WO2023286559A1 - Annular sealing member for scroll compressor - Google Patents
Annular sealing member for scroll compressor Download PDFInfo
- Publication number
- WO2023286559A1 WO2023286559A1 PCT/JP2022/025119 JP2022025119W WO2023286559A1 WO 2023286559 A1 WO2023286559 A1 WO 2023286559A1 JP 2022025119 W JP2022025119 W JP 2022025119W WO 2023286559 A1 WO2023286559 A1 WO 2023286559A1
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- WO
- WIPO (PCT)
- Prior art keywords
- ring
- groove
- seal member
- annular seal
- dynamic pressure
- Prior art date
Links
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Images
Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
- F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
- F04C18/02—Rotary-piston pumps specially adapted for elastic fluids of arcuate-engagement type, i.e. with circular translatory movement of co-operating members, each member having the same number of teeth or tooth-equivalents
Definitions
- the present invention relates to an annular seal member that is attached to the bottom plate of a movable scroll that constitutes a scroll compressor.
- the scroll compressor includes a scroll-type compression mechanism section consisting of a fixed scroll body and a movable scroll body that orbits with respect to the fixed scroll body.
- the fixed scroll body and the movable scroll body each have a bottom plate portion and a spiral wall erected on the surface of the bottom plate portion.
- the compression chamber is moved toward the center of the spiral by the action of the movable scroll revolving around the axis of the fixed scroll, thereby compressing the refrigerant.
- An annular seal member is provided on the back side of the bottom plate of the movable scroll.
- a thrust load is generated on the movable scroll body due to the compression reaction force. Due to this thrust load, frictional force increases between the annular seal member provided on the back side of the orbiting scroll body and the main bearing member that slides therewith, and there is a risk that the annular seal member will be worn. be.
- Patent Literature 2 is a technique that makes it easier to set the intended back pressure in the back pressure chamber when the atmosphere in the back pressure chamber and the suction pressure region is not uniform.
- the load is reduced by inserting a thrust receiving member that receives the thrust force from the bottom plate portion side of the orbiting scroll body toward the main bearing member side as a member separate from the annular seal member.
- a thrust receiving member that receives the thrust force from the bottom plate portion side of the orbiting scroll body toward the main bearing member side as a member separate from the annular seal member.
- JP-A-8-121366 Japanese Patent Application Laid-Open No. 2007-211702 JP 2012-17656 A
- the present invention has been made in view of such circumstances, and an object of the present invention is to provide an annular seal member for a compressor that can exhibit stable low torque performance without impairing durability and sealing function.
- the annular seal member of the present invention comprises: a fixed scroll body having a bottom plate and a spiral wall standing on its surface; a movable scroll body having a bottom plate and a spiral wall standing on its surface; a shaft; a main bearing that rotatably supports the main bearing; and a main bearing member that fixes the main bearing.
- Rotation of the shaft causes the orbiting scroll to revolve around the axis of the fixed scroll, thereby supplying fluid to the compression chamber. and the fluid is supplied to a back pressure chamber on the back side of the orbiting scroll, the fluid being directed toward the back of the bottom plate portion of the orbiting scroll and the orbiting scroll of the main bearing member.
- annular seal member mounted in at least one annular groove formed in one of the end faces and sealing the back pressure chamber, wherein the annular seal member is a sliding member that at least revolves and slides on the ring side surface.
- a dynamic pressure groove is provided on the dynamic surface.
- the dynamic pressure groove is a groove that introduces the fluid generated by the orbiting motion of the orbiting scroll body to generate dynamic pressure. It is a groove that decreases in the direction. The decreasing direction may be in the depth direction of the groove (FIG. 3), the width direction of the groove (FIG. 7), or both.
- the area of the dynamic pressure groove is 5% to 75% of the total area of the ring side surface.
- the shape of the dynamic pressure groove is a substantially V shape recessed in the width direction of the ring along the ring circumferential direction, and the depth from the sliding surface of the dynamic pressure groove is the ring circumferential direction from the deepest part. It becomes shallow toward both ends of the ring and is constant in the radial direction of the ring.
- the dynamic pressure groove is not flush from the sliding surface to the deepest portion, but is connected to the first inclined surface connected to the sliding surface and the deepest portion, and is connected to the sliding surface. and a second inclined surface forming an inclination angle smaller than that of the first inclined surface.
- the inclination angle of the first inclined surface with respect to the sliding surface is 50° to 80°, and the inclination angle of the second inclined surface with respect to the sliding surface is 0.1° to 15°.
- the dynamic pressure groove is characterized in that a boundary portion between the first inclined surface and the second inclined surface is connected by a curved surface.
- a plurality of the dynamic pressure grooves are spaced apart in the ring circumferential direction, and the ring side surface between the adjacent dynamic pressure grooves constitutes a part of the sliding surface.
- the annular seal member is made of synthetic resin, and the synthetic resin is polyphenylene sulfide (hereinafter referred to as PPS) resin or polyether ether ketone (hereinafter referred to as PEEK) resin.
- PPS polyphenylene sulfide
- PEEK polyether ether ketone
- the annular seal member of the present invention comprises: a fixed scroll body having a bottom plate and a spiral wall standing on its surface; a movable scroll body having a bottom plate and a spiral wall standing on its surface; a shaft; a main bearing that rotatably supports the main bearing; and a main bearing member that fixes the main bearing.
- Rotation of the shaft causes the orbiting scroll to revolve around the axis of the fixed scroll, thereby supplying fluid to the compression chamber. and the fluid is supplied to a back pressure chamber on the back side of the orbiting scroll, the fluid being directed toward the back of the bottom plate portion of the orbiting scroll and the orbiting scroll of the main bearing member.
- annular seal member mounted in at least one annular groove formed in one of the end faces and sealing the back pressure chamber, wherein the annular seal member is a sliding member that at least revolves and slides on the ring side surface.
- the sliding surface is provided with a plurality of lubrication grooves that are open to either the outer diameter of the ring or the inner diameter of the ring, and an inclined surface is provided at the boundary between the lubrication groove and the sliding surface.
- the area of the lubricating groove is 5% to 75% of the total area of the ring side surface.
- the lubricating groove is characterized by being substantially mortar-shaped.
- the annular seal member is characterized by having, as the lubricating grooves, an outer lubricating groove opening to the ring outer diameter and an inner lubricating groove opening to the ring inner diameter. Further, a plurality of the outer-diameter-side lubrication grooves and the inner-diameter-side lubrication grooves are provided separately in the ring circumferential direction. It is characterized by being provided alternately in the circumferential direction.
- the inclined surface has an angle of 0.1° to 15° with respect to the sliding surface.
- the annular seal member is made of synthetic resin, and the synthetic resin is PPS resin or PEEK resin.
- the annular seal member of the present invention is mounted in the annular groove formed in either the back surface of the bottom plate of the movable scroll body or the end surface of the main bearing member facing the movable scroll body in the scroll compressor, and At least the sliding surface that revolves and slides on the side surface of the ring is provided with dynamic pressure grooves, so the sliding area can be reduced.
- the sliding torque decreases due to the surface pressure dependence of the coefficient.
- the dynamic pressure grooves the fluid flows into the dynamic pressure grooves, making it easier for the wedge action to occur, leading to a further reduction in torque. As a result, frictional wear characteristics are improved, and stable low torque performance can be exhibited without impairing durability and sealing function without using a thrust receiving member.
- the area of the dynamic pressure groove is 5% to 75% of the total area of the ring side surface, so it is possible to suppress the acceleration of wear while ensuring the torque reduction effect.
- the shape of the groove for hydrodynamic bearing is a substantially V shape recessed in the width direction of the ring along the circumferential direction of the ring. , and is constant in the radial direction of the ring. Therefore, fluid is easily introduced into the dynamic pressure grooves, and a wedge action is likely to occur.
- the hydrodynamic groove is not coplanar from the sliding surface to the deepest part, but is connected to the first inclined surface connected to the sliding surface to the deepest part and has the first inclination with respect to the sliding surface. and a second inclined surface forming an inclination angle smaller than that of the second inclined surface. Even if the groove wears, the opening area of the dynamic pressure generating groove is less reduced (that is, the sliding area is less increased), and the torque is less likely to change. In particular, since the inclination angle of the first inclined surface with respect to the sliding surface is 50° to 80°, an increase in the sliding area can be suppressed while effectively generating the wedge action. Moreover, since the inclination angle of the second inclined surface with respect to the sliding surface is 0.1° to 15°, the wedging effect of the inflowing fluid can be effectively generated.
- the boundary between the first inclined surface and the second inclined surface is connected by a curved surface, and is formed in an R shape. This makes it easier for the fluid to flow out, and the torque can be further reduced.
- the annular seal member of the present invention is mounted in the annular groove formed in either the back surface of the bottom plate of the movable scroll body or the end surface of the main bearing member facing the movable scroll body in the scroll compressor. It is a seal member that seals the pressure chamber, and has a plurality of lubrication grooves that are open to either the outer diameter of the ring or the inner diameter of the ring at least on the sliding surface that revolves and slides on the side surface of the ring. Since an inclined surface is provided at the boundary with the moving surface, the sliding area can be reduced, and by reducing the sliding area, sliding torque is reduced due to the dependence of the friction coefficient on surface pressure. do.
- the area of the lubrication groove is 5% to 75% of the total area of the ring side, so it is possible to suppress the acceleration of wear while ensuring the torque reduction effect.
- the lubricating groove Since the shape of the lubricating groove is approximately mortar-shaped, in applications where the wedge action generating part constantly changes due to the orbital motion, the lubricating groove has less fluid flow than the groove whose groove depth is constant in the radial direction of the ring. It becomes easier to introduce and the wedging action is more likely to occur.
- substantially mortar-shaped means having a portion of a mortar shape
- mortar-shaped means a conical dent, a truncated conical dent, a truncated conical dent, or the like.
- the inclined surface has an inclination angle of 0.1° to 15° with respect to the sliding surface, it is possible to effectively generate a wedge action by the inflowing fluid.
- FIG. 1 is a partial cross-sectional view showing an example of a scroll compressor provided with an annular seal member of the present invention
- FIG. 1 is a perspective view showing a first embodiment of an annular seal member of the present invention
- FIG. FIG. 3 is an enlarged view of part A in FIG. 2 ; It is the figure which looked at the hydrodynamic groove of the annular seal member from the ring internal diameter side.
- FIG. 8 is a diagram showing another example of a substantially V-shaped dynamic pressure groove of the annular seal member
- FIG. 3 is a cross-sectional view showing a state in which the annular seal member of FIG. 2 is incorporated in an annular groove
- FIG. 8 is a diagram showing another example of dynamic pressure grooves of the annular seal member;
- FIG. 1 is a partial cross-sectional view showing an example of a scroll compressor provided with an annular seal member of the present invention
- FIG. 1 is a perspective view showing a first embodiment of an annular seal member of the present invention
- FIG. 8 is a diagram showing another example of dynamic pressure grooves of the annular seal member
- FIG. 5 is a perspective view showing a second embodiment of the annular seal member of the present invention
- FIG. 10 is an enlarged view of a portion D in FIG. 9; It is the figure etc. which looked at the lubrication groove of the annular seal member from the ring internal-diameter side. It is the figure which looked at the lubricating groove of another annular seal member from the ring internal diameter side.
