Rotary compressor

Abstract

A rotary compressor is provided with: a rotary shaft; a bearing part comprising a sliding bearing and having an inner diameter surface for supporting the rotating shaft; a cylinder chamber having a suction chamber for sucking in the refrigerant and a compression chamber for compressing the refrigerant; and a crank section that is attached to the rotating shaft and eccentrically rotates in the cylinder chamber, a gap to which refrigerating oil is supplied is formed between the rotating shaft and the inner diameter surface of the sliding bearing, and the oil film rigidity of the refrigerating oil in the sliding bearing has anisotropy. The angle of action of a gas load generated by compression of a refrigerant accompanying eccentric rotation of the crank portion varies according to the crank angle, and when the range of variation of the angle of action of the gas load is set to a first angle range, the range of variation of the angle of action of the gas load is set to a second angle range. The axial direction of a high oil film rigidity indicating the direction in which the oil film rigidity of the sliding bearing is the highest is disposed within a first angular range.

Description

Technical Field

[0001] This disclosure relates to rotary compressors. Background Technology

[0002] Rotary compressors are one type of refrigerant compressor used in air conditioners and refrigeration equipment. Generally, rotary compressors are known to have a structure comprising a sealed container, an electric motor, a rotating shaft, and a compression mechanism. The compression mechanism includes: an eccentric crank portion disposed on the rotating shaft; a cylinder having a cylinder chamber; a rotary piston fitted with and housed within the cylinder chamber; and blades disposed in blade slots formed radially along the cylinder chamber. The rotary piston rotates together with the eccentric crank portion, compressing the refrigerant within the cylinder chamber. The blades follow the rotary piston, dividing the cylinder chamber into a refrigerant intake chamber and a compression chamber. In a rotary compressor, low-pressure refrigerant is drawn into the compression mechanism, and through rotational compression, it becomes high-pressure refrigerant and is discharged.

[0003] During refrigerant compression, a gas pressure load (hereinafter referred to as gas load) based on the refrigerant acts on the rotating shaft. Generally, in a rotary compressor, the gas load is supported by two sliding bearings located above and below the cylinder chamber of the compression mechanism. The gas load is generated by the volume change of the compression chamber caused by the eccentric rotation of the eccentric crank. Therefore, the magnitude and direction of the gas load change depending on the rotation angle of the rotating shaft. Consequently, in a rotary compressor, the thickness of the oil film formed between the sliding bearing and the rotating shaft changes during one revolution. Especially when the gas load reaches its maximum rotation angle, there is a concern about contact between the rotating shaft and the sliding bearing, which, in the worst case, could lead to the sintering of the sliding bearing.

[0004] As a method for sintering the bearing portion in rotating machinery to avoid the application of variable loads, a method of making the inner diameter shape of the bearing non-circular can be considered (see, for example, Patent Document 1).

[0005] Patent Document 1 describes a bearing that supports a rotating body that may experience load variations. The bearing described in Patent Document 1 has a non-circular shape, such as an ellipse, and is a cylindrical shape extending horizontally along its axial direction, with a rotating body inserted in its central portion. The cylindrical bearing described in Patent Document 1 is divided into an upper bearing section and a lower bearing section.

[0006] In Patent Document 1, the rotating body is, for example, a multi-span rotor used in thermal power plants. Therefore, due to misalignment, variations over time, and conditions during partial load operation, the bearing load may sometimes deviate from the design value. Therefore, in Patent Document 1, by providing a stabilizing groove in the lower bearing, unstable vibrations are prevented from occurring. In Patent Document 1, the load is always applied vertically, and the direction of the load does not change.

[0007] Prior art literature

[0008] Patent documents

[0009] Patent Document 1: Japanese Patent Application Publication No. 2000-145781 Summary of the Invention

[0010] The problem that the invention aims to solve

[0011] When applying the bearing structure described in Patent Document 1, which operates under varying loads with a constant load direction, to a rotary compressor where the magnitude and direction of the load change, it is impossible to uniquely define the load direction for the structure. Furthermore, as a practical problem, depending on the configuration of the non-circular structure in the rotation direction and the different phase angles, the desired hydrodynamic effect may not be achieved for the vector of the varying load, potentially increasing the risk of contact between the rotating shaft and the bearing, as well as the risk of sintering. Moreover, as described in Patent Document 1, when using a multi-arc bearing composed of multiple arcs, or when performing localized grooving on the sliding surface, the anisotropy of the bearing's oil film rigidity increases the likelihood of shaft vibration increasing in the direction of lower rigidity.

[0012] This disclosure was made to solve the problems mentioned above, and its purpose is to suppress the reduction of oil film thickness on the sliding surface of the bearing portion and to suppress sintering in the bearing portion in a rotary compressor subjected to a variable load with varying magnitude and direction of load.

[0013] Methods for solving problems

[0014] The rotary compressor disclosed herein comprises: a rotating shaft; a bearing portion consisting of a sliding bearing and having an inner diameter surface supporting the rotating shaft; a cylinder chamber having an intake chamber for drawing in refrigerant and a compression chamber for compressing the refrigerant; and a crank portion mounted on the rotating shaft and rotating eccentrically within the cylinder chamber. A gap for supplying refrigeration oil is formed between the rotating shaft and the inner diameter surface of the sliding bearing. The oil film rigidity of the refrigeration oil in the sliding bearing is anisotropic. The angle of action of the gas load generated by the compression of the refrigerant associated with the eccentric rotation of the crank portion varies according to the crank angle. When the variation range of the angle of action of the gas load is set as a first angle range, the axial direction of high oil film rigidity, representing the direction of highest oil film rigidity of the sliding bearing, is arranged within the first angle range.

[0015] Invention Effects

[0016] According to the rotary compressor of this disclosure, the axial direction of high oil film rigidity, representing the direction of highest oil film rigidity in the sliding bearing, is positioned within a first angular range that is the variation range of the angle of action of the gas load. This suppresses the reduction of oil film thickness and inhibits sintering in the bearing section. Attached Figure Description

[0017] Figure 1 This is a longitudinal sectional view showing the structure of the rotary compressor 1 according to Embodiment 1.

[0018] Figure 2 This is a plan view showing the structure of the compression mechanism 501 of the rotary compressor 1 provided in Embodiment 1.

[0019] Figure 3 This is an explanatory diagram illustrating the rotational motion of the crank portion 204 and the rotary piston 504 of the rotary compressor 1 provided in Embodiment 1.

[0020] Figure 4 This is an explanatory diagram illustrating the principle of generating a gas load in the rotary compressor 1 of Embodiment 1 by compressing the refrigerant accompanying the rotation of the rotary piston 504.

[0021] Figure 5 This is a diagram illustrating an example of the variation of the gas load GL with respect to the crank angle θ1 in the rotary compressor 1 of Embodiment 1.

[0022] Figure 6 This is a diagram showing an example of the action angle θ2 of the gas load GL in the rotary compressor 1 of Embodiment 1 relative to the crank angle θ1.

[0023] Figure 7This is a plan view showing the structure of a full-circumference bearing.

[0024] Figure 8 This is a plan view showing the structure of a partial bearing.

[0025] Figure 9 This is a plan view showing the structure of a split bearing.

[0026] Figure 10 This is a plan view showing the structure of the matching bearing.

[0027] Figure 11 This is a plan view showing the structure of a stepped thrust bearing.

[0028] Figure 12 This is a plan view showing the structure of a double circular arc bearing.

[0029] Figure 13 This is a plan view showing the structure of a three-circular-arc bearing.

[0030] Figure 14 This is a plan view showing the structure of a double circular arc eccentric bearing.

[0031] Figure 15 This is a plan view showing the structure of the floating bushing bearing.

[0032] Figure 16 This is a plan view showing the structure of the pivot bushing bearing.

[0033] Figure 17 This is a plan view showing the structure of the Michel-Seger bearing.

[0034] Figure 18 This is a plan view showing the structure of a frictionless bearing.

[0035] Figure 19 This is a plan view showing the structure of a porous bearing.

[0036] Figure 20 This is a plan view showing the structure of the foil bearing.

[0037] Figure 21 This is a schematic cross-sectional view showing the structure of a bearing with helical grooves.

[0038] Figure 22 This is a schematic cross-sectional view showing the structure of a spherical bearing.

[0039] Figure 23 This is an explanatory diagram showing the angle ranges A, B1, and C for setting the "axial direction D with high oil film rigidity" in the rotary compressor 1 of Embodiment 1.

[0040] Figure 24This 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 of Embodiment 1.

[0041] Figure 25 This is an explanatory diagram showing the direction in which the "axial direction D of high oil film rigidity" is set for the upper bearing 301 and the lower bearing 302 in the rotary compressor 1 of Embodiment 2.

[0042] Figure 26 This is a plan view showing the structure of the bearing section 300 in the rotary compressor 1 of Embodiment 3.

[0043] Figure 27 This is a plan view showing the structure of the bearing section 300 in the rotary compressor 1 of Embodiment 5. Detailed Implementation

[0044] Hereinafter, embodiments of the rotary compressor 1 of this disclosure will be described with reference to the accompanying drawings. This disclosure is not limited to the following embodiments, and various modifications can be made without departing from the spirit of this disclosure. Furthermore, this disclosure includes all combinations of structures that can be combined in the following embodiments and their modifications. Additionally, in the various figures, components labeled with the same reference numerals are the same or equivalent components, which is consistent throughout the specification. Moreover, in the various figures, the relative dimensional relationships or shapes of the structural members may sometimes differ from actual conditions.

[0045] Implementation method 1.

[0046] Figure 1 This is a longitudinal sectional view showing the structure of the rotary compressor 1 according to Embodiment 1. Figure 2 This is a plan view showing the structure of the compression mechanism section 501 of the rotary compressor 1 in Embodiment 1. Figure 1 and Figure 2 In this example, a single rotary compressor with a single compression mechanism is shown as one example of a rotary compressor 1. However, the rotary compressor 1 is not limited to a single rotary compressor; for example, it could be a rotary compressor with multiple compression mechanisms, such as a double rotary compressor with two compression mechanisms. Figure 1 As shown, the rotary compressor 1 is connected to the suction silencer 3 via the suction pipe 2. Refrigerant is drawn from the suction silencer 3 into the interior of the rotary compressor 1.

