WO2019171427A1

WO2019171427A1 - Compressor

Info

Publication number: WO2019171427A1
Application number: PCT/JP2018/008315
Authority: WO
Inventors: 哲英横山; 岩崎　俊明; 大輔堀口; 伊藤　慎一; 政哉岡本; 修平小山
Original assignee: 三菱電機株式会社
Priority date: 2018-03-05
Filing date: 2018-03-05
Publication date: 2019-09-12
Also published as: JPWO2019171427A1; CN111788394A

Abstract

This compressor is provided with: a compression mechanism section disposed within a hermetic container; an electric motor section which drives the compression mechanism section; a drive shaft which transmits the drive force of the electric motor section to the compression mechanism section; a primary bearing which supports the upper part of the drive shaft; a secondary bearing which supports the lower part of the drive shaft; a radial receiving surface which is provided on the secondary bearing and which supports the radial surface of the drive shaft in a slidable manner; and a thrust receiving surface which supports the thrust surface of the drive shaft in a slidable manner. The compressor has, between the thrust surface and the thrust receiving surface: a contact point where the thrust surface and the thrust receiving surface are in contact with each other through curved surfaces; and a gap which continuously increases radially outward from the contact point.

Description

Compressor

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

One of the components of a refrigeration cycle apparatus such as an air conditioner, a compression mechanism section having an orbiting scroll and a fixed scroll that mesh with each other, an electric motor section that drives the compression mechanism section, and a driving force of the electric motor section is a compression mechanism section There is a scroll compressor provided with a drive shaft for transmitting to the motor. The drive shaft is rotatably supported by a main bearing and a sub bearing provided above and below the motor unit. When the drive shaft is rotationally driven by the electric motor portion, the orbiting scroll installed on the eccentric shaft portion at the upper end portion of the drive shaft revolves. Thereby, the refrigerant is compressed in a compression chamber provided between the swing scroll and the fixed scroll in the compression mechanism section. When the refrigerant is compressed by the compression mechanism portion, a radial gas load acts on the drive shaft, and this gas load is supported by the main bearing and the sub-bearing. The auxiliary bearing supports the rotation of the drive shaft and also supports the weight of the drive shaft in the vertically downward direction.

In such a scroll compressor, a ball bearing is used as a secondary bearing for the purpose of supporting both a radial load (hereinafter referred to as a radial load) and a thrust load (hereinafter referred to as a thrust load) simultaneously. (See, for example, Patent Document 1).

However, ball bearings are expensive and have various problems such as poor long-term reliability because the inner ring and the rolling ball and the outer ring and the rolling ball contact each other at points to support the load. As a countermeasure, there is a scroll compressor in which a slide bearing is applied to the auxiliary bearing (see, for example, Patent Document 2).

In the scroll compressor described in Patent Document 2, a secondary bearing, which is a slide bearing, is configured by a radial bearing and a thrust bearing that individually support a radial load and a thrust load. And the thrust surface provided in the lower end part of the drive shaft is received by the thrust receiving surface provided in the auxiliary bearing, and the radial surface provided in the auxiliary shaft part of the drive shaft is received by the radial receiving surface provided in the auxiliary bearing. .

Japanese Patent Laid-Open No. 04-241786 Japanese Patent No. 4356375

Generally, a compressive load and a centrifugal force act on the drive shaft of a scroll compressor during operation, and a large load acts in the radial direction. For this reason, the shaft center of the drive shaft bends while being inclined with respect to the central axis of the compressor. The central axis of the compressor refers to an axis extending in the vertical direction. The main shaft portion and the sub shaft portion of the drive shaft rotate while being inclined with respect to the main bearing and the sub bearing, respectively. If a ball bearing is adopted as the secondary bearing of the scroll compressor as in Patent Document 1, the inclination of the drive shaft is absorbed by the clearance of the rolling ball and the inner ring and the clearance of the rolling ball and the outer ring. It is easy to secure parallelism between the secondary bearing and the secondary bearing. However, the ball bearing is inferior in terms of cost and long-term reliability as described above.

If a sliding bearing is adopted as a secondary bearing as in Patent Document 2, the cost and long life are superior, but the drive shaft that rotates while tilting hits the thrust receiving surface, and the local Hertz stress is high. Thus, there is a problem that the sliding state on the thrust receiving surface becomes severe. However, in Patent Document 2, no consideration is given to the problem of piece contact with respect to the thrust receiving surface.

The present invention has been made in view of the above points, and provides a compressor capable of ensuring a good sliding state of a thrust receiving surface in a compressor using a slide bearing as a secondary bearing. Objective.

A compressor according to the present invention includes a compression mechanism unit disposed in a sealed container, an electric motor unit that drives the compression mechanism unit, a drive shaft that transmits a driving force of the electric motor unit to the compression mechanism unit, and an upper portion of the drive shaft. The main bearing that supports the shaft, the sub bearing that supports the lower part of the drive shaft, the radial bearing surface that is provided on the sub bearing and slidably supports the radial surface of the drive shaft, and the thrust surface of the drive shaft is slidable A thrust receiving surface that is supported by the contact surface, and a contact point between the thrust surface and the thrust receiving surface that is in contact with the curved surface, and continuously from the contact point toward the radially outer side. And an increasing gap.

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

It is a longitudinal cross-sectional view which shows typically the cross-sectional structure of the scroll compressor which concerns on Embodiment 1 of this invention. It is a figure which shows the principal part of the scroll compressor which concerns on Embodiment 1 of this invention. It is explanatory drawing of the bending state in the ZX cross section of the drive shaft of the scroll compressor which concerns on Embodiment 1 of this invention, and the arrangement position of the circumferential direction of a radial direction oil groove. It is explanatory drawing of the bending state in the YZ cross section of the drive shaft of the scroll compressor which concerns on Embodiment 1 of this invention, and the arrangement position of the circumferential direction of a radial direction oil groove. It is explanatory drawing of the load which acts on the drive shaft of the scroll compressor which concerns on Embodiment 1 of this invention, and the acting direction of a load. It is a figure which shows the modification 1 of the scroll compressor which concerns on Embodiment 1 of this invention. It is a figure which shows the modification 2 of the scroll compressor which concerns on Embodiment 1 of this invention. It is a figure which shows the modification 3 of the scroll compressor which concerns on Embodiment 1 of this invention. It is a figure which shows the principal part of the scroll compressor which concerns on Embodiment 2 of this invention, and is a schematic diagram which expands and shows a lower end part from a subbearing. It is a figure which shows the principal part of the scroll compressor which concerns on Embodiment 3 of this invention, and is a schematic diagram which expands and shows a lower end part from a subbearing.

Hereinafter, a scroll compressor as an example of a compressor according to an embodiment of the present invention will be described with reference to the drawings. Here, in FIG. 1 and the following drawings, the same reference numerals denote the same or corresponding parts, and are common to the whole text of the embodiments described below. And the form of the component represented by the whole specification is an illustration to the last, Comprising: It does not limit to the form described in the specification.

Embodiment 1 FIG.
FIG. 1 is a vertical cross-sectional view schematically showing a cross-sectional configuration of a scroll compressor according to Embodiment 1 of the present invention. In FIG. 1, the thrust load is represented by an arrow a, and the radial load is represented by an arrow b. FIG. 2 is a diagram showing a main part of the scroll compressor according to Embodiment 1 of the present invention. 2A is a schematic diagram showing an enlarged lower end portion of the drive shaft sub-bearing, and FIG. 2B is a schematic diagram of the shape of the thrust surface of the drive shaft viewed from directly below. Moreover, in FIG. 2, the arrow has shown the flow of lubricating oil.

