WO2016143186A1

WO2016143186A1 - Compressor comprising slide bearing

Info

Publication number: WO2016143186A1
Application number: PCT/JP2015/079849
Authority: WO
Inventors: 辰也佐々木; 英人中尾; 祐司 ▲高▼村; 勝紀佐藤
Original assignee: 三菱電機株式会社
Priority date: 2015-03-12
Filing date: 2015-10-22
Publication date: 2016-09-15
Also published as: JP6328322B2; JPWO2016143186A1

Abstract

A compressor comprising a slide bearing comprises: a shaft; a slide bearing that pivotally supports the shaft to rotate freely; and an oil supplying structure that has a pump that is disposed on one end of the shaft and operates by way of shaft rotation and that supplies oil to the slide bearing by passing the oil from the pump through the inside of the shaft. The slide bearing comprises a bearing into which the shaft is rotatably inserted, and a first sliding member that is inserted into a groove formed on the inner circumference of the bearing into which the shaft is inserted or a groove formed on the outer circumference of the shaft and for which the amount of protrusion above the groove increases due to a rise in temperature by having a linear expansion coefficient greater than the bearing.

Description

Compressor with slide bearing

The present invention relates to a compressor having a sliding bearing, and more particularly to a wide range drive scroll compressor that operates from a low speed to a high speed.

The rotating machine is provided with a bearing that supports a load acting on the rotating body. There are two types of bearings: rolling bearings and sliding bearings. A slide bearing has a structure in which an oil film is formed between a shaft and a bearing, and an oil film pressure is generated by the effect of a wedge to support a bearing load. In general, oil film pressure is likely to occur when the rotational speed is high. When oil film pressure is generated, the shaft and the bearing are separated by the oil film and can avoid contact with each other. Does not occur. On the other hand, when the rotational speed is small, the oil film pressure is less likely to be generated, and the oil film thickness becomes thin.

By the way, an air conditioner equipped with a refrigerant compressor is often selected according to the maximum air conditioning load of a space (room) to be air-conditioned. During actual operation (actual operation), the operation in a low load region increases due to the effect of high heat insulation of the building and the heat generated from the internal equipment installed in the air-conditioned room. . In addition, cooling operation with low capacity is required even at low outside temperatures.

For this reason, the compressor used for the air conditioner increases the frequency of operation in the low-speed rotation region. In addition, from the past, we have developed a compressor that can be operated with high efficiency during actual operation by reducing the stroke volume of the compressor and making it a wide range drive that can operate from low speed to ultra high speed equivalent to the maximum load equivalent to the conventional machine. Efforts are being made.

Furthermore, in order to further improve the efficiency in the low load region, ultra-low speed operation lower than the minimum rotation speed (for example, 30 Hz) in the conventional compressor is realized, and friction loss at a sliding portion such as a bearing during low speed operation is achieved. Efforts to reduce this are also being made.

However, in a conventional compressor using a sliding bearing, when operating at a low speed, the effect of generating dynamic pressure in the gap between the shaft and the bearing is reduced, and as a result, the oil film thickness is reduced, resulting in a mixed lubrication region. Easy to transition to. For this reason, the friction coefficient in a bearing rises and there exists a subject that a friction loss cannot be reduced. It is also necessary to improve seizure resistance and wear resistance of the bearing sliding member.

Therefore, what has been proposed to reduce the friction coefficient and improve the wear resistance of the sliding members inside the compressor. In order to ensure the oil film thickness and to operate with fluid lubrication, for example, Patent Document 1 describes that a resin-made dynamic pressure groove is formed on the inner surface of a cylindrical hole of a bearing member into which a shaft is inserted. Yes. Further, in Patent Document 2, a resin having a higher linear expansion coefficient than that of the bearing member is embedded in the inner surface of the cylindrical hole of the bearing member, and the resin is formed into a cylindrical shape due to frictional heat caused by sliding between the shaft and the bearing member. It is described that the sliding can be facilitated by the solid lubrication effect of the resin protruding from the inner surface of the hole.

JP 2013-253650 A Japanese Utility Model Publication No. 58-193122

In the technique described in Patent Document 1, the shaft rotation speed (operation frequency) is not mentioned, and the possibility that the oil film thickness changes between low speed operation and high speed operation is not considered. That is, for example, assuming that the compressor is operated in an ultra-low speed operation range of less than 30 Hz, the oil film thickness is maintained by the action of the dynamic pressure groove, and it is designed to be able to operate with fluid lubrication. The following inconveniences occur during high-speed operation. That is, if the dynamic pressure groove is designed for low-speed operation, the oil film thickness becomes too thick in high-speed operation, and there is a possibility that bearing loss due to the viscous resistance (shear resistance) of the oil film increases. That is, as a result of bearing design adapted to low speed operation, inconvenience occurs in high speed operation.

Compressors are required to ensure fluid lubrication in a wide operating range from low speed to high speed. Conversely, when bearings are designed for high speed operation, seizure occurs due to insufficient oil film pressure at low speed. Is concerned.

Further, in the technique described in Patent Document 2, sliding is facilitated by the solid lubricating effect of the resin. However, since it is a non-lubricated bearing that does not use oil, the coefficient of friction is assumed to be approximately 0.05. Usually, oil is used for bearings arranged in the compressor, and the coefficient of friction is about 0.005. The coefficient of friction due to the solid lubrication effect of the resin is much higher than when oil is used, and an increase in friction loss when oil is not used is inevitable. Moreover, since the technique described in Patent Document 2 does not use oil, cooling of the bearing portion cannot be expected, and frictional heat rises at the bearing portion, which may cause seizure or abnormal wear.

