JP2014148968A - Vane type compressor - Google Patents

Abstract

PROBLEM TO BE SOLVED: To improve the reliability of bush sliding portions by enhancing the oil retaining performance of the bush sliding portions in a vane type compressor.SOLUTION: A vane type compressor 200 comprises: a nearly cylindrical cylinder 1; a rotor shaft 4 including a cylindrical rotor portion 4a rotating in the cylinder 1 and shaft portions 4b, 4c transmitting a rotating force to the rotor portion 4a; and vanes 5, 6 arranged in the rotor portion 4a and including vane end portions 5b, 6b formed in an arc shape on the outside. The rotor portion 4a is provided with bush retaining portions into which a pair of nearly semi-cylindrical bushes 7 or 8 are inserted, and the vanes 5, 6 are sandwiched between the bushes 7 or 8 and are supported so as to be rotatable and movable with respect to the rotor portion 4a. Oil grooves 7f, 7g, 8f, 8g communicating with concave portions 2a, 3a are formed on side surfaces 7b, 7e, 8b, 8e of the bushes 7, 8.

Description

  The present invention relates to a vane type compressor.

  Conventionally, a vane type compressor in which a vane is disposed in a vane groove formed in a rotor portion is known (see, for example, Patent Document 1). In the vane type compressor of Patent Document 1, when the columnar rotor portion rotates in the cylinder, the tip of the vane slides while contacting the inner peripheral surface of the cylinder. Here, since the radius of the vane tip and the radius of the inner peripheral surface of the cylinder are greatly different, an oil film is not formed between the inner peripheral surface of the cylinder and the vane tip, so that the state of boundary lubrication does not become the state of fluid lubrication. become. In general, the friction coefficient in the fluid lubrication state is about 0.001 to 0.005 in the fluid lubrication state, whereas the friction coefficient in the boundary lubrication state is approximately 0.05 or more. For this reason, when the tip of the vane and the inner peripheral surface of the cylinder slide in a boundary lubrication state, the sliding resistance is large and the compressor efficiency is reduced due to an increase in mechanical loss. In addition, the vane tip and the inner peripheral surface of the cylinder are easily worn, and it is difficult to ensure a long life.

  Thus, a vane compressor has been proposed in which the vane rotates on the outer peripheral side of the rotor portion and is held so as to be movable in the radial direction (for example, Patent Document 2 and Patent Document 3). In this vane compressor, the vane is rotatably supported at the center of the inner peripheral surface of the cylinder using a pair of cylindrical bushes with respect to the rotor portion. Thereby, since the longitudinal direction of the vane is always directed toward the center of the cylinder inner peripheral surface, the vane tip rotates along the inner peripheral surface of the cylinder. In addition, since the vane tip and the inner peripheral surface of the cylinder are operated in a non-contact manner while maintaining a minute gap, loss due to sliding at the vane tip does not occur, and the vane tip and the inner peripheral surface of the cylinder wear. Is preventing.

JP-A-10-252675 JP 2000-352390 A International Publication No. 2012/023426

  In the vane type compressors as in Patent Documents 2 and 3, the bush slides with the vane portion on the planar inner surface. The semi-cylindrical bush slides with the rotor portion on the outer surface having a curved shape. Refrigerating machine oil, which is lubricating oil, is supplied to the sliding portion so as to be in a fluid lubrication state. However, when the vane divides the space in the cylinder into two, the lubricating oil in the sliding portion is cut off due to the pressure difference between the two spaces, and the operation reliability of the vane may be impaired.

  The present invention has been made to solve the above-described problems, and provides a vane type compressor that can reliably supply lubricating oil to a sliding portion and improve the reliability of vane operation. It is intended to do.

  A vane-type compressor according to the present invention includes a cylinder having a substantially circumferential inner peripheral surface, open at both axial ends, a cylinder head and a frame closing both axial ends of the cylinder, and a cylinder A cylindrical rotor portion that rotates in the shaft and a shaft portion that transmits the rotational force to the rotor portion, the lower end of the rotor shaft that is immersed in the oil sump, and the center of the inner peripheral surface of the cylinder that is installed in the rotor portion And a vane that is held between a pair of substantially semi-cylindrical bushes in a state of being sandwiched between a pair of substantially semi-cylindrical bushes and that has a substantially cylindrical bushing penetrating in the axial direction provided in the rotor portion. The frame and the cylinder head are formed concentrically with the inner peripheral surface of the cylinder on the surface facing the rotor portion in the cylinder. Jun The bush has a recess filled with oil, and the bush has an oil groove communicating with one or both recesses of the cylinder head and the frame at a predetermined rotational phase on a side surface facing the vane or the rotor portion. It is what.

  According to the vane type compressor of the present invention, since the lubricating oil is supplied from the oil groove to the sliding portion of the side surface of the bush when the vane operates, the reliability of the operation of the vane can be ensured. .

It is a longitudinal cross-sectional view which shows Embodiment 1 of the vane type compressor of this invention. It is a disassembled perspective view of the compression element of the vane type compressor of FIG. It is a schematic diagram which shows an example of the 1st vane and 2nd vane in the vane type compressor of FIG. It is a perspective view which shows an example of the bush in the vane type compressor of FIG. It is II sectional drawing in the vane type compressor of FIG. It is II sectional drawing in the vane type compressor of FIG. It is sectional drawing which shows the rotation operation | movement of the vane aligner part in the vane type compressor of FIG. It is a cross-sectional schematic diagram which shows the periphery site | part of the vane part in FIG. It is a cross-sectional schematic diagram which shows a mode that oil is supplied to an oil groove in the vane type compressor of FIG. It is a cross-sectional schematic diagram which shows a mode that oil is supplied to an oil groove in the vane type compressor of FIG. It is a perspective view which shows the modification of the oil groove of FIG. It is a perspective view which shows the modification of the oil groove of FIG. It is a perspective view which shows the modification of the oil groove of FIG. It is a perspective view which shows the modification of the oil groove of FIG. It is principal part sectional drawing which shows Embodiment 2 of the bush in the vane type compressor of this invention. It is principal part sectional drawing which shows Embodiment 3 of the bush in the vane type compressor of this invention. It is principal part sectional drawing which shows Embodiment 4 of the bush in the vane type compressor of this invention.

