CN111788394A

CN111788394A - Compressor with a compressor housing having a plurality of compressor blades

Info

Publication number: CN111788394A
Application number: CN201880089937.7A
Authority: CN
Inventors: 横山哲英; 岩崎俊明; 堀口大辅; 伊藤慎一; 冈本政哉; 小山修平
Original assignee: Mitsubishi Electric Corp
Current assignee: Mitsubishi Electric Corp
Priority date: 2018-03-05
Filing date: 2018-03-05
Publication date: 2020-10-16
Also published as: WO2019171427A1; JPWO2019171427A1

Abstract

The compressor of the present invention comprises: a compression mechanism unit disposed in the closed container; a motor unit that drives the compression mechanism unit; a drive shaft for transmitting the driving force of the motor to the compression mechanism; a main bearing supporting an upper portion of the drive shaft; a sub-bearing supporting a lower portion of the drive shaft; a radial bearing surface provided on the sub-bearing and supporting a radial surface of the drive shaft to be slidable; and a thrust receiving surface that slidably supports the thrust surface of the drive shaft. The thrust surface and the thrust receiving surface have a contact point between them, which makes the thrust surface and the thrust receiving surface contact with each other in a curved surface, and a gap which continuously increases from the contact point toward the radially outer side.

Description

Compressor with a compressor housing having a plurality of compressor blades

Technical Field

The present invention relates to a compressor used as one of components of a refrigeration cycle apparatus.

Background

A scroll compressor is one of the components of a refrigeration cycle apparatus such as an air conditioner, and includes: a compression mechanism portion having an oscillating scroll and a fixed scroll engaged with each other; a motor unit that drives the compression mechanism unit; and a drive shaft that transmits the driving force of the motor unit to the compression mechanism unit. The drive shaft is rotatably supported by main bearings and sub bearings provided above and below the motor unit. When the drive shaft is rotated by the motor, the orbiting scroll provided on the eccentric shaft portion of the upper end portion of the drive shaft revolves. Thereby, the refrigerant is compressed in the compression chamber between the orbiting scroll and the fixed scroll provided in the compression mechanism portion. When the refrigerant is compressed in the compression mechanism, a radial gas load acts on the drive shaft, and the gas load is supported by the main bearing and the sub-bearing. In addition, the sub-bearing supports the rotation of the drive shaft and supports the weight of the drive shaft in the vertical downward direction.

In such a scroll compressor, a ball bearing is often used as a sub bearing in order to simultaneously support both a load in a radial direction (hereinafter, referred to as a radial load) and a load in a thrust direction (hereinafter, referred to as a thrust load) (see, for example, patent document 1).

However, the ball bearing is expensive, and the inner ring and the rolling ball, and the outer ring and the rolling ball receive the load in point contact, respectively, and therefore, there are various problems such as poor long-term reliability. As a countermeasure, there is a scroll compressor in which a sliding bearing is applied to a sub-bearing (for example, see patent document 2).

In the scroll compressor described in patent document 2, the sub-bearing as the sliding bearing is constituted by a radial bearing and a thrust bearing which support a radial load and a thrust load independently from each other. The thrust surface provided at the lower end portion of the drive shaft is received by the thrust receiving surface provided at the sub-bearing, and the radial surface provided at the sub-shaft portion of the drive shaft is received by the radial receiving surface provided at the sub-bearing.

Patent document 1: japanese laid-open patent publication No. H04-241786

Patent document 2: japanese patent No. 4356375

Generally, a drive shaft of a scroll compressor is subjected to a compression load and a centrifugal force during operation, and a large load is applied in a radial direction. Therefore, the axial center of the drive shaft bends while inclining with respect to the central axis of the compressor. The central axis of the compressor is a shaft extending in the vertical direction. The main shaft portion and the auxiliary shaft portion of the drive shaft rotate while being tilted with respect to the main bearing and the auxiliary bearing, respectively. As in patent document 1, when a ball bearing is used for the sub bearing of the scroll compressor, the inclination of the drive shaft is absorbed by the gap between the rolling ball and the inner ring and the gap between the rolling ball and the outer ring, and thus it is easy to ensure the parallelism between the drive shaft and the sub bearing. However, the ball bearing is inferior in cost and long-term reliability as described above.

As in patent document 2, when a sliding bearing is used for the sub-bearing, although it is advantageous in terms of cost and long life, there are problems as follows: the drive shaft rotating while being inclined is in contact with one side of the thrust receiving surface, so that the local hertzian stress is increased, and the sliding state on the thrust receiving surface becomes severe. However, in patent document 2, no consideration is given to the problem of one-side contact with respect to the thrust receiving surface.

Disclosure of Invention

The present invention has been made in view of the above-described circumstances, and an object thereof is to provide a compressor in which a sliding bearing is used as a sub-bearing, and which can ensure a good sliding state of a thrust receiving surface.

The compressor according to the present invention includes: a compression mechanism unit disposed in the closed container; a motor unit that drives the compression mechanism unit; a drive shaft for transmitting the driving force of the motor to the compression mechanism; a main bearing supporting an upper portion of the drive shaft; a sub-bearing supporting a lower portion of the drive shaft; a radial bearing surface provided on the sub-bearing and supporting a radial surface of the drive shaft to be slidable; and a thrust receiving surface that slidably supports the thrust surface of the drive shaft, and between the thrust surface and the thrust receiving surface, a contact point at which the thrust surface and the thrust receiving surface are brought into contact with each other in a curved surface and a slit that continuously increases from the contact point toward a radially outer side are provided.

According to the present invention, a favorable sliding state of the thrust receiving surface can be ensured.

Drawings

Fig. 1 is a vertical sectional view schematically showing a sectional structure of a scroll compressor according to embodiment 1 of the present invention.

Fig. 2 is a diagram showing a main part of a scroll compressor according to embodiment 1 of the present invention.

Fig. 3 is an explanatory view of a bent state of the drive shaft of the scroll compressor according to embodiment 1 of the present invention in the Z-X cross section and a position of the radial oil groove in the circumferential direction.

Fig. 4 is an explanatory diagram of a state of deflection in the Y-Z cross section and a position of arrangement in the circumferential direction of the radial oil groove of the drive shaft of the scroll compressor according to embodiment 1 of the present invention.

Fig. 5 is an explanatory view of the load acting on the drive shaft and the direction in which the load acts in the scroll compressor according to embodiment 1 of the present invention.

Fig. 6 is a diagram showing a modification 1 of the scroll compressor according to embodiment 1 of the present invention.

Fig. 7 is a diagram showing a modification 2 of the scroll compressor according to embodiment 1 of the present invention.

Fig. 8 is a diagram showing a modification 3 of the scroll compressor according to embodiment 1 of the present invention.

Fig. 9 is a diagram showing a main part of a scroll compressor according to embodiment 2 of the present invention, and is a schematic diagram showing a portion closer to a lower end than a sub bearing in an enlarged manner.

Fig. 10 is a diagram showing a main part of a scroll compressor according to embodiment 3 of the present invention, and is a schematic diagram showing a portion closer to a lower end than a sub bearing in an enlarged manner.

Detailed Description

Hereinafter, a scroll compressor according to an embodiment of the present invention will be described with reference to the drawings as an example of the compressor. Here, including fig. 1, in the following drawings, the same or corresponding portions are denoted by the same reference numerals, and are used in common throughout the embodiments described below. The embodiments of the constituent elements expressed throughout the specification are merely examples, and are not limited to the embodiments described in the specification.

Embodiment mode 1

Fig. 1 is a vertical sectional view schematically showing a sectional structure of a scroll compressor according to embodiment 1 of the present invention. In fig. 1, the thrust load is indicated by arrow a and the radial load is indicated by arrow b. Fig. 2 is a diagram showing a main part of a scroll compressor according to embodiment 1 of the present invention. In fig. 2, (a) is a schematic view showing a portion closer to the lower end than the sub-bearing of the drive shaft in an enlarged manner, and (b) is a schematic view showing the shape of the thrust surface of the drive shaft as viewed from directly below. In fig. 2, arrows indicate the flow of the lubricating oil.

The scroll compressor 100 is mounted as one of refrigeration devices constituting a refrigeration cycle device such as a refrigerator, a freezer, a vending machine, an air conditioner, a refrigeration device, or a water heater, for example.

The scroll compressor 100 includes: a compression mechanism portion A housed in the closed casing 13; a motor section B; and a drive shaft 6 that transmits the driving force of the motor unit B to the compression mechanism unit a. As shown in fig. 1, the compression mechanism portion a is disposed above the closed casing 13, and the motor portion B is disposed below the closed casing 13. When the drive shaft 6 is driven to rotate by the motor portion B, the volume of a compression chamber 5, which will be described later, formed in the compression mechanism portion a is reduced, and the refrigerant in the compression chamber 5 is compressed.