This invention solves the said subject, and aims at obtaining the compressor which has a slide bearing which can ensure low loss and high reliability in a wide operating range.

A compressor having a slide bearing according to the present invention includes a shaft, a slide bearing that rotatably supports the shaft, and a pump that is installed at one end of the shaft and operates by rotation of the shaft. An oil supply structure that passes through the shaft and supplies oil to the slide bearing, and the slide bearing is formed in a bearing portion in which the shaft is rotatably inserted and an inner peripheral portion in which the shaft is inserted in the bearing portion. A first sliding member that is inserted into a recess or a recess formed on the outer periphery of the shaft, has a linear expansion coefficient larger than that of the bearing portion, and increases in height protruding from the recess due to temperature rise. It is.

According to the present invention, it is possible to obtain a compressor having a sliding bearing capable of ensuring low loss and high reliability in a wide operating range.

It is a longitudinal cross-sectional view which shows typically the structure of the scroll compressor in connection with Embodiment 1 of this invention. It is a cross-sectional view of the plain bearing in connection with Embodiment 1 of this invention. It is a figure which shows the state which removed the sliding member from the slide bearing in connection with Embodiment 1 of this invention, (a) is A arrow directional view of FIG. 2, (b) is a cross-sectional expanded view. It is a figure which shows the slide bearing in connection with Embodiment 1 of this invention, (a) is A arrow directional view of FIG. 2, (b) is a cross-sectional expanded view. It is a cross-sectional view of a state in which a shaft is inserted into a general plain bearing. It is the graph which showed the influence which the rotation speed of a shaft exerts on a friction coefficient in a general slide bearing. It is the schematic diagram which showed the operation | movement form of the plain bearing in connection with Embodiment 1 of this invention. FIG. 8 is an enlarged perspective view of a portion surrounded by an alternate long and short dash line in the lower left diagram of FIG. 7. It is a figure which shows the graph which shows the influence which a dynamic pressure groove has on a friction coefficient according to the rotation speed of the scroll compressor in connection with Embodiment 1 of this invention, and the graph which shows the temperature rise according to rotation speed. It is a figure which shows the modification of the plain bearing in connection with Embodiment 1 of this invention. It is a figure which shows the modification of the slide bearing in connection with Embodiment 1 of this invention, (a) is a longitudinal cross-sectional view of a slide bearing, (b) is a cross-sectional expanded view of a slide bearing. FIG. 12 is a cross-sectional development view of the sliding bearing and the shaft when the sliding member of the sliding bearing of FIG. 11 protrudes. It is a figure which shows the slide bearing in connection with Embodiment 2 of this invention, (a) is a longitudinal cross-sectional view of a slide bearing, (b) is a cross-sectional expanded view of a slide bearing. It is a figure which shows the slide bearing in connection with Embodiment 3 of this invention, (a) is a longitudinal cross-sectional view of a slide bearing, (b) is a cross-sectional expanded view of a slide bearing. It is a longitudinal cross-sectional view of the plain bearing in connection with Embodiment 4 of this invention.

Embodiment 1 FIG.
FIG. 1 is a longitudinal sectional view schematically showing a configuration of a scroll compressor according to Embodiment 1 of the present invention.
The scroll compressor 100 has a function of sucking a fluid such as a refrigerant, compressing it, and discharging it in a high temperature and high pressure state. The scroll compressor 100 includes a compression mechanism unit 10 and an electric motor 20 that drives the compression mechanism unit 10 via the main shaft 6 and other components, and these are housed in an airtight container 13 that forms an outer shell. It has a configuration. A lower part of the sealed container 13 is an oil sump 14 for storing lubricating oil.

In the sealed container 13, an upper housing 8a and a lower housing 8b are further arranged. The upper housing 8 a is disposed on the lower side of the compression mechanism unit 10 and is positioned between the compression mechanism unit 10 and the electric motor 20, and the lower housing 8 b is positioned on the lower side of the electric motor 20. The upper housing 8a and the lower housing 8b are fixed to the inner peripheral surface of the sealed container 13 by shrink fitting or welding. A through hole is provided at the center of the upper housing 8a and the lower housing 8b, and the main shaft 6 is rotatably supported by a main bearing 19 and a sub bearing 11 provided in the through hole.

The closed container 13 is provided with a refrigerant suction pipe 15 for sucking fluid and a refrigerant discharge pipe 16 for discharging fluid.

The compression mechanism section 10 has a function of compressing the fluid sucked from the refrigerant suction pipe 15 and discharging it to the high-pressure space 13 a formed above the sealed container 13. The compression mechanism unit 10 includes a fixed scroll 1 and a swing scroll 2. The fixed scroll 1 is fixed to the hermetic container 13 via an upper housing 8a disposed on the lower side. The swing scroll 2 is disposed between the fixed scroll 1 and the upper housing 8a, and is supported on the main shaft 6 so as to be swingable.

The fixed scroll 1 includes a first base plate 1a and a first spiral protrusion 1b erected on one surface of the first base plate 1a. The orbiting scroll 2 includes a second base plate 2a and a second spiral protrusion 2b erected on one surface of the second base plate 2a. The fixed scroll 1 and the orbiting scroll 2 are mounted in the hermetic container 13 with the first spiral protrusion 1b and the second spiral protrusion 2b meshing with each other. A compression chamber 5 is formed between the first spiral protrusion 1b and the second spiral protrusion 2b, the volume of which decreases as the main shaft 6 rotates inward in the radial direction. In the fixed scroll 1, a suction port 3 for supplying the refrigerant sucked from the refrigerant suction pipe 15 into the compression chamber 5 is formed on the outer peripheral side of the compression chamber 5. In addition, a discharge port 4 is formed at the center of the fixed scroll 1 to discharge the compressed and high pressure fluid.