[0047] like Figure 1As shown, the rotary compressor 1 includes a sealed container 100, a rotating shaft 201, a bearing section 300, a motor section 401, and a compression mechanism section 501. The motor section 401, rotating shaft 201, bearing section 300, and compression mechanism section 501 are housed inside the sealed container 100. The motor section 401 is housed in the upper part of the interior of the sealed container 100. The compression mechanism section 501 is housed in the lower part of the interior of the sealed container 100, below the motor section 401. The motor section 401 and the compression mechanism section 501 are connected via the rotating shaft 201. The motor section 401 has a rotor 402 and a stator 403. The rotor 402 is disposed inside the stator 403 and fixed to the rotating shaft 201. The stator 403 is fixed to the sealed container 100. The 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 401 to the compression mechanism 501. The rotating shaft 201 has a crank portion 204 that is eccentric to its central axis. A rotary piston 504 is fitted and mounted in the crank portion 204. The rotary piston 504 is disposed within the cylinder block 502. Figure 2 As shown, the cylinder block 502 is configured to cover the rotating shaft 201 from the outer periphery, with the central axis of the rotating shaft 201 as the center. Alternatively, the crank portion 204 can be the same size as when the rotating piston 504 is mounted on the crank portion 204, and the rotating piston 504 can be omitted. The compression mechanism 501 compresses the refrigerant using the driving force transmitted from the rotating shaft 201. The bearing portion 300 supports the rotating shaft 201 on the upper and lower sides of the compression mechanism 501 and has one or more sliding bearings.

[0048] like Figure 2 As shown, a compression chamber 507 and an intake chamber 507a are formed by separating the cylinder chamber 507 between the rotary piston 504 and the cylinder block 502 using blades 505. When the motor unit 401 rotates, the rotary piston 504 mounted on the crank unit 204 rotates eccentrically within the cylinder block 502. As a result, with the rotation of the rotary piston 504, the internal space of the intake chamber 507a decreases, and the refrigerant in the compression chamber 507b is compressed. Furthermore, the compression chamber 507b and the discharge port 509 (see reference) Figure 1 The refrigerant inside the compression chamber 507b is discharged from the discharge port 509. At this time, a gas load GL based on the compressed refrigerant acts on the rotating shaft 201. Depending on the change in the geometric compression volume of the compression chamber 507b, the gas load GL is adjusted as follows: Figure 2 As shown by arrow W, relative to the rotation axis 201 along Figure 2 The gas load GL operates in a roughly vertically downward direction, approximately at an angle of 120 to 180 degrees. This gas load GL, if subsequently used... Figure 4 The gas load GL acts substantially on the surface of the rotating piston 504 as described, but also acts on the rotating shaft 201 via the rotating piston 504. That is, the gas load GL acts on the rotating shaft 201 via the rotating piston 504 through the oil film located in the gap between the rotating piston 504 and the crank portion 204, or through direct contact between the rotating piston 504 and the crank portion 204. Figure 2 In the example where the gas load GL acts in a direction that is approximately perpendicular and downward relative to the rotation axis 201, such as Figure 2 As indicated by arrow E, the axial position of the rotating shaft 201 changes to a position eccentric relative to the direction of the load acting on the bearing portion 300 along the rotational direction. Furthermore, for illustrative purposes, in Figure 2 The diagram shows a rotary compressor 1 in which the blades 505 and the rotary piston 504 are separately constructed. However, Embodiment 1 is not limited to this and can also be applied to an oscillating compressor in which the blades 505 and the rotary piston 504 are integrated.

[0049] The following is about Figure 1 and Figure 2 The components of the rotary compressor 1 shown are described in detail.

[0050] The sealed container 100 is composed of an upper container 101 and a lower container 102. In addition, the sealed container 100 is not limited to being formed by two structural members, the upper container 101 and the lower container 102, and may also be formed by three or more structural members.

[0051] like Figure 1 As shown, the sealed container 100 is connected to the suction silencer 3 via the suction pipe 2, and gaseous refrigerant is drawn into the interior from the suction silencer 3. The suction silencer 3 is fixed to the outer side of the lower container 102 of the sealed container 100 by welding or the like. The suction silencer 3 is provided for the following purposes: to separate the low-temperature and low-pressure refrigerant supplied from the refrigeration circuit into liquid refrigerant and gaseous refrigerant, to prevent liquid refrigerant from being drawn into the compressor section 501 as much as possible, and to store the separated liquid refrigerant. The reason for this is that if the rotary compressor 1 allows liquid refrigerant to flow into the compressor section 501 and compress it, the pressure inside the cylinder will become abnormally high due to the compression of the liquid refrigerant, which could cause malfunctions of the compressor section 501, such as blade fly-out and damage. In addition, the suction silencer 3 also functions as a silencer to reduce or eliminate the noise generated by the flowing refrigerant.

[0052] A refrigerant discharge pipe 4 is connected to the upper part of the sealed container 100 to discharge the compressed refrigerant. The refrigerant discharge pipe 4 is a refrigerant piping that discharges high-pressure gaseous refrigerant to the outside of the sealed container 100. The refrigerant discharge pipe 4 is joined to the upper container 101, which constitutes the sealed container 100, by means of brazing or resistance welding, for example, while the upper container 101 is being penetrated.

[0053] The interior of the sealed container 100 is filled with high-temperature, high-pressure gaseous refrigerant compressed by the compression mechanism 501, and refrigerant oil 5 for lubricating the compression mechanism 501 is stored at the bottom. The refrigerant oil 5 is mainly used to lubricate the sliding parts of the compression mechanism 501. An oil pump (not shown) is provided at the lower part of the rotating shaft 201. As the rotating shaft 201 rotates, the oil pump draws the refrigerant oil 5 stored at the bottom of the sealed container 100 and supplies it to each sliding part of the compression mechanism 501. The compression mechanism 501 ensures mechanical lubrication by supplying oil to each sliding part.

[0054] like Figure 1 As shown, the motor unit 401 includes: a cylindrical stator 403, which is fixed to the inner wall of the sealed container 100 by thermoforming or the like; and a cylindrical rotor 402, which is rotatably disposed facing the inner surface of the stator 403 and rotates by magnetic action. A rotating shaft 201 is embedded in the center of the rotor 402. The motor unit 401 uses electricity supplied from an external power source to generate a rotational driving force on the rotating shaft 201, and transmits the rotational driving force to the compression mechanism unit 501 via the rotating shaft 201. Furthermore, a brushless DC motor, for example, can be used as the motor unit 401.

[0055] The rotating shaft 201 includes: a main shaft portion 202 fixed to the rotor 402 of the motor portion 401; a secondary shaft portion 203 sandwiching a compression mechanism portion 501 on the opposite side of the main shaft portion 202; and a crank portion 204 disposed between the main shaft portion 202 and the secondary shaft portion 203. The rotating shaft 201 is axially arranged from above to below the sealed container 100, consisting of the main shaft portion 202, the crank portion 204, and the secondary shaft portion 203. The main shaft portion 202 is embedded into the center of the rotor 402 of the motor portion 401 and is fixed by thermoforming or pressing. The central axis of the crank portion 204 is eccentric relative to the central axes of the main shaft portion 202 and the secondary shaft portion 203.

[0056] The compression mechanism 501 utilizes the rotational driving force supplied from the electric motor 401 to compress the low-pressure gaseous refrigerant drawn into the low-pressure space of the sealed container 100 from the suction pipe 2 into a high-pressure gaseous refrigerant. The high-pressure gaseous refrigerant compressed by the compression mechanism 501 is discharged from the top of the compression mechanism 501 into the interior of the sealed container 100. The compression mechanism 501 includes a cylinder 502, an upper bearing 301, a lower bearing 302, a discharge muffler 503, a rotary piston 504, and a vane 505 (see reference). Figure 2 ) and leaf spring 506 (refer to Figure 2 ).

[0057] The outer periphery of the cylinder 502 is fixed to the sealed container 100 by bolts or the like. Figure 1 As shown, the upper surface of cylinder block 502 is designated as one end face side UP, and the lower surface is designated as the other end face side DN. Figure 2 As shown, the cylinder block 502 is a hollow cylindrical shape, and the hollow interior forms the cylinder chamber 507. (As indicated...) Figure 1 As shown, the cylinder chamber 507 is open at both ends along the axial direction of the rotating shaft 201 and is closed by an upper bearing 301 disposed on the upper surface of the cylinder body 502 and a lower bearing 302 disposed on the lower surface of the cylinder body 502. That is, the cylinder chamber 507 is a space surrounded by the inner circumferential surface of the cylinder body 502, the inner wall surface of the upper bearing 301, and the inner wall surface of the lower bearing 302.

[0058] Additionally, a suction port (not shown) is provided in the cylinder body 502, extending from its outer peripheral surface into the cylinder chamber 507, through which gaseous refrigerant from the suction pipe 2 passes. The suction port connects the pipe of the suction pipe 2 to the cylinder chamber 507.

[0059] In addition, such as Figure 2 As shown, a blade groove 508 is formed in the cylinder block 502, communicating with the cylinder chamber 507 and extending radially with the rotation axis 201 as the center. Figure 2 As shown, viewed from a direction where the cylinder body 502 appears circular, the vane groove 508 extends axially from one end face UP towards the other end face DN along the cylinder body 502. Vanes 505 are slidably embedded in the vane groove 508, dividing the cylinder chamber 507 into an intake chamber 507a and a compression chamber 507b. The intake chamber 507a is a low-pressure space and communicates with the intake port. The compression chamber 507b is a high-pressure space and communicates with the discharge port 509 for discharging to the outside of the cylinder chamber 507 (see reference). Figure 1 Connect.

[0060] Furthermore, a blocking portion 508a is formed at the end of the blade groove 508 on the outer peripheral surface side of the cylinder 502. The blocking portion 508a is provided to prevent the blade 505 from moving towards the outer peripheral surface of the cylinder 502 and to restrict the movement of the blade 505 so that the blade 505 does not fly off the outer peripheral surface of the cylinder 502. Additionally, the blocking portion 508a also functions as a back pressure chamber for introducing high-pressure refrigerant. Furthermore, as... Figure 2 As shown, the blocking part 508a is an arc shape that opens only into the blade groove 508 when viewed from one end face UP of the cylinder 502.

[0061] Moreover, such as Figure 2 As shown, a leaf spring receiving hole 508b is formed at the end of the leaf spring slot 508 to serve as a space for accommodating and actuating the leaf spring 506. The leaf spring receiving hole 508b is formed to extend radially along the cylinder body 502. The length of the leaf spring receiving hole 508b is determined according to the shape of the actuating leaf spring 506 or the shape of the cylinder body 502.

[0062] like Figure 1 As shown, the upper bearing 301 is formed into an approximately inverted T-shape when viewed from the side. The upper bearing 301 is located on one end face UP of the cylinder body 502 on the side where the motor unit 401 is disposed, and closes one axial opening of the cylinder chamber 507. Furthermore, the upper bearing 301 engages with the main shaft portion 202 of the rotating shaft 201, supporting the main shaft portion 202 for rotation. The upper bearing 301 and the lower bearing 302 are fixed to the cylinder body 502 together by a common screw 6. The upper bearing 301 is, for example, constructed as a sliding bearing.