Embodiment 1. FIG.
FIG. 1 is a longitudinal sectional view showing Embodiment 1 of the vane type compressor of the present invention, and a vane type compressor 200 will be described with reference to FIG. In FIG. 1, an arrow indicated by a solid line indicates the flow of the fluid (gas refrigerant), and an arrow indicated by a broken line indicates the flow of the refrigerator oil 25. The vane compressor 200 of FIG. 1 includes a sealed container 103, a compression element 101 housed in the sealed container 103, an electric element 102 that is positioned above the compression element 101 and drives the compression element 101, and a sealed container 103. The oil sump 104 is provided at the bottom of the inside and stores the refrigerating machine oil 25.

  The electric element 102 that drives the compression element 101 is composed of, for example, a brushless DC motor. The electric element 102 includes a stator 21 that is fixed to the inner periphery of the hermetic container 103, and a rotor 22 that is disposed inside the stator 21 and uses a permanent magnet. Electric power is supplied to the stator 21 from a glass terminal 23 fixed to the upper surface of the hermetic container 103 by welding. A suction pipe 26 is attached to the side surface of the sealed container 103, and a discharge pipe 24 is attached to the upper surface.

  2 is an exploded perspective view showing an example of the compression element 101 in the vane type compressor 200 of FIG. 1, FIG. 3 is a schematic view showing an example of the first vane and the second vane, and FIG. 4 shows an example of the bush. FIG. 5 is a cross-sectional view taken along the line II in the vane type compressor 200 of FIG. 1, and the compression element 101 will be described with reference to FIGS. In addition, in the vane type compressor 200 of FIGS. 1-5, the case where two vanes are provided is illustrated.

  (1) Cylinder 1: The overall shape is substantially cylindrical, and both ends in the axial direction are open. A part of the cylinder inner peripheral surface 1b is provided with a notch 1c that penetrates in the axial direction and is bent outward, and the suction port 1a that communicates with the suction pipe 26 opens in the notch 1c. Yes. The cylinder 1 is provided with a discharge port 1d for discharging gas (refrigerant). The discharge port 1d is located on the opposite side of the suction port 1a across the closest contact point 32 between the cylinder inner peripheral surface 1b and the rotor portion 4a, and is provided on the side facing the frame 2 in the vicinity of the closest contact point 32 ( (See FIGS. 2 and 5). Further, an oil return hole 1e penetrating in the axial direction is provided in the outer peripheral portion.

  (2) Frame 2: The section is substantially T-shaped, and the portion in contact with the cylinder 1 is substantially disk-shaped, and closes one opening (upper side in FIG. 2) of the cylinder 1. A concave portion 2a formed concentrically with the cylinder inner peripheral surface 1b is formed on the surface of the frame 2 facing the cylinder 1, and a cylindrical main bearing portion 2c is provided at the center of the frame 2. . The concave portion 2a may be a circular hole or a ring-shaped groove. The frame 2 is provided with a discharge port 2d penetrating in the axial direction, and the discharge port 2d communicates with the first discharge port 1d of the cylinder 1. Further, a discharge valve holder 28 for restricting the opening degree of the discharge valve 27 and the discharge valve 27 is attached to the surface of the discharge port 2d opposite to the cylinder 1. Further, a discharge port 2d communicating with the discharge port 1d provided in the cylinder 1 and penetrating in the axial direction is provided, and a discharge valve 27 (shown in FIG. 2) and a discharge port 2d on the surface opposite to the cylinder 1 are provided. A discharge valve presser 28 (shown in FIG. 2) for restricting the opening degree of the discharge valve 27 is attached to the frame 2.

  (3) Cylinder head 3: The cross section is substantially T-shaped, the portion in contact with the cylinder 1 is substantially disk-shaped, and closes the other opening (lower side in FIG. 2) of the cylinder 1. A concave portion 3 a formed concentrically with the cylinder inner peripheral surface 1 b is formed on the surface of the cylinder head 3 facing the cylinder 1, and a cylindrical main bearing portion 3 c is provided at the center of the cylinder head 3. ing. The concave portion 3a may be a circular hole or a ring-shaped groove.

  (4) Rotor shaft 4: A cylindrical rotor portion 4a that rotates in the cylinder 1 and a shaft portion 4b, 4c that transmits rotational force to the rotor portion 4a. The rotor portion 4a and the shaft portions 4b, 4c It has an integrated structure. The shaft portions 4b and 4c are supported by the main bearing portion 2c of the frame 2 and the main bearing portion 3c of the cylinder head 3, respectively. The rotor portion 4a is formed with bush holding portions 4d and 4e and vane relief portions 4f and 4g that are substantially circular in cross section and penetrate in the axial direction. The bush holding portion 4d and the vane relief portion 4f, and the bush holding portion 4e and the vane relief portion 4g communicate with each other. 3 in communication with the recess 3a. Further, the bush holding portion 4d and the bush holding portion 4e, the vane relief portion 4f and the vane relief portion 4g are disposed at substantially symmetrical positions. An oil pump 31 (shown in FIG. 1) using the centrifugal force of the rotor shaft 4 as described in, for example, Japanese Patent Application Laid-Open No. 2009-264175 is provided at the lower end portion of the rotor shaft 4. The oil pump 31 is provided inside the rotor shaft 4 and communicates with an oil supply passage 4h extending in the axial direction. An oil supply passage 4i is provided between the oil supply passage 4h and the recess 2a, and the oil supply passage 4h and the recess An oil supply passage 4j is provided between 3a. Therefore, the lubricating oil (refrigerating machine oil) 25 is supplied to the recess 2a and the recess 3a from the oil supply passages 4i and 4j, respectively. An oil drain hole 4k (shown in FIG. 1) is provided at a position above the main bearing portion 2c of the shaft portion 4b.