The orbiting scroll 2 performs an eccentric turning motion without rotating with respect to the fixed scroll 1. A hollow cylindrical rocking bearing 17 that receives a driving force is disposed at a substantially central portion of a surface (hereinafter referred to as a thrust surface) opposite to the surface on which the second spiral protrusion 2b is formed of the rocking scroll 2. ing. An eccentric shaft 6a (described later) provided at the upper end of the main shaft 6 is fitted (engaged) with the rocking bearing 17. The orbiting scroll 2 is supported by the upper housing 8a through a thrust bearing 18 so as to be slidable from below.

The electric motor 20 includes a rotor 20a and a stator 20b. The stator 20b has a function of rotating the rotor 20a when energized. The stator 20b is formed in a substantially cylindrical shape, and its outer peripheral surface is fixedly supported by the closed container 13 by shrink fitting or the like. The rotor 20a is fixed to the outer periphery of the main shaft 6, has a permanent magnet inside, and is rotatably held inside the stator 20b with a slight gap from the stator 20b. The rotor 20a is rotationally driven when the stator 20b is energized to rotate the main shaft 6. That is, rotation power is transmitted to the compression mechanism 10 via the main shaft 6 by the rotation of the rotor 20a.

The main shaft 6 is rotatably supported by a main bearing 19 provided on the upper housing 8a on the upper side and rotatably supported by a sub bearing 11 provided on the lower housing 8b on the lower side. Further, the main shaft 6 has an eccentric shaft 6 a at the upper end, and the eccentric shaft 6 a is fitted with the rocking bearing 17 of the rocking scroll 2, and the rocking scroll 2 is eccentrically swung by the rotation of the main shaft 6. ing. In addition, a balancer 6b protrudes from the upper portion of the main shaft 6, and an oil supply hole 7a is formed through the main shaft 6 in the axial direction. A pump 7 b is installed at the lower end of the main shaft 6. The pump 7b has an upper surface opening fitted into the lower end portion of the main shaft 6 and a lower end opening immersed in the lubricating oil in the oil sump 14. The oil stored in the oil sump 14 is sucked upward to the oil supply hole 7a. Supply. The oil supply mechanism 7 is constituted by the oil supply hole 7a and the pump 7b.

In the hermetic container 13, an Oldham coupling 12 for preventing the swing scroll 2 from rotating during the eccentric orbiting motion of the swing scroll 2 is disposed between the swing scroll 2 and the upper housing 8a. . The upper housing 8a is connected to an oil return pipe 9 that returns the lubricating oil supplied to the Oldham coupling 12 to the oil sump 14.

Next, the operation will be described. When the main shaft 6 rotates together with the rotor 20 a of the electric motor 20, the orbiting scroll 2 performs a revolving motion while being prevented from rotating by the Oldham joint 12. The gas refrigerant sucked into the sealed container 13 from the refrigerant suction pipe 15 passes through the suction port 3 provided in the fixed scroll 1, and the first spiral protrusion 1 b of the fixed scroll 1 and the second spiral protrusion 2 b of the swing scroll 2. Are taken into the compression chamber 5 formed between the two. And the compression chamber 5 which took in the gas refrigerant | coolant gradually reduces a volume, moving from an outer peripheral part to a center side with the revolution motion of the rocking scroll 2, and compresses refrigerant gas. The compressed refrigerant gas is pumped into the refrigerant pipe outside the apparatus through the discharge port 4 and 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 becomes negative pressure, and the refrigerant from the refrigerant pipe outside the machine is sucked through the refrigerant suction pipe 15 to cool the electric motor 20. Thereafter, the air is sucked into the compression chamber 5 from the suction port 3.

Further, the lubricating oil in the oil sump 14 is sent to the upper end portion of the main shaft 6 through the oil supply hole 7 a by the pump action of the oil supply mechanism 7 to lubricate the sub bearing 11, the main bearing 19, and the rocking bearing 17. Here, the pump 7b is a positive displacement type in which the oil supply flow rate increases in proportion to the rotational speed of the main shaft 6. When the positive displacement type is used, the higher the number of revolutions, the greater the oil supply flow rate, so the bearing cooling effect increases. Although a positive displacement pump is preferably used as the pump 7b from the viewpoint of the bearing cooling effect, the pump 7b is not limited to the positive displacement pump and may be a non-displacement pump. Lubricating oil that has passed through the rocking bearing 17 is supplied to the main bearing 19 and the Oldham coupling 12 to lubricate the sliding portions. The lubricating oil supplied to the Oldham coupling 12 is returned to the oil sump 14 through the oil return pipe 9.

Next, the rocking bearing 17 and the main bearing 19 will be described. Slide bearings are used for the rocking bearing 17 and the main bearing 19, and when lubricating oil is supplied to the slide bearing, an oil film is formed in the gap between the shaft (main shaft 6, eccentric shaft 6a) and the slide bearing. Is done. The shaft and the slide bearing are separated by the oil film, and the slide bearing supports a load acting on the shaft in a non-contact manner.

FIG. 2 is a cross-sectional view of the plain bearing according to the first embodiment of the present invention. FIGS. 3A and 3B are views showing a state in which the sliding member is removed from the slide bearing according to the first embodiment of the present invention, where FIG. 3A is a view as viewed in the direction of arrow A in FIG. 2 and FIG. . The development view shows a state in which a cross section cut at the center in the width direction of the bearing (vertical direction in FIG. 3A) is developed from a position of 0 ° to a plane. 4A and 4B are diagrams showing the plain bearing according to the first embodiment of the present invention, in which FIG. 4A is a view taken in the direction of arrow A in FIG. 2 and FIG.

