate (13) are formed to be relatively thick. An intake pipe (14) is provided at the lower part of the body (11), and a discharge pipe (15) is inserted through the upper part of the upper end plate (12). Lubricating oil is stored at the bottom (lower part) of the casing (10), forming an oil reservoir.

(2-2) Drive shaft The drive shaft (70) is arranged to extend vertically within the casing (10). The upper part of the drive shaft (70) is connected to the rotor (22) of the motor (20), which will be described later. The lower part of the drive shaft (70) has, from top to bottom, an upper main shaft portion (71), an eccentric portion (72), and a lower main shaft portion (73). The axes of the upper main shaft portion (71) and the lower main shaft portion (73) coincide. The eccentric portion (72) is formed to have a larger diameter than the upper main shaft portion (71) and the lower main shaft portion (73), and is eccentric with respect to the axes of the upper main shaft portion (71) and the lower main shaft portion (73). The piston (35) of the compression mechanism (30), which will be described later, is connected to the eccentric portion (72). The drive shaft (70) is driven by the motor (20) and causes the piston (35) to rotate eccentrically.

An oil supply passage (80) is formed inside the lower part of the drive shaft (70) to guide lubricating oil to each sliding part of the compression mechanism (30). The oil supply passage (80) has one main passage (81) extending vertically and multiple branch passages (82) extending radially from the main passage (81) to the outer circumferential surface of the drive shaft (70). In this embodiment 1, one branch passage (82) is provided at the upper side of the compression mechanism (30) at the midpoint of the upper main shaft portion (71) in the vertical direction, and another at the midpoint of the eccentric portion (72) in the vertical direction. An oil tube (not shown) is provided at the lower end of the drive shaft (70) to pump lubricating oil stored at the bottom of the casing (10) into 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). It is positioned above the compression mechanism (30) within the casing (10). The motor (20) has a cylindrical stator (21) along the inner circumferential surface of the body (11) and a rotor (22) positioned inside the stator (21). The stator (21) is fixed to the inner surface of the body (11) by welding. Note that the method of fixing the stator (21) to the body (11) is not limited to welding, but may also be shrink-fitted or press-fitted.

(2-4) Compression Mechanism As shown in Figures 2 to 4, the compression mechanism (30) is housed within the casing (10). The compression mechanism (30) compresses the inhaled refrigerant and discharges it into the internal space (60) of the casing (10). The compression mechanism (30) comprises a cylinder (34), a front head (41), a rear head (42), and a piston (35). The compression mechanism (30) is fixed to the inner surface of the body (11) by welding.

(2-4-1) Cylinder The cylinder (34) is a thick-walled disc-shaped member. As shown in Figure 3, the cylinder (34) has a cylinder bore (31), a vane housing hole (32), and an intake port (55).

The cylinder bore (31) is a circular hole that penetrates the cylinder (34) in the thickness direction. The cylinder bore (31) is formed in the central part of the cylinder (34). The cylinder bore (31) houses the main rollers (36) of the piston (35).

The cylinder (34) has a cylinder chamber (S) inside. Specifically, the cylinder chamber (S) is formed between the inner circumferential surface of the cylinder (34) that demarcates the outer edge of the cylinder bore (31) and the piston (35).

The vane housing hole (32) is a hole that extends radially outward from the inner circumferential surface of the cylinder (34) (i.e., the outer edge of the cylinder bore (31)). This vane housing hole (32) penetrates the cylinder (34) in the thickness direction. The vanes (37) of the piston (35) are housed in the vane housing hole (32).

(2-4-2) Front Head As shown in Figures 2 and 4, the front head (41) is an upper end plate that closes the axial upper end (upper open end face) of the cylinder (34). Specifically, the front head (41) closes the axial end of the cylinder (34) on the motor (20) side (the upper open end face of the cylinder (34) in Figure 1). The front head (41) is an example of an upper end plate (41) of the present disclosure. The front head (41) comprises a first main body portion (41a) and an upper bearing portion (41b). The first main body portion (41a) and the upper bearing portion (41b) are integrally molded.

The first main body (41a) is formed in a generally circular, thick plate shape, with the outer circumference being thicker in the vertical direction than the inner circumference. The lower surface of the first main body (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 main body (41a) toward the motor (20) side (upper side in Figure 1). The upper bearing portion (41b) is positioned in the center of the first main body (41a). The upper bearing portion (41b) rotatably supports the upper main shaft portion (71) of the drive shaft (70).

A discharge port (24) is formed in the front head (41). The discharge port (24) penetrates the first main body portion (41a) of the front head (41) vertically and opens at the top surface. The discharge port (24) communicates with the second chamber (52), described later, in the cylinder chamber (S). A reed valve (not shown) is provided in the discharge port (24). The reed valve is configured such that when the refrigerant pressure in the second chamber (52) exceeds a predetermined value, a portion of the reed valve that closes the opening end face of the discharge port (24) separates from the opening end face, thereby discharging the refrigerant.

The front head (41) is provided with a bowl-shaped discharge muffler (25) positioned above the first main body (41a) and surrounding the upper bearing portion (41b). The discharge muffler (25) is positioned to cover the discharge port (24) and the reed valve (not shown). The discharge muffler (25) has an outlet (not shown) for allowing the refrigerant discharged into the internal muffler space (26) to flow out. The discharged refrigerant, discharged from the discharge port (24) into the muffler space (26) within the discharge muffler (25), flows out through the outlet into the internal space (60) of the casing (10).

(2-4-3) Rear Head The rear head (42) is a lower end plate that closes the axial lower end (lower open end face) of the cylinder (34). Specifically, the rear head (42) closes the end of the cylinder (34) opposite to the motor (20) in the axial direction (the lower open end face of the cylinder (34) in Figure 1). The rear head (42) comprises a second main body (42a) and a lower bearing portion (42b).

The second main body (42a) is formed in a generally circular, thick plate shape. The upper surface of the second main body (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 main body (42a) to the side opposite to the cylinder (34) (the lower side in Figure 2). The lower bearing portion (42b) is positioned in the center of the second main body (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 Figure 3, the piston (35) is housed in the cylinder (34) (cylinder bore (31)) and rotates eccentrically within the cylinder (34). The piston (35) is configured to slide against both the front head (41) and the rear head (42). The piston (35) has rollers (36) and vanes (37).

The roller (36) is formed in an annular shape. Specifically, the roller (36) is formed in a slightly thick-walled cylindrical shape. The eccentric portion (72) of the drive shaft (70) is slidably inserted into it. The roller (36) is configured to revolve along the inner circumferential surface of the cylinder (34) as the drive shaft (70) rotates. The roller (36) is formed such that its height hp is lower than the height hc of the cylinder (34) so that it can slide within the cylinder (34).

