WO2023228343A1

WO2023228343A1 - Rotary compressor

Info

Publication number: WO2023228343A1
Application number: PCT/JP2022/021505
Authority: WO
Inventors: 博之山田
Original assignee: 三菱電機株式会社
Priority date: 2022-05-26
Filing date: 2022-05-26
Publication date: 2023-11-30
Also published as: JP7422950B1; JPWO2023228343A1

Abstract

This rotary compressor includes: a rotational shaft; a bearing that is configured from a slip bearing and that has an inner diameter surface that supports the rotational shaft; a cylinder chamber including a suction chamber into which a refrigerant is sucked in, and a compression chamber in which the refrigerant is compressed; and a crank part that is attached to the rotational shaft and that rotates eccentrically in the cylinder chamber. A gap to which refrigerating machine oil is supplied is formed between the rotational shaft and the inner diameter surface of the slip bearing. The slip bearing has anisotropy with respect to the oil film rigidity of the refrigerating machine oil. A working angle of a gas load generated by compression of the refrigerant caused by the eccentric rotation of the crank part changes in accordance with a crank angle. When a fluctuation range of the working angle of the gas load is a first angle range, an axial direction of high oil film rigidity, indicating the direction in which the oil film rigidity of the slip bearing is highest, is disposed in the first angle range.

Description

rotary compressor

The present disclosure relates to a rotary compressor.

A rotary compressor is one of the refrigerant compressors installed in air conditioners and refrigeration equipment. Generally, rotary compressors are known to have a configuration including a closed container, a motor section, a rotating shaft, and a compression mechanism section. The compression mechanism section includes an eccentric crank section provided on a rotating shaft, a cylinder having a cylinder chamber, a rolling piston fitted into the eccentric crank section and housed in the cylinder chamber, and a vane groove formed in the radial direction of the cylinder. It has a vane provided in the. The rolling piston rotates together with the eccentric crank unit to compress refrigerant within the cylinder chamber. The vane follows the rolling piston and partitions the cylinder chamber into a refrigerant suction chamber and a compression chamber. In a rotary compressor, a low-pressure refrigerant is sucked into a compression mechanism, and the refrigerant is rotationally compressed to become a high-pressure refrigerant and discharged.

During the compression process of the refrigerant, a gas pressure load (hereinafter referred to as gas load) by the refrigerant acts on the rotating shaft. Generally, in a rotary compressor, a gas load is supported by two sliding bearings arranged above and below a cylinder chamber of a compression mechanism. Since the gas load is generated by a change in the volume of the compression chamber due to eccentric rotation of the eccentric crank portion, the magnitude and direction of action of the gas load change depending on the rotation angle of the rotating shaft. Therefore, in a rotary compressor, the thickness of the oil film formed between the sliding bearing and the rotating shaft changes during one rotation, and especially at the rotation angle where the gas load is maximum, the thickness of the oil film formed between the sliding bearing and the rotating shaft changes. There is a concern that contact with

As a method of avoiding seizure in a bearing portion of a rotating machine where a fluctuating load is applied, a method of making the inner diameter shape of the bearing non-perfectly round has been considered (see, for example, Patent Document 1).

The bearing described in Patent Document 1 is a bearing that supports a rotating body where load fluctuations may occur. The bearing described in Patent Document 1 has a non-perfect circular shape such as an elliptical shape, and has a cylindrical shape in which the tube axis direction extends in the horizontal direction, and a rotating body is inserted in the center portion. The cylindrical bearing described in Patent Document 1 is divided into an upper half bearing in the upper half and a lower half bearing in the lower half.

In Patent Document 1, the rotating body is composed of a multi-span rotor used in a thermal power plant, for example. Therefore, the load on the bearing may deviate from the design value due to poor alignment settings, changes over time, conditions during partial load operation, etc. Therefore, in Patent Document 1, a stabilizing groove is provided in the lower half bearing to prevent the occurrence of unstable vibrations. In Patent Document 1, the load is always applied in the vertical direction, and the direction of the load does not change.

Japanese Patent Application Publication No. 2000-145781

When applying the structure of a bearing used under a fluctuating load environment where the direction of load action does not change, as described in Patent Document 1, to a rotary compressor where the magnitude and direction of load change, the structure must be defined. It is not possible to uniquely specify the direction of the load applied. Furthermore, as a practical problem, depending on the arrangement and phase angle of the non-round structure in the direction of rotation, the desired dynamic pressure effect on the variable load vector cannot be obtained, leading to the risk of contact between the rotating shaft and the bearing, and the risk of burnout. may increase the risk of Furthermore, as described in Patent Document 1, when using a multi-arc bearing in which a plurality of arcs are combined, or when partially grooving the sliding surface, anisotropy occurs in the oil film stiffness of the bearing. As a result, there is a high possibility that shaft vibration will increase in the direction of lower rigidity.

The present disclosure has been made in order to solve the above-mentioned problems, and in a rotary compressor that is subjected to a variable load in which the magnitude and direction of load change, the thickness of the oil film on the sliding surface of the bearing part is reduced. The purpose of this is to suppress a decrease in the bearing height and suppress the occurrence of seizure in the bearing portion.

A rotary compressor according to the present disclosure includes a rotating shaft and a sliding bearing, a bearing portion having an inner diameter surface that supports the rotating shaft, a suction chamber into which refrigerant is sucked, and a compression chamber in which the refrigerant is compressed. a cylinder chamber, and a crank portion that is attached to the rotating shaft and rotates eccentrically within the cylinder chamber, and refrigerating machine oil is provided between the rotating shaft and the inner diameter surface of the slide bearing. A gap is formed for oil supply, and the sliding bearing has anisotropy in the oil film stiffness of the refrigerating machine oil, and the sliding bearing is resistant to the action of gas load generated by compression of the refrigerant due to eccentric rotation of the crank section. The angle changes depending on the crank angle, and when the variation range of the action angle of the gas load is set as a first angle range, the axial direction of the high oil film rigidity indicating the direction in which the oil film rigidity of the sliding bearing is the highest is the first angle range. It is arranged in one angular range.

According to the rotary compressor according to the present disclosure, the axial direction of high oil film rigidity, which indicates the direction in which the oil film rigidity of the sliding bearing is highest, is arranged in the first angular range that is the range of variation of the action angle of the gas load. Thereby, it is possible to suppress a decrease in the oil film thickness and suppress the occurrence of seizure in the bearing portion.

1 is a longitudinal cross-sectional view showing the configuration of a rotary compressor 1 according to a first embodiment. FIG. 2 is a plan view showing the configuration of a compression mechanism section 501 provided in the rotary compressor 1 according to the first embodiment. FIG. 3 is an explanatory diagram illustrating rotational motion of a crank portion 204 and a rolling piston 504 provided in the rotary compressor 1 according to the first embodiment. FIG. 3 is an explanatory diagram illustrating the principle of gas load generation due to refrigerant compression accompanying rotational operation of a rolling piston 504 in the rotary compressor 1 according to the first embodiment. 5 is a diagram showing an example of a change in gas load GL with respect to crank angle θ1 in rotary compressor 1 according to Embodiment 1. FIG. FIG. 3 is a diagram showing an example of an action angle θ2 of a gas load GL with respect to a crank angle θ1 in the rotary compressor 1 according to the first embodiment. FIG. 2 is a plan view showing the configuration of a full-circumference bearing. FIG. 3 is a plan view showing the configuration of a partial bearing. FIG. 3 is a plan view showing the configuration of a split bearing. FIG. 2 is a plan view showing the configuration of a fitted bearing. FIG. 3 is a plan view showing the configuration of a step bearing. FIG. 2 is a plan view showing the configuration of a two-arc bearing. FIG. 3 is a plan view showing the configuration of a three-arc bearing. FIG. 2 is a plan view showing the configuration of a two-arc eccentric bearing. FIG. 3 is a plan view showing the configuration of a floating bush bearing. FIG. 2 is a plan view showing the configuration of a pivot pad bearing. FIG. 2 is a plan view showing the configuration of a Michel-Seggell bearing. FIG. 2 is a plan view showing the configuration of the NOUMU bearing. FIG. 2 is a plan view showing the configuration of a porous bearing. FIG. 3 is a plan view showing the configuration of a foil bearing. FIG. 2 is a cross-sectional view schematically showing the configuration of a spiral grooved bearing. FIG. 2 is a cross-sectional view schematically showing the configuration of a spherical bearing. FIG. 2 is an explanatory diagram showing angular ranges A, B1, and C in which "the axial direction D of high oil film rigidity" is set in the rotary compressor 1 according to the first embodiment. FIG. 2 is an explanatory diagram showing a direction in which "an axial direction D of high oil film rigidity" is set in the rotary compressor 1 according to the first embodiment. FIG. 7 is an explanatory diagram showing the direction in which the "axial direction D of high oil film rigidity" of the upper bearing 301 and the lower bearing 302 in the rotary compressor 1 according to the second embodiment is set. FIG. 3 is a plan view showing the configuration of a bearing section 300 in a rotary compressor 1 according to a third embodiment. FIG. 7 is a plan view showing the configuration of a bearing section 300 in a rotary compressor 1 according to a fifth embodiment.

Hereinafter, embodiments of a rotary compressor 1 according to the present disclosure will be described with reference to the drawings. The present disclosure is not limited to the following embodiments, and can be variously modified without departing from the gist of the present disclosure. Furthermore, the present disclosure includes all combinations of configurations that can be combined among the configurations shown in the following embodiments and modifications thereof. Furthermore, in each figure, the same reference numerals are the same or equivalent, and this is common throughout the entire specification. Note that in each drawing, the relative dimensional relationship or shape of each component may differ from the actual one.

Embodiment 1.
FIG. 1 is a longitudinal sectional view showing the configuration of a rotary compressor 1 according to the first embodiment. FIG. 2 is a plan view showing the configuration of the compression mechanism section 501 provided in the rotary compressor 1 according to the first embodiment. 1 and 2, as an example of the rotary compressor 1, a single rotary compressor having a compression mechanism section is shown. Note that the rotary compressor 1 is not limited to a single rotary compressor, and may be a rotary compressor having a plurality of compression mechanisms, such as a twin rotary compressor having two compression mechanisms. As shown in FIG. 1, the rotary compressor 1 is connected to a suction muffler 3 via a suction pipe 2. Refrigerant is taken into the rotary compressor 1 from the suction muffler 3.