The slide bearing 30 includes a bearing portion 31 having a cylindrical inner peripheral portion into which a shaft is rotatably inserted, and a sliding member 32. The bearing 31 has a plurality of V-shaped recesses 31a arranged in parallel in the circumferential direction on the inner peripheral surface as shown in the arrow A view. The height of the recessed part 31a is set to h. A sliding member (first sliding member) 32 having the same shape as the concave portion 31a is inserted into the concave portion 31a.

The sliding member 32 has a height h similar to the depth of the recess 31a, and the inner peripheral surface of the slide bearing 30 is smooth without any step as shown in FIG. That is, the inner peripheral surface of the slide bearing 30 is flush with the concave portion 31a including the portion where the sliding member 32 is installed. The sliding member 32 is made of a material having a higher linear expansion coefficient than the bearing portion 31. The material of the bearing portion 31 is, for example, metal, and the sliding member 32 is, for example, a resin having a higher linear expansion coefficient than that of metal. By applying the resin sliding member 32 having a solid lubricating effect, there is an advantage that a low friction coefficient can be maintained even when the shaft and the sliding member 32 are in contact with each other.

Here, prior to explaining the operation of the sliding bearing 30 of the first embodiment shown in FIGS. 2 to 4, the state of sliding during rotation of a general sliding bearing will be described.

FIG. 5 is a cross-sectional view showing a state in which a shaft is inserted into a general plain bearing.
Assuming that the shaft 40 rotates counterclockwise in FIG. 5, the lubricating oil filled in the gap between the shaft 40 and the slide bearing 41 moves along the rotation of the shaft 40 as indicated by the arrow. On the inner peripheral surface, it is drawn from the position of the A ° phase toward the position of the B ° phase. Here, the flow path of the lubricating oil drawn from the position of the A ° phase toward the position of the B ° phase has a wedge shape in which the cross-sectional area of the flow path gradually narrows, and therefore an oil film pressure is generated. The oil film pressure due to the wedge shape increases as the rotational speed of the shaft 40 increases.

On the other hand, when the rotational speed is reduced, the oil film pressure is reduced. For this reason, when the rotational speed is small, the load acting on the slide bearing 41 cannot be supported, the oil film thickness at the B ° phase position becomes thin, and the shaft 40 comes into contact with the slide bearing 41.

FIG. 6 is a graph showing the influence of the rotational speed of the shaft on the friction coefficient in a general plain bearing. In FIG. 6, the horizontal axis represents the rotational speed and the vertical axis represents the friction coefficient.
Generally, in the sliding bearing 41, when the shaft 40 and the sliding bearing 41 are separated by an oil film, only the viscous resistance of the lubricating oil acts as the frictional force of the shaft 40, and the friction coefficient is 0.001 to 0.009. Is in range. However, when the shaft 40 and the slide bearing 41 come into contact with each other at a low rotational speed and a B ° phase position, the friction coefficient of the solid contact is generally about 0.01 to 0.1, which is a friction coefficient due to viscous resistance. Compared to larger. Here, the friction loss L [W] is expressed by Equation 1.
[Equation 1]
L = μFV (Formula 1)

Here, μ is a friction coefficient, F is a load [N], and V is a sliding speed [m / s].

Thus, when the rotational speed is small, the shaft 40 comes into contact with the slide bearing 41 and heat is generated due to friction loss. Such heat generation raises the temperature of the lubricating oil, the shaft 40, and the slide bearing 41, and causes seizure.

Therefore, in the first embodiment, in order to suppress seizure when the rotational speed is small, that is, when operating at a low speed, the concave portion 31a is provided on the inner peripheral portion of the bearing portion 31 of the slide bearing 30 as described above. In addition, the sliding member 32 having a linear expansion coefficient higher than that of the bearing portion 31 is disposed in the recess 31a. For example, less than 60 rps corresponds to low speed operation, and 60 to 120 rps corresponds to high speed operation.

Since the sliding member 32 is formed of a material having a larger linear expansion coefficient than the bearing portion 31 as described above, if frictional heat is generated by contact with the shaft during operation, the temperature rises due to the frictional heat. The behavior which expand | swells by this and protrudes in the internal peripheral surface direction of the slide bearing 30 from the recessed part 31a is shown. This behavior will be described in detail below with reference to FIGS.

FIG. 7 is a schematic diagram showing an operation mode of the plain bearing according to the first embodiment of the present invention. In FIG. 7, a hollow portion at the center of the shaft 60 (the main shaft 6 and the eccentric shaft 6a) and a portion extending in the radial direction from the hollow portion indicate oil supply holes. Moreover, the part shown with the thin dot in FIG. 7 has shown lubricating oil. FIG. 8 is an enlarged perspective view of a portion surrounded by an alternate long and short dash line in the lower left diagram of FIG. Here, the rotation direction of the shaft 60 is limited to a direction in which the V-shape of the recess 31a is tapered (a direction from left to right in FIG. 8). Therefore, an oil flow is generated in a direction in which the V-shape is tapered.

When the operation is stopped, the sliding member 32 is not deformed, and the inner peripheral surface of the slide bearing 30 is kept smooth.

(Low speed operation)
As shown in FIG. 7 (a), at the beginning of the low speed operation, the flow rate of the lubricating oil is small and the rotational speed of the shaft 60 is small. Arise.