Further, a refrigerant suction pipe 15 for sucking the refrigerant and a refrigerant discharge pipe 16 for discharging the refrigerant are connected to the closed casing 13. The closed casing 13 is a pressure vessel and constitutes an outer shell of the scroll compressor 100. The bottom of the closed casing 13 is an oil storage space 14 in which lubricating oil is stored. The lubricating oil stored in the oil storage space 14 is sucked up by the oil pump 9 provided at the lower end portion of the drive shaft 6, and is supplied to the drive shaft 6 and the sliding portions of the compression mechanism portion a. The oil supply path will be described later.

Hereinafter, each component will be described.

[ compression mechanism part A ]

The compression mechanism portion a has the following functions: the refrigerant sucked from the refrigerant suction pipe 15 is compressed, and is discharged to the outside of the closed casing 13 through a discharge port 4 and a refrigerant discharge pipe 16, which will be described later. The compression mechanism portion a is mainly constituted by the fixed scroll 1, the orbiting scroll 2, and the oldham coupling 25. The fixed scroll 1 includes a platen 1a and a spiral projection 1b provided on a lower surface of the platen 1a. The fixed scroll 1 is fixed to an upper end portion of a main frame 8a, and the main frame 8a is fixed to an inner peripheral surface of the sealed container 13. The fixed scroll 1 may be fixed by a fastening member such as a bolt.

The orbiting scroll 2 is also constituted by a platen 2a and a spiral projection 2b provided on the upper surface of the platen 2a, as in the fixed scroll 1. An eccentric hole 2c is formed near the center of the bottom surface of the platen 2a of the orbiting scroll 2, and the orbiting bearing 17 is pressed into the eccentric hole 2c. An eccentric shaft 6a provided at the upper end of the drive shaft 6 is slidably coupled to the rocking bearing 17. The oscillating scroll 2 performs an orbital motion with respect to the fixed scroll 1 via an oldham coupling 25 provided between the oscillating scroll 2 and the main frame 8a without performing a rotational motion. Further, an oscillating thrust bearing 18 is provided on the lower surface side of the oscillating scroll 2, that is, between the oscillating scroll 2 and the main frame 8a.

The fixed scroll 1 and the orbiting scroll 2 are provided in the sealed container 13 so that their swirl projections mesh with each other. The fixed scroll 1 and the orbiting scroll 2 mesh with each other to form a compression chamber 5 having a relatively variable volume. A suction port 3 for guiding the refrigerant sucked into the closed casing 13 from the refrigerant suction pipe 15 to the compression chamber 5 is formed in the outer peripheral portion of the compression chamber 5, and the refrigerant is sucked into the compression chamber 5 through the suction port 3. The refrigerant is compressed in the compression chamber 5, and the compressed refrigerant is discharged from a discharge port 4 formed in the central portion of the fixed scroll 1. The refrigerant discharged from the discharge port 4 is discharged to the outside of the sealed container 13 through the refrigerant discharge pipe 16.

[ Main frame 8a ]

The fixed scroll 1 is fixed to the upper end of the main frame 8a, and the orbiting scroll 2 is slidably supported from below via an orbiting thrust bearing 18. The main frame 8a is mounted in the sealed container 13 such that the outer peripheral surface thereof contacts the inner peripheral surface of the sealed container 13. A through hole for allowing the drive shaft 6 to pass therethrough is formed near the center of the main frame 8a, and a main bearing 19 for rotatably supporting a main shaft portion 6b of the drive shaft 6, which will be described later, is provided in the through hole. That is, the main frame 8a also has a function of rotatably supporting the drive shaft 6 via the main bearing 19.

[ auxiliary frame 8b ]

The sub-frame 8B is fixed to the inner surface of the side wall of the closed casing 13 below the motor unit B. The auxiliary frame 8b constitutes the housing 8 together with the main frame 8a. The sub-frame 8b includes a cylindrical portion 8ba and a flange portion 8bb extending outward from a lower end portion of the cylindrical portion 8ba.

Further, a through hole through which the drive shaft 6 passes is formed near the center of the sub-frame 8b, and a cylindrical sub-bearing 11 is provided in the through hole. The sub bearing 11 is a radial bearing that rotatably supports a sub shaft portion 6c, which will be described later, of the drive shaft 6 and supports a load in a radial direction. The sub-bearing 11 is formed of a sliding bearing generally called a "bushing (metal)".

The sub-frame 8b receives thrust load generated by the own weight of the drive shaft 6 and the magnetic force of the rotor, in addition to the radial load. This thrust load is received by an upper surface cover 9c of the oil pump 9, which will be described later, disposed below the sub-frame 8b, and an upper surface of the upper surface cover 9c serves as a thrust receiving surface 12.

That is, the scroll compressor 100 according to embodiment 1 is configured such that the radial load acting on the drive shaft 6 is supported by the sub-bearing 11 and the thrust load is received by the thrust receiving surface 12 provided on the upper surface cover 9c of the oil pump 9. The scroll compressor 100 according to embodiment 1 is characterized by a structure that maintains a good sliding state of the thrust receiving surface 12. This structure will be described later.

[ Motor section B ]

The motor section B has a function of driving the orbiting scroll 2 to compress the refrigerant in the compression mechanism section a. The motor unit B is disposed between the main frame 8a and the sub frame 8B. The motor portion B is mainly constituted by a motor 10 having a rotor 10a and a stator 10B. The rotor 10a is fixedly provided on the circumferential surface of the drive shaft 6, and is driven to rotate by starting energization to the stator 10b. The stator 10b is fixed to the inner peripheral surface of the sealed container 13 by shrink fitting or the like, surrounds the rotor 10a with a gap therebetween, and rotates the rotor 10a.

[ drive shaft 6]

The drive shaft 6 transmits the rotation of the rotor 10a of the motor B to the orbiting scroll 2 of the compression mechanism a. The drive shaft 6 includes, in order from above, an eccentric shaft 6a, a main shaft portion 6B fixed to a rotor 10a of the motor portion B, a sub shaft portion 6c, and a pump insertion shaft 6d having a diameter smaller than that of the sub shaft portion 6c. The eccentric shaft 6a is provided eccentrically with respect to the axial center of the drive shaft 6, and is slidably coupled to the rocking bearing 17 as described above. In the main shaft portion 6b, a balancer 26a is provided above the rotor 10a, and a balancer 26b is provided below the rotor 10a.

Further, the drive shaft 6 is formed with a vertical oil supply hole 7a extending in the axial direction at the center of the drive shaft 6, an eccentric shaft horizontal oil supply hole 7b extending in the radial direction while branching from the vertical oil supply hole 7a, a main shaft horizontal oil supply hole 7c, a radial horizontal oil supply hole 7d, and a thrust horizontal oil supply hole 7e. More specifically, an eccentric shaft oil supply lateral hole 7b is formed in the eccentric shaft 6a, a main shaft oil supply lateral hole 7c is formed in the main shaft portion 6b, a radial oil supply lateral hole 7d is formed in the auxiliary shaft portion 6c, and a thrust oil supply lateral hole 7e is formed in the pump insertion shaft 6d.

An axial oil groove 7f extending in the axial direction is formed in the outer peripheral surface of the auxiliary shaft portion 6c of the drive shaft 6, and the axial oil groove 7f communicates with the vertical oil supply hole 7a through a radial horizontal oil supply hole 7d.

The bottom surface of the sub shaft portion 6c is a hollow circular disk shape, and is a thrust surface 6f formed by an orthogonal surface extending radially outward from the upper end of the pump insertion shaft 6d. The weight and magnetic force of the drive shaft 6 act on the lower side of the axial center of the compressor, and the drive shaft 6 rotates while being pressed against the thrust receiving surface 12.

A radial oil groove 7g extending in the radial direction from the inner peripheral end to the outer peripheral end is formed in the thrust surface 6f. Further, the outer diameter side corner portion of the lower end of the auxiliary shaft portion 6c, in other words, the radially outer side end portion of the thrust surface 6f, is formed as a chamfered portion (fillet) 6e formed of a curved surface having a curvature radius Rs1 in order to relax the contact angle with the thrust receiving surface 12 during rotation. The circular arc range of the chamfered portion (fillet) 6 is formed in an angular range of less than 90 degrees.