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

The scroll compressor 100 includes a compression mechanism part A accommodated in the hermetic container 13, an electric motor part B, and a drive shaft 6 that transmits the driving force of the electric motor part B to the compression mechanism part A. As shown in FIG. 1, the compression mechanism part A is disposed on the upper side of the sealed container 13, and the electric motor part B is disposed on the lower side of the sealed container 13. And when the drive shaft 6 is rotationally driven by the electric motor part B, the volume of the later-described compression chamber 5 formed in the compression mechanism part 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 sealed container 13. The hermetic container 13 is a pressure container and forms an outer shell of the scroll compressor 100. The bottom of the hermetic container 13 serves as an oil storage space 14 for storing lubricating oil. The lubricating oil stored in the oil storage space 14 is sucked up by an oil pump 9 provided at the lower end portion of the drive shaft 6 and supplied to the sliding portions of the drive shaft 6 and 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 section A has a function of compressing the refrigerant sucked from the refrigerant suction pipe 15 and discharging the refrigerant to the outside of the hermetic container 13 through the discharge port 4 and the refrigerant discharge pipe 16 described later. The compression mechanism part A is mainly composed of a fixed scroll 1, an orbiting scroll 2, and an Oldham coupling 25. The fixed scroll 1 includes a base plate 1a and a spiral protrusion 1b provided on the lower surface of the base plate 1a. The fixed scroll 1 is fixed to the upper end portion of the main frame 8 a fixed to the inner peripheral surface of the sealed container 13. The fixed scroll 1 is preferably fixed with a fastening member such as a bolt.

As with the fixed scroll 1, the orbiting scroll 2 is composed of a base plate 2a and a spiral projection 2b provided on the upper surface of the base plate 2a. An eccentric hole 2c is formed in the vicinity of the center below the bottom surface of the base plate 2a of the rocking scroll 2, and a rocking bearing 17 is press-fitted into the eccentric hole 2c. An eccentric shaft 6 a provided at the upper end of the drive shaft 6 is slidably connected to the rocking bearing 17. The orbiting scroll 2 revolves without revolving with respect to the fixed scroll 1 by the Oldham coupling 25 provided between the orbiting scroll 2 and the main frame 8a. In addition, a rocking thrust bearing 18 is provided on the lower surface side of the rocking scroll 2, that is, between the rocking scroll 2 and the main frame 8a.

The fixed scroll 1 and the orbiting scroll 2 are provided in the hermetic container 13 so that the spiral protrusions of each other mesh with each other. And the compression chamber 5 in which the volume changes relatively is formed by the meshing of the spiral projections of the fixed scroll 1 and the swing scroll 2. A suction port 3 that guides the refrigerant sucked into the sealed container 13 from the refrigerant suction pipe 15 to the compression chamber 5 is formed on the outer periphery of the compression chamber 5, and the refrigerant enters the compression chamber 5 through the suction port 3. Inhaled. Then, the refrigerant is compressed in the compression chamber 5, and the compressed refrigerant is discharged from the discharge port 4 formed in the center 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.

[Mainframe 8a]
The main frame 8a fixes the fixed scroll 1 at its upper end, and supports the orbiting scroll 2 through the orbiting thrust bearing 18 so as to be slidable from below. The main frame 8 a is attached in the sealed container 13 so that the outer peripheral surface is in contact with the inner peripheral surface of the sealed container 13. In addition, a through hole that allows the drive shaft 6 to pass therethrough is formed near the center of the main frame 8a, and a main bearing 19 that rotatably supports a later-described main shaft portion 6b of the drive shaft 6 is formed in the through hole. Is provided. That is, the main frame 8 a also has a function of rotatably supporting the drive shaft 6 via the main bearing 19.

[Subframe 8b]
The subframe 8b is fixed to the inner surface of the side wall of the hermetic container 13 below the electric motor part B. The subframe 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 the lower end portion of the cylindrical portion 8ba.

Further, a through hole through which the drive shaft 6 passes is formed in the vicinity of the center of the sub frame 8b, and a cylindrical sub bearing 11 is provided in the through hole. The auxiliary bearing 11 is a radial bearing that rotatably supports a later-described auxiliary shaft portion 6c of the drive shaft 6 and supports a load in the radial direction. The auxiliary bearing 11 is a plain bearing commonly called “metal”.

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

That is, the scroll compressor 100 of the first embodiment supports the radial load acting on the drive shaft 6 by the auxiliary bearing 11 and receives the thrust load by the thrust receiving surface 12 provided on the upper surface cover 9c of the oil pump 9. It has a configuration. And the scroll compressor 100 of this Embodiment 1 is characterized by the structure which hold | maintains the favorable sliding state of the thrust receiving surface 12. FIG. This structure will be described later.

[Motor part B]
The electric motor part B has a function of driving the orbiting scroll 2 in order to compress the refrigerant by the compression mechanism part A. The electric motor part B is disposed between the main frame 8a and the subframe 8b. The electric motor part B is mainly composed of an electric motor 10 having a rotor 10a and a stator 10b. The rotor 10a is fixed to the peripheral surface of the drive shaft 6 and is driven to rotate when energization to the stator 10b is started. 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 through a gap, and rotates the rotor 10a.

[Drive shaft 6]
The drive shaft 6 transmits the rotation of the rotor 10a of the electric motor part B to the orbiting scroll 2 of the compression mechanism part A. The drive shaft 6 includes, in order from the top, an eccentric shaft 6a, a main shaft portion 6b fixed to the rotor 10a of the electric motor section B, a sub shaft portion 6c, and a pump insertion shaft 6d having a diameter smaller than that of the sub shaft portion 6c. have. The eccentric shaft 6a is provided eccentrically with respect to the shaft center of the drive shaft 6, and is slidably connected to the rocking bearing 17 as described above. The main shaft portion 6b is provided with a balancer 26a on the upper side of the rotor 10a and a balancer 26b on the lower side of the rotor 10a.

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

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

The bottom surface of the auxiliary shaft portion 6c has a hollow 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 drive shaft 6 rotates while being pressed against the thrust receiving surface 12 due to the weight and magnetic force of the drive shaft 6 acting below the axis of the compressor.

A radial oil groove 7g extending in the radial direction from the inner peripheral end to the outer peripheral end is formed on 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 end portion of the thrust surface 6f is a curved surface having a curvature radius Rs1 in order to relax the contact angle with the thrust receiving surface 12 during rotation. The chamfered portion (fillet) 6e is formed. The arc range of the chamfered portion (fillet) 6 is formed in an angle range smaller than 90 degrees.

[Material hardness of auxiliary bearing 11 of auxiliary shaft portion 6c]
Generally, the main shaft portion 6b and the sub shaft portion 6c of the drive shaft 6 are hardened and used by quenching a base material carbon steel. A portion indicated by dots in FIG. 2A indicates a quenching portion of the auxiliary shaft portion 6c. On the other hand, since the thrust surface 6f is used without being cured, it is necessary to keep a distance from the quenching portion of the countershaft portion 6c. For the radial receiving surface 11a of the auxiliary bearing 11, a metal metal having excellent lubricity is usually used. An aluminum alloy or a copper alloy is used on the inner peripheral surface of the radial receiving surface 11a that slides on the radial surface, and receives the hard countershaft portion 6c. On the other hand, the thrust receiving surface 12 has a fine surface roughness and is made of hardened hard steel material, and does not wear even when a load is applied from the rotating 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 9b, and an upper surface cover 9c that covers the movable portion 9a, and is fixed to the subframe 8b with screws 24 at the main body 9b portion. ing. The upper surface of the upper surface cover 9c is the thrust receiving surface 12 as described above, and the oil pump 9 is attached to the subframe 8b so as to maintain the squareness accuracy and sufficient rigidity between the thrust receiving surface 12 and the auxiliary bearing 11. It is fixed and integrated.

The top cover 9c is made of an annular member, and a through hole 23 is 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 following the through hole 23. The periphery of the upper end opening of the through hole 23 is a chamfered portion 12e formed of a curved surface having a curvature radius Rp0 so as not to contact the pump insertion shaft 6d. The upper surface cover 9c is arrange | positioned under the subshaft part 6c, and is equivalent to the outline member of this invention. Further, the lower end opening 9 d of the oil pump 9 is immersed in the lubricating oil in the oil storage space 14.