- FIG. 10 is a cross-sectional view showing a state in which the annular seal member of FIG. 9 is incorporated in an annular groove;
- FIG. 10 is a diagram showing another example of the lubrication groove of the annular seal member;
- FIG. 10 is a diagram showing another example of the lubrication groove of the annular seal member;
- 1 is a schematic diagram of a thrust test;
- FIG. FIG. 4 is a diagram showing test results of a thrust test;
- FIG. 1 is a partial cross-sectional view of a scroll compressor.
- This scroll compressor is a compressor that compresses a fluid such as a refrigerant such as carbon dioxide gas, a refrigerating machine oil such as polyalkylene glycol oil (PAG oil), or a mixture thereof (hereinafter collectively referred to as refrigerant or the like).
- a refrigerant such as carbon dioxide gas
- a refrigerating machine oil such as polyalkylene glycol oil (PAG oil)
- PAG oil polyalkylene glycol oil
- the compressor 1 has a compression mechanism portion and a motor mechanism portion inside a housing 2, and is connected to the outside through a suction port (not shown) and a discharge port (not shown).
- the compression mechanism section is a section for compressing the refrigerant sucked from the suction port and discharging it from the discharge port.
- the fixed scroll member 3 includes a bottom plate portion 3a and a spiral wall 3b vertically erected from the bottom plate portion 3a, and an opening 3c is provided in the center.
- the movable scroll body 4 also includes a bottom plate portion 4a and a spiral wall 4b vertically erected from the bottom plate portion 4a.
- the fixed scroll body 3 and the movable scroll body 4 are arranged in an eccentric manner, and a compression chamber 5 is formed between the spiral walls 3b and 4b of each scroll body.
- a spiral sealing member (tip seal) is attached to the axial end faces of the spiral walls 3b and 4b of each scroll body. This prevents leakage of refrigerant or the like in the compression chamber.
- the motor mechanism section is a section that applies turning driving force to the movable scroll body 4, and is composed of a stator 6a and a rotor 6b.
- the stator 6a is fixed inside the housing 2 and the rotor 6b is connected to the shaft 7.
- the stator 6a and the rotor 6b constitute an electric motor, and when the stator 6a is energized, the rotor 6b and the shaft 7 rotate together.
- Shaft 7 is rotatably supported via main bearing 9 and sub-bearing 10 .
- An eccentric shaft 7a is formed integrally with one end of the shaft 7, and a balance weight 8 is supported thereon.
- a rotating member is configured by the shaft 7 and the balance weight 8 .
- a boss portion 4c is provided so as to protrude vertically from substantially the center of the back side of the bottom plate portion 4a of the movable scroll body 4, and a turning bearing 11 is press-fitted into the boss portion 4c.
- the eccentric shaft 7 a is supported by the orbiting bearing 11
- the movable scroll body 4 is a mechanism for orbiting by the orbiting bearing 11 .
- the main bearing 9 is fixed to a bearing support portion formed on the central side of the main bearing member 12 .
- the main bearing member 12 is fixed within the housing, and the fixed scroll body 3 is coupled to the main bearing member 12 with bolts or the like.
- a shaft seal 13 is mounted between the outer peripheral surface of the shaft 7 and the main bearing member 12 on the side of the main bearing 9 .
- the shaft seal 13 blocks communication between the motor chamber 14 and the back pressure chamber 15a.
- An annular seal member 16 is provided between the main bearing member 12 and the back surface of the bottom plate portion 4a of the movable scroll body 4.
- an annular seal member 16 is mounted in an annular groove 4d formed in the back surface of the bottom plate portion 4a of the orbiting scroll body 4.
- the annular seal member 16 revolves and slides on the end face of the main bearing member 12 facing the movable scroll body.
- the back pressure chamber 15a is sealed by an annular seal member 16 and a shaft seal 13, and these seal portions, the main bearing member 12, and the bottom plate portion 4a of the movable scroll body 4 form a sealed space.
- the rotor 6b rotates, causing the orbiting scroll 4 to start orbiting.
- Refrigerant or the like that has entered the compression mechanism through the suction port is compressed while moving from the outer periphery to the center of the swirling spiral wall, and is discharged to the outside through the opening 3c of the fixed scroll body 3.
- the back pressure chamber 15a is supplied with pressurized fluid from inside the compression mechanism portion through a pressure introduction hole (not shown) provided in the bottom plate portion 4a of the movable scroll body 4. As shown in FIG.
- the thrust load acting on the orbiting scroll 4 due to the compression reaction force (the force that presses the orbiting scroll 4 toward the main bearing member) is reduced.
- the pressure in the back pressure chamber acts on the orbiting scroll member 4 so as to press the orbiting scroll member 4 against the fixed scroll member.
- the annular seal member 16 partitions the inner back pressure chamber 15a and the outer space 15b. While the space 15b has a pressure value close to the suction pressure, the pressure in the back pressure chamber 15a is higher than that in the space 15b because compressed refrigerant or the like is introduced into the back pressure chamber 15a. becomes. As a result, one ring side surface of the annular seal member 16 comes into sliding contact with the end surface of the main bearing member 12 while revolving.
- the annular seal member 16 is mainly made of resin, whereas the main bearing member 12 is made of metal (made of iron or aluminum die-cast), and there is concern about abrasion of the annular seal member 16 due to sliding contact. In particular, the greater the compression pressure of the fluid, the greater the thrust load acting on the orbiting scroll body 4, and the more likely the annular seal member 16 will wear.
- the sliding surface area is reduced, which in turn reduces the torque.
- the wedge action can further reduce the torque.
- annular seal member of the present invention will be described below.
- FIG. 2 shows a perspective view of an annular seal member.
- the outer diameter dimension ⁇ of the annular seal member 16 is, for example, 50 mm or more, and a preferable range is about 50 mm to 100 mm (the same applies to the annular seal member of the second embodiment described later).
- the annular seal member 16 is an annular body with a substantially rectangular cross section, has a shape that is continuous over the entire circumference, and does not have a joint.
- the annular seal member 16 is provided with a plurality of V-shaped dynamic pressure grooves 18 recessed in the width direction of the ring along the circumferential direction at the inner diameter side end of the ring side surface 17 .
- the corners of the inner peripheral surface 16b and the ring side surfaces 17 (including the dynamic pressure grooves 18) on both sides may be chamfered in a straight line or a curved line.
- a stepped portion 16c may be provided at this portion as a projecting portion from the mold.
- one ring side surface of the annular seal member 16 is a surface that slides against the end surface of the main bearing member facing the orbiting scroll body.
- a V-shaped dynamic pressure groove 18 is formed as a portion.
- the dynamic pressure grooves 18 are recesses that do not communicate in the ring radial direction, and are open only on the ring inner diameter side, which leads to low oil leakage of refrigerant and the like.
- the groove for dynamic pressure should be formed at least on the side surface of the ring on the side of the sliding surface that revolves and slides. It is preferable to form them symmetrically on both sides of the ring.
- a ring side surface between adjacent dynamic pressure grooves is a portion that slides on the main bearing member, and constitutes a part of the sliding surface.
- the area of the dynamic pressure grooves (the total area if there are more than one, the same shall apply hereinafter) is not particularly limited. There is a risk that it will become pressure and wear will be accelerated. From this point of view, the area of the hydrodynamic grooves is preferably 5% to 75%, more preferably 20% to 60%, of the entire area of the ring side surface.
- the total area of the ring side surface is the sliding area including the dynamic pressure groove in a plan view of the ring side surface on the side of the sliding surface on which the annular seal member revolves and slides. The area is the area in the same plan view.
- each hydrodynamic pressure groove in the ring circumferential direction is preferably about 3% to 20% of the ring circumferential length, depending on the number.
- the length of the hydrodynamic groove in the ring radial direction is preferably 10% to 80% of the radial thickness of the sliding surface. Further, since the sliding characteristics are stabilized, it is preferable that all of the dynamic pressure grooves are of the same size, and that a plurality of grooves (13 grooves on one side in FIG. 2) are provided at approximately equal intervals.
- FIG. 3 is an enlarged view of part A in FIG. 2
- FIG. 4(a) is a view of the dynamic pressure groove as viewed from the inner diameter side of the ring
- FIG. 4(b) is an enlarged view of part B
- FIG. ) is an enlarged view of the C part.
- the hydrodynamic grooves 18 are V-shaped and recessed toward the width direction (axial side) of the ring along the circumferential direction of the ring. As shown in FIG.
- the depth from the sliding surface of the groove for hydrodynamic bearing 18 has the deepest portion 18d at the center of the groove for hydrodynamic bearing 18 in the ring circumferential direction, and the depth from the deepest portion 18d to the ring circumferential direction. It becomes shallower towards both ends. That is, the depth becomes shallower in the region closer to the sliding surface in the circumferential direction of the ring. Further, the depth from the sliding surface of the hydrodynamic groove 18 is constant in the radial direction of the ring. In addition, in FIG. 3, the deepest part 18d is formed linearly.
- the dynamic pressure groove 18 has a symmetrical shape centering on the deepest portion 18d. and second inclined surfaces 18b, 18b. Specifically, it is not the same plane from the sliding surface to the deepest portion 18d. and a second inclined surface 18b having a smaller inclination angle than the first inclined surface 18a.
- the first inclined surface 18a is steeper than the second inclined surface 18b.
- the inclination angle ⁇ 1 of the first inclined surface 18a with respect to the sliding surface is not particularly limited, but is preferably 50° to 80°, more preferably 50° to 70°. If the inclination angle ⁇ 1 is less than 50°, the sliding surface area will increase greatly when the sliding surface is worn, and there is concern about torque fluctuations. Moreover, if the inclination angle ⁇ 1 exceeds 80°, the wedge action may become weak.
- the inclination angle ⁇ 2 (see FIG. 4B) of the second inclined surface 18b with respect to the sliding surface is not particularly limited, but it is preferably an acute angle of 0.1° to 15°, and preferably 1° to 10°. ° is more preferred. As a result, the wedge effect of the inflowing fluid can be effectively exhibited. On the other hand, if the inclination angle ⁇ 2 is less than 0.1°, the inflowing fluid becomes difficult to flow to the first inclined surface 18a. By increasing the volume of the pressure groove 18 and distributing the pressure, the wedging effect may be weakened.
- the structure of the boundary between the first inclined surface 18a and the second inclined surface 18b is not particularly limited.
- the first inclined surface 18a and the second inclined surface 18b may be directly connected, or may be connected via a curved surface (R surface) 18c as shown in FIG. 4(c).
- the R surface 18c is a portion having a constant width in the ring circumferential direction, and the radius of curvature of the R surface 18c is, for example, 0.1 to 0.3.
- FIG. 4(c) by forming the boundary between the first inclined surface 18a and the second inclined surface 18b in an R shape, the fluid flows on the sliding surface between the adjacent dynamic pressure grooves. It becomes easy to flow out, and it becomes easy to achieve further low torque.
- the boundary between the end of the first inclined surface 18a in the circumferential direction and the sliding surface can be connected by a curved surface (R shape).
- R shape a curved surface
- the depth from the sliding surface of the deepest portion 18d of the hydrodynamic groove 18 is preferably 45% or less of the total width of the ring, more preferably 30% or less.