  (5) First vane 5: A substantially rectangular plate-like member, which is held by the rotor portion 4a so as to rotate around the center of the inner peripheral surface 1b of the cylinder 1. The first vane 5 is provided on the vane portion 5a, the end surface of the vane portion 5a on the frame 2 side, the partial ring-shaped vane aligner portion 5c, and the end surface of the vane portion 5a on the cylinder head 3 side. And a partial ring-shaped vane aligner portion 5d. And the vane front-end | tip part 5b located in the cylinder internal peripheral surface 1b side is formed in circular arc shape on the outer side, and the radius of the circular arc shape is comprised by the radius substantially equivalent to the radius of the cylinder internal peripheral surface 1b. Further, the vane length direction of the vane portion 5a and the normal direction of the arc of the vane tip portion 5b are formed so as to pass through the centers of the arcs forming the vane aligner portions 5c and 5d.

  (6) Second vane 6: A substantially rectangular plate-like member, which is held by the rotor portion 4a so as to rotate around the center of the inner peripheral surface 1b of the cylinder 1. The second vane 6 is provided on the vane portion 6a, the end surface of the vane portion 6a on the frame 2 side, the partial ring-shaped vane aligner portion 6c, and the end surface of the vane portion 6a on the cylinder head 3 side. And a partial ring-shaped vane aligner portion 6d. And the vane front-end | tip part 6b located in the cylinder internal peripheral surface 1b side is formed in circular arc shape outside, and the radius of the circular arc shape is comprised by the radius substantially equivalent to the radius of the cylinder internal peripheral surface 1b. The vane length direction of the vane portion 6a and the normal direction of the arc of the vane tip portion 6b are formed so as to pass through the centers of the arcs forming the vane aligner portions 6c and 6d. The vane aligner portion 5c and the vane aligner portion 6c described above are accommodated in the recess 2a of the frame 2 and are rotatably supported along the vane aligner bearing portion 2b formed of the outer peripheral surface of the recess 2a. Similarly, the vane aligner portion 5d and the vane aligner portion 6d are accommodated in the concave portion 3a of the cylinder head 3, and are rotatably supported along the vane aligner bearing portion 2b that is the outer peripheral surface of the concave portion 2a.

  (7) Bushes 7 and 8: A substantially semi-cylindrical shape and formed of a pair. A pair of bushes 7 and 8 are inserted into the bush holding portions 4d and 4e of the rotor shaft 4, and slide with the rotor portion 4a on the arc-shaped outer surfaces 7b and 8b. Further, the pair of bushes 7 and 8 hold the vane portion 5a of the plate-like first vane 5 and the vane portion 6a of the second vane 6 on the flat inner side surfaces 7e and 8e side. As a result, the first vane 5 and the second vane 6 are held so as to be rotatable with respect to the rotor portion 4a about the bush centers 7a and 8a and movable in a substantially normal direction (slidable). Oil grooves 7f and 8f are formed on the outer surfaces 7b and 8b, and oil grooves 7g and 8g are formed on the inner surfaces 7e and 8e. Moreover, the bushes 7 and 8 are provided with notched side surfaces 7c, 7d, 8c, and 8d in which a part of the cylinder is notched (see FIG. 4).

  Next, the suction chamber 9, the intermediate chamber 10, and the compression chamber 11 formed in the vane compressor 200 will be described with reference to FIG. In FIG. 5, the rotor portion 4 a and the cylinder inner peripheral surface 1 b are closest to each other (the closest contact point 32), the suction chamber 9 partitioned by the first vane 5 and the closest contact point 32, An intermediate chamber 10 partitioned by the vane 5 and the second vane 6 and a compression chamber 11 partitioned by the second vane 6 and the closest point 32 are formed.

Specifically, in the closest point 32, the space between the cylinder inner peripheral surface 1 b and the outer peripheral surface of the rotor portion 4 a is partitioned into the suction chamber 9 and the compression chamber 11. Further, the vane tip 5b is adjacent to the cylinder inner peripheral surface 1b and partitions the space between the cylinder inner peripheral surface 1b and the outer peripheral surface of the rotor portion 4a. Here, when the radius of the vane aligner bearing portions 2b and 3b is r _a (see FIG. 3) and the radius of the cylinder inner peripheral surface 1b is r _c (see FIG. 5), the vane aligner portion 5c of the first vane 5 is used. The distance r _v (see FIG. 3) between the outer peripheral side of 5d and the vane tip 5b is as follows. In the equation (1), δ represents a gap between the vane tip 5b and the cylinder inner peripheral surface 1b.
r _v = r _c −r _a −δ (1)

By setting the distance r _v by the equation (1), the first vane 5 will rotate without contacting the cylinder inner peripheral surface 1b. Here, setting the distance r _v so that a gap δ is as small as possible to minimize a leakage of the refrigerant from the vane tip 5b. The relationship of the formula (1) is the same for the second vane 6, and the second vane 6 is maintained while maintaining a narrow gap between the vane tip 6b of the second vane 6 and the cylinder inner peripheral surface 1b. Will rotate.

  The first vane 5, the second vane 6, and the cylinder inner peripheral surface 1 b form three spaces in the cylinder 1, that is, a suction chamber 9, an intermediate chamber 10, and a compression chamber 11. In the suction chamber 9, a suction port 1 a communicating with the low pressure side of the refrigeration cycle is opened through the suction pipe 26 through the notch 1 c. The compression chamber 11 communicates with a discharge port 2d provided in the frame 2 that is closed by the discharge valve 27 through a discharge port 1d provided in the cylinder 1 except during discharge. The notch 1c is provided from the vicinity of the closest point 32 to the range of point A where the vane tip 5b of the first vane 5 and the cylinder inner peripheral surface 1b face each other. Accordingly, the intermediate chamber 10 communicates with the suction port 1a up to a rotation angle of 90 °, but thereafter has a range of rotation angles that does not communicate with either the suction port 1a or the discharge port 1d, and then communicates with the discharge port 1d.