As will be described in detail later, as shown in Figure 4, an inclined surface (36b) is formed on the lower end of the roller (36) by chamfering the outer corner. In other words, the lower end surface of the roller (36) is configured to have a horizontal surface (first surface) (36a) parallel to the upper surface of the rear head (42), and an inclined surface (second surface) (36b) that extends continuously from the outer periphery of the horizontal surface (36a) to the outer periphery surface (36o) of the roller (36). On the other hand, no inclined surface is formed on the upper end of the roller (36), 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 Figure 3, the vane (37) is formed integrally with the roller (36). The vane (37) protrudes radially outward from the outer circumferential surface of the roller (36). The vane (37) fits into the vane housing hole (32). The vane (37) is sandwiched between a pair of oscillating bushes (54a, 54b) provided in a bush groove (53) extending radially outward from the inner circumferential surface of the cylinder (34). The vane (37) is configured to restrict the rotation of the roller (36) when the roller (36) revolves. The vane (37) divides the cylinder chamber (S) into a first chamber (51) on the intake side and a second chamber (52) on the discharge side.

(3) Operation In the compressor (1), when the motor (20) is started and the rotor (22) is rotated, the drive shaft (70) rotates as shown in Figure 5, and the eccentric part (72) rotates eccentrically inside the cylinder (34) (cylinder chamber (S)). As the eccentric rotation of the eccentric part (72) occurs, the roller (36) of the piston (35) revolves along the inner surface of the cylinder (34) while restricting its rotation.

The position of the contact point (CP) between the roller (36) and the cylinder (34) will now be explained. As shown in Figure 3, assuming that the angular position of the center line (M) of the vane (37) relative to the center of the roller (36) in the circumferential direction is 0°, and that the angular position of the roller (36) in the circumferential direction increases as it moves in the rotational direction of the drive shaft (70), then, as shown in Figure 5, the roller (36) contacts the cylinder (34) at an angular position equal to the rotation angle of the drive shaft (70). In other words, 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° (Figure 5(A)), the roller (36) contacts the cylinder (34) at the base of the vane (37) at the 0° angular position. When the rotation angle of the drive shaft (70) is 45° (Figure 5(B)), the roller (36) contacts the cylinder (34) at the 45° angular position. When the rotation angle of the drive shaft (70) is 90° (Figure 5(C)), the roller (36) contacts the cylinder (34) at the 90° angular position. When the rotation angle of the drive shaft (70) is 135° (Figure 5(D)), the roller (36) contacts the cylinder (34) at the 135° angular position. When the drive shaft (70) rotates at a 180° angle (Figure 5(E)), the roller (36) contacts the cylinder (34) at an angular position of 180°. When the drive shaft (70) rotates at a 225° angle (Figure 5(F)), the roller (36) contacts the cylinder (34) at an angular position of 225°. When the drive shaft (70) rotates at a 270° angle (Figure 5(G)), the roller (36) contacts the cylinder (34) at an angular position of 270°. When the drive shaft (70) rotates at a 315° angle (Figure 5(H)), the roller (36) contacts the cylinder (34) at an angular position of 315°.

The suction stroke, in which refrigerant is drawn into the cylinder chamber (S), will now be described. When the drive shaft (70) rotates slightly from a state of 0° (state shown in Figure 5(A)) and exceeds the suction start angle, the suction port (55) and the first chamber (51) communicate, and the suction of refrigerant into the first chamber (51) begins. The suction start angle is the rotation angle of the drive shaft (70) when the roller (36) contacts the cylinder (34) at the first position (P1), where the angle of the inner circumference end of the suction port (55) is smallest (when the contact point (CP) between the roller (36) and the cylinder (34) is at the first position (P1)).

Refrigerant is drawn in through the suction pipe (14) and the suction port (55). As the rotation angle of the drive shaft (70) increases, the volume of the first chamber (51) gradually increases, and the amount of refrigerant drawn into the first chamber (51) increases (as shown in Figures 5(B) to (H)). When the rotation angle of the drive shaft (70) exceeds 360° and then the suction termination angle, the first chamber (51) becomes the second chamber (52). The first chamber (51) remains in communication with the suction port (55) until the rotation angle of the drive shaft (70) reaches the suction termination angle. Therefore, the suction stroke continues until the rotation angle of the drive shaft (70) reaches the suction termination angle. The suction termination angle is the rotation angle of the drive shaft (70) when the contact point (CP) between the roller (36) and the cylinder (34) is at the second position (P2), where the angle is maximum at the inner circumference of the suction port (55).

The compression stroke, in which the refrigerant is compressed in the cylinder chamber (S), is described below. When the rotation angle of the drive shaft (70) exceeds the intake termination angle, the process transitions from the intake stroke to the compression stroke. The second chamber (52) is shut off from the intake port (55), and compression of the refrigerant in the second chamber (52) begins. As the rotation angle of the drive shaft (70) increases further, 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 stroke ends, and the process transitions to the discharge stroke.

The discharge stroke for discharging refrigerant from the cylinder chamber (S) is described below. 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. When the reed valve opens, the refrigerant in the second chamber (52) is discharged from the discharge port (24) into the muffler space (26) within the discharge muffler (25). The discharged refrigerant in the muffler space (26) flows into the internal space (60) within the casing (10) from the outlet (not shown) of the discharge muffler (25), and is then discharged to the outside of the compressor (1) via the discharge pipe (15). This refrigerant discharge stroke continues until the rotation angle of the drive shaft (70) reaches 360°.

As described above, in the compressor (1), the refrigerant is continuously compressed in the cylinder chamber (S) by repeating the intake stroke, compression stroke, and discharge stroke.

(4) Regarding the gap between the roller and the cylinder As described above, the roller (36) is formed such that its height hp is lower than the height hc of the cylinder (34) so that it can slide within the cylinder (34). Therefore, a gap is formed 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). This gap formed above and below the roller (36) is sealed by supplying high-pressure lubricating oil (lubricating oil in the oil reservoir) stored at the bottom of the casing (10) via the oil supply passage (80). Specifically, as the drive shaft (70) rotates, lubricating oil is pumped up to the main passage (81) by an oil tube (not shown) provided at the lower end of the drive shaft (70), flows out from a branch passage (82) that opens in the middle of the vertical direction of the eccentric part (72) of the drive shaft (70), and is supplied to the gap.