- the "depth" is the sum of the depths of the recesses on each side. In this case, the depth of the recess on one side is It is 22.5% or less of the total ring width, preferably 15% or less. If it exceeds 45% of the total width of the ring, the strength of the annular seal member may be insufficient and it may be deformed.
- the substantially V-shaped dynamic pressure groove formed on the inner diameter side end of the ring side surface is not limited to the forms shown in FIGS. 3 and 4 above.
- the deepest portion 19a may be arranged at the end portion of the dynamic pressure groove 19 in the ring circumferential direction. Since the orbiting direction (rotational direction) of the orbiting scroll body is unidirectional, an asymmetrical shape can be achieved. In this case, the direction of rotation of the movable scroll body is the X direction. Further, as the bottom surface of the hydrodynamic groove 19, the above-described first inclined surface and second inclined surface can be appropriately employed. Further, as shown in FIG.
- the deepest portion 20a of the hydrodynamic groove 20 may be formed in a plane parallel to the sliding surface. Also, the deepest portion 20a may be curved. Also in this case, the first inclined surface and the second inclined surface as described above can be appropriately employed as the bottom surface of the dynamic pressure generating groove 20 . Further, in these examples, the flat surface forming the bottom surface of the groove for hydrodynamic bearing may be appropriately curved.
- the annular seal member 16 is mounted in an annular groove 4d provided on the back surface of the bottom plate portion 4a of the movable scroll.
- the left side of the drawing is the back pressure chamber 15a side
- the right side of the drawing is the space 15b side.
- the arrows in the drawing indicate the directions in which the pressure from the refrigerant or the like is applied.
- This sealing structure partitions the back pressure chamber 15a and the space 15b.
- a dynamic pressure is generated by introducing the refrigerant or the like into the dynamic pressure grooves 18 due to the flow of the refrigerant or the like caused by co-rotation.
- This dynamic pressure acts on the sliding surface of the annular seal member 16 in a direction away from the main bearing member 12 , thereby further reducing the sliding resistance of the annular seal member 16 with respect to the main bearing member 12 .
- the annular groove may be provided on the main bearing member side instead of the bottom plate portion 4a side of the orbiting scroll.
- the annular seal member mounted in the annular groove is fixed in the annular groove.
- the ring side surface of the annular seal member is in sliding contact with the back surface of the bottom plate portion of the orbiting orbiting scroll body. This ring side surface is provided with the hydrodynamic grooves as described above.
- the lubricating groove described in the second embodiment may be provided.
- the type of refrigerant, etc., is appropriately used according to the application.
- the temperature of the refrigerant or the like is, for example, about -20°C to 140°C.
- the rotational speed of the orbiting movement of the orbiting scroll body is mainly assumed to be about 5000 to 8000 rpm.
- the dynamic pressure grooves provided on the side surface of the ring may be grooves that generate dynamic pressure by introducing the fluid caused by the orbiting motion of the orbiting scroll body, and adopting various shapes.
- a herringbone see FIG. 7(a)
- a spiral see FIG. 7(b)
- FIG. 7 shows the planar shape of the groove for dynamic pressure
- the black portion in the figure is the groove for dynamic pressure.
- the dynamic pressure groove is desirably a groove (non-communication groove) that does not communicate the inner and outer diameters of the ring side surface of the annular seal member. In non-communication grooves, dynamic pressure is likely to occur because the flow of fluid is throttled in the middle.
- the folding position of the groove in the herringbone such as that shown in FIG. 7(a) can be set as appropriate.
- the force directed from the inner diameter side to the outer diameter side increases as the folding position moves toward the outside of the circumference.
- Examples of forming non-communicating dynamic pressure grooves in at least a part of the inner diameter side end of the ring side surface include the examples shown in FIGS. 2 to 5 described above.
- the example of FIG. 8 can be given.
- An annular seal member 21 shown in FIG. 8 is provided with a plurality of substantially V-shaped dynamic pressure grooves 23 recessed in the width direction of the ring along the circumferential direction at the outer diameter side end of the ring side surface 22 .
- FIG. 8 shows a partially enlarged view thereof.
- the hydrodynamic groove 23 has the same structure as the above-described V-shaped hydrodynamic groove 18 (see FIG. 3) except for the formation position on the ring side surface. It should be noted that the dynamic pressure groove 23 can appropriately adopt the configuration of the modified example of the dynamic pressure groove described above.
- the formation position of the non-communicating dynamic pressure grooves is not limited to only the inner diameter side end portion or the outer diameter side end portion of the ring side surface, but may be both the inner diameter side end portion and the outer diameter side end portion of the ring side surface.
- the dynamic pressure grooves on the inner diameter side end and the dynamic pressure grooves on the outer diameter side end may be alternately formed along the ring circumferential direction.
- the dynamic pressure grooves on the inner diameter side end and the dynamic pressure grooves on the outer diameter side end may be formed so as not to overlap in the ring radial direction.
- the annular seal member 31 is an annular body having a substantially rectangular cross section, has a shape that is continuous over the entire circumference, and does not have a joint.
- the ring side surface 32 of the annular seal member 31 is provided with a plurality of lubricating grooves that are open to either the ring outer diameter or the ring inner diameter.
- a plurality of outer lubrication grooves 33 that open to the ring outer peripheral surface 31 a are provided at the outer diameter side end of the ring side surface 32
- the ring inner peripheral lubrication groove 33 is provided at the inner diameter side end of the ring side surface 32
- a plurality of inner lubrication grooves 34 are provided that are open on the surface 31b. The outer lubrication groove 33 and the inner lubrication groove 34 do not penetrate the ring outer peripheral surface 31a and the ring inner peripheral surface 31b.
- the corners between the ring outer peripheral surface 31a and the ring side surfaces 32 on both sides including the outer diameter side lubrication grooves 33), and the ring inner peripheral surface 31b and the ring side surfaces 32 on both sides (including the inner diameter side lubrication grooves 34).
- the corners of and may be chamfered in a straight line or a curved line.
- a stepped portion 31c that protrudes from the mold may be provided at the corners of the ring inner peripheral surface 31b and the ring side surfaces 32 on both sides.
- lubrication grooves inner diameter side lubrication grooves 34 and outer diameter side lubrication grooves 33
- Boundaries (corners) between the lubricating grooves and the sliding surfaces are provided with inclined surfaces as shown in FIG.
- the coolant or the like flows appropriately to the portion that slides on the end surface of the main bearing member, so that the torque can be reduced.
- the lubricating grooves can also function as dynamic pressure grooves.
- a wedge action is generated by the flow of the coolant or the like into the lubricating grooves, which further reduces the torque and improves the low friction and wear resistance characteristics.
- the lubrication groove is a concave portion that does not communicate in the ring radial direction, it also leads to low oil leakage of refrigerant or the like.
- the lubrication grooves should be formed at least on the side of the ring on the side of the sliding surface that revolves and slides. It is preferably formed symmetrically on both sides of the ring.
- a plurality of outer diameter side lubrication grooves 33 and inner diameter side lubrication grooves 34 are provided at equal intervals in the ring circumferential direction.
- the outer lubricating grooves 33 and the inner lubricating grooves 34 are alternately provided along the ring circumferential direction when viewed from the ring side surface 32 .
- a ring side surface between adjacent lubricating grooves is a portion that slides on the main bearing member, and constitutes a part of the sliding surface.
- the area of the lubricating groove (the total area if there are multiple lubricating grooves, hereinafter the same) is not particularly limited, but if the area of the lubricating groove with respect to the ring side surface is too small, the torque reduction effect will be small, and if it is too large, excessive surface pressure will occur. Wear may be accelerated. From this point of view, the area of the lubricating groove is preferably 5% to 75%, more preferably 20% to 60%, of the entire area of the ring side surface.
- the total area of the ring side surface is the area (including the area of the lubrication groove) in a plan view of the ring side surface (one side) on which the annular seal member revolves and slides from the front.
- the area of the groove is the area in the same plan view.
- each lubricating groove in the ring circumferential direction is preferably about 0.5% to 5% of the ring circumferential length, depending on the number of lubrication grooves.
- the length of the lubricating groove in the radial direction of the ring is preferably 10% to 80% of the radial thickness of the sliding surface.
- all the lubrication grooves are of the same size and are spaced apart at approximately equal intervals (in FIG. 9, 12 lubrication grooves on the inner diameter side and 12 lubrication grooves on the outer 24 in total) is preferably provided.
- the number of lubrication grooves on the inner diameter side and the outer diameter side may not be the same.
- the inner diameter side lubrication grooves may be larger than the outer diameter side lubrication grooves.
- FIG. 10 is an enlarged view of part D in FIG.
- the planar shapes of the outer lubricating groove 33 and the inner lubricating groove 34 are arcuate.
- These lubricating grooves 33 and 34 are substantially conical recesses recessed in the width direction side (axial direction side) of the ring along the ring circumferential direction.
- the outer diameter side lubrication groove 33 is a substantially conical recess having a center axis on a circumference concentric with the center point of the ring outer diameter and having a larger diameter than the ring outer diameter.
- the inner lubrication groove 34 is a substantially conical recess having a central axis on a circumference concentric with the center point of the inner diameter of the ring and having a smaller diameter than the inner diameter of the ring.
- the depth of the lubricating groove 34 on the inner diameter side radially becomes shallower from the deepest part located at the inner diameter end part of the substantially central part in the ring circumferential direction to both sides in the ring circumferential direction and toward the ring outer diameter side.
- FIG. 11(a) is a view of the lubricating groove viewed from the inner diameter side of the ring
- FIG. 11(b) is a cross-sectional view taken along the line EE.
- the inner lubrication groove 34 has a substantially conical central axis at the inner diameter end of the sliding surface on the inner diameter side of the ring, and as shown in FIG. It is a concave portion that is not on the circumference of a smaller diameter than .
- the bottom surface of the inner lubricating groove 34 is formed with a surface corresponding to the outer peripheral surface of the cone, and extends linearly toward the central axis. It is formed by an inclined surface 34a.
- the depth from the sliding surface of the inner lubrication groove 34 has the deepest portion 34b at the central portion of the inner lubrication groove 34 in the ring circumferential direction, and becomes shallower radially from the deepest portion 34b. That is, the area closer to the sliding surface in the ring radial direction is shallower (see FIG. 11(b)).
- the inner lubricating groove 34 has a symmetrical shape centering on the deepest portion 34b.
- An inner diameter edge of the inner diameter side lubrication groove 34 is formed in a straight line.
- the bottom surface (inclined surface 34a) of the inner lubricating groove 34 is formed linearly.
- the inclination angle ⁇ (see FIGS. 11(a) and 11(b)) of the inclined surface 34a with respect to the sliding surface is not particularly limited, but is preferably in the range of 0.1° to 15°, more preferably 1° to 10°. ° range is more preferred.
- the coolant or the like tends to flow moderately into the portion that slides on the end surface of the main bearing member, and the wedging effect of the inflowing fluid can be effectively exhibited.
- the inclination angle ⁇ is less than 0.1°, the inflowing fluid becomes difficult to flow toward the sliding surface.
- the angle of inclination ⁇ may be the same or different over the entire boundary portion between the inner diameter side lubrication groove 34 and the sliding surface (the entire circumference of the arc of the lubrication groove in FIG. 10).