In the oil pump 9, the movable portion 9 a is operated in accordance with the rotation of the drive shaft 6, sucks the lubricating oil stored in the oil storage space 14 from the lower end opening 9 d, and supplies it to various portions of the compression mechanism portion A. The various parts of the compression mechanism part A correspond to various bearings such as the rocking bearing 17, the rocking thrust bearing 18, the main bearing 19, the auxiliary bearing 11 and the thrust receiving surface 12, and the sliding part of the Oldham coupling 25. The lubricating oil sucked up by the oil pump 9 is supplied to the rocking bearing 17 and the main bearing 19 from the oil supply vertical hole 7a through the eccentric shaft oil supply horizontal hole 7b and the main shaft oil supply horizontal hole 7c. A part of the lubricating oil flowing out from the upper outlet of the oil supply vertical hole 7 a or the eccentric shaft oil supply horizontal hole 7 b is supplied to the sliding portion of the oscillating thrust bearing 18 and the Oldham joint 25. The lubricating oil sucked up by the oil pump 9 is supplied from the vertical oil supply hole 7a to the auxiliary bearing 11 through the radial oil supply horizontal hole 7d and from the vertical oil supply hole 7a to the thrust receiving surface 12 through the thrust oil supply horizontal hole 7e. Is done. The oil supply to the thrust receiving surface 12 is a characteristic part of the first embodiment, and will be described below again.

[Thrust surface 6f, thrust receiving surface 12]
The thrust receiving surface 12 of the upper surface cover 9c has a flat surface 12b on the radially inner side and an inclined surface 12d on the radially outer side of the flat surface 12b that is gently inclined downward toward the outside. The flat surface 12b is a surface that is perpendicular to the radial receiving surface 11a that is the inner peripheral surface of the auxiliary bearing 11 and has a uniform height. The flat surface 12b of the thrust receiving surface 12 is connected to be in contact with the chamfered portion 12e having the curvature radius Rp0 at the inner peripheral end, and is connected to be in contact with the inclined surface 12d with the curvature radius Rp1 at the flat surface end point 12c. With this configuration, between the thrust surface 6f and the thrust receiving surface 12, a contact point 12f where the thrust surface 6f and the thrust receiving surface 12 are in contact with each other in a curved surface, and continuously from the contact point 12f toward the radially outer side. An increasing gap 20 is formed.

The drive shaft 6 is inclined about 1/1000 [rad] to 2/1000 [rad] in the compressive load direction under the maximum load condition. If the inclination angle of the inclined surface 12d is designed to be the same level as the inclination of the drive shaft 6 under the maximum load condition, the thrust surface 6f is gently in contact with the thrust receiving surface 12 under the maximum load condition under the severest operating conditions. In this state, the wear state can be alleviated.

The position of the contact point 12f that slides with the thrust receiving surface 12 moves in the radial direction from the inner peripheral side to the outer peripheral side as the auxiliary 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, when the radial load acting on the drive shaft 6 is minute and the thrust surface 6f and the flat surface 12b of the thrust receiving surface 12 are substantially parallel, the entire flat surface 12b receives the thrust load acting on the drive shaft 6. The thrust load acting on the drive shaft 6 usually corresponds to the respective weights 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 is applied, the drive shaft 6 is bent and deformed, the thrust surface 6f is tilted, and a curved surface (curvature radius Rp1) near the outside of the flat surface end point 12c of the thrust receiving surface 12 The drive shaft 6 rotates while contacting.

3) When the load is further increased and the compressive load and the centrifugal force are increased, the drive shaft 6 is bent and deformed, the thrust surface 6f is further inclined, and the thrust surface 6f is in gentle contact with the thrust receiving surface 12. Rotate.

4) Finally, under the maximum load condition, when the drive shaft 6 is greatly bent and deformed and the thrust surface 6f is inclined with respect to the flat surface 12b of the thrust receiving surface 12, the curved surface (curvature) of the chamfered portion 6e on the outer peripheral portion of the thrust surface 6f. The radius Rs1) rotates while being in contact with the inclined surface 12d of the thrust receiving surface 12. In this case, by increasing the radius of curvature Rs1 of the chamfered portion 6e, the chamfered portion 12e is prevented from hitting the thrust receiving surface 12 to increase local Hertz stress.

For example, from a formula of Hertz stress when a cylinder with a curvature radius of 1 and a cylinder with a curvature radius of 2 are in contact with each other with a length L, Hertz stress ∝ (load) x (Young's modulus) x L x {1 / (curvature radius 1) ² + (curvature radius 2) ² }
Holds.

The curvature radius Rs1 of the chamfered portion 6e is
Rp0 <Rs1 <Rp1
It is necessary to design Rs1 to be at least larger than Rp0 and to be close to Rp1.
The general R chamfer (fillet) is about R0.2 to R2, whereas the radius of curvature Rs1 of the chamfer 6e is large within the design allowable range, and is usually larger than R3.

In the first embodiment, the sliding portion on which the auxiliary shaft portion 6c and the upper surface cover 9c slide is formed by separating the thrust receiving surface 12 from the thrust surface 6f with the radially outer side of the thrust receiving surface 12 as an inclined surface 12d. Instead of the entire thrust receiving surface 12, the flat surface 12b is formed. Therefore, compared with the case where the entire thrust receiving surface 12 is formed as a flat surface 12b without forming the inclined surface 12d, the outer peripheral position of the sliding portion can be moved radially inward. For this reason, the maximum sliding speed can be suppressed.

Further, after the break-in operation, the corner of the flat surface end point 12c, which is the boundary between the flat surface 12b and the inclined surface 12d, is taken. For this reason, the contact angle between the thrust surface 6f and the thrust receiving surface 12 at the flat surface end point 12c also becomes gentle, and the contact surface pressure of the sliding portion can be relaxed.

[Other configurations]
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 penetrating so as to extend in the axial direction is formed in the stator 10b. . The oil return pipe 29 and the oil return hole 29 a have a function of returning the lubricating oil used in the compression mechanism portion A to the oil storage space 14. In addition, in FIG. 1, although the case where only the oil return pipe | tube 29 and the oil return hole 29a were provided is shown as an example, it is not limited to this. For example, two or more oil return pipes 29 and oil return holes 29a may be provided.

As described above, the scroll compressor 100 has the compression mechanism part A disposed in the upper part of the hermetic container 13 and the motor part B disposed in the lower part, and the driving force of the motor part B is supplied via the drive shaft 6 to the compression mechanism part A. Is transmitted to the orbiting scroll 2, and the orbiting scroll 2 is driven to rotate. In addition, the kind of lubricating oil is not specifically limited, What is necessary is just to be used as lubricating oil of the compression mechanism part A. For example, PAG (polyalkylene glycol) or POE (polyol ester) may be used as the lubricating oil. Also, the type of refrigerant is not particularly limited.

[Description of operation]
Hereinafter, the operation of the scroll compressor 100 will be described.
When energization is started to the stator 10b constituting the electric motor 10, the drive shaft 6 starts to rotate together with the rotor 10a. When the drive shaft 6 starts to rotate, the orbiting scroll 2 connected to the eccentric shaft 6 a performs a revolving motion while being prevented from rotating by the Oldham joint 25. Thereby, the compression chamber 5 moves to the center side while gradually reducing the volume. As a result, the pressure of the refrigerant sucked into the compression chamber 5 from the suction port 3 is gradually increased. Then, the refrigerant whose pressure has increased is discharged out of the apparatus through the discharge port 4 and the refrigerant discharge pipe 16, and is pumped to a refrigerant pipe (not shown) connected to the refrigerant discharge pipe 16.

Since the refrigerant in the sealed container 13 is discharged to the outside in this way, the inside of the sealed container 13 has a negative pressure. For this reason, the refrigerant from the refrigerant pipe (not shown) outside the machine is sucked into the sealed container 13 through the refrigerant suction pipe 15. The refrigerant sucked into the sealed container 13 is sucked into the compression chamber 5 from the suction port 3 after cooling the electric motor 10.