[0063] Furthermore, a discharge port 509 is formed on the upper bearing 301 for discharging the refrigerant compressed in the compression chamber 507b to the outside of the cylinder chamber 507. A discharge valve (not shown) is installed on the discharge port 509. The discharge valve controls the timing of discharging the high-temperature and high-pressure gaseous refrigerant 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 is above a preset pressure, allowing the refrigerant to be discharged into the sealed container 100. Therefore, when the pressure inside the compression chamber 507b is lower than the preset pressure, the discharge valve closes the discharge port 509. Conversely, when the pressure inside the compression chamber 507b is above the preset pressure, the discharge valve is pushed upward by the pressure inside the compression chamber 507b and becomes open. Thus, the refrigerant is discharged from the compression chamber 507b through the discharge port 509 into the sealed container 100.

[0064] like Figure 1As shown, the lower bearing 302 is formed in a roughly T-shape, facing opposite directions when viewed from the side. The lower bearing 302 is located on the opposite end face DN of the cylinder block 502, opposite to the side where the motor unit 401 is located, and closes the axial opening of the cylinder chamber 507. Furthermore, the lower bearing 302 engages with the secondary shaft portion 203 of the rotating shaft 201, supporting the secondary shaft portion 203 for rotation. The lower bearing 302 is, for example, constructed from a sliding bearing.

[0065] like Figure 1 As shown, the exhaust muffler 503 is installed to cover the outside of the upper bearing 301. Inside the compression chamber 507b, the actions of drawing in refrigerant, compressing refrigerant, and discharging refrigerant are repeatedly performed. The compressed gaseous refrigerant is intermittently discharged from the discharge port 509. This sometimes generates pulsating noises such as vibrations from the cylinder block 502. The exhaust muffler 503 is installed to suppress such noises, including pulsating noises, generated from the cylinder block 502.

[0066] Additionally, the exhaust muffler 503 is provided with an exhaust port (not shown) that communicates the space formed by the exhaust muffler 503 and the upper bearing 301 with the interior of the sealed container 100. The gaseous refrigerant discharged from the cylinder 502 via the exhaust port 509 is temporarily discharged into the space formed by the exhaust muffler 503 and the upper bearing 301, and then discharged into the interior of the sealed container 100 through the exhaust port.

[0067] like Figure 2 As shown, the rotary piston 504 is formed as a hollow cylinder, and the crank portion 204 of the rotary shaft 201 can be slidably fitted into the hollow interior. The rotary piston 504 and the crank portion 204 are housed together in the cylinder chamber 507. When the rotary shaft 201 is rotated by the drive of the electric motor 401, the rotary piston 504 rotates along the inner circumferential surface of the cylinder chamber 507 and compresses the refrigerant.

[0068] like Figure 2 As shown, in the refrigerant compression process, the blade 505 slides reciprocally inside the blade slot 508 as the rotating piston 504 rotates, with its leading end abutting against the outer peripheral surface of the rotating piston 504. By having the leading end of the blade 505 abut against the outer peripheral surface of the rotating piston 504, the cylinder chamber 507 is divided into an intake chamber 507a and a compression chamber 507b. The blade 505 is, for example, formed of a non-magnetic material.

[0069] The leaf spring 506 abuts against the back side of the leaf spring 505 and presses the leaf spring 505 so that the front end of the leaf spring 505 abuts against the outer peripheral surface of the rotary piston 504. The leaf spring 506 is housed in the leaf spring housing hole 508b of the cylinder body 502 and is arranged in parallel with the leaf spring 505. The leaf spring 506 is fixed to the cylinder body 502 by having its end opposite to the front end of the rotary piston 504 abut against the inner wall surface of the leaf spring housing hole 508b.

[0070] The operation of the rotary compressor 1 will now be explained. The rotary compressor 1 rotates the rotating shaft 201 by being driven by the electric motor 401, thereby causing the rotary piston 504 and crank 204 to rotate together inside the cylinder chamber 507. Within the cylinder chamber 507, the volume of the suction chamber 507a, separated by vanes 505, increases with the rotation of the rotating shaft 201. Conversely, the volume of the compression chamber 507b, separated by vanes 505, decreases within the cylinder chamber 507.

[0071] In the rotary compressor 1, the suction chamber 507a is connected to the suction port, and low-pressure gaseous refrigerant is drawn into the cylinder chamber 507. Next, the rotating piston 504 closes the connection between the compression chamber 507b and the suction port, causing the volume of the compression chamber 507b to decrease and compressing the gaseous refrigerant inside. Furthermore, the compression chamber 507b is connected to the discharge port 509 of the upper bearing 301. When the gaseous refrigerant inside the compression chamber 507b reaches a preset pressure, the discharge valve located at the discharge port 509 opens. Thus, the compressed, high-pressure, and high-temperature gaseous refrigerant is discharged from the discharge port 509 to the outside of the cylinder chamber 507.

[0072] High-pressure and high-temperature gaseous refrigerant discharged to the outside of cylinder chamber 507 is discharged into the interior of sealed container 100 via discharge muffler 503. Then, the discharged gaseous refrigerant passes through the interior of motor section 401 and rises inside sealed container 100, exiting from refrigerant discharge pipe 4 located at the top of sealed container 100 to the outside of sealed container 100. The refrigerant discharged to the outside of sealed container 100 circulates in the refrigeration circuit and returns to intake muffler 3.

[0073] Figure 3 This is an explanatory diagram illustrating the rotational movement of the crank portion 204 and the rotary piston 504 of the rotary compressor 1 provided in Embodiment 1. As described above, when the motor portion 401 rotates, the rotary piston 504 of the crank portion 204, mounted on the rotating shaft 201, rotates eccentrically within the cylinder body 502. At this time, as... Figure 3 As shown, the rotating piston 504 changes state in the order of (a)→(b)→(c)→(d)→(a)→…….

[0074] Figure 3 State (a) represents the case where the crank angle is 0 degrees. The crank angle θ1 is defined with the top dead center of the blade as the origin and the direction of rotation R (refer to...) Figure 2 This refers to the rotation angle of the crank section 204 in the positive orientation. For example... Figure 3 As shown in (a), the top dead center of the blade refers to the state in which the blade 505 is completely contained within the blade slot 508. The crank angle θ1 at the top dead center of the blade is 0 degrees or 360 degrees.

[0075] Next, the rotary piston 504 turns to Figure 3 State (b). In this state, refrigerant begins to be drawn into the suction chamber 507a, and at the same time, the compression of the refrigerant in the compression chamber 507b begins. Figure 3 The state of (b) is when the crank angle θ1 is 90 degrees.

[0076] Next, the rotary piston 504 turns to Figure 3 In state (c), refrigerant continues to be drawn into the suction chamber 507a, while at the same time, the refrigerant compressed in the compression chamber 507b begins to be discharged. Figure 3 The state of (c) is when the crank angle θ1 is 180 degrees.

[0077] Next, the rotary piston 504 turns to Figure 3 The state is (d). In this state, the intake of refrigerant into the suction chamber 507a continues, and the discharge of the refrigerant after compression in the compression chamber 507b continues. Figure 3 The state of (d) is when the crank angle θ1 is 270 degrees.

[0078] Next, the rotary piston 504 returns to Figure 3 State (a). In this state, the intake of refrigerant into the suction chamber 507a ends, and the discharge of the refrigerant compressed in the compression chamber 507b also ends. Figure 3 The state of (a) is the case where the crank angle θ1 is 360 degrees, that is, the case where the crank angle θ1 is 0 degrees.

[0079] Figure 4 This is an explanatory diagram illustrating the principle of gas load generation in the rotary compressor 1 of Embodiment 1 through refrigerant compression accompanying the rotation of the rotary piston 504. Figure 4 In this context, θ1 represents the crank angle, e represents the radius of the outer diameter of the rotating shaft 201, and r represents the radius of the outer diameter of the rotating piston 504. For example... Figure 4As shown, the rotating piston 504 is subjected to the pressure of the refrigerant from the compression chamber 507b, i.e., the compression chamber pressure Pc. Simultaneously, the rotating piston 504 is subjected to the pressure of the refrigerant from the suction chamber 507a, i.e., the suction pressure Ps. The pressure difference between the compression chamber pressure Pc and the suction pressure Ps is called the "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, as... Figure 4 As shown by the hollow arrow, the direction of the gas load GL vector is from the compression chamber 507b toward the intake chamber 507a. Furthermore, in Embodiment 1, the action angle θ2 of the gas load GL is used as an indicator of the direction of the gas load GL vector. Similar to the crank angle θ1, the action angle θ2 of the gas load GL is defined by the top dead center of the blade (refer to...). Figure 3 With the origin as the reference point and the rotation direction R (refer to...) Figure 2 The positive direction is defined. Since the position of the compression chamber 507b varies according to the crank angle θ1, the action angle θ2 of the gas load GL also varies. Thus, the gas load GL actually acts on the surface of the rotating piston 504, but as mentioned above, it also acts on the rotating shaft 201 via the rotating piston 504.

[0080] Figure 5 This is a diagram illustrating an example of the variation of the gas load GL with respect to the crank angle θ1 in the rotary compressor 1 of Embodiment 1. Furthermore, in Figure 5 In this study, the gas load GL was normalized using its maximum value. Figure 5 In the diagram, the horizontal axis represents the crank angle θ1, and the vertical axis represents the normalized gas load GL. Since the volume of the compression chamber 507b changes with the change in crank angle θ1, therefore, as... Figure 5 As shown, the magnitude of the gas load GL varies with the crank angle θ1. Specifically, as... Figure 5 As shown, the gas load GL reaches its maximum when the crank angle θ1 is approximately 225 degrees. That is, within the range of 0 to 120 degrees of crank angle θ1, the gas load GL increases slowly, while within the range of 120 to 225 degrees of crank angle θ1, the gas load GL increases sharply. Furthermore, within the range of 225 to 360 degrees of crank angle θ1, the gas load GL decreases sharply.