Considering the sliding properties of the roller (36), the gap between the roller (36) and the cylinder (34) needs to be made somewhat large. However, if the gap is too large, the high-pressure lubricating oil supplied to the gap is likely to leak into the cylinder chamber (S) (first chamber (51) and second chamber (52)) between the roller (36) and the cylinder (34). A large amount of refrigerant is dissolved in the high-pressure lubricating oil. If a large amount of such high-pressure lubricating oil containing a large amount of dissolved refrigerant leaks into the relatively low-pressure cylinder chamber (S), a large amount of refrigerant will be separated from the leaked high-pressure lubricating oil. In other words, a large amount of high-pressure refrigerant will flow into the relatively low-pressure cylinder chamber (S) (so-called refrigerant leakage will occur), leading to a significant decrease in compressor efficiency. To achieve both smooth sliding and reduced leakage loss, it is preferable to minimize the gap between the roller (36) and configure the compressor (1) so that the roller (36) floats to an appropriate height during operation, ensuring that the height of the gap between the roller (36) is uniform.

Therefore, in the rotary compressor (1) of this disclosure, an inclined surface (36b) is formed on the lower end of the roller (36) so that the roller (36) floats to an appropriate height (a height at which the height of the gap between the top and bottom of the roller (36) is approximately the same) during operation. The inclined surface (36b) will be described in detail below.

(5) Inclined surface As shown in the enlarged view of Figure 4, an inclined surface (36b) is formed on the lower end of the roller (36) by chamfering the outer corner. This inclined surface (36b) is inclined so that it is located higher as it moves radially outward. The part of the lower end surface of the roller (36) other than the inclined surface (36b) is configured as a horizontal plane (36a) parallel to the upper surface of the rear head (42).

As shown in Figure 3, the inclined surface (36b) is formed around the entire circumference of the roller (36), excluding the angular range where the vane (37) is located, without straddling the vane (37). More specifically, it extends from the angular position of the suction-side surface of the vane (37) in the circumferential direction of the roller (36) (on the line connecting the root end of the suction-side surface and the center of the roller (36) in Figure 3) to the angular position of the discharge-side surface of the vane (37) (on the line connecting the root end of the discharge-side surface and the center of the roller (36) in Figure 3).

As shown in the enlarged view of Figure 4, the inclined surface (36b) extends continuously from the outer circumference of the horizontal surface (36a) to the outer circumference surface (36o) of the roller (36). The cutting angle φ of the inclined surface (36b) (the angle of the inclined surface (36b) relative to the horizontal surface (36a)) is between 10° and 45°. The cutting height X of the inclined surface (36b) (the distance along the axial direction of the drive shaft (70) between the outer end of the inclined surface (36b) and the horizontal surface (36a)) is between 0.01 mm and 5 mm. It is preferable that the cutting height X of the inclined surface (36b) be less than or equal to half the thickness t (radial width) of the roller (36).

(6) Wedge Effect By forming the above-mentioned inclined surface (36b), a wedge-shaped (triangular in vertical cross-section) gap (90) is formed on the outer circumference between the lower end surface of the roller (36) and the upper end surface of the rear head (42). With this configuration, during operation, high-pressure lubricating oil is supplied to each sliding part of the compression mechanism (30), and some of it reaches the cylinder chamber (S) between the cylinder (34) and the roller (36). The lubricating oil that reaches this cylinder chamber (S) is pushed from the outer circumference entrance of the wedge-shaped gap (90) to the back (towards the center of the roller (36)) by the eccentric rotation of the roller (36). The high-pressure lubricating oil pushed inward in this way acts an upward force on the roller (36), causing the roller (36) to lift up. In other words, the high-pressure lubricating oil pushed inward into the wedge-shaped gap (90) acts like a wedge, causing the roller (36) to lift up (wedge effect).

Furthermore, by forming the inclined surface (36b) on the outer circumference of the roller (36), the lubricating oil between the non-rotating cylinder (34) and the eccentrically rotating roller (36) flows into the wedge-shaped gap (90). The relative velocity between the lubricating oil flowing into the gap (90) and the inclined surface (36b), which acts as the surface of action for the lubricating oil, becomes greater compared to when the inclined surface is formed on the inner circumference of the roller (36). Therefore, the upward force acting on the roller (36) due to the wedge effect increases, allowing the roller (36) to be sufficiently lifted.

The rotary compressor (1) of this embodiment 1 is a swing-type compressor (1) equipped with a piston (35) in which a roller (36) and vanes (37) are integrally formed. In the swing-type compressor (1), the roller (36) rotates eccentrically (revolves) within the cylinder (34) as the drive shaft (70) rotates, but does not rotate on its own axis. Therefore, as described above, assuming that the angular position of the center line (M) of the vanes (37) with respect to the center of the roller (36) in the circumferential direction is 0°, and that the angular position of the roller (36) in the circumferential direction increases as it progresses in the direction of rotation of the drive shaft (70), the roller (36) contacts the cylinder (34) at an angular position equal to the rotation angle of the drive shaft (70). In other words, 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).

Incidentally, as mentioned above, if the gap formed above and below the roller (36) during operation is too large, the high-pressure lubricating oil supplied to the gap is more likely to leak into the cylinder chamber (S). This leakage of lubricating oil into the cylinder chamber (S) results in leakage losses of three types of refrigerants (CR (cylinder roller) leakage loss).

The first type of leakage loss is due to leakage from the gap on the intake chamber side, caused by high-pressure lubricating oil leaking into the first chamber (51) through the gaps above and below the roller (36). Specifically, the refrigerant separates from the lubricating oil leaked into the first chamber (51) under low pressure conditions and flows into the first chamber (51), reducing the intake volume and decreasing compressor efficiency.

The second type of leakage loss is due to leakage from the gaps above and below the roller (36) into the relatively lower-pressure second chamber (52). Specifically, the refrigerant separates from the lubricating oil that leaks into the relatively lower-pressure second chamber (52) and flows into the second chamber (52), increasing the compression load and reducing compressor efficiency.

The third type of leakage loss is due to roller trajectory leakage, which occurs because the roller (36) revolves along the inner surface of the cylinder (34), causing the lubricating oil leaked into the second chamber (52) to flow into the first chamber (52) (changing the location of refrigerant leakage within the cylinder chamber (S) from the second chamber (52) to the first chamber (51)). This third type of roller trajectory leakage loss, like the first type of leakage loss due to the gap on the intake chamber side, reduces the intake volume and lowers compressor efficiency.