- the boundary portion between the lubrication groove and the sliding surface may be constituted by an inclined surface, and as shown in FIG. As shown, it may be composed of an inclined surface and a bottom surface parallel to the sliding surface.
- the inclined surface may be connected to the sliding surface in a straight line as shown in FIG. 11, or may be connected in a curved line (R shape) as shown in FIG. 14(b).
- R shape curved line
- the inclined surface 34a of the inner lubricating groove 34 shown in FIG. is also stabilized, and a further reduction in torque can be achieved.
- the rounded shape makes it easier for the refrigerant or the like to flow out from the sliding surface, making it easier to further reduce the torque.
- the depth from the sliding surface of the deepest portion 34b of the inner lubricating groove 34 is preferably 45% or less, more preferably 30% or less, of the total width of the ring. It should be noted that the "depth” here is the sum of the depths of the recesses on each side when lubricating grooves are formed on both sides of the ring. It is 22.5% or less of the total width, preferably 15% or less. If it exceeds 45% of the total width of the ring, the strength of the annular seal member may be insufficient and it may be deformed.
- the substantially mortar-shaped lubrication grooves are not limited to the forms shown in FIGS. 10 and 11.
- the inner diameter lubrication groove 35 may be formed as a substantially truncated conical recess, and the deepest portion 35b of the inner diameter lubrication groove 35 may be formed in a plane parallel to the sliding surface.
- the depth of the groove becomes shallow radially from the deepest portion 35b toward both sides in the ring circumferential direction and toward the ring outer diameter side.
- the deepest portion 35b may be formed with a curved surface.
- the flat surface forming the bottom surface of the lubricating groove may be appropriately curved.
- the annular seal member 31 is mounted in an annular groove 4d provided on the back surface of the bottom plate portion 4a of the movable scroll.
- This sealing structure partitions the back pressure chamber 15a and the space 15b.
- the annular seal member 31 rotates with the ring side surface 32 and comes into sliding contact with the end surface of the main bearing member 12 while revolving and sliding.
- a dynamic pressure is generated by introducing the coolant or the like into the inner diameter side lubrication groove 34 due to the flow of the coolant or the like caused by the co-rotation.
- This dynamic pressure acts on the sliding surface of the annular seal member 31 in a direction away from the main bearing member 12 , thereby further reducing the sliding resistance of the annular seal member 31 with respect to the main bearing member 12 .
- the lubricating groove provided on the side surface of the ring may be a groove that allows the fluid to flow appropriately to the portion that slides on the end surface of the main bearing member due to the flow of fluid generated by the orbiting motion of the orbiting scroll body.
- Various shapes can be employed.
- the bottom surface of the substantially mortar-shaped inner lubrication groove 36 may be formed with an inclined curved surface 36a (substantially spherical segment shape).
- the planar shape of the inner diameter side lubrication groove 37 may be formed in a substantially triangular shape.
- the lubricating groove is preferably a groove (non-communicating groove) that does not communicate with the inner and outer diameters of the ring side surface of the annular seal member. In non-communication grooves, dynamic pressure is likely to occur because the flow of fluid is throttled in the middle.
- non-communicating lubricating grooves in at least a part of the inner diameter side end and the outer diameter side end of the ring side surface include the examples shown in the above figures.
- the lubricating grooves described above may be formed only on the inner diameter side end of the ring side surface serving as the sliding surface, or may be formed only on the outer diameter side end.
- the material of the annular seal member of the present invention is not particularly limited, it is preferably made of synthetic resin.
- Synthetic resins that can be used include, for example, thermosetting polyimide resin, thermoplastic polyimide resin, polyether ketone ether ketone ketone resin, polyether ketone resin, PEEK resin, wholly aromatic polyester resin, polytetrafluoroethylene (hereinafter referred to as PTFE Fluororesins such as resins, PPS resins, polyamideimide resins, and polyamide resins. These resins may be used singly or as a polymer alloy in which two or more kinds are mixed.
- the annular seal member is preferably an injection molded body made by injection molding a synthetic resin. Therefore, it is preferable to use a thermoplastic resin that can be injection molded as the synthetic resin. Among them, it is preferable to use PEEK resin or PPS resin because they are particularly excellent in friction and abrasion properties, flexural modulus, heat resistance, slidability, and the like. These resins have a high modulus of elasticity, can be used even when the temperature of the sealing coolant or the like is high, and are free from solvent cracks.
- the above synthetic resin may be added with fibrous reinforcing materials such as carbon fiber, glass fiber, and aramid fiber, spherical fillers such as spherical silica and spherical carbon, scaly reinforcing materials such as mica and talc, and potassium titanate whiskers.
- fibrous reinforcing materials such as carbon fiber, glass fiber, and aramid fiber
- spherical fillers such as spherical silica and spherical carbon
- scaly reinforcing materials such as mica and talc
- potassium titanate whiskers such as can be blended.
- solid lubricants such as PTFE resin, graphite and molybdenum disulfide, sliding reinforcing materials such as calcium phosphate and calcium sulfate, and pigments such as carbon black and titanium oxide can also be blended. These may be blended singly or in combination.
- PEEK resin or PPS resin containing carbon fiber as a fibrous reinforcing material and PTFE resin as a solid lubricant is preferable because it facilitates obtaining the properties required for the annular seal member of the present invention.
- mechanical strength such as flexural modulus can be improved, and by blending PTFE resin, sliding property can be improved.
- the gate When using a synthetic resin, the above raw materials are melted and kneaded to form pellets for molding, which are then molded into a predetermined shape by a known injection molding method or the like.
- the position of the gate is not particularly limited, but it is preferably provided on the inner peripheral surface of the ring from the viewpoint of ensuring sealing performance and because post-processing is not required.
- the gate positions are multi-point gates (for example, 3 to 6 points) arranged at equal intervals in the circumferential direction, and that the gate positions and the positions of the dynamic pressure generating grooves do not overlap in the ring radial direction.
- the annular seal member has a gate mark on the inner peripheral surface at a position that does not overlap with the dynamic pressure groove in the ring radial direction.
- the sliding area was fixed and the load was divided into three levels and a thrust test was conducted.
- Example A and Comparative Example A Circular test pieces of Example A and Comparative Example A were produced by injection molding using a resin composition (BEAREE AS5302 manufactured by NTN) containing PPS resin as the main material and PTFE resin and carbon fiber.
- the test piece of Comparative Example A had an outer diameter of ⁇ 21 mm, an inner diameter of ⁇ 17 mm, a radial length of 2 mm, and an axial length of 1.6 mm, and no dynamic pressure groove was provided on the ring side surface.
- the test piece of Example A had an outer diameter of ⁇ 21 mm, an inner diameter of ⁇ 17 mm, a radial length of 2 mm, and an axial length of 1.6 mm.
- the deepest groove depth of the dynamic pressure groove is 0.1 mm
- the inclination angle of the first inclined surface with respect to the sliding surface is approximately 65°
- the inclination angle of the second inclined surface with respect to the sliding surface is was about 3°.
- the area of the hydrodynamic grooves was 40% of the total area of the side surface of the ring.
- FIG. 16 A schematic diagram of the thrust tester is shown in Fig. 16.
- a test piece 43 is attached to the tip of a load shaft 41, and a mating member 44 (ADC 12, outer diameter ⁇ 33 mm, thickness 10 mm, sliding surface with the test piece is polished to a Ra of about 0.8 ⁇ m by plane polishing. ) was pressed with a predetermined load F, and a thrust test was performed in the oil 42 under the following conditions. In each test, the dynamic friction coefficient was measured just before the end of the test.
- FIG. 17 shows the relationship between surface pressure and dynamic friction coefficient.
- the coefficient of dynamic friction tends to decrease as the surface pressure (load) increases. Diminished. Accordingly, by forming the hydrodynamic grooves as in the embodiment A, torque reduction can be achieved.
- Example B and Comparative Example B Circular test pieces of Example B and Comparative Example B were produced by injection molding using a resin composition (manufactured by NTN: BEAREE AS5302) containing PPS resin as the main material and PTFE resin and carbon fiber.
- the test piece of Comparative Example B had an outer diameter of ⁇ 21 mm, an inner diameter of ⁇ 17 mm, a radial length of 2 mm, and an axial length of 1.6 mm, and no lubricating groove was provided on the ring side surface.
- the test piece of Example B had an outer diameter of ⁇ 21 mm, an inner diameter of ⁇ 17 mm, a radial length of 2 mm, and an axial length of 1.6 mm. As shown in FIG.
- the deepest groove depth of the lubricating groove was 0.1 mm, and the inclination angle of the inclined surface with respect to the sliding surface was about 3°.
- the area of the lubricating groove was 40% of the total area of the ring side surface.
- Example A A thrust test was performed on these test pieces under the same conditions as in Example A. As a result, as shown in FIG. 17, the coefficient of dynamic friction tends to decrease as the surface pressure (load) increases. The result was that the torque) decreased. Accordingly, by forming the lubricating grooves as in the embodiment B, the torque can be reduced.
- the annular seal member of the present invention can exhibit stable low-torque performance without impairing the durability and sealing function, so it can be widely used as an annular seal member for scroll compressors. Also, it becomes possible to eliminate the thrust receiving member.
Abstract
Provided is an annular sealing member for a compressor that is capable of exhibiting a reliable low-torque property without harming durability or causing deterioration of a sealing function. In a scroll compressor comprising a fixed scroll body 3, a movable scroll body 4, a shaft 7, a main bearing 9 that rotatably supports the shaft 7, and a main bearing member 12 that fixes the main bearing 9, wherein, together with a fluid being compressed in a compression chamber 5 by the rotation of the shaft 7 causing the movable scroll body 4 to revolve about the axis of the fixed scroll body 3, the fluid is supplied to a back pressure chamber 15a on the back surface side of the movable scroll body 4, this annular sealing member 16 is mounted to at least one annular groove 4d formed in the back surface of a bottom plate part 4a of the movable scroll 4, and seals the back pressure chamber 15a. In a ring side surface of the annular sealing member 16, a dynamic pressure groove is provided to at least a sliding surface that revolves and slides.
Description
本発明は、スクロールコンプレッサを構成する可動スクロール体の底板部などに装着される環状シール部材に関する。
The present invention relates to an annular seal member that is attached to the bottom plate of a movable scroll that constitutes a scroll compressor.
スクロールコンプレッサは、固定スクロール体と、該固定スクロール体に対し旋回運動される可動スクロール体とからなるスクロール型の圧縮機構部を備える。固定スクロール体と可動スクロール体はそれぞれ、底板部と該底板部の表面に立設する渦巻壁とを有しており、それぞれ渦巻壁において互いに噛み合わされて、それらの間に圧縮室が形成されている。この圧縮室が固定スクロール体の軸線の周りを公転する可動スクロール体の作用により渦巻中心側に移動して冷媒などの圧縮が行なわれる。
The scroll compressor includes a scroll-type compression mechanism section consisting of a fixed scroll body and a movable scroll body that orbits with respect to the fixed scroll body. The fixed scroll body and the movable scroll body each have a bottom plate portion and a spiral wall erected on the surface of the bottom plate portion. there is The compression chamber is moved toward the center of the spiral by the action of the movable scroll revolving around the axis of the fixed scroll, thereby compressing the refrigerant.