  FIG. 6 is a cross-sectional view taken along the line II of the vane type compressor of FIG. 1, and the manner in which the volumes of the suction chamber 9, the intermediate chamber 10, and the compression chamber 11 change will be described with reference to FIG. In FIG. 6, for the sake of simplicity, the suction port 1a, the cutout portion 1c, and the discharge port 1d are omitted, and the suction port 1a and the discharge port 1d are shown as suction and discharge by arrows, respectively. Further, the rotation angle (rotation phase) of FIG. 6 is such that the closest point 32 where the rotor portion 4a of the rotor shaft 4 and the cylinder inner peripheral surface 1b are closest to each other, the first vane 5 and the cylinder inner peripheral surface 1b. A case where one opposing point coincides is defined as “rotation angle 0 °”. Here, the state after “rotation angle 180 °” is the same as the state where the first vane 5 and the second vane 6 are switched at “rotation angle 0 °”, and thereafter “rotation angle 0 °”. To “Rotation angle 135 °”. Further, the arrow shown in the “rotation angle 0 °” diagram of FIG. 6 indicates the rotation direction of the rotor shaft 4 (clockwise in FIG. 6).

  At “rotation angle 0 °”, the right intermediate chamber 10 partitioned by the closest contact 32 and the second vane 6 communicates with the suction port 1a via the notch 1c, and the intermediate chamber 10 is connected to the suction port. Gas (refrigerant) is sucked from 1a. On the other hand, the left space partitioned by the closest contact 32 and the second vane 6 becomes the compression chamber 11 communicating with the discharge port 1d.

  In the “rotation angle of 45 °”, the space partitioned by the first vane 5 and the closest contact point 32 becomes the suction chamber 9. On the other hand, the intermediate chamber 10 partitioned by the first vane 5 and the second vane 6 communicates with the suction port 1a through the notch 1c, and the volume of the intermediate chamber 10 is “rotation angle 0 °”. The gas (refrigerant) continues to be inhaled. Further, the space partitioned by the second vane 6 and the closest contact point 32 becomes the compression chamber 11. The volume of the compression chamber 11 becomes smaller than when the “rotation angle is 0 °”, and the gas (refrigerant) is compressed and its pressure gradually increases.

  At the “rotation angle of 90 °”, the vane tip 5b of the first vane 5 overlaps with the point A on the cylinder inner peripheral surface 1b, so that the intermediate chamber 10 does not communicate with the suction port 1a. Thereby, the suction of the gas (refrigerant) in the intermediate chamber 10 ends. In this state, the volume of the intermediate chamber 10 is substantially maximum. The volume of the compression chamber 11 becomes even smaller than when the rotation angle is 45 °, and the pressure of the gas (refrigerant) increases. The volume of the suction chamber 9 becomes larger than that at the “rotation angle of 45 °”, and the suction of gas (refrigerant) is continued.

  At the “rotation angle of 135 °”, the volume of the intermediate chamber 10 is smaller than that at the “rotation angle of 90 °”, and the gas (refrigerant) pressure increases. Further, the volume of the compression chamber 11 becomes smaller than that at the “rotation angle of 90 °”, and the pressure of the gas (refrigerant) increases. The volume of the suction chamber 9 becomes larger than that at the “rotation angle of 90 °”, and the suction of gas (refrigerant) is continued.

  Thereafter, the second vane 6 approaches the discharge port 1d, but the discharge valve 27 opens when the pressure in the compression chamber 11 exceeds the high pressure of the refrigeration cycle (including the pressure necessary to open the discharge valve 27), The gas (refrigerant) in the compression chamber 11 passes through the discharge port 1d and the discharge port 2d and is discharged into the sealed container 103 as shown in FIG. The gas (refrigerant) discharged into the sealed container 103 passes through the electric element 102 and is discharged to the outside (high pressure side of the refrigeration cycle) from the discharge pipe 24 fixed (welded) to the upper part of the sealed container 103. Therefore, the pressure in the sealed container 103 is a high discharge pressure.

  After the second vane 6 passes through the discharge port 1d, some high-pressure gas (refrigerant) remains in the compression chamber 11 (a loss occurs). When the compression chamber 11 disappears at a “rotation angle of 180 °” (not shown), the high-pressure refrigerant changes to a low-pressure refrigerant in the suction chamber 9. Note that, at the “rotation angle of 180 °”, the suction chamber 9 moves to the intermediate chamber 10, the intermediate chamber 10 moves to the compression chamber 11, and the compression operation is repeated thereafter.

  As described above, the rotation of the rotor shaft 4 gradually increases the volume of the suction chamber 9 and continues to suck gas (refrigerant). Thereafter, the flow proceeds to the intermediate chamber 10, but the volume gradually increases, and the suction of gas (refrigerant) is continued. On the way, the volume of the intermediate chamber 10 is maximized and is no longer communicated with the suction port 1a. Thus, the suction of gas (refrigerant) is terminated here. Thereafter, the volume of the intermediate chamber 10 gradually decreases and compresses the gas. Thereafter, the intermediate chamber 10 moves to the compression chamber 11 and continues to compress the gas. The gas compressed to a predetermined pressure is discharged into the hermetic container 103 by pushing up the discharge valve 27 through the discharge port 1d and the discharge port 2d.

  FIG. 7 is a cross-sectional view showing an example of the rotation operation of the vane aligner portions 5c and 6c, and an operation example of the vane compressor 200 will be described with reference to FIG. The definition of the rotation angle (rotation phase) shown in FIG. 7 is the same as the definition of the angle shown in FIG. The shaft portion 4 b of the rotor shaft 4 receives rotational power from the drive portion of the electric element 102, and the rotor portion 4 a rotates in the cylinder 1. With the rotation of the rotor part 4a, the bush holding parts 4d and 4e arranged near the outer periphery of the rotor part 4a move on the circumference around the rotor shaft 4. The pair of bushes 7 and 8 held in the bush holding portions 4d and 4e, and the vane portion 5a of the first vane 5 rotatably held between the pair of bushes 7 and 8 are provided. The vane portion 6a of the second vane 6 also rotates together with the rotor portion 4a.