[ hardness of Material of the auxiliary bearing 11 of the auxiliary shaft portion 6c ]

The main shaft portion 6b and the sub shaft portion 6c of the drive shaft 6 are normally hardened by quenching carbon steel as a base material. The portion indicated by a dot in fig. 2 (a) shows a quenched portion of the sub-shaft portion 6c. On the other hand, since the thrust surface 6f is used so as not to be hardened, it is necessary to keep a distance from the hardened portion of the auxiliary shaft portion 6c. A metal bush having excellent lubricity is generally used for the radial receiving surface 11a of the sub-bearing 11. The hard secondary shaft portion 6c is received by an aluminum alloy or a copper alloy on the inner peripheral surface of the radial receiving surface 11a that slides on the radial surface. On the other hand, the thrust receiving surface 12 is made of a hardened steel material having a fine surface roughness and subjected to quenching, and therefore does not wear even when receiving a load from the thrust surface 6f.

[ oil pump 9]

The oil pump 9 includes a movable portion 9a having an inner rotor 9aa and an outer rotor 9ab, a main body base 9b, and an upper surface cover 9c covering the movable portion 9a, and is fixed to the sub-frame 8b by screws 24 at a portion of the main body base 9b. The upper surface of the upper surface cover 9c serves as the thrust receiving surface 12 as described above, and the oil pump 9 is integrally fixed to the sub-frame 8b in order to maintain the accuracy of the right angle between the thrust receiving surface 12 and the sub-bearing 11 and sufficient rigidity.

The upper surface cover 9c is formed of an annular member, and has a through hole 23 formed in the center. A pump insertion shaft 6d of the drive shaft 6 is inserted into an upper end opening formed by the through hole 23 and a space continuing downward from the through hole 23. The periphery of the upper end opening of the through hole 23 is formed as a chamfered portion 12e formed by a curved surface having a curvature radius Rp0 so as not to contact the pump insertion shaft 6d. The upper surface cover 9c is disposed below the secondary shaft portion 6c, and corresponds to the outer shell member of the present invention. Further, the lower end opening 9d of the oil pump 9 is immersed in the lubricating oil in the oil storage space 14.

The oil pump 9 operates as the drive shaft 6 rotates, and sucks up the lubricating oil stored in the oil storage space 14 from the lower end opening 9d and supplies the lubricating oil to various portions of the compression mechanism a. The compression mechanism portion a corresponds to various bearings such as the rocking bearing 17, the rocking thrust bearing 18, the main bearing 19, the sub bearing 11, and the thrust receiving surface 12, and a sliding portion of the oldham coupling 25. The lubricating oil sucked up by the oil pump 9 is supplied from the oil supply vertical hole 7a to the rocking bearing 17 and the main bearing 19 via the eccentric shaft oil supply horizontal hole 7b and the main shaft oil supply horizontal hole 7c. Part of the lubricating oil flowing out of the upper outlet of the oil supply vertical hole 7a or the eccentric shaft oil supply horizontal hole 7b is supplied to the sliding portions of the rocking thrust bearing 18 and the oldham coupling 25. The lubricating oil sucked up by the oil pump 9 is supplied from the vertical oil feed hole 7a to the sub-bearing 11 through the radial horizontal oil feed hole 7d, and is supplied from the vertical oil feed hole 7a to the thrust receiving surface 12 through the thrust horizontal oil feed hole 7e. The supply of oil to the thrust receiving surface 12 is a characteristic portion of embodiment 1, and will be described again below.

[ thrust surface 6 f; thrust receiving surface 12

The thrust receiving surface 12 of the upper cover 9c is a flat surface 12b on the radially inner side, and a portion radially outward of the flat surface 12b is an inclined surface 12d that gradually inclines downward toward the outer side. The flat surface 12b is a surface that is at a right angle and has the same height as the radial receiving surface 11a that is the inner peripheral surface of the sub-bearing 11. The flat surface 12b of the thrust receiving surface 12 is continuously connected to the chamfered portion 12e of the radius of curvature Rp0 at the inner peripheral end, and is continuously connected to the inclined surface 12d at the flat surface end point 12c, forming a radius of curvature Rp 1. According to this configuration, a contact point 12f where the thrust surface 6f and the thrust receiving surface 12 contact each other in a curved surface and a slit 20 that increases continuously from the contact point 12f toward the radial outer side are formed between the thrust surface 6f and the thrust receiving surface 12.

Under the maximum load condition, the drive shaft 6 is tilted in the direction of the compression load from 1/1000 rad to 2/1000 rad. If the inclination angle of the inclined surface 12d is designed to be the same as the inclination level of the drive shaft 6 under the maximum load condition, the thrust surface 6f and the thrust receiving surface 12 can be rotated in a state of being gently in contact under the maximum load condition in which the operating condition is most severe, thereby alleviating the wear state.

The position of the contact point 12f that slides on the thrust receiving surface 12 is shifted in the radial direction from the inner peripheral side to the outer peripheral side as the secondary shaft portion 6c of the drive shaft 6 is inclined.

Here, the inclination of the drive shaft 6 during operation will be described.

1) First, in a state where the radial load acting on the drive shaft 6 is small and the thrust surface 6f is substantially parallel to the flat surface 12b of the thrust receiving surface 12, the thrust load acting on the drive shaft 6 is received by the entire flat surface 12b. The thrust load acting on the drive shaft 6 generally corresponds to the own weight of the drive shaft 6, the rotor 10a, the balancer 26a, and the balancer 26b, and the magnetic force in the thrust downward direction.

2) Next, when the compression operation is started, a radial load acts, the drive shaft 6 undergoes flexural deformation, the thrust surface 6f tilts to contact a curved surface (curvature radius Rp1) in the vicinity of the outer side of the flat surface end point 12c of the thrust receiving surface 12, and the drive shaft 6 rotates.

3) When the load further increases and the compression load and the centrifugal force increase, the drive shaft 6 undergoes flexural deformation, the thrust surface 6f further tilts, and the thrust surface 6f rotates in a state of gradually contacting the thrust receiving surface 12.

4) Finally, under the maximum load condition, when the drive shaft 6 is largely deflected and the thrust surface 6f is inclined from the flat surface 12b of the thrust receiving surface 12, the curved surface (curvature radius Rs1) of the chamfered portion 6e at the outer peripheral portion of the thrust surface 6f is rotated while being in contact with the inclined surface 12d of the thrust receiving surface 12 on one side. In this case, by increasing the curvature radius Rs1 of the chamfered portion 6e, the increase in local hertzian stress due to the contact of the chamfered portion 12e with one side of the thrust receiving surface 12 is suppressed.

For example, according to the equation of Hertz stress when a cylinder of radius curvature 1 and a cylinder of radius curvature 2 are in contact at a length L,

hertz stress ℃ ∈ (load) × (Young's modulus) × L × { 1/(radius of curvature 1)²+ (radius of curvature 2)²The establishment is carried out.

The radius of curvature Rs1 of the chamfered portion 6e is in the size relationship of Rp0 < Rs1 < Rp1, and Rs1 needs to be designed to be at least larger than Rp0 and close to Rp 1.

While the general R chamfer (fillet) is about R0.2 to R2, the radius of curvature Rs1 of the chamfered portion 6e is large within a design tolerance range, and is usually R3 or more.

In embodiment 1, the radial outer side of the thrust receiving surface 12 is formed as the inclined surface 12d and separated from the thrust surface 6f, and the sliding portion where the sub-shaft portion 6c and the upper surface cover 9c slide is not the entire thrust receiving surface 12 but is the flat surface 12b. Therefore, the outer circumferential position of the sliding portion can be made closer to the radially inner side than in the case where the entire thrust receiving surface 12 is made to be the flat surface 12b without forming the inclined surface 12d. Therefore, the maximum sliding speed can be suppressed.

After the running-in operation, the angle of the flat surface end point 12c, which is the boundary between the flat surface 12b and the inclined surface 12d, gradually disappears. Therefore, the contact angle between the thrust surface 6f and the thrust receiving surface 12 at the flat surface end point 12c is also reduced, and the contact surface pressure of the sliding portion can be reduced.

[ other structures ]

An oil return pipe 29 extending in the axial direction of the drive shaft 6 is disposed between the main frame 8a and the stator 10b, and an oil return hole 29a extending in the axial direction is formed in the stator 10b. The oil return pipe 29 and the oil return hole 29a have a function of returning the lubricating oil used in the compression mechanism portion a to the oil storage space 14. In fig. 1, the case where only one oil return pipe 29 and oil return hole 29a are provided is illustrated by way of example, but the present invention is not limited to this case. For example, two or more oil return pipes 29 and oil return holes 29a may be provided.

As described above, in the scroll compressor 100, the compression mechanism section a is disposed at the upper portion in the closed casing 13, the motor section B is disposed at the lower portion, and the driving force of the motor section B is transmitted to the orbiting scroll 2 of the compression mechanism section a via the drive shaft 6 to drive the orbiting scroll 2 to orbit. The type of the lubricating oil is not particularly limited as long as the lubricating oil can be used as the lubricating oil for the compression mechanism portion a. For example, PAG (polyalkylene glycol) or POE (polyol ester) can be used as the lubricating oil. The type of the refrigerant is not particularly limited.