Figures 6A to 6C are graphs showing the change in lubricating oil leakage during one rotation of the drive shaft (70). The dashed line in Figure 6A shows the change in lubricating oil leakage into the first chamber (51) (the sum of leakage due to the intake chamber gap and leakage due to the roller trajectory). The dashed line in Figure 6B shows the change in lubricating oil leakage into the second chamber (52) (leakage due to the compression chamber gap). The dashed line in Figure 6C shows the sum of the leakage amounts shown by the dashed lines in Figures 6A and 6B, i.e., the change in lubricating oil leakage into the cylinder chamber (S).

The amount of leakage from the gap on the intake chamber side, which leads to the first type of leakage loss mentioned above, is determined by the difference between the pressure in the internal space (60) of the casing (10) (the pressure of the lubricating oil) and the internal pressure of the first chamber (51), and the arc length of the portion of the outer surface of the roller (36) facing the first chamber (51). This amount increases as the volume of the first chamber (51) increases (as the rotation angle increases). Similarly, the amount of leakage from the roller trajectory, which leads to the third type of leakage loss mentioned above, also increases as the volume of the first chamber (51) increases (as the rotation angle increases). Therefore, the amount of lubricating oil leaking into the first chamber (51), shown by the dashed line in Figure 6A (the sum of the leakage due to the gap on the intake chamber side and the leakage due to the roller trajectory), increases as the volume of the first chamber (51) increases (as the rotation angle increases). Specifically, the value is smallest when the rotation angle of the drive shaft (70) is near 0°, increases as the rotation angle increases, and reaches its maximum when the rotation angle of the drive shaft (70) is 360°.

On the other hand, the amount of leakage from the compression chamber gap, shown by the dashed line in Figure 6B, which leads to the second type of leakage loss mentioned above, is determined by the difference between the pressure in the internal space (60) of the casing (10) (the pressure of the lubricating oil) and the internal pressure of the second chamber (52), and the arc length of the portion of the outer surface of the roller (36) facing the second chamber (52). This leakage decreases as the volume of the second chamber (52) decreases (as the rotation angle increases). Specifically, it is largest when the rotation angle of the drive shaft (70) is around 0°, decreases as the rotation angle increases, and becomes zero when the rotation angle of the drive shaft (70) is around 180°, at which point the discharge stroke begins.

Therefore, the amount of lubricating oil leakage into the cylinder chamber (S), shown by the dashed line in Figure 6C, is largely influenced by the compression chamber side gap leakage, shown by the dashed line in Figure 6B, from the rotation angle of the drive shaft (70) from 0° to 180°. This leakage is greatest around 0° of the drive shaft (70) rotation angle, decreases as the rotation angle increases, and is lowest when the rotation angle of the drive shaft (70) is around 180°, at which point the discharge stroke begins. On the other hand, from around 180° to 360° of the rotation angle of the drive shaft (70), the compression chamber side gap leakage, shown by the dashed line in Figure 6B, becomes 0, and the amount of lubricating oil leakage into the first chamber (51), shown by the dashed line in Figure 6A, is directly reflected. Specifically, it is smallest when the rotation angle of the drive shaft (70) is around 180°, increases as the rotation angle increases, and is highest when the rotation angle of the drive shaft (70) is 360°. Thus, in a swing-type compressor (rotary compressor) (1), the amount of lubricating oil leaking from the gaps above and below the rollers (36) into the cylinder chamber (S) increases or decreases during one rotation of the drive shaft (70), and is particularly high when the rotation angle of the drive shaft (70) is 0° (360°).

Furthermore, as shown by the dashed line in Figure 6C, although the amount of lubricating oil leaking into the cylinder chamber (S) increases or decreases during one rotation of the drive shaft (70), leakage of lubricating oil into the cylinder chamber (S) occurs as the rotation angle of the drive shaft (70) ranges from 0° to 360°, resulting in leakage losses. Therefore, in this embodiment 1, the inclined surface (36b) is formed around almost the entire circumference of the roller (36) (excluding the angular range where the vanes (37) are located in the circumferential direction of the roller (36)) so that the roller (36) floats to an appropriate height (a height where the height of the gap between the top and bottom of the roller (36) is approximately the same) regardless of the rotation angle of the drive shaft (70).

Incidentally, if an inclined surface (36b) is formed in the circumferential direction of the roller (36) within an angular range from α to α + 180°, the aforementioned wedge effect is effectively obtained when the rotation angle of the drive shaft (70) is α, and the roller (36) effectively levitates. In this embodiment 1, since the inclined surface (36b) is formed over almost the entire circumference of the roller (36), the roller (36) will levitate regardless of the rotation angle of the drive shaft (70).

The graphs shown by solid lines in Figures 6A to 6C show the change in the amount of lubricating oil leakage during one rotation of the drive shaft (70) in the rotary compressor (1) of this embodiment 1, in which the inclined surface (36b) is formed over almost the entire circumference of the roller (36). The solid line in Figure 6A shows the change in the amount of lubricating oil leakage into the first chamber (51) (the sum of leakage due to the gap on the intake chamber side and leakage due to the roller trajectory). The solid line in Figure 6B shows the change in the amount of lubricating oil leakage into the second chamber (52) (leakage due to the gap on the compression chamber side). The solid line in Figure 6C shows the sum of the leakage amounts shown by solid lines in Figures 6A and 6B, i.e., the change in the amount of lubricating oil leakage into the cylinder chamber (S).

As described above, by forming the inclined surface (36b) over almost the entire circumference of the roller (36), as shown in Figure 6A, the amount of lubricating oil leakage into the first chamber (51) (the sum of leakage due to the gap on the intake chamber side and leakage due to the roller trajectory) changes from a dashed line to a solid line, indicating a significant reduction. Furthermore, as shown in Figure 6B, the amount of lubricating oil leakage into the second chamber (51) (leakage due to the gap on the compression chamber side) also changes from a dashed line to a solid line, indicating a significant reduction. Therefore, as shown in Figure 6C, the amount of lubricating oil leakage into the cylinder chamber (S) also changes from a dashed line to a solid line, indicating a significant reduction.

(7) Inclined surface formation method 1
The inventors of the present invention have investigated methods for easily forming an inclined surface (36b) that can raise the roller (36) to an appropriate height during operation for various rotary compressors (1) of different sizes, and have derived the following inclined surface formation method 1.