  The first vane 5 and the second vane 6 receive centrifugal force due to rotation, and the vane aligner portions 5c and 6c and the vane aligner portions 5d and 6d are pressed against the vane aligner bearing portions 2b and 3b, respectively, while sliding. The vane aligner bearing portions 2b and 3b rotate around the center. Here, since the vane aligner bearing portions 2b and 3b and the cylinder inner peripheral surface 1b are concentric, the first vane 5 and the second vane 6 rotate around the center of the cylinder inner peripheral surface 1b. Then, the bushes 7 and 8 are in the bush holding portions 4d and 4e so that the length directions of the vane portion 5a of the first vane 5 and the vane portion 6a of the second vane 6 are directed to the cylinder center. , 8a.

  In the above operation, the side surfaces of the bush 7 and the vane portion 5a of the first vane 5 and the side surfaces of the bush 8 and the vane portion 6a of the second vane 6 slide with each other. Further, the bush holding portion 4d and the bush 7 and the bush holding portion 4e and the bush 8 of the rotor shaft 4 slide on each other.

  FIG. 8 is a schematic cross-sectional view showing the periphery of the vane portion 5a of FIG. 5, and the flow of the refrigerating machine oil (lubricating oil) 25 will be described with reference to FIGS. In FIG. 8, an arrow indicated by a solid line indicates the flow of the refrigerating machine oil 25. As indicated by a broken line in FIG. 1, as the rotor shaft 4 rotates, the refrigeration oil 25 is sucked up from the oil sump 104 by the oil pump 31 and sent out to the oil supply passage 4h. The refrigerating machine oil 25 sent out to the oil supply passage 4h passes through the oil supply passage 4i and is sent out to the recess 3a of the cylinder head 3 through the recess 2a of the frame 2 and the oil supply passage 4j.

  The refrigerating machine oil 25 fed to the recesses 2a and 3a lubricates the vane aligner bearing portions 2b and 3b and is supplied to the vane relief portions 4f and 4g communicating with the recesses 2a and 3a. Here, since the pressure in the sealed container 103 is a high discharge pressure, the pressures in the recesses 2a and 3a and the vane relief portions 4f and 4g are also discharge pressures. A part of the refrigerating machine oil 25 fed to the recesses 2 a and 3 a is supplied to the main bearing portion 2 c of the frame 2 and the main bearing portion 3 c of the cylinder head 3.

  Since the pressure of the vane relief part 4f is a discharge pressure and is higher than the pressures of the suction chamber 9 and the intermediate chamber 10, the refrigerating machine oil 25 has a sliding portion between the outer surfaces 7b and 8b of the bushes 7 and 8 and the rotor part 4a. While lubricating the sliding part between the inner side surfaces 7e, 8e and the vane part 5a, the pressure difference and the centrifugal force are sent to the suction chamber 9 and the intermediate chamber 10. A part of the refrigerating machine oil 25 sent out to the intermediate chamber 10 flows into the suction chamber 9 while sealing the gap between the vane tip 5b and the inner peripheral surface 1b of the cylinder 1.

  In addition, although the case where the space partitioned by the first vane 5 is the suction chamber 9 and the intermediate chamber 10 has been shown, the rotation proceeds and the space partitioned by the first vane 5 is the intermediate chamber 10 and the compression chamber 11. It is the same even if it becomes. Even when the pressure in the compression chamber 11 reaches the same discharge pressure as the pressure of the vane escape portion 4f, the refrigerating machine oil 25 is sent out toward the compression chamber 11 by centrifugal force. Although the above operation is shown for the first vane 5, the same operation is performed for the second vane 6.

  Here, when the vanes 5 and 6 partition the space in the cylinder 1 as described above, the pressure on the rotation direction side becomes larger than the pressure on the counter rotation direction side. Due to the pressure difference generated on both sides of each vane 5, 6, a fluctuating load in the counter-rotating direction side acts on the vanes 5, 6 periodically, and a pair of bushes 7, 8 sandwiching the vanes 5, 6. Among them, the load acting on the bushes 7 and 8 on the counter-rotating direction side becomes large. For this reason, in the bushes 7 and 8 on the counter-rotating direction side, wear may occur or seizure may occur on the sliding surfaces between the inner side surfaces 7e and 8e and the vane portions 5a and 6a. Further, the bushes 7 and 8 are inserted and held with respect to the bush holding portions 4d and 4e with a gap therebetween. The bushes 7 and 8 are rotated concentrically with the rotation center of the rotor portion 4a following the rotation of the rotor portion 4a. At this time, the centrifugal force acting outward with respect to the radial direction of the rotor portion 4a acts on the bushes 7 and 8, and the rotor radial direction reference among the opposing surfaces of the outer surfaces 7b and 8b and the bush holding portions 4d and 4e. As a result, the outer gap is clogged, and wear and seizure are likely to occur.

  As described above, in the bushes 7 and 8, wear may occur on the sliding surfaces between the inner side surfaces 7e and 8e and the vane portions 5a and 6a and the sliding surfaces between the outer side surfaces 7b and 8b and the rotor portion 4a. . On the other hand, when the gap on the friction surface is set large to increase the oil supply amount, the reliability is improved, but the leakage through the bush sliding portion increases, so the loss due to the leakage also increases and the performance decreases. Cause.