As shown in FIG. 1, the rotary compressor 1 includes a closed container 100, a rotating shaft 201, a bearing section 300, an electric motor section 401, and a compression mechanism section 501. Inside the airtight container 100, a motor section 401, a rotating shaft 201, a bearing section 300, and a compression mechanism section 501 are housed. The electric motor section 401 is housed inside the closed container 100 in the upper part. The compression mechanism section 501 is housed in the lower part of the closed container 100, that is, below the electric motor section 401. The electric motor section 401 and the compression mechanism section 501 are connected via the rotating shaft 201. The electric motor section 401 includes a rotor 402 and a stator 403. The rotor 402 is arranged inside the stator 403 and fixed to the rotating shaft 201. Stator 403 is fixed to sealed container 100. The electric motor section 401 rotates the rotating shaft 201 by electromagnetic force generated between the stator 403 and the rotor 402. The rotating shaft 201 transmits the driving force of the electric motor section 401 to the compression mechanism section 501. The rotating shaft 201 has a crank portion 204 that is eccentric with respect to the central axis of the rotating shaft 201 . A rolling piston 504 is fitted and attached to the crank portion 204 . A rolling piston 504 is disposed within cylinder 502. As shown in FIG. 2, the cylinder 502 is arranged around the central axis of the rotating shaft 201 so as to cover the rotating shaft 201 from the outside in the circumferential direction. Note that the dimensions of the crank part 204 may be the same as the dimensions when the rolling piston 504 is attached to the crank part 204, and the rolling piston 504 may not be provided. The compression mechanism section 501 compresses the refrigerant using the driving force transmitted from the rotating shaft 201. The bearing section 300 supports the rotating shaft 201 above and below the compression mechanism section 501, and includes one or more sliding bearings.

As shown in FIG. 2, a cylinder chamber 507 formed between a rolling piston 504 and a cylinder 502 is partitioned by a vane 505 to form a compression chamber 507b and a suction chamber 507a. When the electric motor section 401 rotates, a rolling piston 504 attached to the crank section 204 rotates eccentrically within the cylinder 502. As a result, the internal space of the suction chamber 507a becomes smaller as the rolling piston 504 rotates, and the refrigerant in the compression chamber 507b is compressed in the compression chamber 507b. Then, the compression chamber 507b and the discharge port 509 (see FIG. 1) are connected, and the refrigerant inside the compression chamber 507b is discharged from the discharge port 509. At this time, a gas load GL due to the compressed refrigerant acts on the rotating shaft 201. Due to the change in the geometric compression volume of the compression chamber 507b, the gas load GL is approximately vertically downward to the plane of the paper in FIG. 2 with respect to the rotation axis 201, as shown by the arrow W in FIG. Acts on ~180deg. This gas load GL substantially acts on the surface of the rolling piston 504, as will be described later using FIG. 4, but also acts on the rotating shaft 201 via the rolling piston 504. That is, the gas load GL acts on the rotating shaft 201 via the rolling piston 504 due to the oil film in the gap between the rolling piston 504 and the crank portion 204 or due to direct contact between the rolling piston 504 and the crank portion 204 . In the example of FIG. 2, when the gas load GL acts on the rotating shaft 201 in a generally vertical downward direction, as shown by the arrow E in FIG. The axial center position of the rotating shaft 201 changes to a position eccentric to . Note that, for the sake of explanation, FIG. 2 shows the rotary compressor 1 in which the vane 505 and the rolling piston 504 are configured separately from each other. However, the first embodiment is not limited to this, and can also be applied to a swing type compressor in which the vane 505 and the rolling piston 504 are integrated.

Hereinafter, each component of the rotary compressor 1 shown in FIGS. 1 and 2 will be described in detail.

The airtight container 100 is composed of an upper container 101 and a lower container 102. Note that the sealed container 100 is not limited to being formed from two components, the upper container 101 and the lower container 102, but may be formed from three or more components.

As shown in FIG. 1, the sealed container 100 is connected to a suction muffler 3 via a suction pipe 2, and gas refrigerant is taken into the interior from the suction muffler 3. The suction muffler 3 is fixed to the outer surface of the lower container 102 of the closed container 100 by welding or the like. The suction muffler 3 separates the low temperature and low pressure refrigerant sent from the refrigeration circuit into liquid refrigerant and gas refrigerant, prevents the liquid refrigerant from being sucked into the compression mechanism section 501 as much as possible, and stores the separated liquid refrigerant. It is set up for the purpose of The reason for this is that in the rotary compressor 1, when liquid refrigerant flows into the compression mechanism section 501 and is compressed, the pressure inside the cylinder becomes abnormally high due to the compression of the liquid refrigerant, causing damage to the compression mechanism section 501 such as vane flying damage. This is because it may cause a malfunction. The suction muffler 3 also has a function as a muffler that reduces or eliminates noise generated by the inflowing refrigerant.

A refrigerant discharge pipe 4 for discharging compressed refrigerant is connected to the upper part of the closed container 100. The refrigerant discharge pipe 4 is a refrigerant pipe that discharges high-pressure gas refrigerant to the outside of the closed container 100. The refrigerant discharge pipe 4 passes through the upper container 101 constituting the closed container 100 and is joined to the upper container 101 by, for example, brazing or resistance welding.

The inside of the closed container 100 is filled with high temperature and high pressure gas refrigerant compressed by the compression mechanism section 501, and refrigerating machine oil 5 used for lubricating the compression mechanism section 501 is stored at the bottom. Refrigerating machine oil 5 is mainly used to lubricate the sliding parts of compression mechanism section 501. An oil pump (not shown) is provided below the rotating shaft 201. As the rotating shaft 201 rotates, the oil pump pumps up the refrigerating machine oil 5 stored at the bottom of the closed container 100 and supplies it to each sliding part of the compression mechanism section 501. In the compression mechanism section 501, mechanical lubrication is ensured by supplying oil to each sliding section.

As shown in FIG. 1, the electric motor section 401 includes a cylindrical stator 403 fixed to the inner wall surface of the closed container 100 by shrink fitting or the like, and is rotatably provided opposite to the inner surface of the stator 403, and is rotatably provided with a magnetic effect. and a cylindrical rotor 402 that rotates by. The rotating shaft 201 is fitted into the center of the rotor 402 . The electric motor section 401 generates rotational driving force on the rotating shaft 201 using electric power supplied from an external power source, and transmits the rotational driving force to the compression mechanism section 501 via the rotating shaft 201. Note that the electric motor section 401 uses, for example, a brushless DC motor.

The rotating shaft 201 includes a main shaft section 202 fixed to the rotor 402 of the electric motor section 401, a counter shaft section 203 provided on the opposite side of the main shaft section 202 with the compression mechanism section 501 in between, and the main shaft section 202 and the counter shaft section. 203, and a crank part 204 provided between the crank part 203 and the crank part 204. The rotating shaft 201 is formed in the order of a main shaft part 202, a crank part 204, and a counter shaft part 203 from above to below the closed container 100 in the axial direction. The main shaft portion 202 is fitted into the center of the rotor 402 of the electric motor portion 401 and fixed by shrink fitting or press fitting. The central axis of the crank part 204 is eccentric with respect to the central axes of the main shaft part 202 and the counter shaft part 203.

The compression mechanism section 501 compresses the low-pressure gas refrigerant drawn into the low-pressure space of the closed container 100 from the suction pipe 2 into high-pressure gas refrigerant using the rotational driving force supplied from the electric motor section 401. The high-pressure gas refrigerant compressed by the compression mechanism section 501 is discharged into the closed container 100 from above the compression mechanism section 501. The compression mechanism section 501 includes a cylinder 502, an upper bearing 301, a lower bearing 302, a discharge muffler 503, a rolling piston 504, a vane 505 (see FIG. 2), and a vane spring 506 (see FIG. 2). We are prepared.

The outer peripheral portion of the cylinder 502 is fixed to the closed container 100 with bolts or the like. As shown in FIG. 1, the cylinder 502 has an upper surface as one end surface side UP and a lower surface as the other end surface side DN. As shown in FIG. 2, the cylinder 502 has a hollow cylindrical shape, and a cylinder chamber 507 is formed inside the cylinder. As shown in FIG. 1, the cylinder chamber 507 is open at both ends in the axial direction of the rotating shaft 201, and includes an upper bearing 301 provided on the upper surface of the cylinder 502 and a lower bearing 302 provided on the lower surface of the cylinder 502. It is blocked by. In other words, the cylinder chamber 507 is a space surrounded by the inner peripheral surface of the cylinder 502, the inner wall surface of the upper bearing 301, and the inner wall surface of the lower bearing 302.

Further, the cylinder 502 is provided with a suction port (not shown) through which the gas refrigerant from the suction pipe 2 passes, penetrating into the cylinder chamber 507 from the outer peripheral surface. The suction port allows the pipe line of the suction pipe 2 and the cylinder chamber 507 to communicate with each other.

Further, as shown in FIG. 2, the cylinder 502 is formed with a vane groove 508 that communicates with the cylinder chamber 507 and extends in the radial direction about the rotating shaft 201. As shown in FIG. 2, the vane groove 508 penetrates through the cylinder 502 in the axial direction from one end surface side UP toward the other end surface side DN when viewed from a direction in which the outer shape of the cylinder 502 appears circular. A vane 505 that partitions the cylinder chamber 507 into a suction chamber 507a and a compression chamber 507b is slidably fitted into the vane groove 508. The suction chamber 507a is a low pressure space and communicates with the suction port. The compression chamber 507b is a high pressure space and communicates with a discharge port 509 (see FIG. 1) for discharging to the outside of the cylinder chamber 507.

Furthermore, a stop portion 508a is formed at the end of the vane groove 508 on the outer peripheral surface side of the cylinder 502. The stop portion 508a is provided to limit the movement of the vane 505 by stopping the movement of the vane 505 toward the outer circumferential surface of the cylinder 502 so that the vane 505 does not protrude from the outer circumferential surface of the cylinder 502. Further, the stop portion 508a also has the function of introducing high-pressure refrigerant as a back pressure chamber. Note that, as shown in FIG. 2, the stop portion 508a has an arcuate shape that opens only into the vane groove 508 when viewed from the one end surface side UP of the cylinder 502.

Furthermore, as shown in FIG. 2, a vane spring housing hole 508b is formed at the end of the vane groove 508 as a space for housing the vane spring 506 and operating the vane spring 506. Vane spring storage hole 508b is formed to extend in the radial direction of cylinder 502. The length of the vane spring storage hole 508b is determined depending on the shape of the vane spring 506 to be operated or the shape of the cylinder 502.

As shown in FIG. 1, the upper bearing 301 is formed into a substantially inverted T-shape when viewed from the side. The upper bearing 301 is provided on one end surface side UP of the cylinder 502 on the side where the electric motor section 401 is disposed, and closes one opening of the cylinder chamber 507 in the axial direction. Further, the upper bearing 301 is fitted onto the main shaft portion 202 of the rotating shaft 201, and rotatably supports the main shaft portion 202. The upper bearing 301 and the lower bearing 302 are fixed to the cylinder 502 by a common screw 6. The upper bearing 301 is composed of, for example, a sliding bearing.