[ description of operation ]

The operation of the scroll compressor 100 will be described below.

When the stator 10b constituting the motor 10 starts to be energized, the drive shaft 6 starts to rotate together with the rotor 10a. When the drive shaft 6 starts rotating, the orbiting scroll 2 connected to the eccentric shaft 6a revolves while being prevented from rotating by the oldham coupling 25. Thereby, the compression chamber 5 moves toward the center side while gradually decreasing in volume. As a result, the pressure of the refrigerant sucked into the compression chamber 5 from the suction port 3 gradually increases. The refrigerant having increased pressure is discharged to the outside of the machine through the discharge port 4 and the refrigerant discharge pipe 16, and is pressure-fed to a refrigerant pipe (not shown) connected to the refrigerant discharge pipe 16.

In this way, the refrigerant in the sealed container 13 is discharged to the outside, and therefore the inside of the sealed container 13 becomes negative pressure. Therefore, the refrigerant from the refrigerant pipe (not shown) outside the machine is sucked into the closed casing 13 through the refrigerant suction pipe 15. The refrigerant sucked into the closed casing 13 cools the motor 10 and is then sucked into the compression chamber 5 through the suction port 3.

Further, as the drive shaft 6 rotates, the inner rotor 9aa of the oil pump 9 rotates, and the outer rotor 9ab rotates, whereby the lubricating oil in the oil storage space 14 is sucked upward through the vertical oil supply hole 7a by the pumping action of the oil pump 9. The sucked lubricating oil is distributed to the sub-bearing 11, the main bearing 19, and the rocking bearing 17, respectively, to lubricate the bearings. The lubricating oil passed through the rocking bearing 17 is supplied to the rocking thrust bearing 18 and the oldham ring 25, thereby lubricating these sliding portions. Further, the lubricating oil supplied to the oldham coupling 25 is returned to the oil storage space 14 via the oil return pipe 29.

[ deformation of the cylindrical portion 8ba of the auxiliary frame 8b ]

The secondary bearing 11 supports radial loads generated by operation of the scroll compressor 100. The cylindrical portion 8ba of the sub-frame 8b provided with the sub-bearing 11 has a radial thickness smaller than that of the flange portion 8bb, and is a thin flexible structure portion. By forming the cylindrical portion 8ba as a thin flexible structure portion in this manner, the cylindrical portion 8ba is elastically deformed in accordance with the inclination of the drive shaft 6, and the drive shaft 6 can be prevented from coming into contact with the radial receiving surface 11a on one side. Since fig. 2 is a schematic view, the cylindrical portion 8ba is thinner than the flange portion 8bb, but is illustrated as being thick, and is actually thin so as to be elastically deformable.

[ radial load and thrust load ]

With the operation of the scroll compressor 100, a radial load acts on the auxiliary bearing 11. That is, the sub-bearing 11 is applied with a fluctuating load in synchronization with the rotation of the drive shaft 6 in the radial direction. The sub-bearing 11 is formed of a slide bearing as described above. Since the secondary bearing 11 does not use a ball bearing or the like in this manner, the long-term reliability is ensured, the seizure of the drive shaft 6 is prevented, the increase in wear and friction loss is suppressed, and stable startability is obtained.

In addition, as the scroll compressor 100 operates, a thrust load acts on the thrust receiving surface 12 of the upper surface cover 9c. That is, the self weight of the drive shaft 6 is applied to the thrust receiving surface 12 vertically downward as a thrust load. In order to maintain the thrust surface 6f of the sub-shaft portion 6c of the drive shaft 6 and the thrust receiving surface 12 of the upper surface cover 9c in a good sliding state, it is necessary to ensure a stable oil surface at all times on the thrust receiving surface 12.

Therefore, in embodiment 1, the inclined surface 12d is provided on the outer peripheral side of the thrust receiving surface 12, and the oil storage space 22 is formed between the bottom surface of the sub-bearing 11 and the inclined surface 12d, in the region radially outside the thrust receiving surface 12 and in the region further outside the thrust receiving surface 12. Since the oil flows into the oil storage space 22 from the oil seal portion 11 on the lower end side of the radial receiving surface 11a, the oil is constantly supplied from the outer peripheral side of the thrust receiving surface 12 to the sliding surface by the pressure head of the oil stored up to the upper end 22b. On the other hand, oil is constantly supplied from the thrust oil supply cross hole 7e of the pump insertion shaft 6d to the sliding surface from the inner peripheral side of the thrust receiving surface 12 through the radial oil groove 7g. Hereinafter, the oil supply path for ensuring a stable oil surface at the thrust receiving surface 12 will be described in detail. Here, the region radially outside the thrust receiving surface 12 is the inclined surface 12d in order to form a space as the oil storage space 22, but the inclined surface 12d is not limited thereto, and may be a flat surface formed at a position lower than the height position of the flat surface 12b.

[ oil supply path to the thrust receiving surface 12]

A part of the lubricating oil sucked up by the oil pump 9 flows into the outer peripheral side of the pump insertion shaft 6d and is discharged to the outside in the radial direction of the drive shaft 6 by the centrifugal force. That is, the lubricating oil that has flowed into the outer peripheral side of the pump insertion shaft 6d rises through the gap between the pump insertion shaft 6d and the movable portion 9a of the oil pump 9, and further rises through the gap between the pump insertion shaft 6d and the upper cover 9c. The lubricating oil that has risen in the gap between the pump insertion shaft 6d and the upper surface cover 9c reaches the thrust receiving surface 12. Part of the lubricating oil that has been pressurized by the oil pump 9 and has flowed into the oil feed vertical hole 7a is also discharged radially outward by centrifugal force from the thrust oil feed horizontal hole 7e formed in the pump insertion shaft 6d, and reaches the thrust receiving surface 12. Specifically, the thrust oil supply lateral hole 7e is formed in the pump insertion shaft 6d so as to communicate with the oil supply longitudinal hole 7a at a position lower than the thrust surface 6f, and extend in the radial direction of the drive shaft 6, and supplies the lubricating oil between the thrust surface 6f and the thrust receiving surface 12.

The lubricating oil that has reached the thrust receiving surface 12 flows from the inner peripheral end to the outer peripheral end through the radial oil grooves 7g formed in the thrust surface 6f, and flows into the oil storage space 22. The flow path cross-sectional area obtained by cutting the radial oil groove 7g in the axial direction is smaller than the flow path cross-sectional area obtained by cutting the annular gap between the pump insertion shaft 6d and the upper surface cover 9c in the direction perpendicular to the axial direction. Therefore, the lubricating oil in the annular gap between the pump insertion shaft 6d and the upper surface cover 9c flows intensively into the radial oil groove 7g, and the radial oil groove 7g is filled with the lubricating oil. Further, since the radial oil grooves 7g make one rotation around the circumferential direction by the rotation of the drive shaft 6, the lubricating oil spreads over the entire area of the thrust receiving surface 12.

Further, the lubricating oil supplied from the vertical oil supply hole 7a to the horizontal radial oil supply hole 7d is supplied from the axial oil groove 7f to the gap between the outer peripheral surface of the auxiliary shaft portion 6c and the radial receiving surface 11a. The lubricating oil supplied to the gap between the outer peripheral surface of the auxiliary shaft portion 6c and the radial receiving surface 11a is spread over the entire radial receiving surface 11a of the auxiliary bearing 11, thereby lubricating the auxiliary bearing 11.

Here, in the outer peripheral surface of the auxiliary shaft portion 6c, a region S1 above the axial oil groove 7f in the region facing the auxiliary bearing 11 is longer in axial length than a region S2 below. The gap between the region S1 and the sub bearing 11 and the gap between the region S2 and the sub bearing 11 become oil seal portions due to the stagnation of the lubricating oil. Since the region S1 is longer in axial length than the region S2, the axial length of the upper oil seal portion 11c is longer than the axial length of the lower oil seal portion 11d. Therefore, the amount of the lubricating oil supplied from the radial oil supply cross hole 7d to the axial oil groove 7f that flows out from the upper end side of the sub-bearing 11 and leaks into the closed casing 13 is reduced so that the lubricating oil is stored in the oil storage space 22. In this way, the thrust surface 6f and the thrust receiving surface 12 can be kept in a good sliding state by assuming a state in which a large amount of lubricating oil is present in the oil storing space 22.