Here, as shown in FIG. 7A, when the shaft 60 comes into contact with the slide bearing 30 in the vicinity of the B ° phase, the sliding member 32 disposed in the vicinity of the B ° phase is moved to the position shown in FIG. And as shown in FIG. 8, it expands and protrudes by frictional heat. When the sliding member 32 protrudes, the portion where the sliding member 32 does not protrude becomes a groove, and this groove functions as a dynamic pressure groove. That is, when the sliding member 32 protrudes, first, the surface of the sliding member 32 comes into contact with the shaft 60. At the same time, as shown by the dotted arrow in FIG. Lubricating oil flows toward the center in the width direction. Due to the flow of the lubricating oil, an oil film pressure is generated at the center of the slide bearing 30 in the width direction, and a load acting on the shaft 60 can be supported. For this reason, an oil film is formed between the sliding member 32 and the shaft 60 so as not to contact each other, and the shaft 60 can be slid by fluid lubrication.

For example, when the thermal expansion coefficient of the sliding member 32 is 10 × 10 ⁻⁵ [/ K] and the height h is 1 [mm], if the temperature rise due to frictional heat is 50 [K], the sliding member The protrusion height of 32 is 5 [μm]. If the design is such that the protruding height of the sliding member 32 is 2 to 10 [μm], the shaft 60 can be slid by fluid lubrication. Further, immediately after the sliding member 32 protrudes, it contacts the shaft 60, but wear is suppressed by the solid lubricating effect of the resin. Examples of the resin having a solid lubricating effect include fluorine, POM, PPS, and the like. The sliding

members

33, 34, and 35 to be described later can be made of the same material. Examples of the material for the bearing portion 31 include iron, copper, and aluminum.

As described above, the sliding member 32 expands and protrudes from the inner peripheral surface of the bearing portion 31 toward the shaft 60 under the condition where the shaft 60 and the slide bearing 30 are in contact with each other at low speed operation, and a dynamic pressure groove appears. Can be made. An oil film is formed between the sliding member 32 and the shaft 60 by the action of the dynamic pressure groove, and an increase in friction coefficient, wear and seizure can be suppressed.

(High speed operation)
In high-speed operation, the flow rate of the lubricating oil is large, the rotational speed of the shaft 60 is large, and the oil film pressure is generated due to the effect of the wedge-shaped flow path. For this reason, the contact between the shaft 60 and the plain bearing 30 can be avoided, and high frictional heat is not generated. Therefore, the amount that the sliding member 32 expands due to frictional heat is negligible, and the height of the sliding member 32 remains the same as or slightly higher than the height of the recess 31a. That is, the inner peripheral surface of the slide bearing 30 is formed to have a smooth surface or a slight dynamic pressure groove, and the oil film pressure is generated due to the wedge effect due to the eccentricity of the shaft 60 inside the slide bearing. Can be supported. Further, since the inner peripheral surface of the slide bearing 30 is smooth or slightly formed with a dynamic pressure groove, unnecessary viscous resistance as generated in the conventional dynamic pressure groove can be reduced, and the bearing loss is kept low. it can. FIG. 9 shows the above relationship in a graph.

FIG. 9 is a graph showing the effect of the dynamic pressure groove on the friction coefficient according to the rotation speed of the scroll compressor according to the first embodiment of the present invention, and a graph showing the temperature increase according to the rotation speed. It is. In FIG. 9, the horizontal axis indicates the rotational speed [rps], the left vertical axis indicates the friction coefficient [μ], and the right vertical axis indicates the temperature increase ΔT [K]. 9, (a) to (c) are graphs showing the relationship between the rotational speed of the shaft and the friction coefficient, (a) is the case of the first embodiment, (b) is the case of having the dynamic pressure groove, c) shows the case without a dynamic pressure groove. In FIG. 9, (d) is a graph showing the temperature rise ΔT according to the rotational speed of the shaft.

In low-speed operation, a dynamic pressure groove is formed and an oil film pressure is generated. Therefore, when there is a dynamic pressure groove (FIGS. 9A and 9B), compared to the case without the dynamic pressure groove (FIG. 9C). ) Can maintain a low coefficient of friction from about 15 [rps] to a further low speed operation (arrow A in FIG. 9). In addition, the dynamic pressure groove is not generated or the height is low in high-speed operation. Therefore, the friction coefficient is lower when there is no dynamic pressure groove (FIG. 9C) than when there is a dynamic pressure groove (FIGS. 9A and 9B) (FIG. 9). Arrow B).

In addition, in this Embodiment 1, although the shape of the recessed part 31a and the sliding member 32 was made into V shape, if it is a shape which can express a dynamic pressure effect, it will not be limited to V shape. FIGS. 10 and 11 below show other shape examples of the recess 31a and the sliding member 32. FIG.

FIG. 10 is a view showing a modification of the plain bearing according to the first embodiment of the present invention.
FIG. 10A shows an example in which the shape of the recess 31a is an arc shape protruding in the rotation direction of the shaft. FIG. 10B shows a pair of recesses 31a with the axial center portion of the bearing portion 31 as a boundary, and rotation of the shaft from the axial end portions of the bearing portion 31 toward the central portion. An example is shown in the form of a diagonal line inclined in the direction. In both FIG. 10A and FIG. 10B, the sliding member 32 having the same shape is inserted into the recess 31a. Even if the shape of the sliding member 32 is the shape shown in FIG. 10A and FIG. 10B, the same effect as the case where it is V-shaped can be obtained.