Note that a discharge port 509 is formed in the upper bearing 301 to discharge the refrigerant compressed in the compression chamber 507b to the outside of the cylinder chamber 507. A discharge valve (not shown) is attached to the discharge port 509. The discharge valve controls the timing at which the high temperature and high pressure gas refrigerant is discharged from the compression chamber 507b through the discharge port 509. Specifically, the discharge valve has a leaf spring mechanism that opens when the pressure of the compressed refrigerant in the compression chamber 507b exceeds a preset pressure to discharge the refrigerant into the closed container 100. Therefore, the discharge valve closes the discharge port 509 when the pressure inside the compression chamber 507b is lower than a preset pressure. Further, when the pressure inside the compression chamber 507b exceeds a preset pressure, the discharge valve is pushed upward by the pressure inside the compression chamber 507b and becomes open. Thereby, the refrigerant is discharged from the compression chamber 507b into the closed container 100 via the discharge port 509.

As shown in FIG. 1, the lower bearing 302 is formed in a generally reverse T-shape when viewed from the side. The lower bearing 302 is provided on the other end surface side DN of the cylinder 502, which is opposite to the side where the electric motor section 401 is arranged, and closes the other opening of the cylinder chamber 507 in the axial direction. Further, the lower bearing 302 is fitted into the subshaft portion 203 of the rotating shaft 201, and rotatably supports the subshaft portion 203. The lower bearing 302 is composed of, for example, a sliding bearing.

As shown in FIG. 1, the discharge muffler 503 is attached to cover the outside of the upper bearing 301. Inside the compression chamber 507b, the operations of sucking in the refrigerant, compressing the refrigerant, and discharging the refrigerant are repeated. The compressed gas refrigerant is intermittently discharged from the discharge port 509. As a result, noise such as pulsation noise may be generated from the cylinder 502. The discharge muffler 503 is provided to suppress noise such as pulsation noise generated from the cylinder 502.

Further, the discharge muffler 503 is provided with a discharge hole (not shown) that communicates the space formed by the discharge muffler 503 and the upper bearing 301 with the inside of the closed container 100. The gas refrigerant discharged from the cylinder 502 through the discharge port 509 is once discharged into the space formed by the discharge muffler 503 and the upper bearing 301, and then discharged into the airtight container 100 from the discharge hole.

As shown in FIG. 2, the rolling piston 504 is formed into a hollow cylindrical shape, and the crank portion 204 of the rotating shaft 201 is slidably fitted into the hollow interior. The rolling piston 504 is housed in a cylinder chamber 507 together with the crank portion 204. When the rotating shaft 201 is rotated by the drive of the electric motor section 401, the rolling piston 504 rotates along the inner peripheral surface of the cylinder chamber 507 and compresses the refrigerant.

As shown in FIG. 2, during the refrigerant compression process, the vane 505 reciprocates inside the vane groove 508 following the rotation of the rolling piston 504 while its tip remains in contact with the outer peripheral surface of the rolling piston 504. move. The cylinder chamber 507 is partitioned into a suction chamber 507a and a compression chamber 507b by the tip of the vane 505 coming into contact with the outer peripheral surface of the rolling piston 504. Vane 505 is made of, for example, a non-magnetic material.

The vane spring 506 contacts the back side of the vane 505 and presses the vane 505 so that the tip of the vane 505 contacts the outer peripheral surface of the rolling piston 504. The vane spring 506 is housed in a vane spring housing hole 508b of the cylinder 502, and is arranged in series with the vane 505. The vane spring 506 is fixed to the cylinder 502 by having its end opposite to the tip that contacts the rolling piston 504 contact the inner wall surface of the vane spring storage hole 508b.

Here, the operation of the rotary compressor 1 will be explained. In the rotary compressor 1 , a rolling piston 504 rotates together with a crank part 204 inside a cylinder chamber 507 when a rotating shaft 201 rotates by driving an electric motor part 401 . In the cylinder chamber 507, the suction chamber 507a partitioned by the vane 505 increases in volume as the rotating shaft 201 rotates. Moreover, in the cylinder chamber 507, the volume of the compression chamber 507b partitioned off by the vane 505 is reduced.

In the rotary compressor 1, the suction chamber 507a and the suction port communicate with each other, and low-pressure gas refrigerant is sucked into the cylinder chamber 507. Next, communication between the compression chamber 507b and the suction port is closed by the rolling piston 504, and as the volume of the compression chamber 507b decreases, the gas refrigerant inside the compression chamber 507b is compressed. The compression chamber 507b and the discharge port 509 of the upper bearing 301 communicate with each other, and when the gas refrigerant inside the compression chamber 507b reaches a preset pressure, the discharge valve provided at the discharge port 509 opens. The gas refrigerant thus compressed to a high pressure and high temperature is discharged from the discharge port 509 to the outside of the cylinder chamber 507 .

The high-pressure and high-temperature gas refrigerant discharged to the outside of the cylinder chamber 507 is discharged into the sealed container 100 via the discharge muffler 503. Then, the discharged gas refrigerant passes through the inside of the electric motor section 401, rises inside the closed container 100, and is discharged to the outside of the closed container 100 from the refrigerant discharge pipe 4 provided at the upper part of the closed container 100. be done. The refrigerant discharged to the outside of the sealed container 100 circulates through the refrigeration circuit and returns to the suction muffler 3 again.

FIG. 3 is an explanatory diagram illustrating the rotational movement of the crank portion 204 and rolling piston 504 provided in the rotary compressor 1 according to the first embodiment. As described above, when the electric motor section 401 rotates, the rolling piston 504 attached to the crank section 204 of the rotating shaft 201 rotates eccentrically within the cylinder 502. At this time, the rolling piston 504 changes its state in the order of (a)→(b)→(c)→(d)→(a)→... as shown in FIG.

The state shown in FIG. 3(a) is when the crank angle is 0 degrees. The crank angle θ1 is the rotation angle of the crank portion 204 when the vane top dead center is the origin and the rotation direction R (see FIG. 2) is positive. The vane top dead center means a state in which the vane 505 is completely housed in the vane groove 508, as shown in FIG. 3(a). The crank angle θ1 at the top dead center of the vane is 0 deg or 360 deg.

Next, the rolling piston 504 transitions to the state shown in FIG. 3(b). In this state, suction of the refrigerant into the suction chamber 507a is started, and at the same time, compression of the refrigerant in the compression chamber 507b is started. The state shown in FIG. 3(b) is a case where the crank angle θ1 is 90 degrees.

Next, the rolling piston 504 transitions to the state shown in FIG. 3(c). In this state, the refrigerant continues to be sucked into the suction chamber 507a, and at the same time, the refrigerant compressed in the compression chamber 507b starts to be discharged. The state shown in FIG. 3(c) is a case where the crank angle θ1 is 180 degrees.

Next, the rolling piston 504 transitions to the state shown in FIG. 3(d). In this state, the refrigerant continues to be sucked into the suction chamber 507a, and the refrigerant compressed in the compression chamber 507b continues to be discharged. The state shown in FIG. 3(d) is a case where the crank angle θ1 is 270 degrees.

Next, the rolling piston 504 returns to the state shown in FIG. 3(a). In this state, the suction of the refrigerant into the suction chamber 507a is completed, and the discharge of the refrigerant compressed in the compression chamber 507b is also completed. The state shown in FIG. 3A is a case where the crank angle θ1 is 360 degrees, that is, a case where the crank angle θ1 is 0 degrees.

FIG. 4 is an explanatory diagram illustrating the principle of gas load generation due to refrigerant compression accompanying the rotational operation of the rolling piston 504 in the rotary compressor 1 according to the first embodiment. In FIG. 4, θ1 represents the crank angle, e represents the outer radius of the rotating shaft 201, and r represents the outer radius of the rolling piston 504. As shown in FIG. 4, compression chamber pressure Pc, which is pressure from the refrigerant in compression chamber 507b, is applied to rolling piston 504. At the same time, suction pressure Ps, which is the pressure from the refrigerant in suction chamber 507a, is applied to rolling piston 504. The differential pressure between the compression chamber pressure Pc and the suction pressure Ps is called "gas load GL." Since the internal pressure of the compression chamber 507b is greater than the internal pressure of the suction chamber 507a, the compression chamber pressure Pc is greater than the suction pressure Ps. Therefore, the direction of the vector of the gas load GL is from the compression chamber 507b toward the suction chamber 507a, as shown by the white arrow in FIG. In the first embodiment, the action angle θ2 of the gas load GL is used as an index indicating the direction of the vector of the gas load GL. Like the crank angle θ1, the action angle θ2 of the gas load GL is defined with the vane top dead center (see FIG. 3) as the origin and the rotation direction R (see FIG. 2) as positive. Since the position of the compression chamber 507b changes according to the crank angle θ1, the action angle θ2 of the gas load GL also changes. In this way, the gas load GL substantially acts on the surface of the rolling piston 504, but also acts on the rotating shaft 201 via the rolling piston 504, as described above.

FIG. 5 is a diagram showing an example of a change in the gas load GL with respect to the crank angle θ1 in the rotary compressor 1 according to the first embodiment. In addition, in FIG. 5, the gas load GL is normalized by the maximum value. In FIG. 5, the horizontal axis shows the crank angle θ1, and the vertical axis shows the normalized gas load GL. Since the volume of the compression chamber 507b changes as the crank angle θ1 changes, as shown in FIG. 5, the magnitude of the gas load GL changes in accordance with the change in the crank angle θ1. Specifically, as shown in FIG. 5, when the crank angle θ1 is around 225 degrees, the gas load GL becomes the largest. That is, when the crank angle θ1 is in the range from 0 deg to 120 deg, the gas load GL gradually increases, and when the crank angle θ1 is in the range from 120 deg to 225 deg, the gas load GL increases rapidly. Further, in the range of the crank angle θ1 from 225 degrees to 360 degrees, the magnitude of the gas load GL rapidly decreases.

FIG. 6 is a diagram showing an example of the operating angle θ2 of the gas load GL with respect to the crank angle θ1 in the rotary compressor 1 according to the first embodiment. In FIG. 6, the horizontal axis indicates the crank angle θ1, and the vertical axis indicates the action angle θ2 of the gas load GL. As shown in FIG. 6, as the crank angle θ1 changes, the volume of the compression chamber 507b changes, so the action angle θ2 of the gas load GL changes. In the example of FIG. 6, the range in which the operating angle θ2 of the gas load GL changes, that is, the variation range of the operating angle θ2 of the gas load GL, with respect to the entire range of the crank angle θ1, that is, the range of θ1 from 0 deg to 360 deg. is in the range of 110deg to 330deg. Hereinafter, this range will be referred to as an angular range B1 indicating the variation range of the action angle θ2. Specifically, as shown in FIG. 6, when the crank angle θ1 is around 180 degrees, the action angle θ2 of the gas load GL is the smallest, θ2=110 degrees. That is, θ2=110 degrees is the lower limit value of the angle range B1. Further, when the crank angle θ1 is 360 degrees, the action angle θ2 of the gas load GL is the largest, and θ2=330 degrees. That is, θ2=330 degrees is the upper limit of the angle range B1. Further, when the crank angle θ1 is 0 deg, the action angle θ2 of the gas load GL is 180 deg, and thereafter, in the range of the crank angle θ1 from 0 deg to 30 deg, the action angle θ2 of the load GL decreases. Further, in the range of the crank angle θ1 from 30 degrees to 180 degrees, the action angle θ2 of the load GL remains unchanged or gradually decreases. When the crank angle θ1 is in the range from 180 degrees to 355 degrees, the acting angle θ2 of the load GL gradually increases, and when the crank angle θ1 is in the range from 355 degrees to 360 degrees, the acting angle θ2 of the load GL increases rapidly.