As described above, the lubricating oil passing through the thrust-side horizontal port 7e from the vertical oil feed hole 7a and the lubricating oil passing through the radial-side horizontal port 7d from the vertical oil feed hole 7a merge in the oil reserving space 22. The lubricating oil merged in the oil storage space 22 is returned to the oil storage space 14 through an oil discharge flow path 21, and the oil discharge flow path 21 is constituted by a groove and a hole provided in the attachment surface of the sub-frame 8b to the oil pump 9. The oil discharge flow path 21 has an upstream end opening to the lower end 22a of the oil storage space 22, and a downstream end discharge hole 21a opening downward to the outer surface of the main body 9b of the oil pump 9. The discharge hole 21a is an outlet of the oil discharge flow path 21 and is located below the oil storage space 22. Then, the lubricating oil in the oil storing space 22 is discharged to the outside of the sub-frame 8b through the oil discharge flow path 21. In fig. 2 (a), reference numeral 22b denotes an upper end of the oil storage space 22, and the upper end 22b is disposed above the radial oil groove 7g.

Here, when the oil storage space 22 is newly defined, the sub-bearing 11 is provided inside the cylindrical portion 8ba of the sub-frame 8b as shown in fig. 2, whereby a cylindrical space is formed below the sub-bearing 11 by at least the thickness of the sub-bearing 11. This space becomes a part of the oil reserving space 22.

In fig. 2, a wall surface portion 22d recessed radially outward is formed in a region of the inner peripheral surface 22c of the cylindrical portion 8ba of the sub-frame 8b facing the sub-shaft portion 6c of the drive shaft 6 and located below the sub-bearing 11. In this way, the wall surface 22d is further provided on the inner peripheral surface 22c of the cylindrical portion 8ba of the sub-frame 8b, whereby the oil storing space 22 is further expanded. Further, by providing the wall surface portion 22d, the lower end portion of the sub shaft portion 6c of the drive shaft 6 that rotates while inclining is brought into contact with the thrust receiving surface 12 or the radial receiving surface 11a, so that the risk of damage to the sub shaft portion 6c, the thrust receiving surface 12, or the radial receiving surface 11a can be reduced.

Further, a throttle flow path 21b giving an appropriate flow path resistance is formed in the middle of the oil discharge flow path 21, and by forming the throttle flow path 21b, the lubricating oil can be temporarily stored in the oil storage space 22. When abrasion powder is generated on the radial receiving surface 11a or the thrust receiving surface 12, the abrasion powder needs to be discharged to the oil storage space 14 instead of being stored in the oil storage space 22. That is, it is necessary to discharge the abrasion powder from the oil storing space 22 through the oil discharge passage 21. Therefore, the throttle channel 21b preferably has a channel cross section whose depth and width are set to about 0.2mm to 1mm, respectively.

[ axial positional relationship between the radial oil groove 7g and the oil storing space 22 ]

The radial oil grooves 7g are formed in the thrust surface 6f as described above, but in order to maintain the thrust surface 6f and the thrust receiving surface 12 in a good oil lubrication state, it is necessary to fill the radial oil grooves 7g with the lubricating oil. If the upper surface of the lubricating oil stored in the oil storage space 22 is located below the bottom of the groove constituting the axial oil groove 7f, the radial oil groove 7g cannot be filled with the lubricating oil. Therefore, the upper surface of the oil stored in the oil storage space 22 is located above the bottom of the groove forming the axial oil groove 7f. Specifically, the volume of the oil storage space 22 and the throttle passage 21b may be designed in accordance with the relationship with the amount of oil flowing into the oil storage space 22.

[ circumferential arrangement position of radial oil groove 7g ]

The radial oil groove 7g is formed at one position on the thrust surface 6f, and the circumferential arrangement position thereof on the thrust surface 6f is set in consideration of the deflection direction of the drive shaft 6. The circumferential arrangement position of the radial oil groove 7g will be described below with reference to fig. 3 and 4. Before the description of fig. 3 and 4, first, the load acting on drive shaft 6 and the acting direction of the load will be described with reference to fig. 5.

In the Y axis, when the drive shaft 6 is viewed from the axial direction in plan, a direction (hereinafter referred to as an eccentric direction) in which the axial center of the eccentric shaft 6a is connected to the axial center of the drive shaft 6 is set as +. The Z axis is set to be + above the axial center of the drive shaft 6. The X axis is a coordinate axis orthogonal to the Y axis and the Z axis. The drive shaft 6 rotates counterclockwise in the θ (-X) direction when viewed from above, but receives a gas load in the + X direction on the opposite side. In fig. 5, Fx is a gas load required to compress the refrigerant gas in the compression chamber 5, and mainly acts in the + X direction. In addition, F_1xThe load in the X-axis direction acting on the main bearing 19. F_1yThe load in the Y-axis direction acting on the main bearing 19. F_2xIs a load in the X-axis direction acting on the sub-bearing 11. F_2yIs a load in the Y-axis direction acting on the sub-bearing 11.

In addition, the centrifugal force acts on the driving shaft 6Since the force is a moment that tilts the entire drive shaft 6 in the direction of eccentricity (+ Y), the balancer 26a and the balancer 26b are attached to the upper and lower sides of the rotor 10a so as to cancel the moment. F_BW1Is a load generated by a centrifugal force acting on the drive shaft 6 by the balancer 26a. F_BW2Is a load generated by a centrifugal force acting on the drive shaft 6 by the balancer 26b.

Fig. 3 is an explanatory view of a bent state of the drive shaft of the scroll compressor according to embodiment 1 of the present invention in the Z-X cross section and a position of the radial oil groove in the circumferential direction. In fig. 3, (a) shows a case where the gas load acting on the drive shaft 6 is in the low load operation, and (B) shows a case where the gas load is in the high load operation. In each of (a) and (B), (a) is a diagram showing a state of flexure of the drive shaft 6 in a Z-X cross section, (B) is a plan view of the drive shaft 6 as viewed from above and shows a load acting on the drive shaft 6, and (c) is an explanatory diagram showing a position in the circumferential direction of the radial oil groove 7g as viewed from below. In fig. 3, a double-line arrow in an arc shape indicates a rotation direction of the drive shaft 6.

A centrifugal force Fc acts on the drive shaft 6 in the + Y direction due to the eccentric rotation of the orbiting scroll 2 and the like. However, when the structure is adopted in which the centrifugal force Fc is received by the swirl projection (tip scroll portion) 1b of the fixed scroll 1, the centrifugal force Fc does not act on the drive shaft 6. Here, first, in the Z-X cross section where the centrifugal force can be ignored, a gas load Fx required to compress the refrigerant gas in the compression chamber 5 acts as a main load on the drive shaft 6 in the + X direction, and a reaction force in balance with the main shaft portion 6b and the auxiliary shaft portion 6c acts thereon. Between the main bearing 19 and the main shaft portion 6b, a diameter gap (normally, the right side of the shaft diameter 1/1000) is provided between the sub bearing 11 and the sub shaft portion 6c, respectively, which can be used for fluid lubrication. Therefore, the drive shaft 6 is flexurally deformed, and the shaft center 6h of the sub shaft portion 6c is also inclined with respect to the shaft center (Z-axis) 6g of the compressor so that the main shaft portion 6b approaches in the + X direction at the main bearing 19 and the sub shaft portion 6c approaches in the-X direction at the sub bearing 11 while the shaft is entirely inclined. As a result, during the low load operation in fig. 3 (a), the thrust surface 6f floats in the X-axis negative direction with respect to the thrust receiving surface 12 and is inclined so as to press the X-axis positive direction.

During high load operation, as shown in fig. 3 (B), the drive shaft 6 is more bent and deformed than during low load operation. Therefore, the thrust surface 6f of the sub shaft portion 6c is inclined with respect to the thrust receiving surface 12 so as to float in the + direction of the X axis and press the-direction of the X axis. When the radial oil groove 7g is pressed against the thrust receiving surface 12, the surface pressure locally increases in the vicinity of the edge of the radial oil groove 7g, and there is a problem that the radial oil groove is easily worn. Therefore, the radial oil grooves 7g may be arranged so as to avoid the ± X directions so that the vicinity of the radial oil grooves 7g is not in contact with the thrust receiving surface 12 on one side. On the other hand, the axial oil groove 7f is provided on the + X side (counter load side) because the sub shaft portion 6c is close to the-X direction on the radial receiving surface 11a of the sub bearing 11 and a load acts thereon.

Next, fig. 4 is an explanatory diagram of a state of deflection in the Y-Z cross section and a position of arrangement in the circumferential direction of the radial oil groove of the drive shaft of the scroll compressor according to embodiment 1 of the present invention. In fig. 4, (a) shows a low-speed operation, and (B) shows a high-speed operation. In each of (a) and (B), (a) is a diagram showing a state of flexure of the drive shaft 6 in the Y-Z cross section, (B) is a diagram showing a plan view of the drive shaft 6 as viewed from above and showing a load acting on the drive shaft 6, and (c) is an explanatory diagram showing a position in the circumferential direction of the radial oil groove 7g as viewed from below. In fig. 4, a double-line arrow in an arc shape indicates a rotation direction of the drive shaft 6.