FIG. 11 is a view showing a modification of the slide bearing according to the first embodiment of the present invention, wherein (a) is a longitudinal sectional view of the slide bearing, and (b) is a developed cross-sectional view of the slide bearing. FIG. 12 is a developed cross-sectional view of the sliding bearing and the shaft when the sliding member of the sliding bearing of FIG. 11 protrudes.

11B, the height h1 on the near side with respect to the rotation direction of the shaft 60 in the recess 31a and the sliding member 32 is the height on the back side with respect to the rotation direction of the shaft 60. Lower than h2. With this configuration, when the sliding member 32 expands and protrudes from the inner peripheral surface of the bearing portion 31 toward the shaft 60 side, as shown in FIG. The protrusion height from the inner peripheral surface of the bearing is higher than the front side. With this configuration, the flow path of the lubricating oil is narrowed by the gap between the sliding member 32 and the shaft 60, so that a large oil film pressure can be generated. In a state where the shaft 60 is not rotating (when not rotating), the inner peripheral surface of the bearing portion 31 includes the portion where the sliding member 32 is installed in the recess 31a as shown in FIG. 11B. It is one.

As described above, according to the first embodiment, the sliding member 32 having a linear expansion coefficient larger than that of the bearing portion 31 is inserted into the concave portion 31 a provided in the inner peripheral portion of the bearing portion 31. As the temperature of the member 32 increases, the dynamic pressure groove appears by the expansion of the sliding member 32 so that the dynamic pressure effect is enhanced.

For this reason, in low-speed operation, no oil film pressure is generated at the beginning of operation, the shaft 60 and the bearing 31 come into contact with each other, the frictional heat increases, and the temperature of the sliding member 32 rises. As a result, it is possible to operate with fluid lubrication due to the effect of the dynamic pressure groove, and it is possible to reduce bearing loss and to suppress bearing wear and seizure.

In high-speed operation, the amount of lubricating oil supplied from the pump 7b increases, and heat generated by friction can be suppressed due to the cooling effect of the lubricating oil, so that the deformation amount of the sliding member 32 can be suppressed only slightly. Further, in high-speed operation, the inner peripheral surface of the slide bearing 30 is smooth or slightly protrudes from the sliding member 32, and an oil film pressure is generated due to the wedge effect due to the shaft 60 being eccentric in the slide bearing. Can be supported.

In high-speed operation, the inner peripheral surface of the slide bearing 30 is smooth or the sliding member 32 protrudes only slightly, so that unnecessary viscous resistance caused by the dynamic pressure groove can be reduced and the bearing loss can be kept small. Although not shown, the same effect can be obtained by providing the recess 31 a on the outer periphery of the shaft 60 instead of the inner periphery of the bearing 31 and attaching the sliding member 32 to the recess 31 a provided on the outer periphery of the shaft 60. Is obtained.

From the above, it is possible to obtain the plain bearing 30 capable of ensuring low loss and high reliability in a wide operating range.

Further, the material of the sliding member 32 is resin, and the coefficient of friction when the shaft 60 contacts the sliding member 32 is smaller than that of metal, so that a dynamic pressure effect is generated in low speed operation and the process shifts to fluid lubrication. In the mixed lubrication state up to, friction loss can be reduced.

Embodiment 2. FIG.
In the second embodiment, the height of the sliding member is different from that of the first embodiment. Hereinafter, the difference between the second embodiment and the first embodiment will be mainly described. Note that the modification applied to the same components as those in the first embodiment is similarly applied to the second embodiment.

FIGS. 13A and 13B are diagrams showing a slide bearing according to Embodiment 2 of the present invention, in which FIG. 13A is a longitudinal sectional view of the slide bearing, and FIG. 13B is a developed cross-sectional view of the slide bearing. FIG. 13 shows a state in which the shaft 60 is not rotating. In this state, the sliding member (first sliding member) 33 according to the second embodiment is higher than the inner peripheral surface of the slide bearing 30. This point is different from the first embodiment. And the part which the sliding member 33 does not protrude becomes a groove | channel, and this groove | channel becomes a dynamic pressure groove. Therefore, in the normal operation range of the compressor, for example, in the operation of 60 to 120 rps corresponding to the high speed operation, the slide bearing 30 has the effect of a dynamic pressure groove. Is possible. In addition, the protrusion height h3 of the sliding member 33 from the inner peripheral surface of the sliding bearing 30 in a state where the shaft 60 is not rotated forms an appropriate oil film thickness due to the effect of the dynamic pressure groove during high speed operation. The height is set so that the oil film does not become too thick to increase bearing loss.

On the other hand, in low speed operation (for example, less than 60 rps), the oil film thickness becomes thinner compared to during high speed operation at the beginning of operation, and the shaft 60 contacts the sliding member 32 installed on the inner peripheral portion of the slide bearing 30. Then move. Friction heat is generated by this sliding, and the height of the sliding member 32 is increased by the expansion of the sliding member 32 caused by the frictional heat. In other words, the height of the dynamic pressure groove is higher after the initial operation start of the low speed operation than in the high speed operation. Therefore, in the low speed operation, the effect of the dynamic pressure groove is greater than in the high speed operation, and an oil film pressure is generated, so that fluid lubrication can be maintained. This effect can be obtained in the same way even in an ultra-low speed operation of 30 rps or less, for example.

As described above, according to the second embodiment, it is possible to reduce bearing loss and ensure high reliability in a wide operating range from low speed operation to high speed operation.

Embodiment 3 FIG.
In the third embodiment, the height of the sliding member is different from that of the first embodiment as in the second embodiment, and the height of the sliding member 34 is lower than the inner peripheral surface of the slide bearing 30. Is formed. Hereinafter, the difference between the third embodiment and the first embodiment will be mainly described. Note that the modification applied to the same components as those in the first embodiment is similarly applied to the third embodiment.