As described above, in the rotary compressor 1 according to the first embodiment, as shown in FIGS. 5 and 6, the magnitude of the gas load GL and the operating angle θ2 change as the crank angle θ1 changes.

In the first embodiment, it is assumed that the bearing section 300 is composed of a sliding bearing. That is, at least one of the upper bearing 301 and the lower bearing 302 is configured as a sliding bearing. Further, in the first embodiment, the sliding bearing has anisotropy in oil film rigidity. The oil film rigidity is the stiffness of the refrigerating machine oil interposed between the rotating shaft 201 and the bearing portion 300, and the rotating shaft 201 is supported by the rigidity of the refrigerating machine oil. Therefore, the oil film rigidity functions as bearing support rigidity.

Here, a method for creating anisotropy in the oil film rigidity of a sliding bearing will be explained. One method is to make the shape of the inner diameter surface 300a (for example, see FIG. 12) of the sliding bearing constituting the bearing portion 300 into a non-perfect circular shape. The inner diameter surface 300a is a sliding surface on which the rotating shaft 201 slides. The sliding bearing has a cylindrical shape and is hollow inside. A rotating shaft 201 is arranged within the cavity. Sliding bearings are made of copper alloys such as brass, iron, or special resin. Refrigerating machine oil is injected into a gap 400 (see, for example, FIG. 12) formed between the inner diameter surface 300a of the sliding bearing and the rotating shaft 201.

For example, as shown in FIG. 7, when the shape of the inner diameter surface 300a of the sliding bearing constituting the bearing part 300 is a perfect circle, there is a gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 (see FIG. 7) are uniform in width. In this case, the thickness of the oil film of the refrigerating machine oil filled in the gap 400 is also uniform, and the oil film thickness does not change during one revolution of the rolling piston 504. Therefore, anisotropy does not appear in the oil film stiffness. Here, one rotation of the rolling piston 504 means that the rolling piston 504 rotates in a range of 360 degrees from a state where the crank angle θ1 is 0 degrees until the crank angle θ1 becomes 360 degrees, as shown in FIG. It is.

On the other hand, as shown in FIG. 12, for example, when the shape of the inner diameter surface 300a of the sliding bearing constituting the bearing part 300 is a non-perfect circular shape, a gap is formed between the inner diameter surface 300a and the rotating shaft 201. 400 is not uniform. In other words, the gap 400 includes areas where the gap 400 is narrow and areas where the gap 400 is wide. In that case, the oil film thickness of the refrigerating machine oil filled in the gap 400 is no longer uniform, and the oil film thickness changes during one rotation of the rolling piston 504. Therefore, anisotropy appears in the oil film stiffness. Below, some examples of the shapes of journal type sliding bearings are shown in FIGS. 7 to 22. Hereinafter, the anisotropy of oil film rigidity will be explained in more detail using a specific example of a sliding bearing.

Among FIGS. 7 to 22, examples of the shape of the inner diameter surface having anisotropy include, for example, the two-arc bearing shown in FIG. 12. In a sliding bearing, the size of the gap 400 changes in the rotational direction of the rolling piston 504. As a result, in the regions S1 and S2 where the gap 400 narrows, pressure is generated in the fluid such as refrigerating machine oil flowing through the gap 400, thereby generating oil film rigidity. In the two-arc bearing, as shown in FIG. 12, even when the rotating shaft 201 and the inner diameter surface 300a are concentric, the inner diameter surface 300a is not perfectly circular, so the gap 400 is narrowed in the rotation direction of the rolling piston 504. A distribution occurs. Therefore, in FIG. 12, a high oil film pressure is generated particularly in the direction where the region S1 and the region S2 where the gap 400 is small exist, that is, in the direction perpendicular to the plane of the paper of FIG. Therefore, in a bearing as shown in FIG. 12, the oil film rigidity in the vertical direction where the gap 400 is small is higher than the oil film rigidity in the horizontal direction where the gap 400 is large. Hereinafter, the direction in which the oil film stiffness is highest will be referred to as "the axial direction D of high oil film stiffness." In the case of the two-arc bearing shown in FIG. 12, the axial direction D of high oil film rigidity is the direction in which regions S1 and S2 where the gap 400 is small exist, as shown by thick black arrows in FIG. 24, which will be described later.

Here, in order to make the explanation easier to understand, the two-circle-arc bearing shown in FIG. 12 was explained as an example, but the first embodiment is not limited to that case. In other words, it is not limited to the shape examples shown in FIGS. 7 to 22, but any shape can be used as long as the shape of the inner diameter surface includes a surface structure such as a texture that causes anisotropy in the oil film stiffness of the sliding bearing. , the features of Embodiment 1 can be applied. In addition, as a method that does not change the shape of the inner diameter surface 300a from a perfect circular shape, it is possible to increase the bearing support rigidity by controlling the electromagnetic force of an electromagnet using a magnetic bearing or by controlling the supply pressure of a hydrostatic bearing. It is also conceivable to impart anisotropy to the plain bearing. In addition, when controlling the electromagnetic force of the electromagnet, the electromagnetic force of the electromagnet is not made uniform around the entire circumference of the plain bearing, but the electromagnetic force is strengthened partially, and the anisotropy of the bearing support rigidity is applied to the plain bearing. give. In addition, when controlling the supply pressure of a hydrostatic bearing, the supply pressure is not made uniform around the entire circumference of the plain bearing, but the supply pressure is strengthened in some areas to create anisotropy in the bearing support rigidity of the plain bearing. give.

Hereinafter, the journal type sliding bearing shown in FIGS. 7 to 22 will be explained.

FIG. 7 is a plan view showing the configuration of the full circumference bearing. Full circumference bearings are the most common type of bearing and are suitable for example in the case of dynamic loads. As shown in FIG. 7, in the full circumference bearing, the inner diameter surface 300a of the sliding bearing has a perfect circular shape in plan view. Therefore, when the sliding bearing constituting the bearing portion 300 is a full-circumference bearing, the sliding bearing does not have anisotropy in oil film rigidity. Therefore, when the plain bearing that constitutes the bearing section 300 is a full-circumference bearing, by configuring the plain bearing from a magnetic bearing and controlling the electromagnetic force of the electromagnet, anisotropy of bearing support rigidity is given to the plain bearing. Just do it like this. Alternatively, the sliding bearing may be configured with a static pressure bearing and the supply pressure may be controlled to give the sliding bearing anisotropy in bearing support rigidity.

FIG. 8 is a plan view showing the configuration of the partial bearing. As shown in FIG. 8, the partial bearing supports only a portion of the outer circumference of the rotating shaft 201. As shown in FIG. In the partial bearing, the shape of the inner diameter surface 300a of the sliding bearing is a part of the outer periphery of a perfect circle when viewed from above. Therefore, the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 is formed only over a portion of the outer circumference of the rotating shaft 201. Therefore, when the sliding bearing constituting the bearing portion 300 is a partial bearing, the sliding bearing has anisotropy in oil film rigidity.

FIG. 9 is a plan view showing the configuration of the split bearing. As shown in FIG. 9, in the split bearing, the shape of the inner diameter surface 300a of the sliding bearing is composed of two curved lines with uneven curvatures when viewed from above. Therefore, the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 is not uniform around the entire circumference of the rotating shaft 201. Therefore, when the sliding bearing constituting the bearing portion 300 is a split bearing, the sliding bearing has anisotropy in oil film rigidity. Although FIG. 9 shows a case where the inner diameter surface 300a of the sliding bearing is divided into two parts, the present invention is not limited to this case. That is, the split bearing may be divided into three or more parts. In that case, the inner diameter surface 300a of the sliding bearing is composed of three or more curved lines.

FIG. 10 is a plan view showing the configuration of the fitted bearing. As shown in FIG. 10, the fitted bearing supports only a portion of the outer circumference of the rotating shaft 201. As shown in FIG. The fitted bearing is a type of partial bearing, but compared to the partial bearing shown in FIG. 8, the fitted bearing has almost no gap 400 formed between the inner diameter surface 300a of the sliding bearing and the rotating shaft 201. Therefore, in the fitted bearing, the oil film pressure distribution becomes uniform. However, in the fitted bearing, the shape of the inner diameter surface 300a of the sliding bearing in cross-sectional view is a part of a perfect circle. Therefore, when the sliding bearing that constitutes the bearing portion 300 is a fitted bearing, the sliding bearing has anisotropy in oil film rigidity.

FIG. 11 is a plan view showing the configuration of the step bearing. Step bearings are sometimes called stepped bearings. As shown in FIG. 11, the step bearing has one or more steps 300m formed on the inner diameter surface 300a of the sliding bearing. The step 300m is formed in a concave shape on the inner diameter surface 300a. The step 300m protrudes radially outward with respect to the inner diameter surface 300a of the slide bearing. Therefore, the step 300m is recessed radially outward from the inner diameter surface 300a. Therefore, in the region where the step 300m is provided, the gap 400 is wider than in other parts. In the example of FIG. 11, four steps 300m are formed at intervals in the circumferential direction of the inner diameter surface 300a. The number and shape of steps 300m are not limited to the example of FIG. 11. In the step bearing, since one or more steps 300m are formed on the inner diameter surface 300a, the shape of the inner diameter surface 300a of the sliding bearing is non-circular. Therefore, when the sliding bearing that constitutes the bearing section 300 is a step bearing, the sliding bearing has anisotropy in oil film rigidity.

FIG. 12 is a plan view showing the configuration of a two-arc bearing. As shown in FIG. 12, in the two-arc bearing, the shape of the inner diameter surface 300a of the sliding bearing is composed of two arcs. As shown in FIG. 12, the two circular arcs have the same shape and are arranged line-symmetrically. Furthermore, the curvatures of the two circular arcs are different from the curvature of the outer periphery of the bearing portion 300. Therefore, the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 is not uniform, and includes areas where the gap 400 is narrow and areas where the gap 400 is wide. When the sliding bearing constituting the bearing portion 300 is a two-arc bearing, the sliding bearing has anisotropy in oil film rigidity. The two-arc bearing is sometimes called a multi-arc bearing because the inner diameter surface 300a is composed of a plurality of arc surfaces.

FIG. 13 is a plan view showing the configuration of a three-arc bearing. As shown in FIG. 13, in the three-arc bearing, the shape of the inner diameter surface 300a of the sliding bearing is composed of three arcs. The three arcs have the same shape, as shown in FIG. The three circular arcs are arranged adjacent to each other so as to be convex toward the outside in the radial direction. Furthermore, the curvatures of the three arcs are different from the curvature of the outer periphery of the bearing portion 300. Therefore, the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 is not uniform, and includes areas where the gap 400 is narrow and areas where the gap 400 is wide. When the sliding bearing constituting the bearing portion 300 is a three-arc bearing, the sliding bearing has anisotropy in oil film rigidity. The three-arc bearing is sometimes called a multi-arc bearing because the inner diameter surface 300a is composed of a plurality of arc surfaces.