Further, as the drive shaft 6 rotates, the inner rotor 9aa of the oil pump 9 rotates, and the outer rotor 9ab also rotates accordingly, so that the lubricating oil in the oil storage space 14 is supplied through the oil supply vertical hole by the pump action of the oil pump 9 It is sucked upward through 7a. The sucked lubricating oil is distributed to each of the sub bearing 11, the main bearing 19 and the rocking bearing 17, and lubricates each of these bearings. Lubricating oil passing through the rocking bearing 17 is supplied to the rocking thrust bearing 18 and the Oldham coupling 25 to lubricate these sliding portions. Further, the lubricating oil supplied to the Oldham coupling 25 is returned to the oil storage space 14 through the oil return pipe 29.

[Deformation of the cylindrical portion 8ba of the subframe 8b]
The auxiliary bearing 11 supports a radial load generated when the scroll compressor 100 is operated. The cylindrical portion 8ba of the sub-frame 8b provided with the auxiliary bearing 11 is thinner in the radial direction than the flange portion 8bb, and is a thin flexible structure portion. Thus, by making the cylindrical part 8ba into a thin flexible structure part, the cylindrical part 8ba is elastically deformed following the inclination of the drive shaft 6 and suppresses the single contact of the drive shaft 6 with the radial receiving surface 11a. It is possible. Since FIG. 2 is a schematic diagram, the cylindrical portion 8ba is illustrated as being thicker than the flange portion 8bb, but in reality, the cylindrical portion 8ba is illustrated as being formed thick. Is formed.

[Radial load and thrust load]
A radial load acts on the auxiliary bearing 11 as the scroll compressor 100 is operated. That is, a variable load synchronized with the rotation of the drive shaft 6 is applied to the auxiliary bearing 11 in the radial direction. As described above, the auxiliary bearing 11 is constituted by a plain bearing. As a result of not using a ball bearing or the like for the auxiliary bearing 11 as described above, long-term reliability can be ensured and seizure of the drive shaft 6 can be prevented, and an increase in wear and friction loss can be suppressed and stable start-up can be achieved. Sex can be obtained.

Further, along with the operation of the scroll compressor 100, a thrust load acts on the thrust receiving surface 12 of the top cover 9c. That is, the thrust receiving surface 12 is applied with its own weight as a thrust load vertically downward. In order to keep the thrust surface 6f of the auxiliary 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 always ensure a stable oil surface on the thrust receiving surface 12. is necessary.

Therefore, in the first embodiment, an inclined surface 12d is provided on the outer peripheral side of the thrust receiving surface 12, and a region between the bottom surface of the auxiliary bearing 11 and the inclined surface 12d and radially outside the thrust receiving surface 12 is provided. An oil sump space 22 is formed in a region further outside the thrust receiving surface 12. Since oil flows into the oil sump space 22 from the oil seal portion 11d on the lower end side of the radial receiving surface 11a, the oil head pressure stored up to the upper end 22b causes the oil receiving space 22 to constantly move from the outer peripheral side of the thrust receiving surface 12 to the sliding surface. A stable oil is supplied. On the other hand, a stable oil is always supplied from the inner peripheral side of the thrust receiving surface 12 to the sliding surface through the thrust oil supply lateral hole 7e of the pump insertion shaft 6d through the radial oil groove 7g. Hereinafter, an oil supply path for ensuring a stable oil level on the thrust receiving surface 12 will be described in detail. Here, in order to form a space to be the oil sump space 22, the radially outer region of the thrust receiving surface 12 is the inclined surface 12d. However, the region is not limited to the inclined surface 12d and is higher than the height position of the flat surface 12b. It is good also as a plane formed in the low position.

[Oil supply path to thrust receiving surface 12]
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 centrifugal force. That is, the lubricating oil that has flowed into the outer peripheral side of the pump insertion shaft 6d raises the gap between the pump insertion shaft 6d and the movable portion 9a of the oil pump 9, and further raises the gap between the pump insertion shaft 6d and the top cover 9c. To do. Then, the lubricating oil that has risen through the gap between the pump insertion shaft 6 d and the upper surface cover 9 c reaches the thrust receiving surface 12. Further, part of the lubricating oil that has been pressurized by the oil pump 9 and has flowed into the oil supply vertical hole 7a is also discharged radially outward by centrifugal force from the thrust oil supply lateral hole 7e formed in the pump insertion shaft 6d, and the thrust receiving surface 12 To reach. Specifically, the thrust oil supply lateral hole 7e is formed in the pump insertion shaft 6d so as to communicate with the oil supply vertical hole 7a below the thrust surface 6f and to extend in the radial direction of the drive shaft 6 so that the lubricating oil is supplied to the thrust surface. Supply between 6f and the thrust receiving surface 12.

Then, the lubricating oil that has reached the thrust receiving surface 12 flows through the radial oil groove 7g formed on the thrust surface 6f from the inner peripheral end to the outer peripheral end, and then flows into the oil sump space 22. The flow passage cross-sectional area obtained by cutting the radial oil groove 7g in the axial direction is larger than the flow passage 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. It is formed small. Therefore, the lubricating oil in the annular gap between the pump insertion shaft 6d and the upper surface cover 9c flows in a concentrated manner in the radial oil groove 7g, and the radial oil groove 7g is filled with the lubricating oil. It has become. Since the radial oil groove 7g makes one round in the circumferential direction as the drive shaft 6 rotates, the lubricating oil spreads over the entire thrust receiving surface 12.

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

Here, in the region facing the auxiliary bearing 11 on the outer peripheral surface of the auxiliary shaft portion 6c, the region S1 above the axial oil groove 7f has a longer axial length than the lower region S2. . The gap between each of the regions S1 and S2 and the auxiliary bearing 11 serves as an oil seal portion due to the retention of lubricating oil. Since the region S1 is longer in the axial direction 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. For this reason, the amount of the lubricating oil supplied from the radial oil supply lateral hole 7d to the axial oil groove 7f flows out from the upper end side of the auxiliary bearing 11 and leaks into the sealed container 13 is reduced, and the lubricating oil is accumulated in the oil sump space 22. I am doing so. In this way, the lubricating oil is abundant in the oil sump space 22 so that the thrust surface 6f and the thrust receiving surface 12 can be held in a good sliding state.

As described above, in the oil sump space 22, the lubricating oil that has passed through the vertical oil supply hole 7a through the thrust oil supply horizontal hole 7e and the lubricating oil that has passed through the vertical oil supply hole 7a and the radial oil supply horizontal hole 7d merge. The lubricating oil merged in the oil sump space 22 passes into the oil storage space 14 through the oil discharge passage 21 formed by grooves and holes formed in the mounting surfaces of the subframe 8b and the oil pump 9 on each other. Returned. The oil discharge passage 21 has an upstream end opened to the lower end 22 a of the oil sump space 22, and a downstream end discharge hole 21 a opened downward to the outer surface of the main body 9 b of the oil pump 9. The discharge hole 21 a is an outlet of the oil discharge channel 21 and is located below the oil sump space 22. The lubricating oil in the oil sump space 22 is discharged to the outside of the subframe 8b through the oil discharge channel 21. In addition, the code | symbol 22b of Fig.2 (a) is an upper end of the oil sump space 22, and the upper end 22b is arrange | positioned above the radial direction oil groove 7g.

Here, when the oil sump space 22 is defined again, the auxiliary bearing 11 is provided inside the cylindrical portion 8ba of the subframe 8b as shown in FIG. A cylindrical space is formed below. This space becomes a part of the oil sump space 22.

Further, in FIG. 2, a region of the inner peripheral surface 22 c of the cylindrical portion 8 ba of the sub frame 8 b that is opposed to the sub shaft portion 6 c of the drive shaft 6 and is located below the sub bearing 11. A wall surface portion 22d that is recessed outward in the radial direction is formed. Thus, the oil reservoir space 22 is further expanded by providing the wall surface portion 22d on the inner peripheral surface 22c of the tubular portion 8ba of the subframe 8b. Further, by providing the wall surface portion 22d, the lower end portion of the auxiliary shaft portion 6c of the drive shaft 6 that rotates while tilting comes into contact with the thrust receiving surface 12 or the radial receiving surface 11a, so that the auxiliary shaft portion 6c and the thrust receiving surface are received. The possibility that the surface 12 or the radial receiving surface 11a is damaged can be reduced.