[0081] Figure 6 This is a diagram illustrating an example of the action angle θ2 of the gas load GL relative to the crank angle θ1 in the rotary compressor 1 of Embodiment 1. Figure 6 In the diagram, the horizontal axis represents the crank angle θ1, and the vertical axis represents the acting angle θ2 of the gas load GL. For example... Figure 6As shown, due to the change in the volume of the compression chamber 507b along with the change in the crank angle θ1, the action angle θ2 of the gas load GL also changes. Figure 6 In the example, relative to the entire range of the crank angle θ1, i.e., θ1 is from 0 degrees to 360 degrees, the range of variation of the action angle θ2 of the gas load GL, i.e., the variation amplitude of the action angle θ2 of the gas load GL, is from 110 degrees to 330 degrees. Hereinafter, this range will be referred to as the angle range B1 representing the variation amplitude of the action angle θ2. Specifically, as... Figure 6 As shown, when the crank angle θ1 is approximately 180 degrees, the effective angle θ2 of the gas load GL is at its minimum, becoming θ2 = 110 degrees. That is, θ2 = 110 degrees is the lower limit of the angle range B1. Furthermore, when the crank angle θ1 is 360 degrees, the effective angle θ2 of the gas load GL is at its maximum, becoming θ2 = 330 degrees. That is, θ2 = 330 degrees is the upper limit of the angle range B1. Additionally, when the crank angle θ1 is 0 degrees, the effective angle θ2 of the gas load GL is 180 degrees. Afterward, within the range of 0 to 30 degrees of crank angle θ1, the effective angle θ2 of the load GL decreases. Furthermore, within the range of 30 to 180 degrees of crank angle θ1, the effective angle θ2 of the load GL remains constant or gradually decreases. Furthermore, within the range of crank angle θ1 from 180 degrees to 355 degrees, the effective angle θ2 of load GL gradually increases, while within the range of crank angle θ1 from 355 degrees to 360 degrees, the effective angle θ2 of load GL increases sharply.

[0082] Thus, in the rotary compressor 1 of embodiment 1, as Figure 5 and Figure 6 As shown, the magnitude of the gas load GL and the operating angle θ2 change with the change of the crank angle θ1.

[0083] In Embodiment 1, the bearing section 300 is assumed to be constructed of a sliding bearing. That is, at least one of the upper bearing 301 and the lower bearing 302 is constructed of a sliding bearing. Furthermore, in Embodiment 1, the oil film rigidity of this sliding bearing is anisotropic. Oil film rigidity refers to the rigidity of the refrigeration oil between the rotating shaft 201 and the bearing section 300, through which the rotating shaft 201 is supported. Therefore, the oil film rigidity functions as the bearing support rigidity.

[0084] Here, a method for making the oil film rigidity of a sliding bearing exhibit anisotropy will be described. As one method, the inner diameter surface 300a of the sliding bearing constituting the bearing portion 300 (for example, referring to...) can be used as an example. Figure 12The shape of the sliding bearing is not perfectly circular. The inner diameter surface 300a is the sliding surface for the rotating shaft 201 to slide. The sliding bearing has a cylindrical shape and an internal cavity. The rotating shaft 201 is disposed within this cavity. The sliding bearing is formed of a copper alloy such as brass, iron, or a special resin. A gap 400 is formed between the inner diameter surface 300a of the sliding bearing and the rotating shaft 201 (see, for example, reference). Figure 12 It is injected with refrigeration oil.

[0085] For example, such as Figure 7 As shown, when the inner diameter surface 300a of the sliding bearing constituting the bearing section 300 is circular, a gap 400 is formed between the inner diameter surface 300a and the rotating shaft 201 (see reference). Figure 7 The width of the gap 400 is consistent. In this case, the oil film thickness of the refrigeration oil filling the gap 400 is also consistent, and the oil film thickness does not change during one revolution of the rotating piston 504. Therefore, the oil film rigidity does not exhibit anisotropy. Here, one revolution of the rotating piston 504 refers to... Figure 3 As shown, the rotating piston 504 is rotated within a 360-degree range from a state where the crank angle θ1 is 0 degrees until the crank angle θ1 becomes 360 degrees.

[0086] On the other hand, for example, Figure 12 As shown, when the inner diameter surface 300a of the sliding bearing constituting the bearing section 300 is not circular, the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 is not uniform. In other words, the gap 400 includes a narrower portion and a wider portion. In this case, the oil film thickness of the refrigeration oil filling the gap 400 is also inconsistent, changing as the rotating piston 504 rotates one revolution. Therefore, the oil film rigidity exhibits anisotropy. The following... Figures 7 to 22 A portion of the shape of a journal-type sliding bearing is shown. The following examples of specific sliding bearings further illustrate the anisotropy of oil film stiffness.

[0087] As Figures 7 to 22 The shapes of anisotropic inner diameter surfaces in a given area, for example, can be listed as follows: Figure 12 Double arc bearings, etc. In sliding bearings, the size of the clearance 400 changes in the rotational direction of the rotating piston 504. As a result, pressure is generated on the fluid such as refrigeration oil flowing in the clearance 400 through the narrow regions S1 and S2 of the clearance 400, thereby creating oil film rigidity. In double arc bearings, such as... Figure 12 As shown, even when the rotating shaft 201 and the inner diameter surface 300a are concentric, because the inner diameter surface 300a is not perfectly circular, a gap 400 will still be generated in the rotation direction of the rotating piston 504. Therefore, in Figure 12 In particular, in the direction of regions S1 and S2 where the gap is relatively small (400), that is... Figure 12 A higher oil film pressure will be generated in the direction perpendicular to the paper surface. Therefore, in Figure 12 In bearings like the one shown, the oil film rigidity in the vertical direction, where the clearance 400 is smaller, is higher than the oil film rigidity in the horizontal direction. Hereinafter, the direction with the highest oil film rigidity will be referred to as the "axial direction D with high oil film rigidity." Figure 12 In the case of the double circular arc bearing shown, as described later... Figure 24 As indicated by the thick black arrow, the axial direction D of the high oil film rigidity is the direction of regions S1 and S2 where the gap 400 is relatively small.

[0088] Here, in order to make the explanation easier to understand, Figure 12 The example shown is a double circular arc bearing, but the implementation in Embodiment 1 is not limited to this case. That is, it is not limited to this specific implementation. Figures 7 to 22 The shape example shown, as long as it includes the inner diameter surface shape containing surface structures such as textures that cause anisotropy in the oil film rigidity of the sliding bearing, can apply the features of Embodiment 1 regardless of the shape used. Furthermore, as a method to prevent the shape of the inner diameter surface 300a from changing from a perfect circle, it is also possible to use a magnetic bearing and control the electromagnetic force of the electromagnet, or to control the supply pressure of the hydrostatic bearing, thereby imparting anisotropy to the bearing support rigidity of the sliding bearing. In addition, when controlling the electromagnetic force of the electromagnet, making the electromagnetic force inconsistent throughout the entire circumference of the sliding bearing, and locally strengthening the electromagnetic force, anisotropy is imparted to the sliding bearing support rigidity. Similarly, when controlling the supply pressure of the hydrostatic bearing, making the supply pressure inconsistent throughout the entire circumference of the sliding bearing, and locally strengthening the supply pressure, anisotropy is imparted to the sliding bearing support rigidity.

[0089] The following is about Figures 7 to 22 The journal-type sliding bearing shown is described below.

[0090] Figure 7 This is a plan view showing the structure of a full bearing. Full bearings are the most common type of bearing, used, for example, in dynamic load applications. Figure 7As shown, in a full-circumferential bearing, the inner diameter surface 300a of the sliding bearing has a circular shape when viewed from above. Therefore, when the sliding bearing constituting the bearing section 300 is a full-circumferential bearing, the oil film rigidity of the sliding bearing is not anisotropic. Therefore, when the sliding bearing constituting the bearing section 300 is a full-circumferential bearing, anisotropy of bearing support rigidity can be imparted to the sliding bearing simply by constructing the sliding bearing with a magnetic bearing and controlling the electromagnetic force of the electromagnet. Alternatively, anisotropy of bearing support rigidity can be imparted to the sliding bearing simply by constructing the sliding bearing with a hydrostatic bearing and controlling the supply pressure.

[0091] Figure 8 This is a plan view showing the structure of a partial bearing. For example... Figure 8 As shown, the partial bearing supports only a portion of the outer periphery of the rotating shaft 201. In the partial bearing, the inner diameter surface 300a of the sliding bearing is a portion 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 only formed on a portion of the outer periphery of the rotating shaft 201. Therefore, when the sliding bearing constituting the bearing portion 300 is a partial bearing, the oil film rigidity of the sliding bearing is anisotropic.

[0092] Figure 9 This is a plan view showing the structure of a split bearing. (Example) Figure 9 As shown, in a split bearing, the shape of the inner diameter surface 300a of the sliding bearing, when viewed from above, consists of two curves with inconsistent curvature. Therefore, the clearance 400 formed between the inner diameter surface 300a and the rotating shaft 201 is not consistent relative to the entire circumference of the rotating shaft 201. Consequently, when the sliding bearing constituting the bearing portion 300 is a split bearing, the oil film rigidity of the sliding bearing is anisotropic. Furthermore, in Figure 9 The diagram shows a case where the inner diameter surface 300a of the sliding bearing is divided into two parts, but this is not the only possible scenario. That is, a split bearing can also be divided into three or more parts. In this case, the inner diameter surface 300a of the sliding bearing is composed of three or more curves.

[0093] Figure 10 This is a plan view showing the structure of a fitted bearing. For example... Figure 10 As shown, the matching bearing only supports a portion of the outer circumference of the rotating shaft 201. The matching bearing is a type of partial bearing, but it is different from... Figure 8Compared to the partial bearing, the sliding bearing has almost no gap 400 formed between the inner diameter surface 300a of the sliding bearing and the rotating shaft 201. Therefore, the oil film pressure distribution of the matched bearing is uniform. However, in the matched bearing, the shape of the inner diameter surface 300a of the sliding bearing in cross-section is part of a perfect circle. Therefore, when the sliding bearing constituting the bearing portion 300 is a matched bearing, the oil film rigidity of the sliding bearing is anisotropic.

[0094] Figure 11 This is a plan view showing the structure of a stepped thrust bearing. Stepped thrust bearings are sometimes referred to as stepped bearings. Figure 11 As shown, the stepped thrust bearing has one or more steps 300m formed on the inner diameter surface 300a of the sliding bearing. The step 300m is concave on the inner diameter surface 300a. A portion of the step 300m protrudes radially outward relative to the inner diameter surface 300a of the sliding bearing. Therefore, the step 300m is concave radially outward from the inner diameter surface 300a. Consequently, in the area where the step 300m is provided, the clearance 400 is wider compared to other portions. Furthermore, in... Figure 11 In the example, four steps 300m are formed at intervals along the circumference of the inner diameter surface 300a. The number and shape of the steps 300m are not limited to... Figure 11 For example, in a stepped thrust bearing, one or more steps 300m are formed on the inner diameter surface 300a. Therefore, the shape of the inner diameter surface 300a of the sliding bearing is not a perfect circle. Therefore, when the sliding bearing constituting the bearing part 300 is a stepped thrust bearing, the oil film rigidity of the sliding bearing is anisotropic.

[0095] Figure 12 This is a plan view showing the structure of a double circular arc bearing. (Example) Figure 12 As shown, in a double circular arc bearing, the inner diameter surface 300a of the sliding bearing is composed of two circular arcs. Figure 12 As shown, the two arcs have the same shape and are arranged in a linearly symmetrical manner. Furthermore, the curvature of the two arcs differs 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, including portions with a narrower gap 400 and portions with a wider gap 400. When the sliding bearing constituting the bearing portion 300 is a double-arc bearing, the oil film rigidity of the sliding bearing is anisotropic. Since the inner diameter surface 300a of the double-arc bearing is composed of multiple arc surfaces, it is sometimes referred to as a multi-arc bearing.