FIG. 14 is a plan view showing the configuration of a two-arc eccentric bearing. As shown in FIG. 14, in the two-arc eccentric bearing, the shape of the inner diameter surface 300a of the sliding bearing is composed of two arcs. As shown in FIG. 14, the two circular arcs have the same shape and are arranged point-symmetrically with respect to the axial center of the rotating shaft 201. Furthermore, the curvatures of the two circular arcs are different from the curvature of the outer periphery of the bearing portion 300. Furthermore, the centers of the circles forming the two arcs are shifted from each other in the horizontal direction. Therefore, the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 is not uniform, and includes areas where the gap 400 is narrow and areas where the gap 400 is wide. Further, since the two circular arcs are arranged horizontally shifted, a straight line portion for connecting the two circular arcs is provided between the two circular arcs. That is, the inner diameter surface 300a of the two-arc eccentric bearing is formed from two arcs and two straight parts. When the sliding bearing constituting the bearing section 300 is a two-arc eccentric bearing, the sliding bearing has anisotropy in oil film rigidity. The two-arc eccentric bearing is sometimes called a multi-arc bearing because the inner diameter surface 300a is composed of a plurality of arcuate surfaces.

FIG. 15 is a plan view showing the configuration of the floating bush bearing. Floating bush bearings were originally devised for the purpose of reducing heat generation in high-speed bearings, and are suitable, for example, when the rotating shaft 201 rotates at high speeds. In the floating bush bearing, as shown in FIG. 15, a thin cylindrical bush 300b is inserted into a gap 400 formed between an inner diameter surface 300a and a rotating shaft 201. At this time, when the gap 400 is supplied with oil, the rotating shaft 201 is supported by the double oil film formed in series on both the inside and outside of the bush 300b. When the rotating shaft 201 rotates, the bush 300b floats within the gap 400, and the rotating shaft 201 is rotated by the driving torque of the inner oil film. The rotational speed of the rotating shaft 201 is determined by the balance between the driving torque of the inner oil film and the braking torque of the outer oil film. As shown in FIG. 15, in the floating bush bearing, the inner diameter surface 300a of the sliding bearing has a perfect circular shape. Therefore, when the sliding bearing that constitutes the bearing section 300 is a floating bush bearing, the sliding bearing does not have anisotropy in oil film rigidity. Therefore, when the sliding bearings constituting the bearing section 300 are floating bush bearings, the sliding bearings are constructed from magnetic bearings, or the sliding bearings are constructed from static pressure bearings, so that the sliding bearings have anisotropy of oil film rigidity. All you have to do is give it.

FIG. 16 is a plan view showing the configuration of the pivot pad bearing. A pivot pad bearing is a type of tilting pad bearing. In the pivot pad bearing, as shown in FIG. 16, a plurality of divided pads 300c are arranged in a gap 400 formed between an inner diameter surface 300a and a rotating shaft 201. A fulcrum called a pivot 300d is arranged between each pad 300c and the inner diameter surface 300a. By supporting each pad 300c with a pivot 300d, a wedge film is formed on each pad 300c. In this manner, in the pivot pad bearing, a wedge film effect is obtained in which pressure is generated by forcing the viscous fluid into the wedge-shaped gap 400, which is narrower at the outlet than at the inlet. As shown in FIG. 16, in the pivot pad bearing, the inner diameter surface 300a of the sliding bearing has a perfect circular shape, but since the pivot 300d and the pad 300c are arranged within the gap 400, the thickness of the oil film is uniform. Not like that. Therefore, when the sliding bearing that constitutes the bearing section 300 is a pivot pad bearing, the sliding bearing has anisotropy in oil film rigidity.

FIG. 17 is a plan view showing the configuration of the Michel-Seggell bearing. A Michel-Seggell bearing is a type of tilting pad bearing. In the Michel-Seggell bearing, as shown in FIG. 17, a plurality of pads 300c are arranged in a gap 400 formed between an inner diameter surface 300a and a rotating shaft 201. The difference from the pivot pad bearing shown in FIG. 16 is that the Michelle-Seggell bearing does not use a pivot 300d. In the Michel-Seggell bearing, pad 300c rotates with rotating shaft 201. As shown in FIG. 17, in the Mitchell-Seggell bearing, the inner diameter surface 300a of the plain bearing has a perfect circular shape, but since the pad 300c is arranged within the gap 400, the thickness of the oil film is not uniform. do not have. Therefore, when the sliding bearing constituting the bearing portion 300 is a Mitchell-Seggell bearing, the sliding bearing has anisotropy in oil film rigidity.

FIG. 18 is a plan view showing the configuration of the NOUMU bearing. The Neumu bearing is a type of tilting pad bearing. In the NOUMU bearing, as shown in FIG. 18, a plurality of pads 300c are arranged in a gap 400 formed between an inner diameter surface 300a and a rotating shaft 201. The Noumyu bearing is a bearing named because of its low friction, and the pad surface of the pad 300c has a spherical shape. The Neumu bearing, like the Michel-Seggell bearing, does not use a pivot 300d. Therefore, in the NOUMU bearing, the pad 300c rotates together with the rotating shaft 201. However, in the example shown in FIG. 18, a protrusion 201a is provided on the side surface of the rotating shaft 201. The convex portion 201a protrudes radially outward from the side surface of the rotating shaft 201. The convex portion 201a restricts movement of the pad 300c in the circumferential direction. That is, the pad 300c can move in the circumferential direction only between the two adjacent convex portions 201a. As shown in FIG. 18, the inner diameter surface 300a of the sliding bearing of the NOUMU bearing has a perfect circular shape, but since the pad 300c is disposed within the gap 400, the thickness of the oil film is not uniform. Therefore, when the sliding bearing that constitutes the bearing section 300 is a Neumu bearing, the sliding bearing has anisotropy in oil film rigidity.

FIG. 19 is a plan view showing the configuration of a porous bearing. As shown in FIG. 19, the porous bearing has a thin cylindrical porous member 300e made of a porous material attached to an inner diameter surface 300a. At this time, when oil is supplied to the gap 400 formed between the porous member 300e and the rotating shaft 201, the rotating shaft 201 is supported by the oil film formed between the porous member 300e and the rotating shaft 201. That will happen. A large number of holes 300ee are formed on the surface of the porous member 300e. The hole 300ee is formed in a concave shape in the inner diameter surface 300a. Further, the arrangement of the holes 300ee is not evenly distributed. Therefore, when the refrigerating machine oil enters the hole 300ee, the thickness of the oil film becomes substantially non-uniform. As shown in FIG. 19, in the porous bearing, the inner diameter surface 300a of the sliding bearing has a perfect circular shape, but by providing the porous member 300e on the inner diameter surface 300a, the oil film rigidity has anisotropy. are doing.

FIG. 20 is a plan view showing the configuration of the foil bearing. The foil bearing is a bearing formed of a plurality of foils 300f, as shown in FIG. 20. In the foil bearing, the size of the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 is automatically determined by the generated film pressure. As shown in FIG. 20, the inner surface of the foil 300f constitutes an inner diameter surface 300a of the sliding bearing. The inner surfaces of the three foils 300f have the same shape, as shown in FIG. 20. The inner surfaces of the three foils 300f are arranged adjacent to each other so as to be convex toward the outside in the radial direction. The inner surfaces of the three foils 300f are each formed of a curved surface, but the curvature is not uniform and changes along the way. Therefore, as shown in FIG. 20, the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 is not uniform, and includes areas where the gap 400 is narrow and areas where the gap 400 is wide. When the sliding bearing that constitutes the bearing section 300 is a foil bearing, the sliding bearing has anisotropy in oil film rigidity.

FIG. 21 is a cross-sectional view schematically showing the configuration of a spiral grooved bearing. However, in FIG. 21, for the sake of explanation, the rotating shaft 201 is shown not in cross section but in a side view. In the spiral grooved bearing, the sliding bearing that constitutes the bearing portion 300 has a cylindrical shape. The rotating shaft 201 is arranged in a cylindrical hollow part. Further, in a spiral grooved bearing, a "groove" is provided on either the rotating shaft 201 or the inner diameter surface 300a of the sliding bearing. In the example of FIG. 21, a plurality of thin twisted grooves 300k are provided as the "grooves". The "groove" is formed in a concave shape on the inner diameter surface 300a. Further, in the example of FIG. 21, the twisted groove 300k is provided on the side surface of the rotating shaft 201, but the twisted groove 300k may be provided on the inner diameter surface 300a side. The spiral grooved bearing is a type of bearing that obtains oil pressure by drawing refrigerating machine oil into the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 by a pumping action when the rotating shaft 201 rotates. In spiral grooved bearings, the helix angle of the groove, the depth of the groove, the groove and the land are calculated based on the local characteristics of the stepped bearing, and by applying the theory of infinite grooves to determine the pressure within the gap 400. The twisted groove 300k is formed by determining the optimum value such as the width ratio between the two. In the example of FIG. 21, a plurality of twisted grooves 300k are arranged at intervals in the circumferential direction of the rotating shaft 201, thereby forming a row of twisted grooves 300k. In the example of FIG. 21, two rows of twisted grooves 300k are provided. The two rows of twisted grooves 300k are formed side by side in the axial direction of the rotating shaft 201. Furthermore, the twisting directions of the twisting grooves 300k in the left row of the paper in FIG. 21 and the twisting grooves 300k in the right row are opposite to each other. As shown in FIG. 21, a type in which helical grooves with opposite torsional directions are provided from positions corresponding to both ends of the inner diameter surface 300a and also over the entire inner diameter surface 300a is called a herringbone groove bearing. Sometimes. Note that the number and shape of the twisted grooves 300k are not limited to the example shown in FIG. 21. In the spiral grooved bearing, one or more helical grooves 300k as "grooves" are formed on the rotating shaft 201 or the inner diameter surface 300a. Therefore, the shape of the gap 400 formed between the inner diameter surface 300a of the sliding bearing and the rotating shaft 201 is not uniform. Therefore, when the sliding bearing constituting the bearing portion 300 is a bearing with a spiral groove, the sliding bearing has anisotropy in oil film rigidity.