In the middle of the oil discharge channel 21, a throttle channel 21 b that gives an appropriate channel resistance is formed. By forming the throttle channel 21 b, the lubricating oil is temporarily supplied to the oil sump space 22. It can be stored. When wear powder is generated on the radial receiving surface 11 a or the thrust receiving surface 12, it is necessary to discharge the wear powder to the oil storage space 14 without accumulating the wear powder in the oil sump space 22. That is, it is necessary to discharge the wear powder from the oil reservoir space 22 through the oil discharge channel 21. For this reason, it is preferable that the channel cross section of the throttle channel 21b is a channel cross section in which the depth and width are set to about 0.2 mm to 1 mm, respectively.

[Axial positional relationship between the radial oil groove 7g and the oil sump space 22]
The radial oil groove 7g is formed on 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, the inside of the radial oil groove 7g is lubricated. It needs to be filled with oil. When the upper surface of the lubricating oil accumulated in the oil sump space 22 is positioned below the bottom of the concave groove that constitutes the axial oil groove 7f, the radial oil groove 7g is not filled with the lubricating oil. Therefore, the upper surface of the oil accumulated in the oil sump space 22 is configured to be positioned above the bottom of the concave groove that constitutes the axial oil groove 7f. Specifically, the volume of the oil reservoir space 22 and the throttle channel 21b may be designed from the relationship with the amount of oil flowing into the oil reservoir space 22.

[Disposed position in the circumferential direction of the radial oil groove 7g]
The radial oil groove 7g is formed at one location on the thrust surface 6f, and the circumferential arrangement position on the thrust surface 6f is set in consideration of the bending direction of the drive shaft 6. Hereinafter, the arrangement position in the circumferential direction of the radial oil groove 7g will be described with reference to FIG. 3 and FIG. Prior to the description of FIG. 3 and FIG. 4, first, the load acting on the drive shaft 6 and the acting direction of the load will be described with reference to FIG. 5.

FIG. 5 is an explanatory diagram of the load acting on the drive shaft of the scroll compressor according to Embodiment 1 of the present invention and the acting direction of the load.
With respect to the Y axis, the direction connecting the axis of the drive shaft 6 and the axis of the eccentric shaft 6a (hereinafter referred to as the eccentric direction) is defined as + when the drive shaft 6 is viewed in plan from the axis direction. The Z axis is the axis of the drive shaft 6 and the upward direction is +. 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 as viewed from above, but receives a gas load in the opposite + X direction. In FIG. 5, Fx works mainly in the + X direction with a gas load necessary to compress the refrigerant gas in the compression chamber 5. F _1x is a load in the X-axis direction that acts on the main bearing 19. F _1y is a load in the Y-axis direction that acts on the main bearing 19. F _2x is a load in the X-axis direction that acts on the auxiliary bearing 11. F _2y is a load in the Y-axis direction that acts on the auxiliary bearing 11.

Further, since a moment is exerted on the drive shaft 6 to incline the entire drive shaft 6 by centrifugal force in the eccentric (+ Y) direction, the balancer 26a and the balancer 26b are attached to the top and bottom of the rotor 10a so as to cancel this moment. Yes. _FBW1 is a load due to centrifugal force acting on the drive shaft 6 by the balancer 26a. _FBW2 is a load due to centrifugal force acting on the drive shaft 6 by the balancer 26b.

FIG. 3 is an explanatory diagram of the bending state of the drive shaft of the scroll compressor according to Embodiment 1 of the present invention in the ZX cross section and the circumferential position of the radial oil groove. 3A shows a case where the gas load acting on the drive shaft 6 is in a low load operation, and FIG. 3B shows a case where the gas load is in a high load operation. Further, in each of (A) and (B), (a) is a diagram showing a bending state of the drive shaft 6 in the ZX section, and (b) is a plan view of the drive shaft 6 as viewed from above. The figure which showed the load which acts on the axis | shaft 6, (c) is explanatory drawing of the position of the circumferential direction of the radial direction oil groove 7g seen from the bottom. In FIG. 3, the arcuate double line arrow indicates the rotation direction of the drive shaft 6.

The centrifugal force Fc acts on the drive shaft 6 in the + Y direction due to the eccentric rotation of the swing scroll 2 or the like. However, when adopting a structure in which the centrifugal force Fc is received by the spiral projection (tooth tip wrap) 1 b of the fixed scroll 1, the centrifugal force Fc does not act on the drive shaft 6. Here, first, in the ZX cross section in which the centrifugal force can be ignored, the gas load Fx necessary for compressing the refrigerant gas in the compression chamber 5 acts as a main load in the + X direction on the drive shaft 6, and the main shaft portion 6b. The countershaft portion 6c is subjected to a reaction force commensurate with it. There are diameter gaps (usually about 1/1000 of the shaft diameter) between the main bearing 19 and the main shaft portion 6b and between the sub bearing 11 and the sub shaft portion 6c for enabling fluid lubrication. For this reason, the drive shaft 6 bends and deforms, and the main shaft 19 in the main bearing 19 is inclined in the + X direction, and the auxiliary shaft 11 in the auxiliary bearing 11 is inclined so that the auxiliary shaft portion 6c approaches in the −X direction. The axis 6h of the auxiliary shaft 6c is also inclined with respect to the axis (Z axis) 6g of the compressor. As a result, during the low load operation of FIG. 3A, the thrust surface 6f is inclined with respect to the thrust receiving surface 12 so that the -axis direction of the X axis is lifted and the + direction of the X axis is pressed.

Furthermore, as shown in FIG. 3 (B), during high load operation, the drive shaft 6 bends and deforms more greatly than during low load operation. For this reason, the thrust surface 6f of the auxiliary shaft portion 6c is inclined with respect to the thrust receiving surface 12 so that the + direction of the X axis is lifted and the-direction of the X axis is pressed. When the radial oil groove 7g is pressed against the thrust receiving surface 12, there is a problem that the surface pressure is locally increased near the edge of the radial oil groove 7g and is easily worn. Therefore, it is preferable to arrange the radial oil groove 7g so as to avoid the ± X directions so that the vicinity of the radial oil groove 7g does not come into contact with the thrust receiving surface 12. On the other hand, on the radial receiving surface 11a of the auxiliary bearing 11, the auxiliary shaft portion 6c approaches the -X direction and a load is applied, so that the axial oil groove 7f is provided on the + X side (anti-load side).

Next, FIG. 4 is an explanatory diagram of the bending state in the YZ section of the drive shaft of the scroll compressor according to Embodiment 1 of the present invention and the circumferential position of the radial oil groove. In FIG. 4, (A) shows a low speed operation, and (B) shows a high speed operation. Further, in each of (A) and (B), (a) is a diagram showing a bending state of the drive shaft 6 in the YZ section, and (b) is a plan view of the drive shaft 6 as viewed from above. The figure which showed the load which acts on the axis | shaft 6, (c) is explanatory drawing of the position of the circumferential direction of the radial direction oil groove 7g seen from the bottom. In FIG. 4, the arcuate double line arrow indicates the rotation direction of the drive shaft 6.

When centrifugal force Fc acts on the drive shaft 6 due to eccentric rotation of the orbiting scroll 2 or the like, a large load acts locally on the spiral protrusion (tooth tip wrap) 1b of the fixed scroll 1 or the drive shaft tilts. There is a problem that the rotor swings around. For this reason, as a mechanism for relieving the load caused by the centrifugal force Fc, the balancer 26a and the balancer 26b are attached to the top and bottom of the rotor 10a to balance the force and moment acting on the YZ section of the drive shaft 6. At this time, the inclination angle of the drive shaft 6 that is bent while being bent and deformed changes depending on the presence or absence of the centrifugal force Fc, but the auxiliary shaft portion 6c is inclined with substantially the same tendency.

Since the centrifugal force F _BW1 of the upper balancer 26a of the rotor 10a works in the -Y direction and the centrifugal force F _BW2 of the lower balancer 26b works in the -Y direction, the radial shaft receiving surface 11a of the auxiliary bearing 11 has the auxiliary shaft portion 6c. Tilt closer to the -Y direction on the lower side. As a result, in both the low speed operation shown in FIG. 4A and the high speed operation shown in FIG. 4B, the thrust surface 6f lifts the positive direction of the Y axis against the thrust receiving surface 12 and presses the negative direction of the Y axis. Lean on. Therefore, it is preferable to arrange so as to avoid the −Y direction so that the vicinity of the radial oil groove 7 g does not come into contact with the thrust receiving surface 12.