The inclined surface formation method 1 derives the following formula (1) which shows the range of the cut height X (distance along the axial direction of the drive shaft (70) between the outer edge of the inclined surface (36b) and the horizontal plane (36a) using the weight m of the piston (35) (sum of the weight of the roller (36) and the weight of the vane (37)) and the height difference ΔH between the cylinder (34) and the roller (36), and uses this to form (design) the inclined surface (36b).
(ΔH÷2×0.75)÷(0.0055×m ^-0.893 )≦X≦(ΔH÷2×1.25)÷(0.00006m ^-1.118 )...(1)

When forming the inclined surface (36b), if the cutting angle φ of the inclined surface (36b) is set to an angle of 10° to 45°, and the cutting height X of the inclined surface (36b) is set to satisfy the above formula (1), then a rotary compressor (1) in which the roller (36) can be lifted to an appropriate height during operation can be easily formed.

<Method for deriving formula (1)>
First, for the following five types of rotary compressors (1) with different sizes (different piston (35) masses m), the relationship between the incision height X of the inclined surface (36b) and the amount of buoyancy Y of the piston (35) was estimated when the incision angle φ of the inclined surface (36b) was 10°, 20°, 30°, and 45°. Figure 7A shows the estimated results for type A, which has the lightest piston (35), and Figure 7B shows the estimated results for type E, which has the heaviest piston (35). Note that all five types of rotary compressors (1) have the above configuration, differing only in size (weight).
A: Rotary compressor with piston (35) mass m of 0.01 kg B: Rotary compressor with piston (35) mass m of 0.025 kg C: Rotary compressor with piston (35) mass m of 0.05 kg D: Rotary compressor with piston (35) mass m of 0.075 kg E: Rotary compressor with piston (35) mass m of 0.1 kg

The calculation begins by first determining the upward force py acting on the roller (36) due to the wedge effect described above, while the rotary compressor (1) is in operation and the amount of buoyancy of the piston (35) is zero (Step 1). The upward force py acting on the roller (36) due to the wedge effect is calculated using the following formulas (2) to (5). In the following formulas (2) to (5), x, η, r, a, δ, θ, h, X, and φ are as follows.
x: Distance from the outer surface of the roller (36) in the radial direction (mm)
η: Viscosity of lubricating oil (mPa·s)
r: Amount of eccentricity (mm) of the eccentric section (72)
a: Outer radius of roller (36) (mm)
δ: Angle (rad) of the inclined surface (36b) when the direction of travel of the piston (35) is 0°.
(-π/2 < δ < π/2)
θ: Rotation angle (rad) of the drive shaft (70)
h: Height of the inclined surface (36b) relative to the horizontal plane (36a) at a distance x from the radial outer surface of the roller (36) (distance in the axial direction of the drive shaft (70)) (mm)
X: Height of the inclined surface (36b) on the radial outer surface of the roller (36) relative to the horizontal plane (36a) (distance in the axial direction of the drive shaft (70)) (mm)
φ: Cutting angle (°) of the inclined surface (36b)

Next, the amount of buoyancy y of the piston (35) due to the calculated upward force py is determined (Step 2). Then, with the piston (35) lifted by the amount of buoyancy y, the upward force py acting on the roller (36) due to the wedge effect is calculated (Step 3), and it is determined whether this upward force py balances the weight m of the piston (35) (Step 4).

In step 4, when the upward force py balances with the weight m of the piston (35), the amount of buoyancy y of the piston (35) due to the upward force py is determined, and this amount of buoyancy y is set as the amount of buoyancy Y of the piston (35).

On the other hand, if the upward force py and the weight m of the piston (35) are not balanced in step 4, the process returns to step 2, and the amount of levitation y of the piston (35) due to the upward force py is recalculated (step 2). Then, with the piston (35) lifted by the amount of levitation y, the upward force py acting on the roller (36) due to the wedge effect is recalculated (step 3), and it is determined whether this upward force py and the weight m of the piston (35) are balanced (step 4).

By repeating the above steps, the amount of buoyancy Y of the piston (35) is determined. Then, a relational expression was obtained that expresses the amount of buoyancy Y for each cutting angle φ, as determined by the above calculation, as a function of the cutting height X, and the result is as follows. Figure 7A shows the result for type A, where the piston (35) is lightest, and Figure 7E shows the result for type E, where the piston (35) is heaviest.
(7-1) Type A: When m = 0.01, φ = 10: Y = 0.3313X ... A10
φ=20:Y=0.0938X...A20
φ=30:Y=0.0366X...A30
φ=45:Y=0.0103X...A45
(7-2) Type B: When m = 0.025, φ = 10: Y = 0.1532X ... B10
φ=20:Y=0.0395X...B20
φ=30:Y=0.0149X...B30
φ=45:Y=0.0041X...B45
(7-3) Type C: When m = 0.05, φ = 10: Y = 0.0817X ... C10
φ=20:Y=0.0075X...C20
φ=30:Y=0.0202X...C30
φ=45:Y=0.002X...C45
(7-4) Type D: When m = 0.075, φ = 10: Y = 0.0559X ... D10
φ=20:Y=0.0135X...D20
φ=30:Y=0.0049X...D30
φ=45:Y=0.0009X...D45
(7-5) Type E: When m = 0.1, φ = 10: Y = 0.0424X ... E10
φ=20:Y=0.0101X...E20
φ=30:Y=0.0037X...E30
φ=45:Y=0.0009X...E45

From the calculation results for types A to E above, it can be seen that the lighter the mass m of the piston (35) and the smaller the cutting angle φ of the inclined surface (36b), the greater the slope of the line segment (slope of the function) showing the relationship between the amount of levitation Y and the cutting height X of the inclined surface (36b). Conversely, the heavier 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 (slope of the function) showing the relationship between the amount of levitation Y and the cutting height X of the inclined surface (36b).

Next, for each of the above functions A10 to E45 obtained for each cutting angle φ (10°, 20°, 30°, 45°) for types A to E, we calculated approximate formulas showing the relationship between the inclination I and the weight m of the piston (35) for functions A10, B10, C10, D10, and E10 where the cutting angle φ that maximizes the inclination I is 10°, and for functions A45, B45, C45, D45, and E45 where the cutting angle φ that minimizes the inclination I is 45°.

Specifically, as shown in Figure 8A, the relationship between the inclination I of each function A10, B10, C10, D10, and E10 with a cutting angle φ of 10° and the weight m of the piston (35) is shown. When an approximate formula was obtained in which the inclination I of the functions A10, B10, C10, D10, and E10 with a cutting angle φ of 10° is a function of the weight m of the piston (35), the result was I = 0.0055m ^{- 0.893} .