  Therefore, oil grooves 7f, 7g, 8f, and 8g to which the refrigerating machine oil (lubricating oil) 25 is supplied are provided on the side surfaces 7b, 7e, 8b, and 8e of the bushes 7 and 8 in FIG. The oil grooves 7f, 7g, 8f, and 8g are grooves that extend linearly from the upper surface to the lower surface of the bushes 7 and 8, and when the rotor shaft 4 has a predetermined rotation angle (rotation phase), the recess 2a 3a. Since the recesses 2a and 3a are filled with the refrigerating machine oil 25 through the supply passages 4i and 4j, when the respective oil grooves 7f, 7g, 8f and 8g communicate with the recesses 2a and 3a, oil is supplied from the recesses 2a and 3a. Will be.

  9 and 10 are schematic cross-sectional views showing how oil is supplied to the oil grooves 7f and 7g, respectively. The oil supply to the oil grooves 7f and 7g will be described with reference to FIGS. The oil grooves 7f and 7g of the bush 7 will be described below, but the same applies to the oil grooves 8f and 8g of the bush 8. First, in FIG. 9 and FIG. 10, the refrigerating machine oil 25 is supplied from the oil reservoir 104 through the oil supply passage 4i by the oil pump 31 to the recess 2a of the frame 2 and the recess 3a of the cylinder head. Therefore, the refrigerating machine oil 25 in the recesses 2a and 3a is forcibly supplied to the oil grooves 7f and 7g at a predetermined rotation angle α at which the oil grooves 7f and 7g communicate with the recesses 2a and 3a. On the other hand, oil supply to the oil grooves 7f and 7g is stopped at a rotation angle at which the oil grooves 7f and 7g do not communicate with the recesses 2a and 3a.

  The setting of the rotation angle α at which the oil grooves 7f and 7g communicate with the recesses 2a and 3a is determined by the positions of the oil grooves 7f and 7g, and the positions of the oil grooves 7f and 7g are on the inner peripheral side of the rotor portion 4a. The rotation angle α becomes wider as the value approaches. The predetermined rotation angle α at which the oil grooves 7f, 7g and the recesses 2a, 3a communicate with each other is set in a range in which the pressure difference across the first vane 5 is equal to or greater than the predetermined pressure. Here, when the rotation angle = 0 °, the pressure in the compression chamber 11 on the rotation direction side of the second vane 6 reaches a high pressure (or when the rotation angle = 180 °, the rotation direction with respect to the first vane 5). Assuming that the pressure on the side or the intermediate chamber 10 reaches a high pressure), the oil grooves 7f and 7g are provided at positions that communicate with the recesses 2a and 3a when the rotation angle α = 135 ° to 245 °. It has been. This is because, in the range of the rotation angle = 0 ° to 135 °, the suction chamber 9 and the intermediate chamber 10 straddling the first vane 5 are both low in pressure, and therefore a pressure difference across the first vane 5 is generated. Because there is no. Thus, when the pressure difference is in a low pressure state, by preventing the refrigerating machine oil 25 from being supplied from the oil grooves 7f and 7g, an increase in leakage loss due to excessive oil supply can be suppressed.

  On the other hand, in the range of the rotation angle = 135 ° to 360 °, the oil grooves 7f and 7g communicate with the recesses 2a and 3a in the range of the rotation angle = 135 ° to 245 °, and the rotation angle = 245 ° to 360 °. The oil grooves 7f and 7g do not communicate with the recesses 2a and 3a. Specifically, when the rotation angle is 135 ° to 180 °, the suction chamber 9 on the counter-rotation direction side across the first vane 5 is low pressure, and the intermediate chamber 10 on the rotation direction side is intermediate pressure to high pressure. Therefore, a pressure difference is generated on the side surface of the first vane 5. Further, at the rotation angle = 180 °, the suction chamber 9 shifts to the intermediate chamber 10, the intermediate chamber 10 shifts to the compression chamber 11, and the first vane 5 is increased by the pressure of the compression chamber 11 on the rotation direction side. The pressure difference across is larger than the rotation angle = 135 °. For this reason, the load acting on the bush 7 on the counter-rotating direction side of the pair of bushes 7 sandwiching the first vane 5 becomes large, and the oil film on the sliding surfaces of the outer side surface 7b and the inner side surface 7e tends to disappear. . Therefore, in the range of the rotation angle α = 135 ° to 180 °, the oil grooves 7f and 7g of the bush 7 communicate with the recesses 2a and 3a, and the refrigerating machine oil 25 in the recesses 2a and 3a is forcedly supplied to the oil grooves 7f and 7g. To be.

  Further, in the range of the rotation angle = 180 ° to 360 °, the pressure on the rotation direction side of the bush 7 is high. On the other hand, the pressure on the anti-rotation direction side of the bush 7 changes from a low pressure to a high pressure in the range of the rotation angle = 180 ° to 270 °, and from a low pressure to a high pressure in the range of the rotation angle = 270 ° to 360 °. In the relationship between the pressure on the rotation direction side and the pressure on the counter rotation direction side, the pressure difference in the space on both sides of the first vane 5 becomes high in the range of the rotation angle α = 180 ° to 245 °. For this reason, the oil grooves 7f and 7g communicate with the recesses 2a and 3a in the range of the rotation angle = 180 ° to 245 °, and do not communicate in the range of the rotation angle = 245 ° to 360 °.

  That is, the pressure difference across the first vane 5 is minimum in the range of the rotation angle = 0 ° to 90 °, increases in the range of the rotation angle = 90 ° to 180 °, and the rotation angle = 180 ° to 270 °. And the rotation angle decreases in the range of 270 ° to 360 °. Among them, the oil grooves 7f and 7g communicate with the recesses 2a and 3a in the range of the rotation angle = 180 ° to 245 ° where the pressure difference becomes a predetermined high pressure state.

  As described above, for the sake of simplification, this explanation is that the pressure in the compression chamber 11 on the rotation direction side of the second vane 6 reaches a high pressure at the rotation angle = 0 ° (or the first angle at the rotation angle = 180 °). It is assumed that the pressure in the intermediate chamber 10 on the rotation direction side of the vane 5 reaches a high pressure), but the fluctuation of the pressure difference across the vane 5 shows a similar tendency under other operating conditions.