[0096] Figure 13 This is a plan view showing the structure of a three-circular-arc bearing. (Example) Figure 13 As shown, in a three-circular-arc bearing, the inner diameter surface 300a of the sliding bearing is composed of three circular arcs. Figure 13 As shown, the three arcs have the same shape. The three arcs are arranged adjacent to each other, each protruding radially outward. Furthermore, the curvature of the three arcs differs 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, including portions with narrower gaps and portions with wider gaps. When the sliding bearing constituting the bearing portion 300 is a three-arc bearing, the oil film rigidity of the sliding bearing is anisotropic. Since the inner diameter surface 300a of the three-arc bearing is composed of multiple arc surfaces, it is sometimes referred to as a multi-arc bearing.

[0097] Figure 14 This is a plan view showing the structure of a double circular arc eccentric bearing. (Example) Figure 14 As shown, in a double-circular-arc eccentric bearing, the inner diameter surface 300a of the sliding bearing is composed of two circular arcs. Figure 14 As shown, the two arcs have the same shape and are arranged symmetrically with respect to the center of the rotation axis 201. Furthermore, the curvature of the two arcs differs from the curvature of the outer periphery of the bearing portion 300. Moreover, the centers of the circles constituting the two arcs are located at positions offset from each other in the horizontal direction. Therefore, the gap 400 formed between the inner diameter surface 300a and the rotation axis 201 is not uniform, including a narrower gap 400 and a wider gap 400. Additionally, since the two arcs are arranged offset in the horizontal direction, a straight section connecting the two arcs is provided between them. That is, the inner diameter surface 300a of the double arc eccentric bearing is formed by two arcs and two straight sections. When the sliding bearing constituting the bearing portion 300 is a double arc eccentric bearing, the oil film rigidity of the sliding bearing is anisotropic. Since the inner diameter surface 300a of the double arc eccentric bearing is composed of multiple arc surfaces, it is sometimes referred to as a multi-arc bearing.

[0098] Figure 15 This is a plan view showing the structure of a floating bushing bearing. Floating bushing bearings were originally designed to reduce heat generation in high-speed bearings, for example, in applications where the rotating shaft 201 rotates at high speed. Figure 15 As shown, a floating bushing bearing has a thin-walled cylindrical bushing 300b inserted into a gap 400 formed between the inner diameter surface 300a and the rotating shaft 201. When oil is supplied to the gap 400, the rotating shaft 201 is supported by a double oil film formed on both the inner and outer sides of the bushing 300b. As the rotating shaft 201 rotates, the bushing 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. Figure 15As shown, in a floating bushing bearing, the inner diameter surface 300a of the sliding bearing is a perfect circle. Therefore, when the sliding bearing constituting the bearing section 300 is a floating bushing bearing, the oil film rigidity of the sliding bearing is not anisotropic. Therefore, when the sliding bearing constituting the bearing section 300 is a floating bushing bearing, it is sufficient to impart anisotropy to the oil film rigidity of the sliding bearing by using either a magnetic bearing or a hydrostatic bearing.

[0099] Figure 16 This is a plan view showing the structure of a pivot-pad bearing. A pivot-pad bearing is a type of tilting pad bearing. (Example...) Figure 16 As shown, the pivot bushing bearing has a plurality of segmented bushings 300c arranged in a gap 400 formed between the inner diameter surface 300a and the rotating shaft 201. A pivot point, referred to as a pivot 300d, is arranged between each bushing 300c and the inner diameter surface 300a. By supporting each bushing 300c using the pivot 300d, a wedge-shaped film is formed on each bushing 300c. Thus, in the pivot bushing bearing, a wedge-shaped film effect can be obtained, where pressure is generated by forcing viscous fluid into the wedge-shaped gap 400, where the outlet is narrower than the inlet. Figure 16 As shown, in a pivot-shielded bearing, the inner diameter surface 300a of the sliding bearing is circular, but a pivot 300d and a shim 300c are arranged within the clearance 400, resulting in inconsistent oil film thickness. Therefore, when the sliding bearing constituting the bearing portion 300 is a pivot-shielded bearing, the oil film rigidity of the sliding bearing exhibits anisotropy.

[0100] Figure 17 This is a plan view showing the structure of a Michell-Seggell bearing. The Michell-Seggell bearing is a type of tilting pad bearing. (As shown...) Figure 17 As shown, the Michel-Seger bearing has multiple gaskets 300c arranged in the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201. Figure 16 The difference in the pivot bushing bearing shown is that pivot 300d is not used in the Michel-Seger bearing. In the Michel-Seger bearing, bushing 300c rotates together with the rotating shaft 201. Figure 17 As shown, in a Michel-Seger bearing, the inner diameter surface 300a of the sliding bearing is circular, but a shim 300c is disposed within the clearance 400, resulting in inconsistent oil film thickness. Therefore, when the sliding bearing constituting the bearing section 300 is a Michel-Seger bearing, the oil film rigidity of the sliding bearing exhibits anisotropy.

[0101] Figure 18 This is a plan view showing the structure of a frictionless bearing. Frictionless bearings are a type of tilting pad bearing. For example... Figure 18 As shown, the frictionless bearing has multiple gaskets 300c arranged in the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201. Frictionless bearings are named for their low friction, and the gaskets 300c have a spherical surface. Like the Michel-Seger bearing, the frictionless bearing does not use a pivot 300d. Therefore, in the frictionless bearing, the gaskets 300c rotate together with the rotating shaft 201. However, in… Figure 18 In the example, a protrusion 201a is provided on the side of the rotating shaft 201. The protrusion 201a protrudes radially outward from the side of the rotating shaft 201. The protrusion 201a restricts the circumferential movement of the gasket 300c. That is, the gasket 300c can only move circumferentially between two adjacent protrusions 201a. Figure 18 As shown, in a frictionless bearing, the inner diameter surface 300a of the sliding bearing is circular, but a shim 300c is disposed within the clearance 400, resulting in inconsistent oil film thickness. Therefore, when the sliding bearing constituting the bearing portion 300 is a frictionless bearing, the oil film rigidity of the sliding bearing exhibits anisotropy.

[0102] Figure 19 This is a plan view showing the structure of a porous bearing. (For example...) Figure 19 As shown, a porous bearing has a thin-walled cylindrical porous member 300e made of porous material mounted on its inner diameter surface 300a. 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. The porous member 300e has a large number of holes 300ee formed on its surface. The holes 300ee are concave on the inner diameter surface 300a. Furthermore, the holes 300ee are not uniformly distributed. Therefore, by allowing refrigeration oil to enter the holes 300ee, the thickness of the oil film becomes substantially inconsistent. Figure 19 As shown, in a porous bearing, the inner diameter surface 300a of the sliding bearing is circular, but by providing a porous component 300e on the inner diameter surface 300a, the oil film rigidity becomes anisotropic.

[0103] Figure 20 This is a plan view showing the structure of the foil bearing. (Example) Figure 20 As shown, the foil belt bearing is a bearing formed by multiple foil belts 300f. In the foil belt 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 membrane pressure. Figure 20As shown, the inner surface of the foil strip 300f is used to form the inner diameter surface 300a of the sliding bearing. Figure 20 As shown, the inner surfaces of the three foil strips 300f have the same shape. The inner surfaces of the three foil strips 300f are arranged adjacent to each other in a manner that bulges outward in a radial direction. The inner surfaces of the three foil strips 300f are each composed of curved surfaces, but their curvature is not uniform; rather, it changes midway. Therefore, as... Figure 20 As shown, the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 is not uniform, but includes a narrower gap 400 and a wider gap 400. When the sliding bearing constituting the bearing section 300 is a foil bearing, the oil film rigidity of the sliding bearing is anisotropic.

[0104] Figure 21 This is a schematic cross-sectional view illustrating the structure of a bearing with helical grooves. However, for illustrative purposes, in Figure 21 The section showing the rotating shaft 201 is not shown; instead, its side surface is depicted. In the helical groove bearing, the sliding bearing constituting the bearing portion 300 has a cylindrical shape. The rotating shaft 201 is disposed within the cylindrical cavity portion. Furthermore, in the helical groove bearing, a groove is provided on either the rotating shaft 201 or the inner diameter surface 300a of the sliding bearing. Figure 21 In the example, multiple finer spiral grooves 300k are provided as the "groove". This "groove" is formed concavely on the inner diameter surface 300a. Furthermore, in... Figure 21 In the example, a helical groove 300k is provided on the side of the rotating shaft 201, but it can also be provided on the inner diameter surface 300a. The bearing with helical grooves is a bearing that provides hydraulic pressure by pumping refrigerant oil into the gap 400 formed between the inner diameter surface 300a and the rotating shaft 201 during the rotation of the rotating shaft 201. In the bearing with helical grooves, the optimal values ​​of the groove torsion angle, groove depth, and the width ratio of the groove to the shoulder are determined using narrow groove theory, which solves for the pressure within the gap 400 based on the characteristics of locally stepped bearings, thus forming the helical groove 300k. Figure 21 In the example, a column with helical grooves 300k is formed by arranging multiple helical grooves 300k spaced apart from each other in the circumferential direction of the rotating shaft 201. Figure 21 In this example, two columns of helical grooves 300k are provided. These two columns of helical grooves 300k are arranged axially along the rotation shaft 201. Furthermore, Figure 21 The twisting direction of the spiral groove 300k in the left column of the paper is opposite to that of the spiral groove 300k in the right column. For example... Figure 21As shown, a type of bearing in which spiral grooves with opposite directions of twist are arranged from positions corresponding to both ends of the inner diameter surface 300a and are arranged in a manner that corresponds to the entire inner diameter surface 300a is sometimes referred to as a herringbone groove bearing. Furthermore, the number and shape of the spiral grooves 300k are not limited to... Figure 21 For example, in a bearing with helical grooves, one or more helical grooves 300k are formed as "grooves" 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 consistent. Therefore, when the sliding bearing constituting the bearing portion 300 is a bearing with helical grooves, the oil film rigidity of the sliding bearing is anisotropic.