When a centrifugal force Fc acts on the drive shaft 6 due to the eccentric rotation of the orbiting scroll 2 or the like, a large load locally acts on the spiral projection (tip spiral portion) 1b or the like of the fixed scroll 1, or the rotor oscillates due to the inclination of the drive shaft, or the like. Therefore, as a mechanism for relaxing the load due to the centrifugal force Fc, a balancer 26a and a balancer 26b are attached to the upper and lower sides of the rotor 10a, and the forces and moments acting on the Y-Z section of the drive shaft 6 are balanced with each other. At this time, the drive shaft 6 inclined while being deformed is inclined, but the auxiliary shaft portion 6c is inclined in substantially the same tendency, although the inclination angle is changed by the presence or absence of the centrifugal force Fc.

Centrifugal force F of balancer 26a on upper side of rotor 10a_BW1Centrifugal force F of the lower balancer 26b acting in the-Y direction_BW2Since the radial bearing surface 11a of the sub-bearing 11 acts in the-Y direction, the radial bearing surface is inclined so as to be closer to the-Y direction below the sub-shaft portion 6c. As a result, in both the low-speed operation in fig. 4 (a) and the high-speed operation in fig. 4 (B), the thrust surface 6f is inclined so as to float in the + direction of the Y axis and press the-direction of the Y axis against the thrust receiving surface 12. Therefore, it is preferable to arrange the radial oil grooves 7g so as not to contact the thrust receiving surface 12 on one side, so as to avoid the-Y direction.

As described above with reference to fig. 3 and 4, although the thrust surface 6f is inclined by applying a gas load and a centrifugal force, which are main loads, to the drive shaft 6, the radial oil groove 7g is hard to be in one-side contact with the Y-axis + direction, that is, the eccentric direction, under all operating conditions. Therefore, the radial oil groove 7g is preferably arranged in the eccentricity direction (+ Y direction).

Under the most severe high-speed high-load operating conditions, as shown in fig. 3 (B) and 4 (B), the centrifugal force F of the balancer 26a on the upper side of the rotor is the gas load Fx, which is the main load, and_BW1centrifugal force F of balancer 26b on the lower side of rotor acting in-Y direction all the time_BW2Always acting in the + Y direction. Gas load Fx and centrifugal force F_BW1The resultant force Fr acts in an angular direction slightly shifted from the + X axis in the clockwise direction (counter-rotational direction) when viewed from above, and the drive shaft 6 is deflected and deformed, and therefore the angle at which the thrust surface 6f contacts on one side and the direction of the radial load on the counter shaft also slightly shift, and in consideration of the above, when the position of the radial oil groove 7g of the thrust surface 6f is shifted by an angle α from the + direction (eccentric direction) of the Y axis to the counterclockwise direction (counter-rotational direction) when viewed from below, as shown in fig. 3 (B) and fig. 4 (B), higher durability can be obtained, and the angle α is an acute angle range of 0deg to 45 deg.

(Effect of the invention)

According to embodiment 1 described above, the gap 20 that continuously increases from the contact point 12f toward the radially outer side is provided between the thrust surface 6f and the thrust receiving surface 12. Further, the structure is as follows: an oil storage space 22 is provided on the outer peripheral side of the thrust surface 6f, and after the lubricating oil is supplied to the thrust receiving surface 12 through the radial oil grooves 7g and the slits 20, the oil is stored in the oil storage space 22. By providing the oil storage space 22 in this way, the flow of the lubricating oil between the thrust surface 6f and the thrust receiving surface 12 can be ensured at all times. As a result, a good oil lubrication state and a good sliding state can be ensured in the thrust receiving surface 12, and abnormal wear of the drive shaft 6, seizure of the drive shaft 6, and the like can be suppressed. As described above, in embodiment 1, the life and reliability of the compressor can be improved by a relatively simple means of providing the gap and the oil storing space 22, which become larger toward the radial outer side, between the thrust surface 6f and the thrust receiving surface 12.

Further, by providing the throttle flow path 21b in the middle of the oil discharge flow path 21, the lubricating oil can be easily temporarily held in the oil storing space 22, and it is effective in obtaining good oil lubrication of the thrust receiving surface 12.

Further, as a specific configuration of the oil storing space 22, a portion of the thrust receiving surface 12 radially outward of the radially inward side may be formed as the inclined surface 12d inclined downward toward the outward side, and thus, the configuration can be simplified.

Further, the oil storing space 22 is present radially outside the thrust receiving surface 12, so that the radial position of the sliding portion where the thrust surface 6f and the thrust receiving surface 12 contact each other can be located radially inward. Therefore, when the high-speed operation is performed under a high load, which is a severe sliding condition in which the inclination angle of the drive shaft 6 becomes large, the sliding speed at the flat surface end point 12c, which is a contact point, can be reduced. Alternatively, the contact angle can be suppressed to be small.

Further, since the periphery of the upper surface opening of the through hole 23 of the upper surface cover 9c is formed as the chamfered portion 12e formed by a curved surface, it is possible to avoid contact with the pump insertion shaft 6d at the time of rotation of the drive shaft 6.

Further, since the radially outer end portion of the thrust surface 6f of the drive shaft 6 is formed as the chamfered portion 6e formed by a curved surface, the contact angle with the thrust receiving surface 12 can be relaxed, and a favorable sliding state can be maintained.

The compressor is not limited to the structure shown in fig. 1 to 5, and various modifications can be made without departing from the scope of the present invention, for example, as follows.

< modification 1 of embodiment 1 >

Fig. 6 is a diagram showing a modification 1 of the scroll compressor according to embodiment 1 of the present invention, and is a schematic diagram showing a portion of the scroll compressor closer to a lower end than the sub bearing in an enlarged manner.

The modification 1 is different from the basic structure shown in fig. 2 in that the thrust oil supply cross hole 7e of the pump insertion shaft 6d is eliminated. The thrust receiving surface 12 is also in a state after hardening by quenching, and the distance between the thrust receiving surface 12 and the quenched portion of the auxiliary shaft portion 6c (radial surface) is closer than in the basic structure shown in fig. 2. Further, the space of the oil storing space 22 is also different from the basic structure shown in fig. 2 in that the volume is reduced because the inner wall surface 22ca is not recessed as in fig. 2.

Under high-speed and high-load operating conditions, the thrust surface 6f is greatly inclined and slides while making contact with a chamfered portion (fillet) 6e having a large radius of curvature. Even when the thrust oil supply cross hole 7e is removed from the pump insertion shaft 6d in this manner, oil flows into the oil storage space 22 from the oil seal portion 11d on the lower end side of the radial receiving surface 11a. Therefore, a large amount of oil that is always stable can be supplied to the outside of the thrust receiving surface 12 by the pressure head of the oil stored up to the upper end 22b. On the other hand, the lubricating oil leaking from the oil pump 9 can be supplied from the gap between the pump insertion shaft 6d and the through hole 23 to the chamfered portion 6e from the inner peripheral side of the thrust receiving surface 12 through the radial oil groove 7g. However, the amount of oil supply due to the latter oil pump leakage is reduced as the number of revolutions is lower and the pressure head is smaller, and therefore, there is a limit to the low-speed possible range.

As described with reference to fig. 2, the configuration in which the thrust oil supply cross hole 7e is provided in the pump insertion shaft 6d is effective for supplying a large amount of lubricating oil to the thrust receiving surface 12 constantly and stably via the radial oil groove 7g of the thrust surface 6f by the centrifugal pump. However, in addition to the problem in the low speed operation, the same effect as that in the present modification 1 can be obtained.

< modification 2 of embodiment 1 >

Fig. 7 is a diagram showing a modification 2 of the scroll compressor according to embodiment 1 of the present invention, and is a schematic diagram showing a portion of the scroll compressor closer to the lower end than the sub-bearing in an enlarged manner.

Modification 2 differs from the basic structure shown in fig. 2 in the following two points. That is, the flat surface 12b is extended to the end of the radial outside without the inclined surface 12d on the thrust receiving surface 12 side. On the other hand, the oil reserving space 22 is formed so that the radius of curvature Rs1 of the chamfered portion 6e at the radially outer end of the thrust surface 6f is larger than that in fig. 2. The thrust receiving surface 12 is also in a state after hardening by quenching, and the distance between the thrust receiving surface 12 and the quenched portion of the auxiliary shaft portion 6c is close, and the space of the oil storing space 22 is also smaller than that in fig. 2.