FIG. 22 is a cross-sectional view schematically showing the configuration of a spherical bearing. However, in FIG. 22, for the sake of explanation, the rotating shaft 201 is shown not in cross section but in a side view. In the spherical bearing, as shown in FIG. 22, the sliding bearing is composed of an inner ring 300g and an outer ring 300h. Both the inner ring 300g and the outer ring 300h have a cylindrical shape. The inner ring 300g is arranged concentrically inside the outer ring 300h. The shape of the outer peripheral surface of the inner ring 300g is spherical, as shown in FIG. Moreover, the shape of the inner circumferential surface of the outer ring 300h is spherical, as shown in FIG. Therefore, the inner ring 300g and the outer ring 300h are in spherical contact. The contact surface between the inner ring 300g and the outer ring 300h may be specially processed or a member called a liner may be attached to the contact surface to improve slippage. In the spherical bearing, when the rotating shaft 201 rotates, the inner ring 300g tilts with respect to the outer ring 300h. Since the spherical bearing has a structure in which the inner ring 300g and the outer ring 300h are in spherical contact, it is a self-aligning bearing that can simultaneously bear a radial load in the radial direction and an axial load in the axial direction. In the spherical bearing, the inner diameter surface 300a of the sliding bearing is constituted by the inner circumferential surface of the inner ring 300g. The inner ring 300g is inclined with respect to the axial direction of the rotating shaft 201 as the rotating shaft 201 rotates. Therefore, the shape of the gap 400 formed between the inner diameter surface 300a of the sliding bearing and the rotating shaft 201 is not uniform. Therefore, when the sliding bearing that constitutes the bearing section 300 is a spherical bearing, the sliding bearing has anisotropy in oil film rigidity.

When the sliding bearing has anisotropy in oil film rigidity as shown in FIGS. 7 to 22 described above, it is possible to determine "the axial direction D of high oil film rigidity" which indicates the direction in which the oil film rigidity is highest. can. Since "the axial direction D of high oil film rigidity" differs depending on the type of sliding bearing that constitutes the bearing part 300, for example, by experiment or simulation, "the axial direction D of high oil film rigidity" has been determined for each type of sliding bearing. All you have to do is decide. For example, as shown in FIG. 12, when the plain bearing constituting the bearing part 300 is a two-arc bearing, the "axial direction D of high oil film rigidity" indicating the direction in which the oil film rigidity is highest is the area S1 and the area S2. is the direction in which it exists. That is, to explain using FIG. 24, which will be described later, "the axial direction D of high oil film rigidity" is the direction in which regions S1 and S2 where the gap 400 becomes narrow exist, as shown by the thick black arrow in FIG. . In the first embodiment, as shown in FIG. 24, "the axial direction D of high oil film rigidity" is set to a direction opposite to the action angle θ2 where the gas load GL acts largely.

FIG. 23 is an explanatory diagram showing angular ranges A, B1, and C in which the "axial direction D of high oil film rigidity" is set in the rotary compressor 1 according to the first embodiment. FIG. 24 is an explanatory diagram showing the direction in which "the axial direction D of high oil film rigidity" is set in the rotary compressor 1 according to the first embodiment.

As explained using FIG. 6 above, the angular range B1 indicating the variation range of the action angle θ2 of the gas load GL is 110 deg to 330 deg. FIG. 23 illustrates the angular range B1. As can be seen from the angle range B1 shown in FIG. 23, it can be seen that the action angle θ2 of the gas load GL generally acts in a vertically downward direction. In the sliding bearing, the position of the axial center of the rotating shaft 201 changes to a position eccentric in the rotational direction with respect to the acting direction of the gas load GL acting on the bearing portion 300 due to oil film pressure generation due to rotation. To explain using the example of FIG. 2 described above, when the gas load GL acts in a vertically downward direction as shown by arrow W, the axis center is at a position of 180 deg to 270 deg in the third quadrant, as shown by arrow E. It is approximately eccentric in the angular range of .

In the sliding bearing that constitutes the bearing section 300 of the rotary compressor 1, factors that influence the size of the minimum oil film thickness and the position where the minimum oil film is formed include "the size of the gas load GL" and "the size of the gas load GL". There is an "action angle θ2" and an "eccentric direction of the shaft center" of the rotating shaft 201. In addition, there are also influencing factors such as the viscosity coefficient of the refrigerating machine oil and the rotation speed of the rotating shaft 201. However, in the first embodiment, for the purpose of increasing the minimum oil film thickness under a fluctuating load, the above three points, ie, "size of gas load GL", "action angle θ2 of gas load GL", and "axis center This will be explained with respect to the eccentric direction of.

As shown in the graph of FIG. 5, in the rotary compressor 1, the magnitude of the gas load GL due to refrigerant compression changes during one rotation in which the crank angle θ1 changes from 0 degrees to 360 degrees. In the bearing portion 300 where the oil film rigidity is anisotropic, as described above, there are a direction in which the oil film rigidity is high and a direction in which the oil film rigidity is low.

In the rotary compressor 1, the thickness of the oil film of the refrigerating machine oil supplied to the gap 400 formed between the bearing portion 300 and the rotating shaft 201 changes during one rotation of the rolling piston 504. In particular, at the crank angle θ1 where the gas load GL is maximum, there is a concern that the rotating shaft 201 and the bearing portion 300 will come into contact, and in the worst case, seizure may occur. Therefore, for example, if a direction with low oil film rigidity is placed in the direction of crank angle θ1 where the gas load GL is maximum, there is a concern that the rotating shaft 201 will come into contact with the bearing part 300, and in the worst case, it may lead to seizure. There is. Therefore, in the first embodiment, the "axis direction D of high oil film rigidity" indicating the direction in which the oil film rigidity is highest is arranged in the direction of the crank angle θ1 where the gas load GL is large. Note that the contact between the rotating shaft 201 and the bearing portion 300 is contact between metals, and is therefore sometimes referred to as metal contact.

Therefore, in the first embodiment, for example, the angular range of the crank angle θ1 is such that a gas load GL (normal load = 0.5 [-]) of 50% or more with respect to the maximum load (normal load = 1 [-]) acts. The “axis direction D of high oil film rigidity” of the bearing portion 300 having anisotropic oil film rigidity is arranged at . By doing so, it is possible to suppress a decrease in the oil film thickness. According to the graph of FIG. 5, the range of the crank angle θ1 that includes the region where the normal load is 0.5 or more is the angle range A of 180 degrees to 330 degrees. Therefore, by arranging "the axial direction D of high oil film rigidity" within the angular range A, it is possible to suppress a decrease in the oil film thickness. Specifically, "the axial direction D of high oil film rigidity" indicated by the thick black arrow shown in FIG. 24 is arranged in any direction within the angular range A of FIG. 23. Since the angular range A is the angular range determined from the normalized graph of FIG. 5, it is sometimes referred to as the "gas load absolute value reference angular range." Further, the angular range A is sometimes referred to as a "second angular range."

Furthermore, according to the graph of FIG. 6, the angular range B1 indicating the variation range of the operating angle θ2 of the gas load GL is 110 degrees to 330 degrees, as described above. Therefore, "the axial direction D of high oil film rigidity" may be arranged within the angular range B1. Even in that case, reduction in oil film thickness can be suppressed. Specifically, "the axial direction D of high oil film rigidity" indicated by the thick black arrow shown in FIG. 24 is arranged in any direction within the angular range B1 of FIG. 23. Since the angular range B1 is an angular range determined based on the graph of FIG. 6 showing the acting angle θ2 of the gas load GL, it is sometimes referred to as a "gas load acting direction reference angular range." Further, the angular range B1 is sometimes referred to as a "first angular range."

Furthermore, according to the graph in FIG. 6, the action angle θ2 of the gas load GL corresponding to the angle range A in FIG. 5 is in the angle range B2 of 110 degrees to 180 degrees. Therefore, "the axial direction D of high oil film rigidity" may be arranged within the angular range B2. Even in that case, reduction in oil film thickness can be suppressed. Further, the angular range B2 is sometimes referred to as a "third angular range".

As described above, the acting angle θ2 of the gas load GL generally acts in a downward direction perpendicular to the paper plane of FIG. In the example of FIG. 2 described above, when a load is applied in a vertically downward direction, the axis is eccentric in an angular range C of 180 degrees to 270 degrees, which is the position of the third quadrant. Therefore, "the axial direction D of high oil film rigidity" may be arranged in the angular range C of 180 degrees to 270 degrees. Specifically, "the axial direction D of high oil film rigidity" indicated by the thick black arrow shown in FIG. 24 is arranged in any direction within the angular range C of FIG. 23. The angle range C is sometimes referred to as "the eccentricity characteristic reference angle range of the sliding bearing." Further, the angular range C is sometimes referred to as a "fourth angular range."

FIGS. 23 and 24 respectively show angular ranges A, B1, and C corresponding to each of the above criteria, and an example of bearing arrangement using a two-arc bearing as an example. The angular range of the crank angle θ1 in which the “axial direction D of high oil film rigidity” is arranged may be selected by comprehensively considering each criterion shown in FIG. 23.

Furthermore, we disassembled a rotary compressor using a conventional perfectly circular plain bearing after it was put into operation, measured the wear range of the shaft surface or bearing surface, that is, the metal contact range, and determined that this area had a high oil film rigidity. It is also conceivable to arrange it in the axial direction D.

The arrangement of “high oil film rigidity in the axial direction D” described in the first embodiment is applicable to both or one of the upper bearing 301 and the lower bearing 302. Furthermore, one of the bearings may be a sliding bearing having anisotropy in oil film stiffness, and the other bearing may be a conventional perfectly circular sliding bearing.

As described above, the rotary compressor 1 according to Embodiment 1 includes the rotating shaft 201 and the bearing portion 300 that is configured from a sliding bearing and has an inner diameter surface 300a that supports the rotating shaft 201. Furthermore, the sliding bearing that constitutes the bearing portion 300 has anisotropy in oil film rigidity. The "high oil film rigidity axial direction D" indicating the direction in which the oil film rigidity of the sliding bearing is highest is arranged in an angular range B1 in which the crank angle of the crank portion 204 of the rotating shaft 201 is from 110 degrees to 330 degrees. Thereby, a decrease in the oil film thickness of the refrigerating machine oil filled in the gap 400 formed between the rotating shaft 201 and the bearing portion 300 can be suppressed. In the rotary compressor 1, the oil film thickness changes during one rotation of the rolling piston 504. By setting "the axial direction D of high oil film rigidity" within the angular range A, even at the crank angle θ1 (= around 225 deg) where the gas load GL is maximum, the rotating shaft 201 and the inner diameter surface 300a of the bearing part 300 metal contact can be avoided. As a result, the occurrence of seizure of the inner diameter surface 300a of the bearing portion 300 can be suppressed.

Embodiment 2.
In the first embodiment described above, in the rotary compressor 1, even when the magnitude of the gas load GL is large, the decrease in the oil film thickness and the occurrence of metal contact caused by it are suppressed with respect to the fluctuating gas load GL. are doing. For this purpose, in the first embodiment, a sliding bearing having anisotropy in oil film rigidity is used. Then, taking into consideration the size of the gas load GL, the operating angle θ2 of the gas load GL, the eccentricity characteristics of the plain bearing, etc., the "axial direction D of high oil film rigidity" is arranged within a preset angular range A. There is.

However, in this case, in the axial direction D of high oil film rigidity, it is possible to suppress the decrease in oil film thickness and the accompanying metal contact, and avoid seizure, which was a conventional problem, but in the axial direction of low oil film rigidity, In this case, the shaft vibration may increase due to the fluctuating gas load GL.