[0105] Figure 22 This is a schematic cross-sectional view illustrating the structure of a spherical bearing. However, for illustrative purposes, in Figure 22 The section regarding the rotating shaft 201 is not shown; instead, the side view is shown. In spherical bearings, such as... Figure 22 As shown, the sliding bearing consists of an inner ring 300g and an outer ring 300h. Both the inner ring 300g and the outer ring 300h are cylindrical. The inner ring 300g is concentrically arranged inside the outer ring 300h. Figure 21 As shown, the outer circumference of the inner 300g ring is spherical. Additionally, as... Figure 21 As shown, the inner circumferential surface of the outer ring 300h is spherical. Therefore, the inner ring 300g and the outer ring 300h make spherical contact. To improve sliding, the contact surfaces of the inner ring 300g and the outer ring 300h can be specially machined or a component called a liner can be installed. In the spherical bearing, the inner ring 300g tilts relative to the outer ring 300h as the rotating shaft 201 rotates. Because the spherical bearing is constructed to make spherical contact between the inner ring 300g and the outer ring 300h, it is a self-aligning type bearing capable of simultaneously bearing radial loads and axial loads. In the spherical bearing, the inner diameter surface 300a of the sliding bearing is formed by the inner circumferential surface of the inner ring 300g. The inner ring 300g tilts axially relative to 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 constituting the bearing section 300 is a spherical bearing, the oil film rigidity of the sliding bearing is anisotropic.

[0106] In the above Figures 7 to 22When the oil film stiffness of the sliding bearing shown is anisotropic, the "axial direction D of high oil film stiffness," representing the direction with the highest oil film stiffness, can be determined. Since the "axial direction D of high oil film stiffness" varies depending on the type of sliding bearing constituting the bearing section 300, it is possible to determine the "axial direction D of high oil film stiffness" according to the type of sliding bearing, for example, through experimentation or simulation. For example, as... Figure 12 As shown, when the sliding bearing constituting the bearing section 300 is a double circular arc bearing, the "axial direction D of high oil film rigidity," representing the direction of highest oil film rigidity, is the direction where regions S1 and S2 exist. That is, if the following is used... Figure 24 To explain, as follows Figure 24 As indicated by the thick black arrow, "Axial direction D of high oil film rigidity" refers to the direction of regions S1 and S2 where the gap narrows by 400. In Embodiment 1, as... Figure 24 As shown, the "axis direction D with high oil film rigidity" is set to the opposite direction to the angle θ2 of the gas load GL.

[0107] Figure 23 This is an explanatory diagram showing the angle ranges A, B1, and C for setting the "axial direction D with high oil film rigidity" in the rotary compressor 1 of Embodiment 1. Figure 24 This 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 of Embodiment 1.

[0108] If using the above Figure 6 As explained, the angular range B1 representing the variation of the action angle θ2 of the gas load GL is 110 degrees to 330 degrees. Figure 23 The middle diagram shows the angle range B1. (From...) Figure 23 As shown in the angle range B1, the gas load GL acts approximately in a vertically downward direction at the angle θ2. In the sliding bearing, due to the rotational oil film pressure, the position of the axis of rotation of the shaft 201 changes to an off-center position relative to the direction of action of the gas load GL acting on the bearing portion 300 in the rotational direction. If the above... Figure 2 To illustrate with an example, when the gas load GL acts in a vertically downward direction as shown by arrow W, the axis, as shown by arrow E, is eccentric within an angle range of approximately 180 to 270 degrees, which is located in the third quadrant.

[0109] In the sliding bearings of the bearing section 300 constituting the rotary compressor 1, factors affecting the minimum oil film thickness and the position where the minimum oil film is formed include the magnitude of the gas load GL, the angle θ2 of the gas load GL, and the eccentricity direction of the shaft center of the rotating shaft 201. In addition, factors such as the viscosity coefficient of the refrigeration oil and the rotational speed of the rotating shaft 201 also play a role. However, in Embodiment 1, with the aim of increasing the minimum oil film thickness under variable loads, the above three points, namely the magnitude of the gas load GL, the angle θ2 of the gas load GL, and the eccentricity direction of the shaft center, are explained.

[0110] like Figure 5 As shown in the graph, in the rotary compressor 1, the magnitude of the gas load GL accompanying refrigerant compression changes during one revolution of the crank angle θ1 from 0 degrees to 360 degrees. In the bearing section 300 where the oil film rigidity is anisotropic, as described above, directions with higher oil film rigidity and directions with lower oil film rigidity will be observed.

[0111] In the rotary compressor 1, the thickness of the oil film supplied to the refrigerant oil formed in the gap 400 between the bearing portion 300 and the rotating shaft 201 changes during one revolution of the rotary piston 504. In particular, when the gas load GL reaches its maximum crank angle θ1, there is a concern that the rotating shaft 201 may come into contact with the bearing portion 300, potentially leading to sintering in the worst case. Therefore, for example, if the direction with lower oil film rigidity is positioned in the direction where the gas load GL reaches its maximum crank angle θ1, there is a concern that the rotating shaft 201 may come into contact with the bearing portion 300, potentially leading to sintering in the worst case. Therefore, in Embodiment 1, the "axis direction D of high oil film rigidity," representing the direction with the highest oil film rigidity, is positioned in the direction where the gas load GL increases at the crank angle θ1. Furthermore, since the contact between the rotating shaft 201 and the bearing portion 300 is a metal-to-metal contact, it is sometimes referred to as metal-to-metal contact.

[0112] Therefore, in Embodiment 1, for example, a bearing portion 300 with anisotropic oil film rigidity is arranged in an "axial direction D of high oil film rigidity" within the angle range of the crank angle θ1 acting relative to the gas load GL (normalized load = 0.5[-]) which is more than 50% of the maximum load (normalized load = 1[-]). This suppresses the reduction in oil film thickness. Figure 5 The curve shows that the crank angle θ1, including the region where the normalized load is 0.5 or higher, falls within the angle range A of 180 degrees to 330 degrees. Therefore, when the "axial direction D with high oil film rigidity" is configured within angle range A, the reduction in oil film thickness can be suppressed. Specifically, in Figure 23 Configured in any direction within the angle range A Figure 24The thick black arrow indicates the "axial direction D of high oil film rigidity". The angle range A is based on a normalized... Figure 5 The angle range is determined by the curve of the gas load and is therefore sometimes referred to as the "gas load absolute value reference angle range". In addition, angle range A is sometimes referred to as the "second angle range".

[0113] In addition, according to Figure 6 The curve, as described above, shows that the angular range B1 representing the variation of the action angle θ2 of the gas load GL is 110 degrees to 330 degrees. Therefore, the "axial direction D with high oil film rigidity" can also be configured within the angular range B1. In this case, the reduction in oil film thickness can also be suppressed. Specifically, in Figure 23 Configured in any direction within the angle range B1 Figure 24 The thick black arrow indicates the "axial direction D of high oil film rigidity". Since the angle range B1 is based on the action angle θ2 representing the gas load GL... Figure 6 The angle range is determined by the curve diagram, and therefore is sometimes referred to as the "reference angle range for the direction of gas load action". In addition, the angle range B1 is sometimes referred to as the "first angle range".

[0114] In addition, according to Figure 6 The curve, and Figure 5 The angle range A corresponds to the action angle θ2 of the gas load GL, which is the angle range B2 from 110 degrees to 180 degrees. Therefore, the "axial direction D with high oil film rigidity" can also be configured within the angle range B2. Even in this case, the reduction in oil film thickness can be suppressed. In addition, the angle range B2 is sometimes referred to as the "third angle range".

[0115] As mentioned above, the angle θ2 of the gas load GL is approximately along... Figure 2 The action is performed in the direction perpendicular to the downward surface of the paper. In the above... Figure 2 In the example, when the load acts in a vertically downward direction, the shaft center is eccentric within an angular range C of 180 to 270 degrees, which is located in the third quadrant. Therefore, the "axis direction D with high oil film rigidity" can also be configured within this angular range C of 180 to 270 degrees. Specifically, in Figure 23 Configured in any direction within the angle range C Figure 24 The thick black arrow indicates the "axial direction D of high oil film rigidity". The angle range C is sometimes referred to as the "reference angle range of the eccentricity characteristic of the sliding bearing". In addition, the angle range C is sometimes referred to as the "fourth angle range".

[0116] exist Figure 23 and Figure 24The diagram shows the angle ranges A, B1, and C corresponding to the aforementioned references, and an example of a bearing configuration using a double circular arc bearing. A comprehensive approach can also be considered. Figure 23 The references shown are used to select the crank angle θ1 range for the configuration of "high oil film rigidity axial direction D".

[0117] Furthermore, it is also possible to disassemble the rotary compressor that uses the conventional circular sliding bearing after actual operation, and measure the wear range of the shaft surface or bearing surface, i.e. the metal contact range, and configure "high oil film rigidity axial direction D" within this range.

[0118] The "axial direction D with high oil film rigidity" configuration described in Embodiment 1 can be applied to either the upper bearing 301 or the lower bearing 302, or either the upper bearing 301 or the lower bearing 302. Furthermore, it can also be a structure where one bearing is a sliding bearing with anisotropic oil film rigidity and the other is a conventional circular sliding bearing.

[0119] As described above, the rotary compressor 1 of Embodiment 1 includes a rotating shaft 201 and a bearing portion 300. The bearing portion 300 is composed of a sliding bearing and has an inner diameter surface 300a that supports the rotating shaft 201. Furthermore, the oil film rigidity of the sliding bearing constituting the bearing portion 300 is anisotropic. The "axial direction D of high oil film rigidity," representing the direction with the highest oil film rigidity of the sliding bearing, is arranged within an angle range B1 of 110 degrees to 330 degrees for the crank angle of the crank portion 204 of the rotating shaft 201. This suppresses the reduction in the oil film thickness of the refrigeration oil filling the gap 400 formed between the rotating shaft 201 and the bearing portion 300. In the rotary compressor 1, the oil film thickness changes during one revolution of the rotary piston 504. By setting the "axial direction D with high oil film rigidity" within the angle range A, metal-to-metal contact between the rotating shaft 201 and the inner diameter surface 300a of the bearing portion 300 can be avoided even when the gas load GL reaches its maximum crank angle θ1 (approximately 225 degrees). As a result, sintering of the inner diameter surface 300a of the bearing portion 300 can be suppressed.

[0120] Implementation method 2.

[0121] In Embodiment 1 described above, in the rotary compressor 1, even when the magnitude of the fluctuating gas load GL is large, the reduction in oil film thickness and the associated metal-to-metal contact can be suppressed. Therefore, a sliding bearing with anisotropic oil film rigidity is used in Embodiment 1. Furthermore, considering the magnitude of the gas load GL, the operating angle θ2 of the gas load GL, and the eccentric characteristics of the sliding bearing, the "axial direction D with high oil film rigidity" is configured within a predetermined angle range A.

[0122] However, in this case, although it is possible to suppress the reduction of oil film thickness and the associated metal contact in the "axial direction D with high oil film rigidity" and avoid sintering, which has been a previous problem, shaft vibration may increase due to the changing gas load GL in the axial direction with low oil film rigidity.