In fig. 2, a flat surface end point 12c of the thrust receiving surface 12 is a contact point between the thrust receiving surface 12 and the thrust surface 6f. In contrast, in modification 2, the entire thrust receiving surface 12 is made flat, so that when the drive shaft 6 is inclined, the point on the thrust receiving surface 12 that comes into contact with the chamfering start position 6ea of the chamfered portion 6e becomes the contact point 12f. Therefore, by increasing the radius Rs of the chamfered portion 6e, the contact point 12f is located radially inward of the thrust receiving surface 12 than when the radius Rs is small, and therefore the sliding speed at the contact point 12f can be reduced. Further, by increasing the radius Rs of the chamfered portion 6e, the contact angle can be reduced as compared with the case where the radius Rs is smaller even if the entire thrust receiving surface 12 is the flat surface 12b.

In modification 2, a circle (hereinafter, contact sliding circle) having a radius of a line segment that descends perpendicularly from the contact point 12f to the axis of the drive shaft 6 is configured to be larger than the movable portion 9a as viewed in the axial direction. With this configuration, the following effects can be obtained. That is, the upper surface cover 9c is in a so-called cantilever shape in which the outer peripheral portion is supported by the main body base 9b and the inner peripheral portion is not supported but is in a floating state. Therefore, if the contact sliding circle is located inward of the movable portion 9a of the oil pump 9, the inner peripheral portion side of the upper surface cover 9c is pressed down by the drive shaft 6, and the upper surface cover 9c may be deflected. However, since the contact sliding circle is larger than the movable portion 9a, the upper surface cover 9c can be prevented from being bent.

< modification 3 of embodiment 1 >

Fig. 8 is a diagram showing a modification 3 of the scroll compressor according to embodiment 1 of the present invention. In fig. 8, (a) is an enlarged schematic view of a portion of the scroll compressor closer to the lower end than the sub-bearing, and (b) is a view of the annular steel plate as the thrust receiving surface in modification 3, as viewed from directly above.

Modification 3 differs from modification 2 in the following points. The thrust receiving surface 12 is different from the annular steel plate 12g having a small thickness and a small surface roughness. The annular steel plate 12g is placed on the upper surface cover 9c of the oil pump 9. Here, "thin" means, for example, a thickness of 1mm or less, and "thin surface roughness" means, for example, a surface roughness of z1 or less. For example, a commercially available quenched thin steel sheet such as a PK steel sheet having a thickness of 0.5mm can be used for the annular steel sheet 12g. The annular steel sheet 12g corresponds to an example of the annular member of the present invention, and is finished by polishing the surface of the quenched steel strip.

The annular steel plate 12g has an outwardly projecting lug-shaped protrusion 12ga at a portion of the circular outer periphery, and is fixed so as not to rotate by inserting the lug-shaped protrusion 12ga into a cutout of the main body base 9b of the oil pump 9.

In modification 3, the use of the annular steel plate 12g improves the sliding property because the annular steel plate 12g follows the behavior of the thrust surface 6f. Further, since a commercially available quenched thin steel plate is used as the annular steel plate 12g, there is an advantage that the processing can be easily completed.

Embodiment mode 2

The basic configuration of embodiment 2 differs from that shown in fig. 2 in that: the formation position of the radial oil groove 7g is replaced from the thrust surface 6f side to the thrust receiving surface 12 side, and the number of the radial oil grooves 7g is increased to a plurality, preferably 3 or more. The thrust receiving surface 12 side is formed only by a flat surface 12b orthogonal to the axial center 6g of the compressor, and does not have the inclined surface 12d. Instead, the thrust surface 6f side is different in point by an inclined surface 6fa inclined from an orthogonal surface orthogonal to the shaft center 6g (shaft center reference of the compressor). The thrust surface 6f is inclined upward as going radially outward from the axial center 6g. By designing the inclination angle of the thrust surface 6f to be the same as the inclination angle θ s of the axial center 6h of the secondary shaft portion 6c that flexes under high load and high speed operation, the contact angle between the thrust surface 6f and the thrust receiving surface 12 can be kept small. Wherein the "inclination angle of the shaft center 6g deflected at high load and high speed operation" is about 1/1000 to 2/1000 rad.

As shown in fig. 2, in the configuration in which the radial oil groove 7g is formed in the thrust surface 6f, the rotation of the drive shaft 6 causes the radial oil groove 7g to rotate, thereby generating a centrifugal pump action, and the lubricating oil is diffused over the entire thrust receiving surface 12. In contrast, in the configuration in which the radial oil grooves 7g are formed on the thrust receiving surface 12 as in embodiment 2, since the positions of the radial oil grooves 7g are not rotated, the lubricating oil in the radial oil grooves 7g is less likely to flow and is less likely to spread on the thrust receiving surface 12. For this reason, in embodiment 2, the number of radial oil grooves 7g is increased to a plurality, so that the flow path resistance can be reduced, thereby obtaining an effect of spreading the lubricating oil over the entire thrust receiving surface 12.

However, the radial oil groove 7g has a flow path cross-sectional area smaller than that of a flow path formed by an annular gap between the pump insertion shaft 6d and the upper surface cover 9c. This allows the lubricating oil flowing through the radial oil grooves 7g to overflow to the upper surface side of the thrust receiving surface 12, and to spread over the entire surface by the rotation of the thrust surface 6f, thereby maintaining a satisfactory lubricating state.

Further, by providing the inclined surface 6fa, the contact point 12f is located radially inward as compared with a case where the inclined surface 6fa is not provided and both the thrust surface 6f and the thrust receiving surface 12 are flat surfaces. Therefore, reduction in the sliding speed at the contact point 12f and alleviation of the contact angle at the contact point 12f can be achieved.

The formation position of the radial oil groove 7g is also set to the thrust surface 6f as described in fig. 2, and a large amount of lubricant can be supplied to the thrust receiving surface 12 by the centrifugal force. However, even if the position where the radial oil groove 7g is formed is the thrust receiving surface 12 as in embodiment 2, the radial oil groove 7g can be designed appropriately, and the same effect as that of embodiment 1 can be obtained.

In addition, at least one of both end portions of the radial receiving surface 11a of the sub-bearing 11 in the axial direction of the drive shaft 6 is formed of a curved surface. This can suppress damage caused by contact between the end portion formed by the curved surface and the drive shaft 6. Fig. 9 shows a structure in which both end portions of the radial receiving surface 11a of the sub-bearing 11 in the axial direction of the drive shaft 6 are formed of curved surfaces.

Embodiment 3

The sub-frame 8b receives thrust load generated by the self-weight of the drive shaft 6 and the rotor magnetic force in addition to the radial load. In fig. 2, the thrust load is received by the upper surface cover 9c of the oil pump 9 disposed below the sub-frame 8b, and the upper surface of the upper surface cover 9c serves as a thrust receiving surface 12. In embodiment 3, the difference is that the thrust load is received by the bottom plate 8bc fixed to the bottom of the sub-frame 8b by the bolts 40, and the upper surface of the bottom plate 8bc becomes the thrust receiving surface 12. The shape of the thrust receiving surface 12 is substantially the same as that of fig. 2. The shape of the thrust surface 6f is also the same as that of fig. 2.

The bottom plate 8bc is an annular circular platen having a through hole formed in the center, and the outer diameter of the bottom plate 8bc is larger than the outer diameter of the cylindrical portion 8ba of the sub-frame 8b. The bottom plate 8bc is provided with a plurality of screw holes, and the bottom plate 8bc is fixed to the auxiliary frame 8b by screwing bolts 40 into the screw holes provided in the auxiliary frame 8b through the screw holes. Further, the length of the pump insertion shaft 6d is greater than that of fig. 2 by the thickness of the bottom plate 8bc. The bottom plate 8bc corresponds to an example of the ring member of the present invention.

According to embodiment 3, the thrust receiving surface 12 is formed by the bottom plate 8bc that is separate from the upper surface cover 9c of the oil pump 9, and therefore the material and thickness of the bottom plate 8bc can be selected without being restricted by the design of the oil pump 9. Therefore, the surface hardness and the bending strength of the thrust receiving surface 12 can be easily enhanced without being restricted by the design of the oil pump 9.

Further, the characteristic configurations of embodiment 1, modification 1 to modification 3, embodiment 2, and embodiment 3 may be appropriately combined without departing from the scope of the present invention. For example, the structure in which the plurality of radial oil grooves 7g are provided on the thrust receiving surface 12, which is the characteristic configuration of embodiment 2, and the structure in which the inclined surface 6fa is provided on the thrust surface 6f are used in combination, but not limited thereto. For example, the inclined surface 12d may be provided on the thrust receiving surface 12 as shown in fig. 2. Further, the configuration of enlarging the radius of curvature Rs1 of the chamfered portion 6e in modification 2 may be combined with embodiment 2. The annular steel plate 12g of modification 3 may be combined with

embodiments

2 and 3.