Similarly, as shown in Figure 8B, the relationship between the inclination I of functions A45, B45, C45, D45, and E45 with a cutting angle φ of 45° and the weight m of the piston (35) is shown. When an approximate formula was obtained in which the inclination I of functions A45, B45, C45, D45, and E45 with a cutting angle φ of 45° is a function of the weight m of the piston (35), the result was I = 0.00006m ^{- 1.118} .

Next, using the approximate formulas I = 0.0055m ^{- 0.893} for the slope I of functions A10, B10, C10, D10, and E10 with a cutting angle φ of 10°, and I = 0.00006m ^{- 1.118} for the slope I of functions A45, B45, C45, D45, and E45 with a cutting angle φ of 45°, we determined the range of cutting heights X for which the amount of floating Y can be an appropriate value.

Specifically, the appropriate range for the floating amount Y is set to ΔH÷2×0.75≦Y≦ΔH÷2×1.25, and as shown in Figure 9, line segment L1 (Y = 0.0055 × m ^{- 0.893}・X) shows the relationship between the floating amount Y and the cutting height X when the cutting angle φ is 10°, and line segment L2 (Y = 0.00006 m) shows the relationship between the floating amount Y and the cutting height X when the cutting angle φ is 45°. The range of cut height X that can fall within the range enclosed by ^-1.118 * X, line segment L3 (Y = ΔH ÷ 2 × 0.75) which indicates the lower limit of the appropriate range for the amount of floating Y, and line segment L4 (Y = ΔH ÷ 2 × 1.25) which indicates the upper limit of the appropriate range for the amount of floating Y (the shaded area in Figure 9) is derived (formula (1) above, X1 ≤ X ≤ X2 in Figure 9), and this range is defined as the range of cut height X for which the amount of floating Y can be an appropriate value.

(8) Inclined surface formation method 2
Furthermore, the inventors of the present application investigated a method for easily forming an inclined surface (36b) that can raise the roller (36) to an appropriate height during operation for various rotary compressors (1) of different sizes, using a method different from the inclined surface forming method 1 described above, and derived the following inclined surface forming method 2.

Method 2 for forming an inclined surface involves deriving a relationship between the weight m of the piston (35), the cutting angle φ, and the cutting height X, such that the amount of float Y of the roller (36) during operation falls within an appropriate range (ΔH ÷ 2 × 0.75 ≤ Y ≤ ΔH ÷ 2 × 1.25), and using this relationship to design the inclined surface (36b).

First, for three types of rotary compressors (1) (Type A, Type C, and Type E of Inclined Surface Forming Method 1) with piston (35) weights m of 0.01 kg, 0.05 kg, and 0.1 kg, the appropriate range of the floating amount Y (ΔH ÷ 2 × 0.75 ≤ Y ≤ ΔH ÷ 2 × 1.25) was determined and found to be as follows. When m = 0.01 kg, ΔH = 0.015; when m = 0.05 kg, ΔH = 0.022; and when m = 0.1 kg, ΔH = 0.028.
(8-1) Type A: When m = 0.01
0.00563 ≤ Y ≤ 0.009375
(8-2) Type C: When m = 0.05
0.00825 ≤ Y ≤ 0.0138
(8-3) Type E: When m = 0.1
0.0105 ≤ Y ≤ 0.0175

Next, we determined the range of cutting height X and cutting angle φ that can fall within the appropriate range of floating amount Y in the relational expression for each cutting angle φ of type A shown in Figure 7A, which is a function of the cutting height X, (8-1) the range of cutting height X and cutting angle φ that can fall within the appropriate range of floating amount Y in the relational expression for each cutting angle φ of type C, (8-2) the range of cutting height X and cutting angle φ that can fall within the appropriate range of floating amount Y, and in the relational expression for each cutting angle φ of type E shown in Figure 7B, which is a function of the cutting height X, (8-3) the range of cutting height X and cutting angle φ that can fall within the appropriate range of floating amount Y, and the results were as follows.
(8-4) When m = 0.01
10° ≤ φ ≤ 20°, and 0.017 ≤ X ≤ 0.10
20° < φ ≤ 30°, and 0.060 ≤ X ≤ 0.26
30° < φ ≤ 45°, and 0.15 ≤ X ≤ 0.91
(8-5) When m = 0.05
10° < φ ≤ 20°, and 0.10 ≤ X ≤ 0.68
20° < φ ≤ 30°, and 0.41 ≤ X ≤ 1.84
30° < φ ≤ 45°, and 1.10 ≤ X ≤ 5.00
(8-6) When m = 0.1
10° < φ ≤ 20°, and 0.25 ≤ X ≤ 1.73
20° < φ ≤ 30°, and 1.04 ≤ X ≤ 4.70
30° < φ ≤ 45°, and 2.84 ≤ X ≤ 5.00

From the above results, the following relational equations (8-7) to (8-9) were derived as the relationship between the weight m of the piston (35), the cutting angle φ, and the cutting height X, such that the amount of floating Y falls within the appropriate range. Furthermore, if the relational equations (8-7) to (8-9) are satisfied when forming the inclined surface (36b), a rotary compressor (1) can be formed in which the roller (36) can float to an appropriate height during operation.
(8-7) When m ≤ 0.03
10° ≤ φ ≤ 45°, and 0.017 ≤ X ≤ 0.91
(8-8) 0.03 < m ≤ 0.06
10° ≤ φ ≤ 45°, and 0.10 ≤ X ≤ 5.00
(8-9) When m > 0.06
10° ≤ φ ≤ 45°, and 0.25 ≤ X ≤ 5.00

Furthermore, from the above results, it is possible to derive the following more detailed relational expressions (8-10) to (8-12) than those shown in (8-7) to (8-9). If the relational expressions (8-10) to (8-12) are satisfied when forming the inclined surface (36b), a rotary compressor (1) can be formed in which the roller (36) can float to a more appropriate height during operation.
(8-10)m ≤ 0.03
10° ≤ φ ≤ 20°, and 0.017 ≤ X ≤ 0.10
20° < φ ≤ 30°, and 0.060 ≤ X ≤ 0.26
30° < φ ≤ 45°, and 0.15 ≤ X ≤ 0.9
(8-11) When 0.03 < m ≤ 0.06
10° < φ ≤ 20°, and 0.10 ≤ X ≤ 0.68
20° < φ ≤ 30°, and 0.41 ≤ X ≤ 1.84
30° < φ ≤ 45°, and 1.1 ≤ X ≤ 5.0
(8-12) When m > 0.06
10° < φ ≤ 20°, and 0.25 ≤ X ≤ 1.73
20° < φ ≤ 30°, and 1.04 ≤ X ≤ 4.7
30° < φ ≤ 45°, and 2.84 ≤ X ≤ 5.0