可動スクロール体の底板部の背面側には環状シール部材が設けられている。このようなスクロールコンプレッサにおいて、冷媒などが圧縮されると、その圧縮反力によって可動スクロール体にスラスト荷重が発生する。このスラスト荷重に起因して、可動スクロール体の背面側に設けられた環状シール部材とそれと摺動する主軸受部材との間で摩擦力が大きくなり、環状シール部材の摩耗などが発生するおそれがある。
An annular seal member is provided on the back side of the bottom plate of the movable scroll. In such a scroll compressor, when a refrigerant or the like is compressed, a thrust load is generated on the movable scroll body due to the compression reaction force. Due to this thrust load, frictional force increases between the annular seal member provided on the back side of the orbiting scroll body and the main bearing member that slides therewith, and there is a risk that the annular seal member will be worn. be.
このような環状シール部材の摩擦摩耗の対策として、オイルなどの潤滑剤を使用して摩擦摩耗の低減を図る方法が知られている（特許文献１参照）。
As a countermeasure against such frictional wear of the annular seal member, there is known a method of using a lubricant such as oil to reduce the frictional wear (see Patent Document 1).
また、別の方法として、可動スクロール体から主軸受部材へ一方的にかかるスラスト荷重を低減させる目的として、吐出圧領域と背圧室とを圧力導入孔を介して接続する方法が知られている。さらにこの方法において、環状シール部材の側面に、径方向に連通した溝を設けることで、背圧室と吸入圧領域を連通させることも知られている（特許文献２参照）。しかし、この特許文献２は、背圧室および吸入圧領域の雰囲気が一様ではない場合において、背圧室で意図する背圧を設定しやすくするという技術である。
Another known method is to connect the discharge pressure area and the back pressure chamber via a pressure introduction hole for the purpose of reducing the thrust load that is unilaterally applied from the movable scroll body to the main bearing member. . Furthermore, in this method, it is also known that a back pressure chamber and a suction pressure area are communicated by providing a groove communicating in the radial direction on the side surface of the annular seal member (see Patent Document 2). However, Patent Literature 2 is a technique that makes it easier to set the intended back pressure in the back pressure chamber when the atmosphere in the back pressure chamber and the suction pressure region is not uniform.
さらに、別の方法として、上記の環状シール部材とは別の部材として、可動スクロール体の底板部側から主軸受部材側へのスラスト力を受けるスラスト受け部材を介装することで、荷重を低減する手段も知られている（特許文献３参照）。
Furthermore, as another method, the load is reduced by inserting a thrust receiving member that receives the thrust force from the bottom plate portion side of the orbiting scroll body toward the main bearing member side as a member separate from the annular seal member. There is also known means for doing so (see Patent Document 3).
環状シール部材の摩擦摩耗の対策として、例えば潤滑剤などを使用することで、摩擦摩耗の低減が図れるものの、この方法では摺動面を常に良い潤滑状態とする必要がある。そのため、局所的に潤滑剤切れが発生した場合などはトルクが安定しないことから、コンプレッサ自体の安定した性能を維持させることに懸念がある。一方、環状シール部材とは別の部材として、スラスト受け部材を介装する場合は、その分、部品点数が多くなり、ユニット全体のコストアップに繋がるおそれがある。
As a countermeasure against frictional wear of the annular seal member, it is possible to reduce frictional wear by using, for example, a lubricant, but this method requires that the sliding surface is always in a well-lubricated state. Therefore, when the lubricant runs out locally, the torque becomes unstable, and there is concern about maintaining stable performance of the compressor itself. On the other hand, if a thrust receiving member is interposed as a separate member from the annular seal member, the number of parts increases accordingly, which may lead to an increase in the cost of the entire unit.
本発明はこのような事情に鑑みてなされたものであり、耐久性やシール機能の低下を損なうことなく、安定した低トルク性を発揮できるコンプレッサの環状シール部材を提供することを目的とする。
The present invention has been made in view of such circumstances, and an object of the present invention is to provide an annular seal member for a compressor that can exhibit stable low torque performance without impairing durability and sealing function.
本発明の環状シール部材は、底板部とその表面に立設する渦巻壁を有する固定スクロール体と、底板部とその表面に立設する渦巻壁を有する可動スクロール体と、シャフトと、該シャフトを回転可能に支持する主軸受と、該主軸受を固定する主軸受部材とを備え、上記シャフトの回転により、上記可動スクロール体を上記固定スクロール体の軸線の周りで公転させて流体を圧縮室にて圧縮するとともに、上記流体が上記可動スクロール体の背面側の背圧室に供給されるスクロールコンプレッサにおいて、上記可動スクロール体の上記底板部の背面と、上記主軸受部材の上記可動スクロール体に向く端面のいずれか一方の面に形成された少なくとも１個の環状溝に装着され、上記背圧室をシールする環状シール部材であって、上記環状シール部材は、リング側面において少なくとも公転摺動する摺動面に動圧溝が設けられていることを特徴とする。
The annular seal member of the present invention comprises: a fixed scroll body having a bottom plate and a spiral wall standing on its surface; a movable scroll body having a bottom plate and a spiral wall standing on its surface; a shaft; a main bearing that rotatably supports the main bearing; and a main bearing member that fixes the main bearing. Rotation of the shaft causes the orbiting scroll to revolve around the axis of the fixed scroll, thereby supplying fluid to the compression chamber. and the fluid is supplied to a back pressure chamber on the back side of the orbiting scroll, the fluid being directed toward the back of the bottom plate portion of the orbiting scroll and the orbiting scroll of the main bearing member. An annular seal member mounted in at least one annular groove formed in one of the end faces and sealing the back pressure chamber, wherein the annular seal member is a sliding member that at least revolves and slides on the ring side surface. A dynamic pressure groove is provided on the dynamic surface.
本発明において、動圧溝は、可動スクロール体の旋回運動によって生じる流体の流れにより該流体を導入して動圧を発生させる溝であり、流体が絞り込まれるように溝の断面積が流体の流動方向に向かって減少する溝である。減少する方向は、溝の深さ方向（図３）であっても溝の幅方向（図７）であっても、あるいはその両方であってもよい。
In the present invention, the dynamic pressure groove is a groove that introduces the fluid generated by the orbiting motion of the orbiting scroll body to generate dynamic pressure. It is a groove that decreases in the direction. The decreasing direction may be in the depth direction of the groove (FIG. 3), the width direction of the groove (FIG. 7), or both.
上記動圧溝の面積は上記リング側面の全体の面積に対して５％～７５％であることを特徴とする。
The area of the dynamic pressure groove is 5% to 75% of the total area of the ring side surface.
上記動圧溝の形状は、リング周方向に沿ってリングの幅方向側に凹んだ略Ｖ字状であり、上記動圧溝の上記摺動面からの深さは、最深部からリング周方向の両端部に向けて浅くなり、リング径方向には一定であることを特徴とする。
The shape of the dynamic pressure groove is a substantially V shape recessed in the width direction of the ring along the ring circumferential direction, and the depth from the sliding surface of the dynamic pressure groove is the ring circumferential direction from the deepest part. It becomes shallow toward both ends of the ring and is constant in the radial direction of the ring.
上記動圧溝は、上記摺動面から上記最深部に至るまで同一平面ではなく、上記摺動面に接続される第１の傾斜面と、上記最深部に接続され、上記摺動面に対して、上記第１の傾斜面よりも小さな傾斜角度をなす第２の傾斜面とを有することを特徴とする。
The dynamic pressure groove is not flush from the sliding surface to the deepest portion, but is connected to the first inclined surface connected to the sliding surface and the deepest portion, and is connected to the sliding surface. and a second inclined surface forming an inclination angle smaller than that of the first inclined surface.
上記摺動面に対する上記第１の傾斜面の傾斜角度が５０°～８０°であり、上記摺動面に対する上記第２の傾斜面の傾斜角度が０．１°～１５°であることを特徴とする。
The inclination angle of the first inclined surface with respect to the sliding surface is 50° to 80°, and the inclination angle of the second inclined surface with respect to the sliding surface is 0.1° to 15°. and
上記動圧溝において、上記第１の傾斜面と上記第２の傾斜面の境界部が曲面で接続されていることを特徴とする。
The dynamic pressure groove is characterized in that a boundary portion between the first inclined surface and the second inclined surface is connected by a curved surface.
上記動圧溝がリング周方向で離間して複数個設けられ、隣り合う動圧溝同士の間のリング側面が上記摺動面の一部を構成することを特徴とする。
A plurality of the dynamic pressure grooves are spaced apart in the ring circumferential direction, and the ring side surface between the adjacent dynamic pressure grooves constitutes a part of the sliding surface.
上記環状シール部材は合成樹脂製であり、該合成樹脂がポリフェニレンサルファイド（以下、ＰＰＳと記す）樹脂またはポリエーテルエーテルケトン（以下、ＰＥＥＫと記す）樹脂であることを特徴とする。
The annular seal member is made of synthetic resin, and the synthetic resin is polyphenylene sulfide (hereinafter referred to as PPS) resin or polyether ether ketone (hereinafter referred to as PEEK) resin.
本発明の環状シール部材は、底板部とその表面に立設する渦巻壁を有する固定スクロール体と、底板部とその表面に立設する渦巻壁を有する可動スクロール体と、シャフトと、該シャフトを回転可能に支持する主軸受と、該主軸受を固定する主軸受部材とを備え、上記シャフトの回転により、上記可動スクロール体を上記固定スクロール体の軸線の周りで公転させて流体を圧縮室にて圧縮するとともに、上記流体が上記可動スクロール体の背面側の背圧室に供給されるスクロールコンプレッサにおいて、上記可動スクロール体の上記底板部の背面と、上記主軸受部材の上記可動スクロール体に向く端面のいずれか一方の面に形成された少なくとも１個の環状溝に装着され、上記背圧室をシールする環状シール部材であって、上記環状シール部材は、リング側面において少なくとも公転摺動する摺動面に、リング外径およびリング内径のいずれか一方に開口した潤滑溝が複数設けられており、上記潤滑溝と上記摺動面との境界部には傾斜面が設けられていることを特徴とする。
The annular seal member of the present invention comprises: a fixed scroll body having a bottom plate and a spiral wall standing on its surface; a movable scroll body having a bottom plate and a spiral wall standing on its surface; a shaft; a main bearing that rotatably supports the main bearing; and a main bearing member that fixes the main bearing. Rotation of the shaft causes the orbiting scroll to revolve around the axis of the fixed scroll, thereby supplying fluid to the compression chamber. and the fluid is supplied to a back pressure chamber on the back side of the orbiting scroll, the fluid being directed toward the back of the bottom plate portion of the orbiting scroll and the orbiting scroll of the main bearing member. An annular seal member mounted in at least one annular groove formed in one of the end faces and sealing the back pressure chamber, wherein the annular seal member is a sliding member that at least revolves and slides on the ring side surface. The sliding surface is provided with a plurality of lubrication grooves that are open to either the outer diameter of the ring or the inner diameter of the ring, and an inclined surface is provided at the boundary between the lubrication groove and the sliding surface. and
上記潤滑溝の面積は上記リング側面の全体の面積に対して５％～７５％であることを特徴とする。
The area of the lubricating groove is 5% to 75% of the total area of the ring side surface.