As described above with reference to FIGS. 3 and 4, the driving shaft 6 is subjected to the gas load, which is the main load, and the centrifugal force, and the thrust surface 6f is tilted. The groove 7g is less likely to hit one side in the + direction of the Y axis, that is, the eccentric direction. Therefore, the radial oil groove 7g is preferably arranged in the eccentric direction (+ Y direction).

Furthermore, under the most severe high-speed and high-load operating conditions, as shown in FIGS. 3B and 4B, in addition to the main load gas load Fx, the centrifugal force of the balancer 26a on the upper side of the rotor is always maintained. F _BW1 works in the −Y direction, and the centrifugal force F _BW2 of the balancer 26b on the lower side of the rotor always works in the + Y direction. The resultant force Fr of the gas load Fx and the centrifugal force F _BW1 works in an angular direction slightly shifted clockwise (counter-rotation direction) from the + X axis, and the drive shaft 6 is bent and deformed. The angle at which 6f hits one side and the direction of the secondary shaft radial load also work slightly deviated. Considering this, as described in (c) of FIG. 3 (B) and (c) of FIG. 4 (B), the position of the radial oil groove 7g on the thrust surface 6f is set in the + direction (eccentricity) of the Y axis. Further durability can be obtained when the angle α is shifted counterclockwise (counter-rotation direction) as viewed from below (direction). The angle α is an acute angle range of 0 deg to 45 deg.

(Effect of the invention)
According to the first embodiment, 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. An oil sump space 22 is provided on the outer peripheral side of the thrust surface 6f. Lubricating oil is supplied to the thrust receiving surface 12 through the radial oil groove 7g and the gap 20, and then stored in the oil sump space 22. It has the composition to be. Thus, by providing the oil sump space 22, it is possible to always ensure the flow of the lubricating oil between the thrust surface 6 f and the thrust receiving surface 12. As a result, a good oil lubrication state and a sliding state can be secured on the thrust receiving surface 12, and abnormal wear of the drive shaft 6 and seizure of the drive shaft 6 can be suppressed. As described above, in the first embodiment, the compressor of the compressor is provided by a relatively simple means that merely provides the gap and the oil sump space 22 that increase radially outward between the thrust surface 6f and the thrust receiving surface 12. It is possible to improve the life and reliability.

Further, by providing the throttle passage 21b in the middle of the oil discharge passage 21, it is possible to make it easy to once hold the lubricating oil in the oil sump space 22, and good oil lubrication of the thrust receiving surface 12 is achieved. Is effective to get.

Further, as a specific configuration of the oil sump space 22, the thrust receiving surface 12 may be configured such that the radially outer side of the thrust receiving surface 12 is an inclined surface 12 d that is inclined downward as it goes outward. Can be configured.

Further, since the oil sump space 22 is provided on the radially outer side of the thrust receiving surface 12, the radial position of the sliding portion where the thrust surface 6f and the thrust receiving surface 12 are brought into contact can be moved inward. For this reason, the sliding speed at the flat surface end point 12c, which is the contact point, can be lowered during high load and high speed operation, which is a severe sliding condition in which the inclination angle of the drive shaft 6 becomes large. Or a contact angle can be restrained small.

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

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

It should be noted that the compressor is not limited to the structure shown in FIGS. 1 to 5, and various modifications can be implemented as follows, for example, without departing from the gist of the present invention.

<Modification 1 of Embodiment 1>
FIG. 6 is a diagram illustrating a first modification of the scroll compressor according to the first embodiment of the present invention, and is a schematic diagram in which a lower end portion is enlarged from a secondary bearing of the scroll compressor.
The modification 1 is different from the basic configuration shown in FIG. 2 in that the thrust oil supply lateral hole 7e of the pump insertion shaft 6d is eliminated. The thrust receiving surface 12 is also hardened and hardened, and the distance between the thrust receiving surface 12 and the quenching portion of the auxiliary shaft portion 6c (radial surface) is closer than the basic configuration shown in FIG. Also, the space of the oil sump space 22 is different from the basic configuration shown in FIG. 2 because the inner wall surface 22ca has no recess as shown in FIG.

Under high-speed and high-load operation conditions, the thrust surface 6f is greatly inclined and contacts and slides at a chamfered portion (fillet) 6e having a large curvature radius. Thus, even when the thrust oil supply lateral hole 7e is eliminated from the pump insertion shaft 6d, the oil flows into the oil sump space 22 from the oil seal portion 11d on the lower end side of the radial receiving surface 11a. For this reason, it is possible to supply a stable amount of oil to the outside of the thrust receiving surface 12 abundantly by the head pressure of the oil stored up to the upper end 22b. On the other hand, the lubricating oil leaking from the oil pump 9 is supplied to the chamfered portion 6e from the inner peripheral side of the thrust receiving surface 12 through the radial oil groove 7g from the gap between the pump insertion shaft 6d and the through hole 23. Is possible. However, since the amount of oil supply due to the latter oil pump leakage decreases as the rotational speed decreases, the head becomes smaller and decreases, so there is a limit to the range where low speed can be achieved.

As described above with reference to FIG. 2, when the thrust oil supply lateral hole 7e is provided on the pump insertion shaft 6d, the lubricating oil is supplied to the thrust receiving surface 12 more and more constantly through the radial oil groove 7g of the thrust surface 6f by the centrifugal pump action. This is an effective configuration for stable supply. However, except for the problem at the time of low-speed operation, the present modification 1 can achieve the same effect.

<Modification 2 of Embodiment 1>
FIG. 7 is a diagram showing a second modification of the scroll compressor according to the first embodiment of the present invention, and is a schematic diagram in which the lower end portion is enlarged from the auxiliary bearing of the scroll compressor.
Modification 2 differs from the basic configuration shown in FIG. 2 in the following two points. That is, one is that there is no inclined surface 12d on the thrust receiving surface 12 side, and the flat surface 12b is extended to the radially outer end. The other is that the oil sump space 22 is formed by making the radius of curvature Rs1 of the chamfered portion 6e at the radially outer end of the thrust surface 6f larger than that in FIG. The thrust receiving surface 12 is also hardened and hardened, the distance between the thrust receiving surface 12 and the quenching portion of the auxiliary shaft portion 6c is close, and the space of the oil sump space 22 is also smaller than that in FIG.

In FIG. 2, the flat surface end point 12c of the thrust receiving surface 12 is a contact point with the thrust surface 6f. On the other hand, in Modification 2, the entire thrust receiving surface 12 is a flat surface, so that when the drive shaft 6 is inclined, the point on the thrust receiving surface 12 that contacts the chamfering start position 6ea of the chamfered portion 6e is reduced. It becomes the contact point 12f. Therefore, by increasing the radius Rs of the chamfered portion 6e, the contact point 12f is located on the radially inner side of the thrust receiving surface 12 as compared with the case where the radius Rs is small, so that the sliding speed at the contact point 12f can be reduced. it can. Further, by increasing the radius Rs of the chamfered portion 6e, the contact angle can be relaxed as compared with the case where the radius Rs is small, even if the thrust receiving surface 12 as a whole is the flat surface 12b.

In the second modification, a circle whose radius is a line segment perpendicular to the axis of the drive shaft 6 from the contact point 12f (hereinafter referred to as a contact sliding circle) is larger than the movable portion 9a when viewed in the axial direction. It has a configuration. With this configuration, the following effects can be obtained. In other words, the upper surface cover 9c has an outer peripheral portion supported by the main body body 9b and an inner peripheral portion side which is not supported and is floated, and has a so-called cantilever shape. For this reason, if the contact sliding circle is inside the movable portion 9a of the oil pump 9, the inner peripheral side of the upper surface cover 9c may be pushed down by the drive shaft 6 and the upper surface cover 9c may be bent. . However, the configuration in which the contact sliding circle is larger than the movable portion 9a has an effect of preventing the upper surface cover 9c from being bent.