Therefore, in the second embodiment, both the upper bearing 301 and the lower bearing 302 are constructed of sliding bearings having anisotropy in oil film rigidity. The upper bearing 301 is arranged so that the "axial direction D of high oil film rigidity" and the "axial direction D of high oil film rigidity" of the lower bearing 302 are different from each other. Since the other configurations are the same as those in Embodiment 1, their description will be omitted here.

FIG. 25 is an explanatory diagram showing the direction in which the "axial direction D of high oil film rigidity" of the upper bearing 301 and the lower bearing 302 in the rotary compressor 1 according to the second embodiment is set. In FIG. 25, D1 indicates "the axial direction of low oil film rigidity" which indicates the direction in which the oil film rigidity is the smallest in the upper bearing 301.

A preferred form is shown in FIG. 25, for example. In the example of FIG. 25, "the axial direction D of high oil film rigidity" is arranged in the angular range A of the crank angle θ1 in which the upper bearing 301 aims to avoid seizure. The "axial direction D of high oil film rigidity" of the lower bearing 302 is arranged in a direction parallel to the "axial direction D1 of low oil film rigidity" of the upper bearing 301. Thereby, even if the shaft vibration increases in the axial direction D1 with the low oil film rigidity of the upper bearing 301, the shaft vibration can be suppressed due to the high oil film rigidity of the lower bearing 302.

As described above, in the second embodiment, when the rotary compressor 1 has a plurality of sliding bearings, the axial direction D of the "high oil film rigidity" of at least two or more sliding bearings among the plurality of sliding bearings. ” are arranged so that they are different from each other. That is, the "high oil film rigidity axial direction D" of at least two or more slide bearings are arranged in a direction that intersects with each other. In the example of FIG. 25, "the axial direction D of high oil film rigidity" of the upper bearing 301 and the "axial direction D of high oil film rigidity" of the lower bearing 302 are orthogonal to each other. This makes it possible to suppress an increase in shaft vibration due to the fluctuating gas load GL in the axial direction with low oil film rigidity.

Note that, as described above, the sliding bearing has the characteristic of being eccentric in the rotational direction due to the rotational movement of the rotating shaft 201. Therefore, as shown in FIG. 25, even if the "axial direction D of high oil film rigidity" of the lower bearing 302 is arranged in the "axial direction D1 of low oil film rigidity" of the upper bearing 301, the amount of shaft vibration can be reduced. , there is a possibility that sufficient effects may not be obtained.

In that case, measure the amount of shaft vibration using a displacement sensor, etc., and set the lower bearing 302 so that the effect of reducing the amount of shaft vibration is maximized, or so that the measured amount of shaft vibration is equal to or less than a preset amount of vibration. The angular arrangement of "the axial direction D1 of low oil film rigidity" may be finely adjusted.

Further, since the rotating shaft 201 of the rotary compressor 1 is housed in the closed container 100 made of a pressure container, it is not easy to directly measure the amount of vibration. Therefore, for example, a vibration meter such as an acceleration sensor is attached to the closed container 100, and the vibration of the rotary compressor 1 is measured. may be determined.

The configuration of Embodiment 2 can be applied to a rotary compressor 1 having multiple sliding bearings. Further, the number of upper bearings 301 arranged at an angle for the purpose of avoiding seizure, and the number of lower bearings 302 arranged at an angle for the purpose of vibration reduction can be divided arbitrarily. Furthermore, if there are a plurality of upper bearings 301 arranged at an angle for the purpose of avoiding seizure, the angular arrangement of each upper bearing 301 may be the same as long as it is within the angular range A, B1, or C described in the first embodiment. However, they can also be different. Furthermore, even when there are a plurality of lower bearings 302 arranged at an angle for the purpose of reducing vibration, the angular arrangement of each lower bearing 302 may be the same or different.

In the second embodiment, for convenience, the upper bearing 301 is described as a bearing with an angular arrangement for the purpose of avoiding seizure, and the lower bearing 302 is described as a bearing with an angular arrangement for the purpose of reducing vibration. However, it is not limited to this case. That is, the lower bearing 302 may be used as an angularly positioned bearing for the purpose of avoiding seizure. Additionally, the upper bearing 301 may be used as an angularly positioned bearing for vibration reduction purposes.

As described above, in the second embodiment, when the bearing section 300 has a plurality of sliding bearings, the "axial direction D of high oil film rigidity" of two or more sliding bearings among the plurality of sliding bearings are arranged so that they are different from each other. Thereby, it is possible to obtain both the effect of avoiding seizure of the bearing portion 300 and the effect of reducing vibration.

Embodiment 3.
FIG. 26 is a plan view showing the configuration of the bearing section 300 in the rotary compressor 1 according to the third embodiment. Embodiment 3 is characterized in that the inner diameter surface 300a of the bearing portion 300 is composed of a plurality of arcs in plan view. The other configurations are the same as in Embodiment 1, so their explanation will be omitted here.

As described above, in the third embodiment, as shown in FIG. 26, the inner diameter surface 300a of the bearing portion 300 is composed of a plurality of circular arcs in plan view. That is, the inner diameter surface 300a is composed of a plurality of circular arc surfaces. Here, in order to simplify the explanation, a bearing shape in which the inner diameter surface 300a is constituted by two circular arc surfaces shown in FIG. 26 will be described as an example. Hereinafter, the two arcuate surfaces will be referred to as a first arcuate surface portion 300a-1 and a second arcuate surface portion 300a-2, respectively.

Since the first circular arc surface section 300a-1 and the second circular arc surface section 300a-2 are of the same type, the configuration of the first circular arc surface section 300a-1 will be described below, and the description of the second circular arc surface section 300a-2 will be omitted. .

The shape of the circular arc surface of the first circular arc surface portion 300a-1 is determined by the preload coefficient mp defined by the following equation (1).

mp=1-Cb/Cp (1)

Here, mp: preload coefficient, Cb: assembly radius clearance, Cp: machining radius clearance.

The assembly radius gap Cb is the size of the gap 400 formed between the rotating shaft 201 and the first circular arc surface portion 300a-1 on the imaginary line L1 extending in the first direction. In FIG. 26, the first direction is the vertical direction. In the example of FIG. 26, the assembly radius gap Cb is the size of the gap 400 at the position where the gap 400 is the narrowest among the sizes of the gap 400. In other words, in the example of FIG. 26, the assembly radius gap Cb is the minimum value of the size of the gap 400.

Next, the machining radius gap Cp will be explained. As shown in FIG. 26, first, a virtual circle 700 is defined as a virtual circle having the same curvature as the curvature of the arc forming the first circular arc surface portion 300a-1. Next, a virtual rotation axis obtained by shifting the rotation axis 201 in the first direction so as to be concentric with the virtual circle 700 is defined as a virtual rotation axis 701. Therefore, the virtual rotation axis 701 is a rotation axis at a position concentric with the virtual circle 700. At this time, the machining radius gap Cp is the size of the virtual gap 702 formed between the virtual rotating shaft 701 and the first circular arc surface portion 300a-1, on the virtual line L1 extending in the first direction. .

In the above description, a case has been described in which the preload coefficient mp of the first circular arc surface portion 300a-1 and the preload coefficient mp of the second circular arc surface portion 300a-2 are the same, but the present invention is not limited to that case. That is, the preload coefficient mp of the first circular arc surface portion 300a-1 and the preload coefficient mp of the second circular arc surface portion 300a-2 may be the same or different.

In addition, in the above description, since the case of a two-arc bearing composed of two arcuate surfaces shown in FIG. 26 is taken as an example, the angle in the circumferential direction of the first arcuate surface portion 300a-1 Although the case has been described in which the angles in the circumferential direction of the arcuate surface portion 300a-2 are the same and the first arcuate surface portion 300a-1 and the second arcuate surface portion 300a-2 face each other, the present invention is limited to that case. Not done. That is, the central angle of the arc forming the first arc surface section 300a-1 and the center angle of the arc forming the second arc surface section 300a-2 may be the same or different. In this way, the two arcuate surfaces forming the two-arc bearing do not have to be equally spaced. Furthermore, even in the case of a bearing composed of three or more arcuate surfaces, these arcuate surfaces do not have to be equally distributed. Specifically, the three circular arc bearings shown in Fig. 13, the split bearings with three or more divisions, the step bearings shown in Fig. 11, the pivot pad bearings shown in Fig. 16, the Michel-Seggell bearings shown in Fig. 17, and the Neumu bearings shown in Fig. 18. Similarly, in the case of the foil bearing shown in FIG. 20, the plurality of surfaces constituting the inner diameter surface 300a do not need to be equally distributed.

Furthermore, in the above description of the two-arc bearing using FIG. Although a case has been described in which there is a line symmetry with respect to the virtual line L2 extending in two directions, the present invention is not limited to that case. Note that the second direction is a direction that intersects the first direction. For example, the second direction is perpendicular to the first direction. That is, the circumferential position at which the first arcuate surface portion 300a-1 starts from the arcuate arrangement and the circumferential position from which the arcuate arrangement of the second arcuate surface portion 300a-2 starts from the axis of the bearing portion 300. In O, it is not necessary to arrange line symmetrically. That is, the positions in the circumferential direction that are the starting points of the arcuate arrangement of these arcuate surfaces do not have to be arranged at equal intervals. Furthermore, even in the case of a bearing composed of three or more arcuate surfaces, the positions in the circumferential direction from which the arcuate arrangement of the arcuate surfaces starts do not have to be arranged at equal intervals. Specifically, for example, in the three-arc bearing shown in FIG. 13, the center angle of each arc surface is 180 degrees, the center angle of the second arc is 120 degrees, and the center angle of the third arc is 60 degrees. The positions in the circumferential direction that are the starting points of the arc arrangement do not need to be arranged at equal intervals. However, it is desirable that the position in the circumferential direction, which is the starting point of the circular arc arrangement of each circular arc surface, be a position where the desired oil film rigidity can be obtained. Also, split bearings with three or more parts, step bearings shown in Fig. 11, pivot pad bearings shown in Fig. 16, Michel-Seggell bearings shown in Fig. 17, Neumu bearings shown in Fig. 18, foil bearings shown in Fig. 20, etc. Similarly, the positions in the circumferential direction, which are the starting points of the arcuate arrangement of the plurality of arcuate surfaces constituting the inner diameter surface 300a, do not have to be arranged at equal intervals. In these cases as well, it is desirable that the position in the circumferential direction, which is the starting point of the circular arc arrangement of each circular arc surface, be a position where the desired oil film rigidity can be obtained.

Further, in the above description, a case has been described in which the inner diameter surface 300a is composed of two circular arc surfaces, but the present invention is not limited to that case. That is, the inner diameter surface 300a may be composed of three or more circular arc surface portions. FIG. 13 described above shows a case where the inner diameter surface 300a is composed of three arcuate surface portions.