上記潤滑溝は略すり鉢状であることを特徴とする。
The lubricating groove is characterized by being substantially mortar-shaped.
上記環状シール部材は、上記潤滑溝として、リング外径に開口した外径側潤滑溝とリング内径に開口した内径側潤滑溝とを有することを特徴とする。さらに、上記外径側潤滑溝および上記内径側潤滑溝はそれぞれ、リング周方向で離間して複数個設けられ、上記リング側面から見て上記外径側潤滑溝と上記内径側潤滑溝とがリング周方向に交互に設けられていることを特徴とする。
The annular seal member is characterized by having, as the lubricating grooves, an outer lubricating groove opening to the ring outer diameter and an inner lubricating groove opening to the ring inner diameter. Further, a plurality of the outer-diameter-side lubrication grooves and the inner-diameter-side lubrication grooves are provided separately in the ring circumferential direction. It is characterized by being provided alternately in the circumferential direction.
上記傾斜面は上記摺動面に対する角度が０．１°～１５°であることを特徴とする。
The inclined surface has an angle of 0.1° to 15° with respect to the sliding surface.
上記環状シール部材は合成樹脂製であり、該合成樹脂がＰＰＳ樹脂またはＰＥＥＫ樹脂であることを特徴とする。
The annular seal member is made of synthetic resin, and the synthetic resin is PPS resin or PEEK resin.
本発明の環状シール部材は、スクロールコンプレッサにおいて可動スクロール体の底板部の背面と、主軸受部材の可動スクロール体に向く端面のいずれか一方の面に形成された環状溝に装着され、背圧室をシールするシール部材であり、リング側面において少なくとも公転摺動する摺動面に動圧溝が設けられているので、摺動面積を小さくすることができ、摺動面積が小さくなることで、摩擦係数の面圧依存性により、摺動トルクが低下する。さらに、動圧溝とすることで、その動圧溝に流体が流入し、くさび作用が発生しやすくなり、一層の低トルク化に繋がる。これにより、摩擦摩耗特性が向上し、スラスト受け部材を用いなくても、耐久性やシール機能の低下を損なうことなく、安定した低トルク性を発揮できる。
The annular seal member of the present invention is mounted in the annular groove formed in either the back surface of the bottom plate of the movable scroll body or the end surface of the main bearing member facing the movable scroll body in the scroll compressor, and At least the sliding surface that revolves and slides on the side surface of the ring is provided with dynamic pressure grooves, so the sliding area can be reduced. The sliding torque decreases due to the surface pressure dependence of the coefficient. Furthermore, by forming the dynamic pressure grooves, the fluid flows into the dynamic pressure grooves, making it easier for the wedge action to occur, leading to a further reduction in torque. As a result, frictional wear characteristics are improved, and stable low torque performance can be exhibited without impairing durability and sealing function without using a thrust receiving member.
動圧溝の面積はリング側面の全体の面積に対して５％～７５％であるので、トルク低減効果を確保しつつ、摩耗の促進を抑えられる。
The area of the dynamic pressure groove is 5% to 75% of the total area of the ring side surface, so it is possible to suppress the acceleration of wear while ensuring the torque reduction effect.
動圧溝の形状は、リング周方向に沿ってリングの幅方向側に凹んだ略Ｖ字状であり、動圧溝の摺動面からの深さは、最深部からリング周方向の両端部に向けて浅くなり、リング径方向には一定であるので、動圧溝に流体を導入しやすくなり、くさび作用を発生させやすくなる。
The shape of the groove for hydrodynamic bearing is a substantially V shape recessed in the width direction of the ring along the circumferential direction of the ring. , and is constant in the radial direction of the ring. Therefore, fluid is easily introduced into the dynamic pressure grooves, and a wedge action is likely to occur.
動圧溝は、摺動面から最深部に至るまで同一平面ではなく、摺動面に接続される第１の傾斜面と、最深部に接続され、摺動面に対して、第１の傾斜面よりも小さな傾斜角度をなす第２の傾斜面とを有し、第１の傾斜面は、第２の傾斜面よりも摺動面に対して急勾配に形成されているので、摺動面が摩耗した場合でも動圧溝の開口面積の減少が小さく（つまり、摺動面積の増加が小さく）、トルクの変化を生じにくくできる。特に、摺動面に対する第１の傾斜面の傾斜角度が５０°～８０°であるので、くさび作用を効果的に発生させつつ、摺動面積の増加を抑えることができる。また、摺動面に対する第２の傾斜面の傾斜角度が０．１°～１５°であるので、流入した流体によるくさび作用を効果的に発生させることができる。
The hydrodynamic groove is not coplanar from the sliding surface to the deepest part, but is connected to the first inclined surface connected to the sliding surface to the deepest part and has the first inclination with respect to the sliding surface. and a second inclined surface forming an inclination angle smaller than that of the second inclined surface. Even if the groove wears, the opening area of the dynamic pressure generating groove is less reduced (that is, the sliding area is less increased), and the torque is less likely to change. In particular, since the inclination angle of the first inclined surface with respect to the sliding surface is 50° to 80°, an increase in the sliding area can be suppressed while effectively generating the wedge action. Moreover, since the inclination angle of the second inclined surface with respect to the sliding surface is 0.1° to 15°, the wedging effect of the inflowing fluid can be effectively generated.
動圧溝において、第１の傾斜面と第２の傾斜面の境界部が曲面で接続されており、Ｒ状に形成されているので、例えば、隣り合う動圧溝同士の間の摺動面に流体を流出させやすくなり、更なる低トルク化が図れる。
In the groove for hydrodynamic bearing, the boundary between the first inclined surface and the second inclined surface is connected by a curved surface, and is formed in an R shape. This makes it easier for the fluid to flow out, and the torque can be further reduced.
また、本発明の環状シール部材は、スクロールコンプレッサにおいて可動スクロール体の底板部の背面と、主軸受部材の可動スクロール体に向く端面のいずれか一方の面に形成された環状溝に装着され、背圧室をシールするシール部材であり、リング側面において少なくとも公転摺動する摺動面に、リング外径およびリング内径のいずれか一方に開口した潤滑溝が複数設けられており、当該潤滑溝と摺動面との境界部には傾斜面が設けられているので、摺動面積を小さくすることができ、摺動面積が小さくなることで、摩擦係数の面圧依存性により、摺動トルクが低下する。さらに、潤滑溝と摺動面との境界部には傾斜面が設けられているので、潤滑溝に流入した流体により、くさび作用（動圧効果）が発生しやすくなり、一層の低トルク化に繋がる。これにより、摩擦摩耗特性が向上し、スラスト受け部材を用いなくても、耐久性やシール機能の低下を損なうことなく、安定した低トルク性を発揮できる。
Further, the annular seal member of the present invention is mounted in the annular groove formed in either the back surface of the bottom plate of the movable scroll body or the end surface of the main bearing member facing the movable scroll body in the scroll compressor. It is a seal member that seals the pressure chamber, and has a plurality of lubrication grooves that are open to either the outer diameter of the ring or the inner diameter of the ring at least on the sliding surface that revolves and slides on the side surface of the ring. Since an inclined surface is provided at the boundary with the moving surface, the sliding area can be reduced, and by reducing the sliding area, sliding torque is reduced due to the dependence of the friction coefficient on surface pressure. do. Furthermore, since an inclined surface is provided at the boundary between the lubrication groove and the sliding surface, the wedge action (dynamic pressure effect) is likely to occur due to the fluid that has flowed into the lubrication groove, further reducing torque. Connect. As a result, frictional wear characteristics are improved, and stable low torque performance can be exhibited without impairing durability and sealing function without using a thrust receiving member.
潤滑溝の面積はリング側面の全体の面積に対して５％～７５％であるので、トルク低減効果を確保しつつ、摩耗の促進を抑えられる。
The area of the lubrication groove is 5% to 75% of the total area of the ring side, so it is possible to suppress the acceleration of wear while ensuring the torque reduction effect.
潤滑溝の形状は、略すり鉢状であるので、旋回運動によってくさび作用発生部が常に変化する用途では、溝深さがリング径方向に対して一定である溝に比べて、潤滑溝に流体を導入しやすくなり、くさび作用がより発生しやすくなる。なお、略すり鉢状とは、すり鉢状の一部分を有するという意味であり、すり鉢状とは、円錐状の凹み、円錐台状の凹み、球欠状の凹みなどを意味する。
Since the shape of the lubricating groove is approximately mortar-shaped, in applications where the wedge action generating part constantly changes due to the orbital motion, the lubricating groove has less fluid flow than the groove whose groove depth is constant in the radial direction of the ring. It becomes easier to introduce and the wedging action is more likely to occur. Note that the term “substantially mortar-shaped” means having a portion of a mortar shape, and the term “mortar-shaped” means a conical dent, a truncated conical dent, a truncated conical dent, or the like.
また、外径側および内径側の両方に、かつ、リング周方向に交互に複数個の潤滑溝を設けることで、旋回運動する摺動面の全域において潤滑効果を得ることができる。
In addition, by alternately providing a plurality of lubricating grooves in both the outer diameter side and the inner diameter side in the circumferential direction of the ring, it is possible to obtain a lubricating effect over the entire sliding surface in orbital motion.
また、傾斜面は摺動面に対する傾斜角度が０．１°～１５°であるので、流入した流体によるくさび作用を効果的に発生させることができる。
In addition, since the inclined surface has an inclination angle of 0.1° to 15° with respect to the sliding surface, it is possible to effectively generate a wedge action by the inflowing fluid.
本発明の環状シール部材を備えるスクロール型コンプレッサの一例を図１に基づいて説明する。図１はスクロール型コンプレッサの一部断面図である。このスクロール型コンプレッサは、炭酸ガスなどの冷媒、ポリアルキレングリコール油（ＰＡＧ油）などの冷凍機油、またはこれらの混合物など（以下、まとめて冷媒等と称す）の流体を圧縮する圧縮機である。
An example of a scroll compressor provided with the annular seal member of the present invention will be described with reference to FIG. FIG. 1 is a partial cross-sectional view of a scroll compressor. This scroll compressor is a compressor that compresses a fluid such as a refrigerant such as carbon dioxide gas, a refrigerating machine oil such as polyalkylene glycol oil (PAG oil), or a mixture thereof (hereinafter collectively referred to as refrigerant or the like).
図１において、コンプレッサ１は、ハウジング２の内部に圧縮機構部とモータ機構部とを有し、吸入口（図示省略）および吐出口（図示省略）によって外部と接続されている。圧縮機構部は、吸入口より吸入した冷媒等を圧縮して吐出口より吐出する部分であり、固定スクロール体３と可動スクロール体４とから構成されている。固定スクロール体３は、底板部３ａと、この底板部３ａから垂直に立設した渦巻壁３ｂとを備え、中心に開口部３ｃが設けられている。また、可動スクロール体４は、底板部４ａと、この底板部４ａから垂直に立設した渦巻壁４ｂとを備える。固定スクロール体３および可動スクロール体４は偏心状態にかみ合わされて配置され、各スクロール体の渦巻壁３ｂ、４ｂの間に圧縮室５が形成されている。
In FIG. 1, the compressor 1 has a compression mechanism portion and a motor mechanism portion inside a housing 2, and is connected to the outside through a suction port (not shown) and a discharge port (not shown). The compression mechanism section is a section for compressing the refrigerant sucked from the suction port and discharging it from the discharge port. The fixed scroll member 3 includes a bottom plate portion 3a and a spiral wall 3b vertically erected from the bottom plate portion 3a, and an opening 3c is provided in the center. The movable scroll body 4 also includes a bottom plate portion 4a and a spiral wall 4b vertically erected from the bottom plate portion 4a. The fixed scroll body 3 and the movable scroll body 4 are arranged in an eccentric manner, and a compression chamber 5 is formed between the spiral walls 3b and 4b of each scroll body.