14A and 14B are diagrams showing a sliding bearing according to Embodiment 3 of the present invention, in which FIG. 14A is a longitudinal sectional view of the sliding bearing, and FIG. 14B is a developed sectional view of the sliding bearing. FIG. 14 shows a state where the shaft 60 is not rotating. In this state, the sliding member (first sliding member) 34 of the third embodiment has a height lower than the inner peripheral surface of the slide bearing 30. This point is different from the first embodiment. And the part where the sliding member 34 is low becomes a groove | channel, and this groove | channel becomes a dynamic pressure groove | channel. Therefore, in the normal operation range of the compressor, for example, in the operation of 60 to 120 rps corresponding to the high speed operation, the slide bearing 30 has the effect of a dynamic pressure groove. Is possible. The recess height h4 of the sliding member 33 from the inner peripheral surface of the sliding bearing 30 in a state where the shaft 60 is not rotating is such that the sliding member 34 slightly expanded during high-speed operation is the inner periphery of the bearing portion 31. It is set so as to be flush with the surface, so that the oil film becomes too thick to increase bearing loss.

On the other hand, in low-speed operation (for example, less than 60 rps), the oil film thickness becomes thinner in the initial stage of operation than in high-speed operation, and the shaft 60 slides in contact with the inner peripheral surface of the bearing portion 31. This sliding generates frictional heat, and the sliding member 32 is expanded by the expansion of the sliding member 32 due to the frictional heat and protrudes from the inner peripheral surface of the bearing portion 31. Therefore, in the low speed operation, the effect of the dynamic pressure groove is greater than in the high speed operation, and an oil film pressure is generated, so that fluid lubrication can be maintained. This effect can be obtained in the same way even in an ultra-low speed operation of 30 rps or less, for example.

As described above, according to the third embodiment, it is possible to reduce bearing loss and ensure high reliability in a wide operation range from low speed operation to high speed operation.

Embodiment 4 FIG.
In the fourth embodiment, the sliding bearing 30 of the first embodiment is further provided with a sliding member. Hereinafter, the difference between the fourth embodiment and the first embodiment will be mainly described. Note that the modification applied to the same components as those in the first embodiment is similarly applied to the fourth embodiment.

FIG. 15 is a longitudinal sectional view of a plain bearing according to Embodiment 4 of the present invention.
The sliding bearing 30 according to the fourth embodiment is a ring-shaped sliding member (second sliding member) formed of the same material as the sliding member 32 at both ends in the axial direction of the sliding bearing 30 according to the first embodiment. 35 is inserted.

In the slide bearing 30 according to the fourth embodiment configured as described above, under the condition of, for example, 60 to 120 rps corresponding to high speed operation within the normal operation range of the compressor, the oil film pressure is caused by the effect of the wedge as in the first embodiment. And the lubricating oil supplied from the pump 7b increases, and the heat generated by friction can be suppressed by the cooling effect of the lubricating oil. For this reason, the frictional heat due to the contact between the shaft 60 and the slide bearing 30 can be kept small. Therefore, the height of the sliding member 32 is the same as or slightly higher than the height of the recess 31a. Therefore, unnecessary bearing loss that occurs in the dynamic pressure groove does not occur, and the loss can be kept low.

Further, in the initial stage of operation at low speed operation (for example, less than 60 rps), the oil film pressure is small and the oil film pressure is less likely to be generated, so the load acting on the slide bearing 30 cannot be supported, and the oil film thickness becomes thin. The shaft contacts the slide bearing 30. Both the V-shaped sliding member 32 and the ring-shaped sliding member 35 gradually protrude toward the inside of the sliding bearing 30 due to the temperature rise of the sliding bearing 30 caused by frictional heat at the time of contact.

As the sliding member 32 protrudes, a dynamic pressure groove is formed between the sliding members 32. Further, when the sliding member 35 protrudes, both ends of the dynamic pressure groove are closed by the sliding member 35. Accordingly, the amount of lubricating oil flowing out from both ends in the axial direction of the slide bearing 30 to the outside of the slide bearing is reduced. As a result, since the dynamic pressure effect can be largely ensured, fluid lubrication can be ensured in low speed operation.

As described above, according to the fourth embodiment, the same effect as in the first embodiment can be obtained, and the ring-shaped sliding members 35 are provided at both ends in the axial direction of the slide bearing 30. Further, the following effects can be obtained. That is, in low speed operation, both ends of the dynamic pressure groove are closed by the sliding member 35, and the amount of lubricating oil flowing out from the both ends in the axial direction of the slide bearing 30 to the outside of the slide bearing is reduced. As a result, since the dynamic pressure effect can be largely ensured, fluid lubrication can be ensured in low speed operation.

The material of the sliding member 35 is the same as that of the sliding member 32, but is not limited to the same material and may be a different material.

In the fourth embodiment, the configuration in which the sliding member 35 is provided to the sliding bearing 30 of the first embodiment has been described. However, the sliding movement of the sliding bearing 30 of the second and third embodiments is described. The member 35 may be provided.