[0123] Therefore, in Embodiment 2, both the upper bearing 301 and the lower bearing 302 are constructed from sliding bearings with anisotropic oil film rigidity. Furthermore, 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 configured to be different directions. Since the other structures are the same as in Embodiment 1, their description is omitted here.

[0124] Figure 25 This is an explanatory diagram showing the direction in which the "axial direction D of high oil film rigidity" is set for the upper bearing 301 and the lower bearing 302 in the rotary compressor 1 of Embodiment 2. Figure 25 In the upper bearing 301, D1 shows the "axial direction of low oil film rigidity", which means the direction of minimum oil film rigidity.

[0125] For example, in Figure 25 The preferred configuration is shown in the diagram. Figure 25 In this example, the upper bearing 301 is positioned within the crank angle θ1 range A, which is intended to avoid sintering, along its "high oil film rigidity axial direction D". Furthermore, the lower bearing 302's "high oil film rigidity axial direction D" is positioned parallel to the upper bearing 301's "low oil film rigidity axial direction D1". Therefore, even if shaft vibration increases along the lower oil film rigidity axial direction D1 of the upper bearing 301, the high oil film rigidity of the lower bearing 302 can suppress the shaft vibration.

[0126] Thus, in Embodiment 2, when the rotary compressor 1 has multiple sliding bearings, the bearings are configured such that the "axial direction D of high oil film rigidity" of at least two or more of the sliding bearings is different from each other. That is, the "axial direction D of high oil film rigidity" of at least two or more sliding bearings are arranged in mutually intersecting directions. Figure 25 In the example, the "axial direction D of high oil film rigidity" of the upper bearing 301 is orthogonal to the "axial direction D of high oil film rigidity" of the lower bearing 302. Therefore, it is possible to suppress the increase in shaft vibration due to varying gas load GL in the axial direction of low oil film rigidity.

[0127] Furthermore, 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... Figure 25 As shown, even if the "axial direction D with high oil film rigidity" of the lower bearing 302 is positioned on the "axial direction D1 with low oil film rigidity" of the upper bearing 301, it may not be possible to achieve a sufficient effect in reducing shaft vibration.

[0128] In this case, displacement sensors or similar devices can be used to actually measure the shaft vibration, and the angle configuration of the "low oil film rigidity axial direction D1" of the lower bearing 302 can be finely adjusted to maximize the reduction effect of shaft vibration, or to make the actual measured shaft vibration below the preset vibration amount.

[0129] Furthermore, since the rotating shaft 201 of the rotary compressor 1 is housed within a sealed container 100 made of a pressure vessel, it is not easy to directly measure the amount of vibration. Therefore, for example, a vibration meter such as an accelerometer can be installed in the sealed container 100 to measure the vibration of the rotary compressor 1 and determine the angular configuration of the lower bearing 302 so that the vibration level of the rotary compressor 1 is below the desired vibration level.

[0130] The structure of Embodiment 2 can be applied to a rotary compressor 1 having multiple sliding bearings. Furthermore, the number of upper bearings 301 angled for preventing sintering and the number of lower bearings 302 angled for reducing vibration can be arbitrarily allocated. In the case of multiple upper bearings 301 angled for preventing sintering, the angle configuration of each upper bearing 301 only needs to be within the angle range A, B1, or C described in Embodiment 1; they can be the same or different from each other. Similarly, in the case of multiple lower bearings 302 angled for reducing vibration, the angle configuration of each lower bearing 302 can also be the same or different from each other.

[0131] Furthermore, in Embodiment 2, for convenience, the upper bearing 301 was described as an angled bearing for the purpose of avoiding sintering, and the lower bearing 302 was described as an angled bearing for the purpose of reducing vibration. However, this is not a limitation. That is, the lower bearing 302 may also be used as an angled bearing for the purpose of avoiding sintering. Alternatively, the upper bearing 301 may also be used as an angled bearing for the purpose of reducing vibration.

[0132] As described above, in Embodiment 2, when the bearing section 300 has multiple sliding bearings, the bearings are configured such that the "axial direction D of high oil film rigidity" of two or more of the multiple sliding bearings are different from each other. This allows for the combined effect of preventing sintering and reducing vibration in the bearing section 300.

[0133] Implementation method 3.

[0134] Figure 26 This is a plan view showing the structure of the bearing section 300 in the rotary compressor 1 of Embodiment 3. In Embodiment 3, the inner diameter surface 300a of the bearing section 300 is composed of multiple arcs when viewed from above. Since other structures are the same as in Embodiment 1, their description is omitted here.

[0135] As described above, in embodiment 3, such as Figure 26 As shown, the inner diameter surface 300a of the bearing portion 300, when viewed from above, is composed of multiple arcs. That is, the inner diameter surface 300a is composed of multiple arc surfaces. For simplicity, here, we will use the example of the inner diameter surface 300a being composed of... Figure 26 The bearing shape formed by the two arc surfaces shown is illustrated as an example. Hereinafter, the two arc surfaces will be referred to as the first arc surface 300a-1 and the second arc surface 300a-2, respectively.

[0136] Since the first arcuate face 300a-1 and the second arcuate face 300a-2 have the same shape, the structure of the first arcuate face 300a-1 will be described below, and the description of the second arcuate face 300a-2 will be omitted.

[0137] The shape of the arc surface of the first arc surface 300a-1 is determined according to the preload coefficient mp defined by the following formula (1).

[0138] mp=1-Cb / Cp (1)

[0139] Here, mp: preload coefficient, Cb: assembly radius clearance, and Cp: machining radius clearance.

[0140] The assembly radius gap Cb is the size of the imaginary line L1 extending along the first direction within the size of the gap 400 formed between the rotation axis 201 and the first arcuate surface 300a-1. Figure 26 In this context, the first direction is the vertical direction. Furthermore, in... Figure 26 In the example, the assembly radius gap Cb is the size of the gap 400 at the narrowest point of the gap 400. In other words, in Figure 26 In the example, the assembly radius gap Cb is the minimum value of the gap size of 400.

[0141] Next, the machining radius clearance Cp will be explained. For example... Figure 26 As shown, firstly, an imaginary circle with the same curvature as the arc forming the first arcuate surface 300a-1 is designated as imaginary circle 700. Next, an imaginary rotation axis that offsets the rotation axis 201 along the first direction in a manner concentric with the imaginary circle 700 is designated as imaginary rotation axis 701. Therefore, imaginary rotation axis 701 is a rotation axis located concentric with the imaginary circle 700. At this time, the machining radius gap Cp is the size of the imaginary gap 702 formed between the imaginary rotation axis 701 and the first arcuate surface 300a-1 at the imaginary line L1 extending along the first direction.

[0142] In the above description, the case where the preload coefficient mp of the first arc-shaped surface 300a-1 is the same as the preload coefficient mp of the second arc-shaped surface 300a-2 was described, but it is not limited to this case. That is, the preload coefficient mp of the first arc-shaped surface 300a-1 and the preload coefficient mp of the second arc-shaped surface 300a-2 can be the same or different.

[0143] Additionally, in the above explanation, it is said that... Figure 26 The example shown is a double-arc bearing composed of two arc surfaces. Therefore, it illustrates the case where the circumferential angle of the first arc surface 300a-1 is the same as the circumferential angle of the second arc surface 300a-2, and the first arc surface 300a-1 and the second arc surface 300a-2 face each other, but it is not limited to this case. That is, the central angle of the arc constituting the first arc surface 300a-1 and the central angle of the arc constituting the second arc surface 300a-2 can be the same or different. Thus, the two arc surfaces constituting the double-arc bearing may not be equally distributed. Moreover, even in the case of a bearing composed of three or more arc surfaces, these arc surfaces may not be equally distributed. Specifically, in... Figure 13 The three-circular-arc bearing shown, and the split bearing divided into three or more parts, Figure 11 The stepped thrust bearing shown Figure 16 The pivot bushing bearing shown Figure 17 The Michel-Seger bearing shown Figure 18 The frictionless bearing shown Figure 20 In the case of foil belt bearings, etc., the multiple surfaces constituting the inner diameter surface 300a can also be made not equally distributed among each other.

[0144] Moreover, after using Figure 26In the above description of the double arc bearing, it was explained that the circumferential position of the first arc surface 300a-1 and the circumferential position of the second arc surface 300a-2 are linearly symmetrical with respect to the imaginary line L2 extending along the second direction, but this is not a limitation. Furthermore, the second direction is a direction that intersects the first direction. For example, the second direction is orthogonal to the first direction. That is, the circumferential position of the starting point of the arc arrangement of the first arc surface 300a-1 and the circumferential position of the starting point of the arc arrangement of the second arc surface 300a-2 may not be linearly symmetrically arranged at the axis O of the bearing portion 300. That is, the circumferential positions of the starting points of the arc arrangements of these arc surfaces may not be equally spaced. Moreover, even in the case of a bearing composed of three or more arc surfaces, the circumferential positions of the starting points of the arc arrangements of these arc surfaces may not be equally spaced. Specifically, for example, in... Figure 13 In the three-arc bearing shown, if the central angle of the first arc is 180 degrees, the central angle of the second arc is 120 degrees, and the central angle of the third arc is 60 degrees, the circumferential positions that serve as the starting points of the arc configuration for each arc surface may not be equally spaced. However, it is desirable that the circumferential positions that serve as the starting points of the arc configuration for each arc surface are positions where the desired oil film rigidity can be obtained. Furthermore, in split bearings divided into three or more parts... Figure 11 The stepped thrust bearing shown Figure 16 The pivot bushing bearing shown Figure 17 The Michel-Seger bearing shown Figure 18 The frictionless bearing shown Figure 20 In the case of foil-belt bearings, etc., the circumferential positions of the starting points of the arc configurations that form the multiple arc surfaces constituting the inner diameter surface 300a may not be arranged at equal intervals. In these cases, it is also desirable that the circumferential positions of the starting points of the arc configurations of each arc surface are positions from which the desired oil film rigidity can be obtained.

[0145] Furthermore, while the above description illustrates the case where the inner diameter surface 300a is composed of two arcuate surfaces, it is not limited to this case. That is, the inner diameter surface 300a can also be composed of three or more arcuate surfaces. Figure 13 The case where the inner diameter surface 300a is composed of three arcuate surfaces is shown.

[0146] As described above, in Embodiment 3, the shape of the inner diameter surface 300a of the sliding bearing constituting the bearing section 300 is composed of multiple arc surfaces when viewed from above. Therefore, the oil film rigidity of the sliding bearing constituting the bearing section 300 has anisotropy. Thus, as explained in Embodiment 1, for example, by using an angle range A of 180 degrees to 330 degrees (refer to...) Figure 5The bearing portion 300 is configured with a "high oil film rigidity axial direction D", thereby suppressing the reduction of oil film thickness. As a result, since the oil film thickness is always maintained, metal-to-metal contact between the rotating shaft 201 and the bearing portion 300 can be prevented. Consequently, sintering of the bearing portion 300 can be avoided.