なお、図示は省略するが、各スクロール体の渦巻壁３ｂ、４ｂの軸方向端面には渦巻き状のシール部材（チップシール）が装着されている。これにより、圧縮室内の冷媒等の漏洩を防止する。
Although not shown, a spiral sealing member (tip seal) is attached to the axial end faces of the spiral walls 3b and 4b of each scroll body. This prevents leakage of refrigerant or the like in the compression chamber.
モータ機構部は、可動スクロール体４に旋回駆動力を与える部分であり、ステータ６ａとロータ６ｂとから構成されている。ステータ６ａは、ハウジング２の内側に固定されており、ロータ６ｂはシャフト７に結合している。ステータ６ａおよびロータ６ｂは電動機を構成し、ステータ６ａへの通電によりロータ６ｂおよびシャフト７が一体回転する。シャフト７は主軸受９および副軸受１０を介して回転可能に支持されている。シャフト７の一端側には偏心軸７ａが一体に形成され、これにバランスウエイト８が支持されている。シャフト７およびバランスウェイト８によって回転部材が構成されている。
The motor mechanism section is a section that applies turning driving force to the movable scroll body 4, and is composed of a stator 6a and a rotor 6b. The stator 6a is fixed inside the housing 2 and the rotor 6b is connected to the shaft 7. As shown in FIG. The stator 6a and the rotor 6b constitute an electric motor, and when the stator 6a is energized, the rotor 6b and the shaft 7 rotate together. Shaft 7 is rotatably supported via main bearing 9 and sub-bearing 10 . An eccentric shaft 7a is formed integrally with one end of the shaft 7, and a balance weight 8 is supported thereon. A rotating member is configured by the shaft 7 and the balance weight 8 .
可動スクロール体４の底板部４ａの背面側の略中央にはボス部４ｃが垂直に突出するように設けられ、このボス部４ｃ内に旋回軸受１１が圧入されている。旋回軸受１１に偏心軸７ａが支持されており、可動スクロール体４は、旋回軸受１１により旋回運動する機構となっている。
A boss portion 4c is provided so as to protrude vertically from substantially the center of the back side of the bottom plate portion 4a of the movable scroll body 4, and a turning bearing 11 is press-fitted into the boss portion 4c. The eccentric shaft 7 a is supported by the orbiting bearing 11 , and the movable scroll body 4 is a mechanism for orbiting by the orbiting bearing 11 .
主軸受９は、主軸受部材１２の中央側に形成された軸受支持部に固定されている。主軸受部材１２は、ハウジング内に固定されており、主軸受部材１２には固定スクロール体３がボルトなどによって結合されている。また、主軸受９の側方であって、シャフト７の外周面と主軸受部材１２との間にはシャフトシール１３が装着されている。このシャフトシール１３によって、モータ室１４と背圧室１５ａとの連通が遮断されている。
The main bearing 9 is fixed to a bearing support portion formed on the central side of the main bearing member 12 . The main bearing member 12 is fixed within the housing, and the fixed scroll body 3 is coupled to the main bearing member 12 with bolts or the like. A shaft seal 13 is mounted between the outer peripheral surface of the shaft 7 and the main bearing member 12 on the side of the main bearing 9 . The shaft seal 13 blocks communication between the motor chamber 14 and the back pressure chamber 15a.
ここで、主軸受部材１２と、可動スクロール体４の底板部４ａの背面との間には環状シール部材１６が設けられている。図１では、可動スクロール体４の底板部４ａの背面に形成された環状溝４ｄに、環状シール部材１６が装着されている。この構造では、環状シール部材１６は、主軸受部材１２の可動スクロール体に向く端面に対して公転摺動する。背圧室１５ａは、環状シール部材１６とシャフトシール１３とによってシールされ、これらシール部と、主軸受部材１２と、可動スクロール体４の底板部４ａとの間で密封空間を形成している。
An annular seal member 16 is provided between the main bearing member 12 and the back surface of the bottom plate portion 4a of the movable scroll body 4. In FIG. 1, an annular seal member 16 is mounted in an annular groove 4d formed in the back surface of the bottom plate portion 4a of the orbiting scroll body 4. As shown in FIG. In this structure, the annular seal member 16 revolves and slides on the end face of the main bearing member 12 facing the movable scroll body. The back pressure chamber 15a is sealed by an annular seal member 16 and a shaft seal 13, and these seal portions, the main bearing member 12, and the bottom plate portion 4a of the movable scroll body 4 form a sealed space.
コンプレッサ１が運転を開始すると、ロータ６ｂの回転により可動スクロール体４が旋回運動を始める。吸入口より圧縮機構部に入った冷媒等は、旋回する渦巻壁の外周から中心に移動しながら圧縮され、固定スクロール体３の開口部３ｃより外部に吐出される。一方、背圧室１５ａには、圧縮機構部内から加圧された流体が、可動スクロール体４の底板部４ａに設けられた圧力導入孔（図示省略）を通して供給されるようになっている。背圧室１５ａに加圧流体を導入することにより、圧縮反力によって可動スクロール体４に作用するスラスト荷重（可動スクロール体４を主軸受部材側に押し付けようとする力）を低減するように、または、可動スクロール体４を固定スクロール体側に押し付けるように、背圧室内の圧力が可動スクロール体４に作用することになる。
When the compressor 1 starts operating, the rotor 6b rotates, causing the orbiting scroll 4 to start orbiting. Refrigerant or the like that has entered the compression mechanism through the suction port is compressed while moving from the outer periphery to the center of the swirling spiral wall, and is discharged to the outside through the opening 3c of the fixed scroll body 3. As shown in FIG. On the other hand, the back pressure chamber 15a is supplied with pressurized fluid from inside the compression mechanism portion through a pressure introduction hole (not shown) provided in the bottom plate portion 4a of the movable scroll body 4. As shown in FIG. By introducing the pressurized fluid into the back pressure chamber 15a, the thrust load acting on the orbiting scroll 4 due to the compression reaction force (the force that presses the orbiting scroll 4 toward the main bearing member) is reduced. Alternatively, the pressure in the back pressure chamber acts on the orbiting scroll member 4 so as to press the orbiting scroll member 4 against the fixed scroll member.
環状シール部材１６は、内側の背圧室１５ａと外側の空間１５ｂとを仕切っている。空間１５ｂは、吸入圧に近い圧力値を有しているのに対して、背圧室１５ａには圧縮された冷媒等が導入されることから、空間１５ｂよりも背圧室１５ａの方が高圧となる。その結果、環状シール部材１６の一方のリング側面が主軸受部材１２の端面に公転しながら摺動接触する。環状シール部材１６は主に樹脂製であるのに対して、主軸受部材１２は金属製（鉄製やアルミダイカスト製）であり、摺動接触によって環状シール部材１６の摩耗などが懸念される。特に、流体の圧縮圧力が大きくなるほど、可動スクロール体４に作用するスラスト荷重も大きくなり、環状シール部材１６が摩耗しやすくなる。
The annular seal member 16 partitions the inner back pressure chamber 15a and the outer space 15b. While the space 15b has a pressure value close to the suction pressure, the pressure in the back pressure chamber 15a is higher than that in the space 15b because compressed refrigerant or the like is introduced into the back pressure chamber 15a. becomes. As a result, one ring side surface of the annular seal member 16 comes into sliding contact with the end surface of the main bearing member 12 while revolving. The annular seal member 16 is mainly made of resin, whereas the main bearing member 12 is made of metal (made of iron or aluminum die-cast), and there is concern about abrasion of the annular seal member 16 due to sliding contact. In particular, the greater the compression pressure of the fluid, the greater the thrust load acting on the orbiting scroll body 4, and the more likely the annular seal member 16 will wear.
本発明では、環状シール部材１６のリング側面に動圧溝を設けることで、摺動面積を低下させ、ひいては低トルク化を図っている。また、くさび作用により更なる低トルク化を図ることができる。
In the present invention, by providing a dynamic pressure groove on the ring side surface of the annular seal member 16, the sliding surface area is reduced, which in turn reduces the torque. In addition, the wedge action can further reduce the torque.
以下には、本発明の環状シール部材について説明する。
The annular seal member of the present invention will be described below.
（第１実施形態）
本発明の環状シール部材の第１実施形態を図２に基づいて説明する。図２は環状シール部材の斜視図を示す。コンプレッサの吐出量を確保する観点から、環状シール部材１６の外径寸法φは例えば５０ｍｍ以上であり、好ましい範囲としては５０ｍｍ～１００ｍｍ程度である（後述する第２実施形態の環状シール部材も同様）。 (First embodiment)
A first embodiment of the annular seal member of the present invention will be described with reference to FIG. FIG. 2 shows a perspective view of an annular seal member. From the viewpoint of securing the discharge amount of the compressor, the outer diameter dimension φ of theannular seal member 16 is, for example, 50 mm or more, and a preferable range is about 50 mm to 100 mm (the same applies to the annular seal member of the second embodiment described later). .
本発明の環状シール部材の第１実施形態を図２に基づいて説明する。図２は環状シール部材の斜視図を示す。コンプレッサの吐出量を確保する観点から、環状シール部材１６の外径寸法φは例えば５０ｍｍ以上であり、好ましい範囲としては５０ｍｍ～１００ｍｍ程度である（後述する第２実施形態の環状シール部材も同様）。 (First embodiment)
A first embodiment of the annular seal member of the present invention will be described with reference to FIG. FIG. 2 shows a perspective view of an annular seal member. From the viewpoint of securing the discharge amount of the compressor, the outer diameter dimension φ of the
図２に示すように、環状シール部材１６は、断面略矩形の環状体であり、全周にわたって繋がった形状であり、合い口を有していない。図２において、環状シール部材１６には、リング側面１７の内径側端部に、周方向に沿ってリングの幅方向側に凹んだＶ字状の動圧溝１８が複数設けられている。また、内周面１６ｂと両側のリング側面１７（動圧溝１８を含む）との角部は直線状、曲線状の面取りが設けられていてもよく、環状シール部材を射出成形で製造する場合、該部分に金型からの突出し部分となる段部１６ｃを設けてもよい。
As shown in FIG. 2, the annular seal member 16 is an annular body with a substantially rectangular cross section, has a shape that is continuous over the entire circumference, and does not have a joint. In FIG. 2, the annular seal member 16 is provided with a plurality of V-shaped dynamic pressure grooves 18 recessed in the width direction of the ring along the circumferential direction at the inner diameter side end of the ring side surface 17 . In addition, the corners of the inner peripheral surface 16b and the ring side surfaces 17 (including the dynamic pressure grooves 18) on both sides may be chamfered in a straight line or a curved line. A stepped portion 16c may be provided at this portion as a projecting portion from the mold.