<Modification 3 of Embodiment 1>
FIG. 8 is a diagram showing a third modification of the scroll compressor according to the first embodiment of the present invention. 8A is a schematic diagram in which the lower end portion is enlarged from the auxiliary bearing of the scroll compressor, and FIG. 8B is a diagram of the annular steel plate that is the thrust receiving surface of the third modification viewed from directly above.
The third modification differs from the second modification in the following points. The thrust receiving surface 12 is different in that an annular steel plate 12g having a small thickness and a small surface roughness is used. The annular steel plate 12g is used on the top cover 9c of the oil pump 9. In addition, “thickness” means, for example, that the thickness is 1 mm or less, and “fine surface roughness” means that the surface roughness is, for example, z1 or less. For the annular steel plate 12g, for example, a commercially-quenched thin steel plate material such as a PK steel plate having a thickness of 0.5 mm can be used. The annular steel plate 12g corresponds to an example of the annular member of the present invention, and is obtained by polishing the surface of a quenched steel strip.

The annular steel plate 12g has an outer protrusion 12ga projecting outwardly at a part of the circular outer periphery, and the outer protrusion 12ga is formed at the notched portion of the main body 9b of the oil pump 9. Inserted and fixed so as not to rotate.

In Modification 3, by using the annular steel plate 12g, the annular steel plate 12g follows the behavior of the thrust surface 6f, so that the slidability is improved. Moreover, there is an advantage that processing is simple by using a commercially available quenched steel sheet as the annular steel plate 12g.

Embodiment 2. FIG.
FIG. 9 is a diagram showing a main part of the scroll compressor according to Embodiment 2 of the present invention, and is a schematic diagram showing an enlarged lower end portion from the auxiliary bearing.
In the second embodiment, compared with the basic configuration shown in FIG. 2, the formation position of the radial oil groove 7g is changed from the thrust surface 6f side to the thrust receiving surface 12 side, and the number of radial oil grooves 7g The difference is that the number is increased to a plurality, preferably 3 or more. Further, the thrust receiving surface 12 side is formed only by the flat surface 12b orthogonal to the compressor shaft 6g, and there is no inclined surface 12d. Instead, the thrust surface 6f side is different in that it is formed by an inclined surface 6fa inclined from a plane orthogonal to the shaft center 6g (compressor shaft center reference). The thrust surface 6f is inclined upward as it goes radially outward from the shaft center 6g. The contact angle between the thrust surface 6f and the thrust receiving surface 12 is designed by designing the inclination angle of the thrust surface 6f to be approximately the same as the inclination angle θs of the shaft center 6h of the countershaft portion 6c bent under high load and high speed operation. Can be kept small. The “inclination angle of the axis 6g deflected by 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 on the thrust surface 6f, the rotation of the drive shaft 6 causes the position of the radial oil groove 7g to rotate, thereby generating a centrifugal pump action. 12 had the effect of spreading the lubricant over the entire surface. On the other hand, in the configuration in which the radial oil groove 7g is formed on the thrust receiving surface 12 as in the second embodiment, the position of the radial oil groove 7g does not rotate, so the lubricating oil in the radial oil groove 7g It is difficult to flow and spread to the thrust receiving surface 12. Therefore, in the second embodiment, the flow resistance is reduced by increasing the number of the radial oil grooves 7g to a plurality, and the effect of spreading the lubricating oil over the entire thrust receiving surface 12 can be obtained.

However, the flow passage cross-sectional area of the radial oil groove 7g is made smaller than the flow passage cross-sectional area of the flow passage formed by the annular gap between the pump insertion shaft 6d and the upper surface cover 9c. As a result, the lubricating oil flowing through the radial oil groove 7g overflows to the upper surface side of the thrust receiving surface 12, and the thrust surface 6f rotates to spread over the entire surface, so that the lubrication state can be kept good.

Further, by providing the inclined surface 6fa, the contact point 12f is located radially inward as compared to the case where the inclined surface 6fa is not provided and the thrust surface 6f and the thrust receiving surface 12 are both flat surfaces. For this reason, it is possible to reduce the sliding speed at the contact point 12f and reduce the contact angle at the contact point 12f.

When the radial oil groove 7g is formed at the thrust surface 6f as described above with reference to FIG. 2, the lubricating oil can be supplied to the thrust receiving surface 12 more abundantly by centrifugal action. However, even when the radial oil groove 7g is formed at the thrust receiving surface 12 as in the second embodiment, if the radial oil groove 7g is appropriately designed, the same effect as in the first embodiment can be obtained.

Further, at least one end of the radial receiving surface 11a of the auxiliary bearing 11 in the axial direction of the drive shaft 6 is formed as a curved surface. Thereby, the damage by the contact of the end part formed with the curved surface and the drive shaft 6 can be suppressed. FIG. 9 shows a configuration in which both ends of the radial receiving surface 11a of the auxiliary bearing 11 in the axial direction of the drive shaft 6 are curved.

Embodiment 3 FIG.
FIG. 10 is a diagram illustrating a main part of the scroll compressor according to the third embodiment of the present invention, and is a schematic diagram illustrating an enlarged lower end portion of the auxiliary bearing.
In addition to the radial load, the subframe 8b receives a thrust load generated by the weight of the drive shaft 6 and the rotor magnetic force. In FIG. 2, this thrust load is received by the upper surface cover 9 c of the oil pump 9 disposed below the subframe 8 b, and the upper surface of the upper surface cover 9 c is the thrust receiving surface 12. The third embodiment is different in that the thrust load is received by the bottom plate 8bc fixed to the bottom of the subframe 8b with the bolt 40, and the upper surface of the bottom plate 8bc is the thrust receiving surface 12. The shape of the thrust receiving surface 12 is substantially the same as in FIG. The shape of the thrust surface 6f is the same as that in FIG.

The bottom plate 8bc is an annular circular base plate in which a through hole is formed at the center, and the outer diameter of the bottom plate 8bc is larger than the outer diameter of the cylindrical portion 8ba of the subframe 8b. The bottom plate 8bc has a plurality of bolt holes. The bolt 40 is passed through the bolt holes and screwed into the screw holes provided in the subframe 8b, so that the bottom plate 8bc is fixed to the subframe 8b. . Moreover, the pump insertion shaft 6d is longer than FIG. 2 by the thickness of the bottom plate 8bc. The bottom plate 8bc corresponds to an example of the annular member of the present invention.

According to the third embodiment, since the thrust receiving surface 12 is formed by the bottom plate 8bc separate from the top cover 9c of the oil pump 9, the material and thickness of the bottom plate 8bc are not subject to the design restrictions of the oil pump 9. Can be selected. Therefore, it is easy to easily increase the surface hardness and bending strength of the thrust receiving surface 12 without being restricted by the design of the oil pump 9.

Further, the characteristic configurations of the first embodiment, the first modification to the third modification, the second embodiment, and the third embodiment may be appropriately combined without departing from the gist of the present invention. For example, the configuration in which a plurality of radial oil grooves 7g are provided on the thrust receiving surface 12, which is a characteristic configuration of the second embodiment, is not limited to use in combination with the configuration in which the inclined surface 6fa is provided on the thrust surface 6f. . For example, you may combine with the structure which provides 12 d of inclined surfaces shown in FIG. In addition, the configuration in which the radius of curvature Rs1 of the chamfered portion 6e of Modification 2 is increased may be combined with Embodiment 2. Moreover, you may combine the cyclic | annular steel plate 12g of the modification 3 with Embodiment 2 and Embodiment 3. FIG.

In the first to third embodiments, the sub-bearing 11 and the sub-frame 8b are separately formed. However, the sub-bearing 11 and the sub-frame 8b are integrally formed. It is good also as a secondary bearing which supports a surface.