In embodiments 1 to 3, the sub-bearing 11 and the sub-frame 8b are separately formed, but may be integrally formed, and an integrated component may be used as the sub-bearing that supports the radial surface of the sub-shaft portion 6c of the drive shaft 6.

In embodiments 1 to 3, the member forming the thrust receiving surface is the upper surface cover 9c of the oil pump 9, the annular steel plate 12g, or the bottom plate 8bc fixed to the bottom of the sub-frame 8b, and is configured separately from the sub-bearing 11 forming the radial receiving surface 11a. However, in addition to integrating the sub-bearing 11 with the sub-frame 8b as described above, it is also possible to integrate a member forming the thrust receiving surface and use the integrated component as a sub-bearing that supports the radial surface and the thrust surface of the sub-shaft portion 6c of the drive shaft 6.

In the above description, the configuration in which the compressor is a scroll compressor has been described, but the present invention is not limited to this, and other types of compressors such as a rotary compressor may be used.

Description of the reference numerals

A fixed scroll; a bedplate; a vortex projection; an oscillating scroll; a bedplate; vortex projections (addendum vortex portions); an eccentric bore; a suction inlet; a discharge port; a compression chamber; a drive shaft; 6a.. eccentric shaft; a main shaft portion; 6c.. the secondary shaft portion; a pump insertion shaft; chamfering; 6ea.. chamfer start position; a thrust surface; 6fa... inclined face; an axial center of the compressor; the axis of the auxiliary shaft part; oil supply longitudinal holes; 7b.. oil supply cross holes; 7c. oil supply cross holes; 7d.. oil supply cross holes; oil supply cross holes; an axial oil groove; a radial oil groove; a housing; a main frame; an auxiliary frame; 8ba.. cylindrical portion (secondary bearing pad); 8bb. A bottom plate; an oil pump; a movable portion; 9aa. 9ab.. an outer rotor; a body substrate; 9c.. upper surface mask; 9d.. lower end opening; 10.. a motor; a rotor; a stator; a secondary bearing; a radial bearing surface; an oil seal; an oil seal; a thrust receiving surface; a flat face; a flat face termination point; an inclined surface; chamfering; a contact point; an annular steel plate; 12ga.. raised portion; sealing the container; an oil storage space; a refrigerant suction tube; a refrigerant discharge tube; a wobble bearing; oscillating a thrust bearing; a main bearing; an oil drain flow path; a discharge hole; throttling a flow path; an oil holding space; a lower end; an upper end; inner peripheral surface; 22ca... wall face portion; a through hole; a screw; an oldham coupling; a balancer; a balancer; an oil return pipe; an oil return hole; a bolt; a scroll compressor; a compression mechanism portion; a motor portion.

Claims

1. A compressor, wherein,

the compressor is provided with:

a compression mechanism unit disposed in the closed container;

a motor unit that drives the compression mechanism unit;

a drive shaft that transmits a driving force of the motor unit to the compression mechanism unit;

a main bearing supporting an upper portion of the drive shaft;

a sub-bearing supporting a lower portion of the drive shaft;

a radial bearing surface provided on the sub-bearing and slidably supporting a radial surface of the drive shaft; and

a thrust receiving surface that slidably supports a thrust surface of the drive shaft,

between the thrust surface and the thrust receiving surface: a contact point at which the thrust surface is brought into contact with the thrust receiving surface in a curved surface, and a slit that continuously increases from the contact point toward a radially outer side.

2. The compressor of claim 1,

an oil storing space for storing lubricating oil is formed radially outside the thrust receiving surface.

3. The compressor of claim 2,

the hermetic container has an oil storage space inside,

and an oil pump connected to a pump insertion shaft formed on the drive shaft at a lower portion of the sub-bearing, for sucking up the lubricating oil in the oil storage space by rotation of the drive shaft,

a radial oil groove extending in a radial direction of the drive shaft is provided between the thrust surface and the thrust receiving surface,

the lubricating oil sucked up by the oil pump and passing through the radial oil groove is stored in the oil storage space.

4. The compressor of claim 3,

the compressor includes an auxiliary frame disposed in the hermetic container and provided with the auxiliary bearing,

the oil pump is attached to the sub-frame, an oil discharge flow path for discharging the lubricating oil in the oil storage space from an outlet to the oil storage space is formed by a groove and a hole formed in one or both of attachment surfaces of the sub-frame and the oil pump, and a throttle flow path serving as a flow path resistance is formed in the middle of the oil discharge flow path.

5. The compressor of claim 4,

the outlet of the oil discharge flow path is disposed below the oil storage space, and the upper end of the oil storage space is disposed above the radial oil groove.

6. A compressor according to any one of claims 3 to 5,

the drive shaft is provided at an upper end thereof with an eccentric shaft eccentric from an axial center of the drive shaft,

the radial oil groove is formed to extend in an eccentric direction, which is a direction connecting the axial center of the eccentric shaft to the axial center of the drive shaft.

7. The compressor of claim 6,

the radial oil groove is disposed on the thrust surface of the drive shaft in a range of 0deg to 45deg in a rotational direction of the drive shaft from the eccentric direction.

8. A compressor according to any one of claims 3 to 7,

an annular member having a through hole in a central portion into which the pump insertion shaft is inserted is disposed below the thrust surface, and the thrust receiving surface is formed on an upper surface side of the annular member.

9. The compressor of claim 8,

the annular member is a profile member of the oil pump.

10. The compressor of claim 8,

the annular member is a bottom plate fixed to a lower side of the auxiliary frame on which the auxiliary bearing is provided.

11. The compressor according to any one of claims 8 to 10,

the radial outer side of the thrust receiving surface of the annular member is formed by an inclined surface inclined downward as it faces outward, thereby forming the oil storage space.

12. The compressor according to any one of claims 8 to 11,

an inner peripheral portion of the thrust receiving surface on the upper surface side of the annular member, which is formed by the through hole, is a chamfered portion having a curvature radius Rp 0.

13. The compressor of claim 8,

the ring-shaped member is a member obtained by grinding the surface of a quenched steel strip.

14. The compressor of any one of claims 3 to 13,

the plurality of radial oil grooves are formed on the thrust receiving surface.

15. The compressor according to any one of claims 2 to 6,

the oil storage space is formed by an inclined surface inclined upward as the thrust surface of the drive shaft is directed outward in the radial direction.

16. The compressor of any one of claims 1 to 15,

a chamfered portion that comes into contact with the thrust receiving surface at the contact point is formed in a radially outer peripheral portion of the thrust surface of the drive shaft, and a curvature radius Rs1 of the chamfered portion is formed in a circular arc angle range of less than 90 degrees.

17. The compressor of claim 16,

the radius of curvature Rs1 of the chamfered portion at the radially outer peripheral portion of the thrust surface is larger than the radius of curvature Rp0 at the inner peripheral side of the thrust receiving surface, that is, Rs1 > Rp0 is provided.

18. The compressor of any one of claims 1 to 17,

the drive shaft is provided with:

an oil supply longitudinal hole extending in the axial direction of the drive shaft, through which lubricating oil flows; and

and a thrust oil supply lateral hole that communicates with the oil supply longitudinal hole at a position lower than the thrust surface and extends in a radial direction of the drive shaft, and that supplies the lubricating oil between the thrust surface and the thrust receiving surface.

19. The compressor of claim 18,

the drive shaft is provided with:

a radial oil supply lateral hole that communicates with the oil supply longitudinal hole at a position above the thrust surface, extends in a radial direction of the drive shaft, and supplies the lubricating oil between the radial surface and the radial receiving surface; and

an axial oil groove which is communicated with the radial oil supply transverse hole, is formed on the outer peripheral surface of the driving shaft, and extends along the axial direction of the driving shaft,

the axial oil groove faces the radial receiving surface, and a length of the drive shaft in the axial direction is longer in a region of the outer peripheral surface of the drive shaft facing the radial receiving surface, the region being located above the axial oil groove than in a region located below the axial oil groove.

20. A compressor, wherein,

the compressor is provided with:

a compression mechanism unit disposed in the closed container;

a motor unit that drives the compression mechanism unit;

a main bearing supporting an upper portion of the drive shaft;

a sub-bearing supporting a lower portion of the drive shaft;

a thrust receiving surface that slidably supports a thrust surface of the drive shaft; and

a radial bearing surface that slidably supports a radial surface of the drive shaft,

the drive shaft is provided with:

an oil supply longitudinal hole extending in the axial direction of the drive shaft, through which the lubricating oil flows; and