As a specific configuration when the sliding member 35 is provided for the sliding bearing 30 according to the second embodiment, the sliding member 35 has an inner peripheral surface of the sliding bearing 30 in a state where the shaft 60 is not rotating. The projection height h3 (see FIG. 13B) is the same as the projection height of the sliding member 35 in the same state. As a result, the same effect as in the second embodiment can be obtained, and by providing the sliding member 35, both ends of the dynamic pressure groove are blocked by the sliding member 35 in both the low speed operation and the high speed operation. As a result, the outflow amount of the lubricating oil from both ends in the axial direction of the slide bearing 30 to the outside of the slide bearing becomes small. As a result, a large dynamic pressure effect can be ensured, so that fluid lubrication can be further ensured in both low speed operation and high speed operation compared to the case where the sliding member 35 is not provided.

Further, as a specific configuration when the sliding member 35 is provided for the sliding bearing 30 of the third embodiment, the inner periphery of the sliding bearing 30 of the sliding member 35 in a state where the shaft 60 is not rotating. The recess height h4 from the surface (see FIG. 14B) is the same as the recess height of the sliding member 33 in the same state. As a result, the same effects as in the third embodiment are obtained, and the sliding member 35 is provided, so that both ends of the dynamic pressure groove are closed by the sliding member 35 in the low speed operation. Thereby, the outflow amount of lubricating oil from both ends in the axial direction of the slide bearing 30 to the outside of the slide bearing is reduced. As a result, a large dynamic pressure effect can be ensured, so that fluid lubrication can be further ensured in the low speed operation as compared with the case where the sliding member 35 is not provided.

In the first to fourth embodiments described above, 60 to 120 rps of the normal operating range of the compressor has been described as high-speed operation, less than 60 rps as low-speed operation, and less than 30 rps as ultra-low speed operation, but this is an example. However, it is not limited to this value.

In the first to fourth embodiments, the configuration in which the compressor is a scroll compressor has been described. However, the slide bearing of the present invention can be applied to other types of compressors such as a rotary compressor. .

Further, although the embodiments 1 to 4 have been described as different embodiments, the sliding bearings may be configured by appropriately combining the characteristic configurations and modifications of the embodiments. For example, the modified example shown in FIG. 10 of the first embodiment and the fourth embodiment may be combined so that the sliding member 32 in FIG. 15 has an arc shape in FIG. 10 (a) or a hatched shape in FIG. 10 (b). Good.

1 fixed scroll, 1a first base plate, 1b first spiral protrusion, 2 swing scroll, 2a second base plate, 2b second spiral protrusion, 3 inlet, 4 outlet, 5 compression chamber, 6 main shaft, 6a eccentric Shaft, 6b balancer, 7 oil supply mechanism, 7a oil supply hole, 7b pump, 8a upper housing, 8b lower housing, 9 oil return pipe, 10 compression mechanism, 11 sub bearing, 12 Oldham joint, 13 sealed container, 13a high pressure space, 14 Oil sump, 15 Refrigerant suction pipe, 16 Refrigerant discharge pipe, 17 Rocking bearing, 18 Thrust bearing, 19 Main bearing, 20 Motor, 20a rotor, 20b Stator, 30 Sliding bearing, 31 Bearing part, 31a Recess, 32 Sliding Member, 33 sliding member, 34 sliding member, 35 sliding member, 40 shaft, 41 Bearing 60 axes, 100 scroll compressor.

Claims

The axis,
A plain bearing that rotatably supports the shaft;
It has a pump installed at one end of the shaft and operated by rotation of the shaft, and an oil supply structure that supplies oil to the slide bearing through the shaft from the pump,
The slide bearing is inserted into a bearing portion into which the shaft is rotatably inserted, a recess formed in an inner peripheral portion of the bearing portion into which the shaft is inserted, or a recess formed in an outer peripheral portion of the shaft. A compressor having a slide bearing including a first sliding member having a linear expansion coefficient larger than that of the bearing portion and increasing in height protruding from the recess due to a temperature rise.
The slide bearing according to claim 1, wherein the height of the recess and the height of the first sliding member are the same, and the inner peripheral surface of the bearing portion is flush with the shaft when the shaft is not rotating. Compressor.
Each of the height of the recess and the height of the first sliding member is configured such that the front side in the rotation direction of the shaft is lower than the back side in the rotation direction of the shaft, and the bearing is not rotated when the shaft is not rotating. The compressor having a slide bearing according to claim 1, wherein an inner peripheral surface of the portion is flush.
The height of the said 1st sliding member is comprised higher than the height of the said recessed part, and the said 1st sliding member protrudes from the internal peripheral surface of the said bearing part at the time of the non-rotation of the said axis | shaft. Compressor with plain bearing.
The height of the said 1st sliding member is comprised lower than the height of the said recessed part, and the said 1st sliding member is low from the internal peripheral surface of the said bearing part at the time of the said shaft not rotating. Compressor with plain bearing.
The compressor having a slide bearing according to any one of claims 1 to 5, wherein a material of the bearing portion is a metal, and a material of the first sliding member is a resin.
The compressor having a slide bearing according to any one of claims 1 to 6, wherein a shape of the first sliding member is a V-shape that tapers in a rotation direction of the shaft.
The compressor having a slide bearing according to any one of claims 1 to 6, wherein a shape of the first sliding member is an arc shape protruding in a rotation direction of the shaft.
The first sliding member is configured as a pair with the axial center portion of the bearing portion as a boundary, and is inclined in the rotational direction of the shaft from the axial end portion of the bearing portion toward the central portion. A compressor having a slide bearing according to any one of claims 1 to 6.
The compressor having a slide bearing according to any one of claims 1 to 9, wherein a ring-shaped second sliding member is inserted into both axial ends of the bearing portion.
The compressor having a slide bearing according to claim 10, wherein a material of the second sliding member is resin.
The compressor having a slide bearing according to any one of claims 1 to 11, wherein the pump is a positive displacement pump.