  As described above, the rotation angle (rotation phase) α at which the pressure difference across the vane 5 increases, that is, the range near the rotation angle α at which the oil film on the sliding surface tends to disappear is limited to the oil grooves 7f and 7g. Thus, the reliability of the vane operation can be improved while suppressing an increase in leakage loss due to excessive oiling. Further, by providing oil grooves 7f and 7g on the bushes 7 and 8 side instead of providing oil grooves on the vane parts 5a and 6a side, it is possible to prevent the strength of the vane parts 5a and 6a from being reduced due to the formation of the grooves. Can do.

  In addition, although illustrated about the case where the oil grooves 7f and 7g of the bush 7 and the recessed parts 2a and 3a communicate with each other at a predetermined rotation angle = 135 ° to 245 °, the rotation in which the pressure difference across the first vane 5 increases. As long as the oil grooves 7f and 7g of the bush 7 and the recesses 2a and 3a are set to communicate with each other in the vicinity of the phase, the rotation angle α is not limited. Further, the positional relationship between the oil grooves 7f and 7g of the bush 7 and the recesses 2a and 3a has been described. It is done.

  11A to 11D are perspective views showing modifications of the oil grooves 7f and 7g. The oil grooves 7f and 7g of the bush 7 will be described below, but the oil grooves 8f and 8g of the bush 8 can have the same configuration. In the bush 7 of FIG. 11A, the oil grooves 7f, 7g extend in a straight line along the axial direction and are divided into two parts without penetrating the upper and lower end surfaces of the bush 7. 11B, the oil grooves 7f and 7g extend linearly along the axial direction, but are divided into two parts without penetrating the upper and lower end surfaces of the bush 7, as in FIG. 11A. Further, a plurality of oil grooves 7f and 7g are formed on the side surfaces 7b and 7e, respectively. In the bush 7 of FIG. 11C, the oil grooves 7f and 7g are inclined in the axial direction and extend linearly, and are divided into two parts without penetrating the upper and lower end surfaces of the bush 7. In the bush 7 of FIG. 11D, the bush 7 is provided with oil grooves 7f and 7g in a curved shape with a predetermined curvature.

  Even in the oil grooves 7f and 7g having the shapes shown in FIGS. 11A to 11D, the rotational phase at which the pressure difference across the vane 5 increases, that is, the rotational angle (rotational phase) α where the oil film on the sliding surface tends to disappear. By restricting the oil pressure to the oil grooves 7f and 7g in the vicinity, the reliability of the operation of the vanes 5 and 6 can be improved while suppressing an increase in leakage loss due to excessive oil supply. In addition, in the bush 7 of FIG. 11A-FIG. 11D, although illustrated about the case where the oil grooves 7f and 7g formed in the outer surface 7b and the inner surface 7e have the same pattern, it differs in FIG. 11A-11D, respectively. You may make it combine a pattern.

  FIGS. 12 to 14 are schematic views showing Embodiments 2 to 4 of the bushes 7 and 8 in the vane type compressor of the present invention. Embodiments 2 to 4 of the bushes 7 and 8 are described with reference to FIGS. Will be described. 12 to 14, parts having the same configuration as in FIG. 8 are denoted by the same reference numerals and description thereof is omitted.

  In FIG. 12, of the pair of bushes 7 with the first vane 5 interposed therebetween, only the bushes 7 in the counter-rotating direction are provided with oil grooves 7f and 7g. As described above, since the load due to the pressure difference across the vane portion 5a acts on the anti-rotation direction side, the oil grooves 7f and 7g only on the anti-rotation direction side bush 7 side where it is difficult to ensure the reliability of the sliding surface. Is provided. Thereby, the reliability of the sliding surface in the bush 7 can be improved, excessive oil supply on the side surface of the bush 7 on the rotational direction side can be suppressed, and an increase in loss due to leakage can be suppressed.

  In FIG. 13, among the pair of bushes 7 sandwiching the first vane 5, the oil groove 7f is provided only on the outer side surface 7b side, and is not provided on the inner side surface 7e side. In this configuration, it is possible to suppress wear and seizure of the outer surface 7b by positively supplying oil to the outer surface 7b side where the gap is easily clogged by the centrifugal force acting on the bush 7. In addition, although the oil groove 7f on the outer surface of the bush 7 has been described, the same effect can be obtained by configuring the oil groove 8f of the bush 8 in the same manner.

  In FIG. 14, the oil groove 7g is provided only on the inner side surface 7e side of the bush 7 and is not provided on the inner side surface 7e side. The pressure in the vane relief portion 4f is always high, and the pressure in the suction chamber 9 and the intermediate chamber 10 changes from low pressure to high pressure depending on the rotation phase. Therefore, when neither of the oil grooves 7f and 7g is provided, the pressure of the oil passing through the outer surface 7b and the inner surface 7e becomes an intermediate pressure between the vane relief portion 4f and the suction chamber 9 (or the intermediate chamber 10). At this time, as shown in FIG. 14, by providing an oil groove 7g on the inner surface 7e and directly supplying high-pressure oil, the pressure is higher than that on the outer surface 7b, and the entire surface of the outer surface 7b is placed on the rotor. It can be pressed against the part 4a.

  As a result, it becomes possible to press the gap of the outer surface 7b, which was easily clogged only by the radial outer gap of the rotor portion 4a, by the centrifugal force over the entire sliding surface, so that the outer surface 7b is worn or seized. Can be suppressed. Although the oil groove 7g of the bush 7 has been described here, the same effect can be obtained by configuring the oil groove 8g of the bush 8 in the same manner.

  Embodiments of the present invention are not limited to the above embodiments. For example, in each of the above embodiments, the case where the number of vanes is two has been described, but the number of vanes may be one or three or more.

  Further, in each of the first to fourth embodiments, the oil pump 31 using the centrifugal force of the rotor shaft 4 has been described. However, the oil pump 31 may be in any form, for example, a volume described in Japanese Patent Application Laid-Open No. 2009-62820. A shape pump may be used as the oil pump 31.