As described above, in the third embodiment, the shape of the inner diameter surface 300a of the sliding bearing that constitutes the bearing portion 300 is composed of a plurality of arcuate surfaces in plan view. As a result, the sliding bearing that constitutes the bearing portion 300 has anisotropy in oil film rigidity. Therefore, as explained in the first embodiment, by arranging the "axial direction D of high oil film rigidity" of the bearing part 300 in the angular range A of 180 degrees to 330 degrees (see FIG. 5), the oil film thickness can be reduced. The decline can be suppressed. Thereby, the oil film thickness is always ensured, so that it is possible to prevent metal contact between the rotating shaft 201 and the bearing portion 300. As a result, seizure of the bearing portion 300 can be avoided.

Embodiment 4.
Embodiment 4 is characterized in that a dynamic pressure generating section 600 that generates dynamic pressure of refrigerating machine oil is provided on the inner diameter surface 300a of the bearing section 300. The other configurations are the same as in Embodiment 1, so their explanation will be omitted here.

The dynamic pressure generating section 600 is, for example, a step 300m shown in FIG. 11. Alternatively, the dynamic pressure generating section 600 is, for example, a twisted groove 300k shown in FIG. 21. Alternatively, the dynamic pressure generating section 600 is, for example, a hole 300ee shown in FIG. 19. Although FIG. 19 shows an example in which a porous member 300e having holes 300ee is attached to the inner diameter surface 300a, the holes 300ee may be formed directly on the inner diameter surface 300a.

As described above, in the fourth embodiment, at least one of a step, a groove, or a hole is formed on the inner diameter surface 300a of the bearing section 300 as the dynamic pressure generating section 600. The step, groove, or hole is formed in a concave shape on the inner diameter surface 300a. As the rotating shaft 201 and the rolling piston 504 rotate, refrigerating machine oil flows into the dynamic pressure generating section 600 and flows out from the dynamic pressure generating section 600. Therefore, compared to the case where the dynamic pressure generating section 600 is not provided, the flow of the refrigerating machine oil becomes more active, and dynamic pressure of the refrigerating machine oil is generated.

Note that the steps, grooves, and holes may be formed on the entire inner diameter surface 300a of the bearing section 300, or may be formed only on a portion of the inner diameter surface 300a of the bearing section 300.

As described above, according to the fourth embodiment, at least one of a step, a groove, or a hole is formed on the inner diameter surface 300a of the bearing portion 300. As a result, refrigerating machine oil flows into and out of the steps, grooves, or holes, and dynamic pressure of the refrigerating machine oil is generated. Therefore, even if the rotational arrangement and phase angle of the bearing portion 300 having a non-perfect circular structure change and the direction of the vector of the gas load GL changes, the desired dynamic pressure effect can always be obtained. As a result, the thickness of the oil film in the gap 400 formed between the rotating shaft 201 and the inner diameter surface 300a is ensured, and seizing of the bearing portion 300 due to the rigidity of the oil film can be prevented. Further, since the refrigerating machine oil is stored inside the steps, grooves, or holes, the refrigerating machine oil can be held on the surface of the rotating shaft 201. Further, by forming steps, grooves, or holes in the inner diameter surface 300a of the bearing portion 300, an effect of trapping foreign matter from the outside can be expected.

Embodiment 5.
Embodiment 5 is characterized in that the inner diameter surface 300a of the bearing portion 300 has a non-perfect circular shape, and the gap distribution of the gap 400 is continuous. The other configurations are the same as in Embodiment 1, so their explanation will be omitted here.

FIG. 27 is a plan view showing the configuration of the bearing section 300 in the rotary compressor 1 according to the fifth embodiment. FIG. 27 shows an example in which the inner diameter surface 300a of the bearing portion 300 is configured by adding up Fourier series in plan view.

The circumferential gap distribution shape of the gap 400 on the inner diameter surface 300a of the bearing portion 300 may be expressed as, for example, the addition of polynomials to the rotation angle or the addition of Fourier series. This makes it possible to create a complex shape on the inner diameter surface 300a that provides the desired oil film rigidity. Further, the above-mentioned complex shape may not only change in the circumferential direction of the bearing section 300 but also change in the axial direction of the bearing section 300.

In addition, when expressing the shape of the inner diameter surface 300a by a polynomial, for example, one or more polynomials shown in the following equation (2) can each represent one or more curves forming the inner diameter surface 300a. Therefore, the inner diameter surface 300a is divided into one or more ranges, a polynomial expressed by the following formula (2) is generated for each range, and the coefficient a _k (k is an integer from 0 to n) of the polynomial is determined. do. In this way, the entire curve constituting the inner diameter surface 300a is generated using one or more polynomials shown in Equation (2) below.

y=a _n x ⁿ +a _n-1 x ^n-1 +...+a ₁ x+a ₀
=Σa _k x ^k (2)

Further, when expressing the shape of the inner diameter surface 300a using a Fourier series, for example, one or more curves constituting the inner diameter surface 300a can be represented by the Fourier series shown in the following equation (3). Therefore, the inner diameter surface 300a is divided into one or more ranges, a Fourier series expressed by the following formula (3) is generated for each range, and the coefficients a ₀ , a _n , b _n (n is (an integer between 1 and ∞). In this way, the entire curve forming the inner diameter surface 300a is generated by adding up the Fourier series shown in equation (3) below.

y=a ₀ /2+Σ(a _n cos(nx)+b _n sin(nx))
(3)

As described above, in the fifth embodiment, the inner diameter surface 300a of the sliding bearing forming the bearing portion 300 has a non-perfect circular shape in plan view, and the gap distribution of the gap 400 is continuous. Thereby, the inner diameter surface 300a can be formed into a complex shape that can obtain the desired oil film rigidity. In that case, a necessary oil film thickness can always be ensured in the gap 400 formed between the rotating shaft 201 and the bearing part 300, and metal contact between the rotating shaft 201 and the bearing part 300 can be prevented. . As a result, seizure of the bearing portion 300 can be avoided.

Embodiment 6.
Bearing section 300 according to Embodiment 6 has two or more of the features of bearing section 300 shown in Embodiments 3, 4, and 5 above. The other configurations are the same as in any of the first to fifth embodiments, so their explanation will be omitted here.

In Embodiment 6, for example, in the two-arc bearing shown in FIG. 26 (see Embodiment 3), the first arc surface portion 300a-1 has the characteristics of Embodiment 4, and the second arc surface portion 300a-2 However, this embodiment has the characteristics of the fifth embodiment.

That is, the first arcuate surface portion 300a-1 is provided with a dynamic pressure generating portion 600 consisting of a step 300m, a groove 300k, or a hole 300ee, which is a feature of the fourth embodiment.

Further, the second arcuate surface portion 300a-2 has an inner diameter surface shape expressed by a polynomial, which is an example of the fifth embodiment.

In this way, the bearing section 300 according to the sixth embodiment has a combination of a plurality of features among the features of the bearing section 300 shown in the third, fourth, and fifth embodiments. Thereby, in the sixth embodiment, effects obtained from each of the plurality of combined features can be obtained. That is, in Embodiment 6, when Embodiments 3, 4, and 5 are combined as in the above example, all the effects obtained from Embodiments 3, 4, and 5 can be obtained.

1 rotary compressor, 2 suction pipe, 3 suction muffler, 4 refrigerant discharge pipe, 5 refrigeration oil, 6 screw, 100 sealed container, 101 upper container, 102 lower container, 201 rotating shaft, 201a convex part, 202 main shaft part, 203 Deputy axis, 204 cranks, 300 Axial portion, 300A inner diameter, 300A -1 1st arc, 300A -2 2nd arcs, 300B bush, 300C pad, 300D pivot, 300E porous member, 300EE holes, 300F holes, 300F. Foil, 300g inner ring, 300h outer ring, 300k groove, 300m step, 301 upper bearing, 302 lower bearing, 400 gap, 401 electric motor, 402 rotor, 403 stator, 501 compression mechanism, 502 cylinder, 503 Discharge muffler, 504 rolling piston , 505 vane, 506 venue spring, 507 cylinder room, 507A inhalation room, 507B compression room, 508 vane groove, 508A stop, 508B venue spring storage hole, 509 discharge port, 600 vibration pressure generator, 700 virtual yen, 701 virtual yen Rotation axis, 702 virtual gap, A angular range, B1 angular range, B2 angular range, C angular range, D axial direction of high oil film rigidity.

Claims

a rotating shaft;
a bearing portion configured from a sliding bearing and having an inner diameter surface that supports the rotating shaft;
a cylinder chamber having a suction chamber into which a refrigerant is drawn, and a compression chamber into which the refrigerant is compressed;
a crank portion attached to the rotating shaft and rotating eccentrically within the cylinder chamber;
Equipped with
A gap is formed between the rotating shaft and the inner diameter surface of the slide bearing, and a gap is formed into which refrigerating machine oil is supplied.
The sliding bearing has anisotropy in oil film rigidity of the refrigerating machine oil,
The angle of action of the gas load generated by compression of the refrigerant due to eccentric rotation of the crank portion changes depending on the crank angle,
When the variation range of the action angle of the gas load is defined as a first angle range,
An axial direction of high oil film rigidity indicating a direction in which the oil film rigidity of the sliding bearing is highest is arranged in the first angular range,
rotary compressor.
comprising a vane that separates the suction chamber and the compression chamber of the cylinder chamber,
When the crank angle of the vane top dead center of the vane is 0deg,
The first angle range is an angle range in which the crank angle is from 110 degrees to 330 degrees.
The rotary compressor according to claim 1.
The bearing section has a plurality of the sliding bearings,
Of the plurality of slide bearings, two or more of the slide bearings have different axial directions of the high oil film rigidity,
The rotary compressor according to claim 1 or 2.
The shape of the inner diameter surface of the slide bearing is composed of a plurality of arcuate surfaces when viewed from above,
The rotary compressor according to any one of claims 1 to 3.
The inner diameter surface of the slide bearing is provided with a dynamic pressure generating portion that generates a dynamic pressure on the inner diameter surface,
The dynamic pressure generating portion is configured from at least one of a step, a groove, or a hole formed in a concave shape on the inner diameter surface.
The rotary compressor according to any one of claims 1 to 4.
The inner diameter surface of the slide bearing has a non-perfect circular shape in plan view, and the gap distribution of the gap formed between the rotating shaft and the inner diameter surface of the slide bearing is continuous.
The rotary compressor according to any one of claims 1 to 5.
The magnitude of the gas load generated by compression of the refrigerant due to eccentric rotation of the crank portion changes depending on the crank angle,
When the angular range of the crank angle in which the magnitude of the gas load is 50% or more of the maximum value of the gas load is defined as a second angular range,
An axial direction of high oil film rigidity indicating a direction in which the oil film rigidity of the sliding bearing is highest is arranged in the second angular range instead of the first angular range,
The rotary compressor according to claim 1.
comprising a vane that separates the suction chamber and the compression chamber of the cylinder chamber,
When the crank angle of the vane top dead center of the vane is 0deg,
The second angle range is an angle range in which the crank angle is from 180 degrees to 330 degrees.
The rotary compressor according to claim 7.