- Effects of Embodiment 1 -
In the rotary compressor (1) of this embodiment 1, a portion of the outer circumference of the lower end surface of the roller (36) is configured as an inclined surface (second surface) (36b) that slopes upward as it moves radially outward. With this configuration, a wedge-shaped gap (90) is formed on the outer circumference between the lower end surface of the roller (36) and the rear head (lower end plate) (42). By forming such a gap (90), during operation, high-pressure lubricating oil equivalent to the pressure inside the casing (10) is supplied to each sliding part of the compression mechanism (30), and some of it reaches the space between the cylinder (34) and the roller (36) (cylinder chamber (S)). As the roller (36) rotates eccentrically inside the cylinder (34), the lubricating oil that reaches this cylinder chamber (S) is pushed from the outer circumference entrance to the back (towards the center of the roller (36)) in the wedge-shaped gap (90). In this way, the high-pressure lubricating oil that is pushed in exerts an upward force on the roller (36), causing the roller (36) to lift off the ground (wedge effect). Therefore, according to this embodiment 1, it is possible to provide a rotary compressor (1) that can levitate the roller (36) to an appropriate height during operation. Furthermore, if the roller (36) can be levitated to an appropriate height during operation, the amount of lubricating oil leaking from the gaps formed above and below the roller (36) into the cylinder chamber (S) can be reduced, thereby reducing leakage losses.

Furthermore, in the rotary compressor (1) of this embodiment 1, by forming the inclined surface (36b) on the outer circumference side of the roller (36), lubricating oil between the non-rotating cylinder (34) and the eccentrically rotating roller (36) flows into the wedge-shaped gap (90). The relative velocity between the lubricating oil flowing into the gap (90) and the surface acting on it (inclined surface) becomes greater compared to when the inclined surface is formed on the inner circumference side of the roller (36). Therefore, the upward force acting on the roller (36) increases due to the wedge effect described above, allowing the roller (36) to be sufficiently lifted. Thus, according to this embodiment 1, a rotary compressor (1) capable of lifting the roller (36) to an appropriate height during operation can be provided.

Furthermore, in this embodiment 1, for various rotary compressors (1) of different sizes, the range of height (cut height) X of the wedge-shaped gap (90) that can raise the roller (36) to an appropriate height during operation is estimated. The above-mentioned formula (1), expressed using the weight of the piston (35) (the sum of the weight of the roller (36) and the weight of the vane (37)) m and the height difference ΔH between the cylinder (34) and the roller (36), is derived, and the inclined surface (36b) is formed to satisfy the above formula (1). By forming (designing) the inclined surface (36b) in this way, rotary compressors (1) of various sizes that can raise the roller (36) to an appropriate height during operation can be easily constructed.

Furthermore, in this embodiment 1, the rotary compressor (1) is a swing-type compressor equipped with a piston (35) in which the roller (36) and vane (37) are integrally formed. In a swing-type compressor, as shown in Figure 6C, when the rotation angle of the drive shaft (70) during one rotation of the drive shaft (70) is 0° (360°), the amount of lubricating oil leaking from the gaps above and below the roller (36) into the cylinder chamber (S) is particularly large.

As mentioned above, by forming an inclined surface (36b) in the circumferential direction of the roller (36) within an angular range from α to α + 180°, the wedge effect described above is effectively obtained when the rotation angle of the drive shaft (70) is α, and the roller (36) effectively levitates.

Therefore, in this embodiment 1, an inclined surface (36b) is formed over a range of at least 180° from the angular position of the intake-side surface of the vane (37) in the circumferential direction of the roller (36). With this configuration, a wedge effect is effectively generated when the rotation angle of the drive shaft (70) is 0°, and the roller (36) can be reliably levitated when the rotation angle of the drive shaft (70), which is particularly high in terms of lubricating oil leakage into the cylinder chamber (S), is 0°. Consequently, according to this embodiment 1, the amount of lubricating oil leakage into the cylinder chamber (S) can be reduced, and leakage losses can be reduced.

Furthermore, in this embodiment 1, the inclined surface (36b) is formed around almost the entire circumference of the roller (36), except for a certain angular range where the circumferential vanes (37) of the roller (36) are located. Since the roller (36) can be levitated regardless of the rotation angle of the drive shaft (70), the amount of lubricating oil leaking into the cylinder chamber (S) can be reduced, thereby reducing leakage losses.

Embodiment 2
Embodiment 2 is a modification of the configuration of the inclined surface (36b) of the rotary compressor (1) in Embodiment 1.

Specifically, as shown in Figure 10, in Embodiment 2, the inclined surface (36b) is not formed over almost the entire circumference of the roller (36) in the circumferential direction (excluding the angular range where the vanes (37) are located in the circumferential direction of the roller (36)), as in Embodiment 1. Instead, it is formed over a range from the angular position of the suction-side side of the vanes (37) in the circumferential direction of the roller (36) up to 270°. Therefore, in Embodiment 2, the inclined surface (36b) is not formed in the range from 270° in the circumferential direction of the roller (36) to the angular position of the discharge-side side of the vanes (37).

Incidentally, as mentioned above, from the viewpoint of reducing the amount of lubricating oil leaking into the cylinder chamber (S) from the gaps formed above and below the roller (36) by levitating the roller (36) to an appropriate height, it is preferable to make the range of inclined surface (36b) formation (the angular range in which the inclined surface (36b) is formed in the circumferential direction of the roller (36)) as wide as possible (almost the entire circumference as in Embodiment 1). On the other hand, if the outer peripheral corner of the lower end of the roller (36) is chamfered to form an inclined surface (36b), and a wedge-shaped gap (90) is formed between the inclined surface (36b) and the upper surface of the rear head (42), the amount of lubricating oil leaking into the cylinder chamber (S) from the gaps above and below the roller (36) can be reduced by levitating the roller (36), but there is a risk that the refrigerant compressed in the second chamber (52) will leak into the first chamber (51) through the gap (90), leading to a decrease in performance. The wider the area where the inclined surface (36b) is formed, the more likely refrigerant leakage through the gap (90) will occur, increasing the risk of performance degradation. Therefore, from the viewpoint of reducing refrigerant leakage from the second chamber (52) to the first chamber (51) through the gap (90), it is preferable to make the area where the inclined surface (36b) is formed as narrow as possible.