  Further, in each of the first to fourth embodiments described above, the case where the partial ring-shaped vane aligner is integrally attached to the vane has been described. For example, the inner side of the rotor shaft is configured to be hollow and the vane fixing shaft is disposed therein. Even when the vane is rotatably attached to the fixed shaft, the same effect can be obtained.

  Moreover, in each said embodiment, although the case where the vane aligner parts 5c, 5d, 6c, 6d are the shape of a partial ring is illustrated, when there is one vane in the rotor part 4a, the vane aligner part is a ring. It may be a shape.

  Further, in each of the above embodiments, the oil grooves 7f, 7g, 8f, 8g are illustrated as being formed in so-called U-shaped grooves, but have a rectangular cross section or a triangular cross section. May be. Moreover, any of the shapes shown in FIGS. 11A to 11D may be adopted as the oil grooves 7f, 7g, 8f, and 8g shown in FIGS.

1 cylinder, 1a suction port, 1b cylinder inner surface, 1c notch, 1d discharge port, 1e oil return hole, 2 frame, 2a recess, 2b vane aligner bearing, 2c main bearing, 2d discharge port, 3 cylinder Head, 3a recess, 3b vane aligner bearing part, 3c main bearing part, 4 rotor shaft, 4a rotor part, 4b, 4c shaft part, 4d, 4e bush holding part, 4f, 4g vane relief part, 4h, 4i, 4j Path, 4k oil drain hole, 5 first vane, 6 second vane, 5a, 6a vane part, 5b, 6b vane tip part, 5c, 6c, 5d, 6d vane aligner part, 7, 8 bushing, 7a, 8a Bush center, 7b, 8b Outer side, 7c, 7d, 8c, 8d Notched side, 7e, 8e Inner side, 7f, 7g, 8f, 8g Oil groove, 9 Suction chamber 10 Intermediate chamber, 11 Compression chamber, 21 Stator, 22 Rotor, 23 Glass terminal, 24 Discharge pipe, 25 Refrigerator oil (lubricating oil), 26 Suction pipe, 27 Discharge valve, 28 Discharge valve retainer, 31 Oil pump, 32 closest point, 101 compression element 102 electric element 103 closed vessel, 104 oil clasp, 200 vane compressor, _{r v} distance, alpha predetermined rotation angle (rotational phase), [delta] clearance.

Claims (16)

A cylinder having a substantially circumferential inner circumferential surface and having both axial ends open;
A cylinder head and a frame for closing both axial ends of the cylinder;
A cylindrical rotor portion that rotates in the cylinder and a shaft portion that transmits rotational force to the rotor portion, and a rotor shaft whose lower end is immersed in an oil sump;
A vane installed on the rotor portion and held to rotate around the center of the inner peripheral surface of the cylinder;
The vane is inserted into a substantially cylindrical bush holding portion penetrating in the axial direction provided in the rotor portion while being sandwiched between a pair of substantially semi-cylindrical bushes, whereby the rotor portion Is supported to be rotatable and movable with respect to
The frame and the cylinder head have a concave portion filled with lubricating oil formed concentrically with the inner peripheral surface of the cylinder on a surface facing the rotor portion in the cylinder,
The bush has an oil groove communicating with the concave portion of one or both of the cylinder head and the frame at a predetermined rotational phase on a side surface facing the vane or the rotor portion.   The vane rotates slidably along a vane portion that partitions a space between an inner peripheral surface of the cylinder and an outer peripheral surface of the rotor portion, and an outer peripheral surface of the recess of the frame and the cylinder head. The vane type compressor according to claim 1, further comprising a vane aligner portion that supports the vane portion.   3. The vane compressor according to claim 1, wherein the rotor shaft has oil supply means and an oil supply passage for sending the refrigeration oil in the oil sump to the recess.   The oil bushing is provided in the bush so that the oil groove is located in a region where the recess is formed in a rotational phase in which a pressure difference between spaces on both sides of the vane is a predetermined pressure difference. The vane type compressor of any one of 1-3.   5. The vane-type compression according to claim 1, wherein the oil groove is provided only in the bush on the counter-rotating direction side of the rotor portion among the pair of bushes. Machine.   The vane compressor according to any one of claims 1 to 5, wherein the oil groove is provided only on an outer surface of the bush facing the rotor portion.   The vane compressor according to any one of claims 1 to 5, wherein the oil groove is provided only on an inner surface of the bush facing the vane.   The vane compressor according to any one of claims 1 to 7, wherein a plurality of the oil grooves are provided on a side surface facing the vane or a side surface facing the rotor portion.   The vane compressor according to any one of claims 1 to 8, wherein the oil groove has a linear shape along an axial direction of the rotor portion.   The vane compressor according to any one of claims 1 to 9, wherein the oil groove has a linear shape inclined with respect to the axial direction of the rotor portion.   The vane type compressor according to any one of claims 1 to 10, wherein the oil groove has a curved shape.   The predetermined rotation phase is approximately 135 ° to 245 ° when the rotation phase is set to 0 ° at the closest point where the inner peripheral surface of the cylinder and the outer peripheral surface of the rotor portion are closest to each other. The vane type compressor according to any one of claims 1 to 11.   The tip end portion of the vane is formed in an arc shape on the outside, and the radius of the arc shape is substantially equal to the radius of the inner peripheral surface of the cylinder. The vane type compressor as described.   The said vane aligner part supports the said vane so that a clearance gap may be maintained between the front-end | tip part of the said vane, and the said internal peripheral surface of the said cylinder, The any one of Claims 2-13 characterized by the above-mentioned. Vane type compressor.   The vane type compressor according to any one of claims 2 to 14, wherein the vane aligner portion is integrally attached to the vane portion or formed integrally with the vane portion.   The vane type compressor according to any one of claims 2 to 15, wherein the vane aligner portion has a partial ring shape or a ring shape.

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