[0147] Implementation method 4.

[0148] In Embodiment 4, the feature is that a dynamic pressure generating part 600 for generating dynamic pressure of refrigeration oil is provided on the inner diameter surface 300a of the bearing part 300. Since the other structures are the same as in Embodiment 1, their description is omitted here.

[0149] For example, the dynamic pressure generating unit 600 is Figure 11 The step shown is 300m. Alternatively, the dynamic pressure generating unit 600 is, for example... Figure 21 The spiral groove 300k is shown. Alternatively, the dynamic pressure generating unit 600 is, for example, […]. Figure 19 The hole shown is 300ee. In Figure 19 The image shows an example of a porous component 300e with a hole 300ee mounted on an inner diameter surface 300a, but the hole 300ee can also be formed directly on the inner diameter surface 300a.

[0150] Thus, in Embodiment 4, as the dynamic pressure generating section 600, at least one of a step, a groove, and a hole is formed on the inner diameter surface 300a of the bearing section 300. The step, groove, or hole is formed in a concave shape on the inner diameter surface 300a. With the rotation of the rotating shaft 201 and the rotating piston 504, refrigerant oil flows into and out of the dynamic pressure generating section 600. Therefore, compared to the case where the dynamic pressure generating section 600 is not provided, the flow of refrigerant oil becomes more active, and dynamic pressure of the refrigerant oil is generated.

[0151] Furthermore, steps, grooves, and holes can be formed integrally on the inner diameter surface 300a of the bearing portion 300, or they can be formed only on a portion of the inner diameter surface 300a of the bearing portion 300.

[0152] As described above, according to Embodiment 4, at least one of a step, a groove, and a hole is formed on the inner diameter surface 300a of the bearing portion 300. This allows refrigerant oil to flow into and out of the step, groove, or hole, generating dynamic pressure of the refrigerant oil. Therefore, even if the direction of the gas load GL vector changes due to the rotational direction configuration and phase angle variation of the non-circular bearing portion 300, the desired dynamic pressure effect can always be obtained. As a result, the oil film thickness formed in the gap 400 between the rotating shaft 201 and the inner diameter surface 300a can be ensured, and sintering of the bearing portion 300 caused by oil film rigidity can be prevented. Furthermore, since the refrigerant oil is stored inside the step, groove, or hole, it can be retained on the surface of the rotating shaft 201. Moreover, by forming a step, groove, or hole on the inner diameter surface 300a of the bearing portion 300, the effect of capturing foreign matter from the outside can be expected.

[0153] Implementation method 5.

[0154] In Embodiment 5, the inner diameter surface 300a of the bearing portion 300 is non-circular, and the gap distribution of the gap 400 is continuous. Since other structures are the same as in Embodiment 1, their description is omitted here.

[0155] Figure 27 This is a plan view showing the structure of the bearing section 300 in the rotary compressor 1 of embodiment 5. Figure 27 The diagram shows an example of the inner diameter surface 300a of the bearing section 300 as formed by summing Fourier series when viewed from above.

[0156] For example, the circumferential distribution shape of the gap 400 in the inner diameter surface 300a of the bearing portion 300 can be expressed as a summation of a polynomial or a Fourier series with respect to the rotation angle. This allows for the formation of a complex shape in the inner diameter surface 300a that provides the desired oil film rigidity. Furthermore, this complex shape can vary not only in the circumferential direction of the bearing portion 300 but also in the axial direction.

[0157] Furthermore, when expressing the shape of the inner diameter surface 300a using a polynomial, for example, one or more curves constituting the inner diameter surface 300a can be represented by one or more polynomials as shown in equation (2) below. Therefore, the inner diameter surface 300a is divided into one or more regions, and the polynomial shown in equation (2) below is generated for each region, and the coefficients a of the polynomial are determined. k (k is an integer from 0 to n). Thus, the curve constituting the inner diameter surface 300a is generated by one or more polynomials as shown in equation (2) below.

[0158] y = a n x n+a n-1 x n-1 +……+a1x+a0

[0159] =Σa k x k (2)

[0160] Furthermore, when using Fourier series to express the shape of the inner diameter surface 300a, for example, one or more curves constituting the inner diameter surface 300a can be represented by the Fourier series shown in equation (3) below. Therefore, the inner diameter surface 300a is divided into one or more ranges, and the Fourier series shown in equation (3) below is generated for each range, and the coefficients a0, a10, and a20 of the Fourier series are determined. n b n (n is an integer from 1 to ∞). Thus, the curve constituting the inner diameter surface 300a is generated by summing the Fourier series shown in the following equation (3).

[0161] y = a0 / 2 + Σ(a n cos(nx)+b n sin(nx) ) (3)

[0162] As described above, in Embodiment 5, the inner diameter surface 300a of the sliding bearing constituting the bearing portion 300 has a non-circular shape when viewed from above, and the gap distribution of the gap 400 is continuous. Therefore, the inner diameter surface 300a can be formed into a complex shape that achieves the desired oil film rigidity. In this case, the necessary oil film thickness can always be ensured in the gap 400 formed between the rotating shaft 201 and the bearing portion 300, preventing metal-to-metal contact between the rotating shaft 201 and the bearing portion 300. As a result, sintering of the bearing portion 300 can be avoided.

[0163] Implementation method 6.

[0164] The bearing portion 300 of Embodiment 6 has two or more of the features of the bearing portion 300 shown in Embodiments 3, 4, and 5. Since other structures are the same as any of Embodiments 1 to 5, their description is omitted here.

[0165] In implementation 6, for example in Figure 26 In the double arc bearing shown (refer to embodiment 3), the first arc portion 300a-1 has the features of embodiment 4, and the second arc portion 300a-2 has the features of embodiment 5.

[0166] That is, a dynamic pressure generating part 600, which is a feature of embodiment 4, is provided on the first arc surface 300a-1 and is composed of a step 300m, a groove 300k, or a hole 300ee.

[0167] In addition, the second arcuate face 300a-2 has an inner diameter face shape expressed by a polynomial as an example of embodiment 5.

[0168] Thus, the bearing portion 300 of Embodiment 6 has a structure that combines multiple features of the bearing portion 300 shown in Embodiments 3, 4, and 5. Therefore, in Embodiment 6, the effects obtained from each of the combined features can be obtained. That is, in Embodiment 6, when Embodiments 3, 4, and 5 are combined as in the examples described above, all the effects obtained from Embodiments 3, 4, and 5 can be obtained.

[0169] Explanation of reference numerals in the attached figures

[0170] 1. Rotary compressor; 2. Suction piping; 3. Suction silencer; 4. Refrigerant discharge pipe; 5. Refrigeration oil; 6. Screw; 100. Sealed container; 101. Upper container; 102. Lower container; 201. Rotating shaft; 201a. Protrusion; 202. Main shaft; 203. Secondary shaft; 204. Crank; 300. Bearing; 300a. Inner diameter surface; 300a-1. First arcuate surface; 300a-2. Second arcuate surface; 300b. Bushing; 300c. Gasket; 300d. Pivot; 300e. Porous component; 300ee. Hole; 300f. Foil strip; 300g. Inner ring; 300h. Outer ring; 300k. Groove; 300m. Step; 3 01 Upper bearing, 302 Lower bearing, 400 Clearance, 401 Motor section, 402 Rotor, 403 Stator, 501 Compression mechanism section, 502 Cylinder block, 503 Exhaust muffler, 504 Rotary piston, 505 Blade, 506 Blade spring, 507 Cylinder chamber, 507a Intake chamber, 507b Compression chamber, 508 Blade groove, 508a Blocking section, 508b Blade spring receiving hole, 509 Exhaust port, 600 Dynamic pressure generating section, 700 Imaginary circle, 701 Imaginary rotating shaft, 702 Imaginary clearance, A Angle range, B1 Angle range, B2 Angle range, C Angle range, D Axial direction with high oil film rigidity.

Claims

1. A rotary compressor, wherein: The rotary compressor comprises: Axis of rotation; a bearing portion, the bearing portion being composed of a sliding bearing and having an inner diameter surface for supporting the rotating shaft; a cylinder chamber having a suction chamber for sucking refrigerant and a compression chamber for compressing the refrigerant; and a crank portion, the crank portion being mounted on the rotating shaft and rotating eccentrically in the cylinder chamber, A gap to which refrigeration oil is supplied is formed between the rotating shaft and the inner diameter surface of the sliding bearing. The oil film rigidity of the refrigerator oil of the sliding bearing is anisotropic, The angle at which the gas load generated by the compression of the refrigerant accompanying the eccentric rotation of the crank portion acts changes according to the crank angle. When the variation range of the action angle of the gas load is set to a first angle range, the axial direction of high oil film rigidity indicating the direction in which the oil film rigidity of the sliding bearing is highest is arranged in the first angle range.

2. The rotary compressor according to claim 1, wherein: The rotary compressor includes a vane that divides the suction chamber and the compression chamber of the cylinder chamber. When the crank angle of the blade top dead center of the blade is set to 0 degrees, the first angle range is an angle range in which the crank angle is 110 degrees to 330 degrees.

3. The rotary compressor according to claim 1 or 2, wherein: The bearing portion includes a plurality of the sliding bearings. The axial directions of the high oil film rigidity of two or more of the plurality of sliding bearings are different from each other.

4. The rotary compressor according to any one of claims 1 to 3, wherein: The inner diameter surface of the sliding bearing has a shape composed of a plurality of arc surfaces in a plan view.

5. The rotary compressor according to any one of claims 1 to 4, wherein: A dynamic pressure generating portion for generating dynamic pressure on the inner diameter surface of the sliding bearing is provided on the inner diameter surface. The dynamic pressure generating portion is formed by at least any one of a step, a groove, and a hole formed in a concave shape on the inner diameter surface.

6. The rotary compressor according to any one of claims 1 to 5, wherein: The inner diameter surface of the sliding bearing has a non-circular shape in a plan view, and the gap distribution of the gap formed between the rotating shaft and the inner diameter surface of the sliding bearing is continuous.

7. The rotary compressor according to claim 1, wherein: The magnitude of the gas load generated by the compression of the refrigerant accompanying the eccentric rotation of the crank portion varies according to the crank angle. When the crank angle range in which the gas load is 50% or more of the maximum value of the gas load is set as a second angle range, the high oil film rigidity axial direction indicating the direction in which the oil film rigidity of the sliding bearing is highest is arranged in the second angle range instead of the first angle range.

8. The rotary compressor according to claim 7, wherein: The rotary compressor includes a vane that divides the suction chamber and the compression chamber of the cylinder chamber. When the crank angle of the blade top dead center of the blade is set to 0 degrees, the second angle range is an angle range in which the crank angle is 180 degrees to 330 degrees.

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