Therefore, in Embodiment 2, as described above, the inclined surface (36b) is formed only in the range from the angular position of the suction-side side of the vane (37) in the circumferential direction of the roller (36) up to 270°. By forming the inclined surface (36b) in this way, the roller (36) can be raised when the rotation angle of the drive shaft (70), where the amount of lubricating oil leakage into the cylinder chamber (S) (mainly the second chamber (52)) is large, is between 0° and 180°. This significantly reduces the amount of lubricating oil leakage into the cylinder chamber (S) and also significantly reduces leakage losses. On the other hand, by not forming the inclined surface (36b) in the range from 270° in the circumferential direction of the roller (36) to the angular position of the discharge-side side of the vane (37), refrigerant leakage from the second chamber (52) to the first chamber (51) through the wedge-shaped gap (90) created by forming the inclined surface (36b) can be reduced.

Other embodiments
In embodiments 1 and 2 described above, a swing-type compressor equipped with a piston (35) in which a roller (36) and a vane (37) are integrally formed was described as an example of the rotary compressor (1) of the present disclosure. However, 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 vane separate from the roller contacts the roller while the roller rotates eccentrically, or a so-called hinge vane type rotary compressor in which the tip of the vane is rotatably fitted into a recess on the outer circumferential surface of the roller while the roller rotates eccentrically.

Furthermore, while embodiments 1 and 2 described above described a so-called single-cylinder rotary compressor (1) equipped with one cylinder (34) as an example of the rotary compressor (1) of this disclosure, the rotary compressor (1) of this disclosure is not limited to a single-cylinder type. The rotary compressor (1) of this disclosure may be a so-called multi-cylinder rotary compressor (1) equipped with multiple cylinders (34). In the case of a multi-cylinder rotary compressor (1), the lower end surface of each roller (36) that rotates eccentrically within each cylinder (34) should be configured to have an inclined surface (second surface) (36b). With this configuration, each roller (36) will float to an appropriate height within each cylinder (34) during operation.

Furthermore, in the above embodiment 2, the inclined surface (36b) was formed only in the range from the angular position of the suction-side side of the vane (37) in the circumferential direction of the roller (36) up to 270°, and not in the range from 270° to the angular position of the discharge-side side of the vane (37). When the inclined surface (36b) is formed only in a limited angular range in the circumferential direction of the roller (36) in this way, it is preferable that the starting end of the inclined surface (36b) be provided at the angular position of the suction-side side of the vane (37) in the circumferential direction of the roller (36), as in embodiment 2, but the ending end of the inclined surface (36b) can be provided at a predetermined angular position between 180° and 270° in the circumferential direction of the roller (36), and is not limited to the 270° position. By positioning the starting end of the inclined surface (36b) at an angular position on the suction-side side of the vane (37) in the circumferential direction of the roller (36), and positioning the ending end of the inclined surface (36b) at a predetermined angular position of 180° to 270° in the circumferential direction of the roller (36), the same effects as in Embodiment 2 can be achieved.

While embodiments and modifications have been described above, it will be understood that various changes in form and detail are possible without departing from the spirit and scope of the claims. Furthermore, these embodiments and modifications may be combined or substituted as appropriate, as long as they do not impair the function of the subject matter of this disclosure. The terms "first," "second," etc., used above are used to distinguish the phrases to which they are assigned, and do not limit the number or order of such phrases.

As described above, this disclosure is useful for rotary compressors and refrigeration cycle systems.

1. Rotary Compressor
9. Refrigerant Circuit
10 Casing
20 Motor
30 Compression mechanism
34 cylinders
35 Pistons
36 Laura
36a Horizontal surface (first surface)
36b Slanted surface (second surface)
36o outer circumferential surface
37. Bane
41 Front head (upper plate)
42 Rear head (lower end plate)
51 Room 1
52 2nd room
70 Drive shaft 100 Refrigeration cycle device

Claims (6)

Casing (10),
A drive shaft (70) is provided within the casing (10) and extends in the vertical direction,
A motor (20) is provided within the casing (10) and drives the drive shaft (70),
The device comprises a compression mechanism (30) provided within the casing (10), connected to the drive shaft (70), and rotationally driven by the motor (20) to compress the refrigerant.
The above compression mechanism (30) is
A cylindrical cylinder (34) and
The upper end plate (41) and lower end plate (42) close the upper and lower open end faces of the cylinder (34),
A cylindrical roller (36) is attached to the drive shaft (70) and rotates eccentrically within the cylinder (34),
The space between the cylinder (34) and the roller (36) is divided into a first chamber (51) on the intake side and a second chamber (52) on the discharge side by a vane (37).
The lower end surface of the roller (36) has a first surface (36a) parallel to the upper surface of the lower end plate (42), and a second surface (36b) that is continuous with the outer circumference of the first surface (36a) and extends to the outer circumference surface (36o) of the roller (36).
The second surface (36b) described above is inclined such that it is positioned higher as it moves radially outward.
The second surface (36b) above is,
The sum of the weight of the roller (36) and the weight of the vane (37) is m.
Let X be the distance between the outer edge of the second surface (36b) and the first surface (36a) along the axial direction of the drive shaft (70).
When the difference between the height of the cylinder (34) and the height of the roller (36) is ΔH,
The following equation (1) is satisfied: (ΔH ÷ 2 × 0.75) ÷ (0.0055 × m ^{- 0.893} ) ≤ X ≤ (ΔH ÷ 2 × 1.25) ÷ (0.00006 m ^{- 1.118} ) ... (1)
Rotary compressor. The upper and lower ends of the roller (36) described above have different shapes.
The rotary compressor according to claim 1. The rotary compressor described above is a swing-type compressor equipped with a piston (35) in which the roller (36) and the vane (37) are integrally formed.
When 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 is set to 0°, and the angular position of the roller (36) in the circumferential direction increases as it moves in the rotational direction of the drive shaft (70),
The rotary compressor according to claim 1 or 2, wherein the second surface (36b) is formed over a range of at least 180° from the angular position of the suction-side surface of the vane (37) in the circumferential direction of the roller (36). The rotary compressor according to claim 3, wherein the second surface (36b) is formed over a range from the angular position of the suction-side surface of the vane (37) to the angular position of the discharge-side surface of the vane (37) in the circumferential direction of the roller (36). The rotary compressor according to claim 3, wherein the second surface (36b) is formed over a range from an angular position of the suction-side surface of the vane (37) in the circumferential direction of the roller (36) to a predetermined angular position of 180° to 270°. A refrigeration cycle apparatus comprising a rotary compressor according to claim 1 or 2 .