In the first to third embodiments, the member on which the thrust receiving surface is formed is the top cover 9c of the oil pump 9, the annular steel plate 12g, or the bottom plate 8bc fixed to the bottom of the subframe 8b. It was comprised separately from the subbearing 11 in which the receiving surface 11a is formed. However, as described above, the auxiliary bearing 11 is integrated with the subframe 8b, and the member on which the thrust receiving surface is formed is also integrated, and the integrated component is a radial of the auxiliary shaft portion 6c of the drive shaft 6. It is good also as a subbearing which supports a surface and a thrust surface.

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

1 fixed scroll, 1a base plate, 1b spiral projection, 2 swing scroll, 2a base plate, 2b spiral projection (tooth tip wrap), 2c eccentric hole, 3 suction port, 4 discharge port, 5 compression chamber, 6 drive shaft, 6a Eccentric shaft, 6b Main shaft portion, 6c Sub shaft portion, 6d Pump insertion shaft, 6e Chamfered portion, 6ea Chamfer start position, 6f Thrust surface, 6fa inclined surface, 6g Compressor shaft center, 6h Sub shaft portion shaft center, 7a oil supply vertical hole, 7b oil supply horizontal hole, 7c oil supply horizontal hole, 7d oil supply horizontal hole, 7e oil supply horizontal hole, 7f axial oil groove, 7g radial oil groove, 8 housing, 8a main frame, 8b subframe, 8ba cylindrical part (sub bearing) Metal), 8bb flange, 8bc bottom plate, 9 oil pump, 9a movable part, 9aa inner rotor, 9ab outer rotor, 9b Body, 9c Top cover, 9d Bottom opening, 10 Motor, 10a Rotor, 10b Stator, 11 Sub bearing, 11a Radial receiving surface, 11c Oil seal part, 11d Oil seal part, 12 Thrust receiving surface, 12b Flat surface, 12c Flat surface End point, 12d inclined surface, 12e chamfered part, 12f contact point, 12g annular steel plate, 12ga protruding portion, 13 sealed container, 14 oil storage space, 15 refrigerant suction pipe, 16 refrigerant discharge pipe, 17 rocking bearing, 18 rocking thrust Bearing, 19 main bearing, 21 oil discharge flow path, 21a discharge hole, 21b throttle flow path, 22 oil sump space, 22a lower end, 22b upper end, 22c inner peripheral surface, 22ca wall surface, 23 through hole, 24 screw, 25 Oldham Joint, 26a balancer, 26b balancer, 29 oil return pipe, 2 a Kaeaburaana, 40 volts, 100 scroll compressor, A compression mechanism unit, B motor unit.

Claims

A compression mechanism disposed in a sealed container;
An electric motor that drives the compression mechanism;
A drive shaft that transmits the driving force of the electric motor unit to the compression mechanism unit;
A main bearing that supports an upper portion of the drive shaft;
A sub-bearing supporting the lower portion of the drive shaft;
A radial receiving surface provided on the auxiliary bearing and slidably supporting a radial surface of the drive shaft;
A thrust receiving surface that slidably supports the thrust surface of the drive shaft,
Between the thrust surface and the thrust receiving surface, there is a contact point where the thrust surface and the thrust receiving surface are in contact with each other on a curved surface, and a gap continuously increasing from the contact point toward the radially outer side. Compressor.
The compressor according to claim 1, wherein an oil sump space for storing lubricating oil is formed on a radially outer side of the thrust receiving surface.
The sealed container has an oil storage space inside,
An oil pump connected to a pump insertion shaft formed on the drive shaft at a lower portion of the sub-bearing, and sucking up 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 between the thrust surface and the thrust receiving surface;
The compressor according to claim 2, wherein the lubricating oil sucked up by the oil pump and passed through the radial oil groove is stored in the oil sump space.
A sub-frame disposed in the sealed container and provided with the sub-bearing;
The oil pump is attached to the subframe, and the oil in the oil sump space is supplied to the oil by a groove and a hole formed in one or both of the attachment surfaces of the subframe and the oil pump. The compressor according to claim 3, wherein an oil discharge passage for discharging from an outlet is formed in the storage space, and a throttle passage serving as a passage resistance is formed in the middle of the oil discharge passage.
The compressor according to claim 4, wherein an outlet of the oil discharge passage is disposed below the oil sump space, and an upper end of the oil sump space is disposed above the radial oil groove.
The drive shaft includes an eccentric shaft that is eccentric from an axis of the drive shaft at an upper end,
The radial oil groove is formed so as to extend in an eccentric direction that is a direction connecting the axis of the eccentric shaft from the axis of the drive shaft. Compressor.
The compressor according to claim 6, wherein the radial oil groove is disposed on the thrust surface of the drive shaft in a range from 0 deg to 45 deg from the eccentric direction to the rotation direction of the drive shaft.
The annular member which has the through-hole in which the said pump insertion shaft is inserted in the center part is arrange | positioned under the said thrust surface, The said thrust receiving surface is formed in the upper surface side of the said annular member. The compressor according to any one of claims 7 to 9.
The compressor according to claim 8, wherein the annular member is an outer member of the oil pump.
The compressor according to claim 8, wherein the annular member is a bottom plate fixed to a lower side of the sub frame provided with the sub bearing.
11. The oil sump space is formed by forming a radially outer side of the thrust receiving surface of the annular member as an inclined surface that is inclined downward as it goes outward. The compressor according to item.
The compressor according to any one of claims 8 to 11, wherein an inner peripheral portion of a thrust receiving surface formed by the through hole on the upper surface side of the annular member is a chamfered portion having a curvature radius Rp0. .
The compressor according to claim 8, wherein the annular member is obtained by polishing a surface of a hardened steel strip.
The compressor according to any one of claims 3 to 13, wherein a plurality of the radial oil grooves are formed on the thrust receiving surface.
The oil sump space is formed by forming a radially outer side of the thrust surface of the drive shaft as an inclined surface that is inclined upward as it goes outward. The compressor described in 1.
A chamfered portion that is in contact with the thrust receiving surface at the contact point is formed on a radially outer peripheral portion of the thrust surface of the drive shaft, and a radius of curvature Rs1 of the chamfered portion is within an arc angle range smaller than 90 degrees. The compressor according to any one of claims 1 to 15, wherein the compressor is formed.
The radius of curvature Rs1 of the chamfered portion in the radially outer peripheral portion of the thrust surface is:
The compressor according to claim 16, wherein the compressor has a relationship that is larger than a radius of curvature Rp0 on the inner peripheral side of the thrust receiving surface, that is, a relationship of Rs1> Rp0.
The drive shaft includes
An oil supply vertical hole extending in the axial direction of the drive shaft and through which lubricating oil flows;
A thrust oil lateral hole that extends in the radial direction of the drive shaft and communicates with the oil supply vertical hole below the thrust surface and supplies the lubricating oil between the thrust surface and the thrust receiving surface is formed. The compressor according to any one of claims 1 to 17.
The drive shaft includes
A radial oil supply lateral hole that communicates with the oil supply vertical hole above the thrust surface and extends in the radial direction of the drive shaft, and supplies the lubricating oil between the radial surface and the radial receiving surface;
An axial oil groove formed in the outer peripheral surface of the drive shaft in communication with the radial oil supply lateral hole and extending in the axial direction of the drive shaft;
The axial oil groove is opposed to the radial receiving surface, and a region above the axial oil groove is a lower region in a region facing the radial receiving surface on an outer peripheral surface of the drive shaft. The compressor according to claim 18, wherein the length of the drive shaft in the axial direction is longer than that of the compressor.
A compression mechanism disposed in a sealed container;
An electric motor that drives the compression mechanism;
A drive shaft that transmits the driving force of the electric motor unit to the compression mechanism unit;
A main bearing that supports an upper portion of the drive shaft;
A sub-bearing supporting the lower portion of the drive shaft;
A thrust receiving surface that slidably supports the thrust surface of the drive shaft;
A radial receiving surface that slidably supports the radial surface of the drive shaft,
The drive shaft includes
An oil supply vertical hole that extends in the axial direction of the drive shaft and through which the lubricating oil flows,
A thrust oil supply lateral hole that communicates with the oil supply vertical hole below the thrust surface, extends in the radial direction of the drive shaft, and supplies the lubricating oil between the thrust surface and the thrust receiving surface is formed. Compressor.