図２に示すように、環状シール部材１６は、一方のリング側面が主軸受部材の可動スクロール体に向く端面と摺動する側の面となり、このリング側面に主軸受部材の端面との非接触部となるＶ字状の動圧溝１８が形成されている。この動圧溝１８を設けることで、冷媒等が該動圧溝に流入して、くさび作用が発生し、また冷媒等が主軸受部材の端面と摺動する部分に適度に流出することで、低トルク化が図れ、低摩擦耐摩耗特性を向上させることができる。また、図２の構成では、動圧溝１８はリング径方向に非連通の凹部であり、リング内径側にのみ開口していることから、冷媒等の低オイルリーク性にも繋がる。
As shown in FIG. 2, one ring side surface of the annular seal member 16 is a surface that slides against the end surface of the main bearing member facing the orbiting scroll body. A V-shaped dynamic pressure groove 18 is formed as a portion. By providing the dynamic pressure grooves 18, the coolant or the like flows into the dynamic pressure grooves to generate a wedge effect, and the coolant or the like flows appropriately to the portion where the end face of the main bearing member slides, Low torque can be achieved, and low friction and wear resistance characteristics can be improved. In addition, in the configuration of FIG. 2, the dynamic pressure grooves 18 are recesses that do not communicate in the ring radial direction, and are open only on the ring inner diameter side, which leads to low oil leakage of refrigerant and the like.
図２において、動圧溝は少なくとも公転摺動する摺動面側のリング側面に形成すればよいが、組み付け方向の依存性がなく、重量バランスにも優れることから、反摺動面側を含めた両側のリング側面に対称に形成することが好ましい。
In FIG. 2, the groove for dynamic pressure should be formed at least on the side surface of the ring on the side of the sliding surface that revolves and slides. It is preferable to form them symmetrically on both sides of the ring.
また、図２に示すように、動圧溝１８はリング周方向で等間隔に離間して複数個設けることが好ましい。隣り合う動圧溝同士の間のリング側面は主軸受部材に対して摺動する部分となり、摺動面の一部を構成する。動圧溝の面積（複数個の場合は合計の面積、以下同じ）は特に限定されないが、リング側面に対する動圧溝の面積が小さくなりすぎるとトルク低減効果が小さくなり、大きくなりすぎると過剰面圧となり摩耗が促進されるおそれがある。このような観点から、動圧溝の面積はリング側面の全体の面積に対して５％～７５％であることが好ましく、２０％～６０％であることがより好ましい。なお、リング側面の全体の面積とは、環状シール部材の公転摺動する摺動面側のリング側面を正面から見た平面視における動圧溝を含んだ摺動面積であり、動圧溝の面積は同平面視における面積である。
Also, as shown in FIG. 2, it is preferable to provide a plurality of dynamic pressure grooves 18 at equal intervals in the ring circumferential direction. A ring side surface between adjacent dynamic pressure grooves is a portion that slides on the main bearing member, and constitutes a part of the sliding surface. The area of the dynamic pressure grooves (the total area if there are more than one, the same shall apply hereinafter) is not particularly limited. There is a risk that it will become pressure and wear will be accelerated. From this point of view, the area of the hydrodynamic grooves is preferably 5% to 75%, more preferably 20% to 60%, of the entire area of the ring side surface. The total area of the ring side surface is the sliding area including the dynamic pressure groove in a plan view of the ring side surface on the side of the sliding surface on which the annular seal member revolves and slides. The area is the area in the same plan view.
動圧溝のそれぞれのリング周方向の長さは、個数に応じて、リング円周長さの約３％～２０％とすることが好ましい。動圧溝のリング径方向の長さは、摺動面の径方向厚みの１０％～８０％とすることが好ましい。また、摺動特性が安定することから、動圧溝は全て同サイズとし、略等間隔で離間して複数個（図２では片面１３個）設けることが好ましい。
The length of each hydrodynamic pressure groove in the ring circumferential direction is preferably about 3% to 20% of the ring circumferential length, depending on the number. The length of the hydrodynamic groove in the ring radial direction is preferably 10% to 80% of the radial thickness of the sliding surface. Further, since the sliding characteristics are stabilized, it is preferable that all of the dynamic pressure grooves are of the same size, and that a plurality of grooves (13 grooves on one side in FIG. 2) are provided at approximately equal intervals.
図３および図４を用いて、Ｖ字状の動圧溝について説明する。図３は、図２のＡ部の拡大図であり、図４（ａ）は動圧溝をリング内径側から見た図であり、図４（ｂ）はＢ部拡大図、図４（ｃ）はＣ部拡大図である。図３および図４に示すように、動圧溝１８は、リング周方向に沿ってリングの幅方向側（軸方向側）に凹んだＶ字状である。図４（ａ）に示すように、動圧溝１８の摺動面からの深さは、動圧溝１８のリング周方向の中央部に最深部１８ｄがあり、最深部１８ｄからリング周方向の両端部に向けて浅くなる。すなわち、リング周方向で摺動面に近い領域程浅くなる。また、動圧溝１８の摺動面からの深さは、リング径方向には一定である。なお、図３において最深部１８ｄは線状に形成されている。
The V-shaped dynamic pressure grooves will be described with reference to FIGS. 3 and 4. FIG. 3 is an enlarged view of part A in FIG. 2, FIG. 4(a) is a view of the dynamic pressure groove as viewed from the inner diameter side of the ring, FIG. 4(b) is an enlarged view of part B, and FIG. ) is an enlarged view of the C part. As shown in FIGS. 3 and 4, the hydrodynamic grooves 18 are V-shaped and recessed toward the width direction (axial side) of the ring along the circumferential direction of the ring. As shown in FIG. 4(a), the depth from the sliding surface of the groove for hydrodynamic bearing 18 has the deepest portion 18d at the center of the groove for hydrodynamic bearing 18 in the ring circumferential direction, and the depth from the deepest portion 18d to the ring circumferential direction. It becomes shallower towards both ends. That is, the depth becomes shallower in the region closer to the sliding surface in the circumferential direction of the ring. Further, the depth from the sliding surface of the hydrodynamic groove 18 is constant in the radial direction of the ring. In addition, in FIG. 3, the deepest part 18d is formed linearly.
図４（ａ）に示すように、動圧溝１８は、最深部１８ｄを中心に対称形状になっており、動圧溝１８の底面は、一対の第１の傾斜面１８ａ、１８ａと一対の第２の傾斜面１８ｂ、１８ｂとを有している。具体的には、摺動面から最深部１８ｄに至るまで同一平面ではなく、摺動面に接続される第１の傾斜面１８ａと、最深部１８ｄに接続され、摺動面に対して、第１の傾斜面１８ａよりも小さな傾斜角度をなす第２の傾斜面１８ｂとを有している。
As shown in FIG. 4A, the dynamic pressure groove 18 has a symmetrical shape centering on the deepest portion 18d. and second inclined surfaces 18b, 18b. Specifically, it is not the same plane from the sliding surface to the deepest portion 18d. and a second inclined surface 18b having a smaller inclination angle than the first inclined surface 18a.
図４（ｃ）に示すように、第１の傾斜面１８ａは、第２の傾斜面１８ｂに比べて、摺動面に対して急勾配に形成されている。これにより、摺動面が摩耗した場合でも動圧溝１８の開口面積の減少が小さく、トルクの変化が生じにくい。第１の傾斜面１８ａの摺動面に対する傾斜角度θ１は特に限定されないが、５０°～８０°が好ましく、５０°～７０°がより好ましい。傾斜角度θ１が５０°未満であると、摺動面が摩耗した場合、摺動面積の増加が大きくなり、トルク変動が懸念される。また、傾斜角度θ１が８０°を超えると、くさび作用が小さくなるおそれがある。
As shown in FIG. 4(c), the first inclined surface 18a is steeper than the second inclined surface 18b. As a result, even if the sliding surface is worn, the opening area of the hydrodynamic groove 18 is less reduced, and the torque is less likely to change. The inclination angle θ 1 of the first inclined surface 18a with respect to the sliding surface is not particularly limited, but is preferably 50° to 80°, more preferably 50° to 70°. If the inclination angle θ1 is less than 50°, the sliding surface area will increase greatly when the sliding surface is worn, and there is concern about torque fluctuations. Moreover, if the inclination angle θ1 exceeds 80°, the wedge action may become weak.
一方、第２の傾斜面１８ｂの摺動面に対する傾斜角度θ２（図４（ｂ）参照）は特に限定されないが、０．１°～１５°と鋭角であることが好ましく、１°～１０°であることがより好ましい。これにより、流入してきた流体によるくさび作用を効果的に発揮できる。一方、傾斜角度θ２が０．１°未満であると流入した流体が第１の傾斜面１８ａに流れにくくなり、また１５°を超えると動圧溝１８の最深部１８ｄが深くなり、該動圧溝１８の容積が増加し、圧力が分散することで、くさび作用が薄れるおそれがある。
On the other hand, the inclination angle θ 2 (see FIG. 4B) of the second inclined surface 18b with respect to the sliding surface is not particularly limited, but it is preferably an acute angle of 0.1° to 15°, and preferably 1° to 10°. ° is more preferred. As a result, the wedge effect of the inflowing fluid can be effectively exhibited. On the other hand, if the inclination angle θ2 is less than 0.1°, the inflowing fluid becomes difficult to flow to the first inclined surface 18a. By increasing the volume of the pressure groove 18 and distributing the pressure, the wedging effect may be weakened.
第１の傾斜面１８ａと第２の傾斜面１８ｂの境界部の構成は特に限定されない。例えば、第１の傾斜面１８ａと第２の傾斜面１８ｂを直接接続させてもよく、図４（ｃ）に示すように曲面（Ｒ面）１８ｃを介して接続させてもよい。Ｒ面１８ｃは、リング周方向に一定の幅を持つ部分であり、Ｒ面１８ｃの曲率半径は、例えば０．１～０．３である。図４（ｃ）に示すように、第１の傾斜面１８ａと第２の傾斜面１８ｂの境界部をＲ状に形成することで、隣り合う動圧溝同士の間の摺動面に流体が流出しやすくなり、更なる低トルク化を図りやすくなる。
The structure of the boundary between the first inclined surface 18a and the second inclined surface 18b is not particularly limited. For example, the first inclined surface 18a and the second inclined surface 18b may be directly connected, or may be connected via a curved surface (R surface) 18c as shown in FIG. 4(c). The R surface 18c is a portion having a constant width in the ring circumferential direction, and the radius of curvature of the R surface 18c is, for example, 0.1 to 0.3. As shown in FIG. 4(c), by forming the boundary between the first inclined surface 18a and the second inclined surface 18b in an R shape, the fluid flows on the sliding surface between the adjacent dynamic pressure grooves. It becomes easy to flow out, and it becomes easy to achieve further low torque.
また、第１の傾斜面１８ａの周方向の端部と摺動面との境界部は、曲面（Ｒ状）で接続することができる。当該境界部をＲ状に形成することで、冷媒等が摺動面により流出しやすくなり、更なる低トルク化を図りやすくなる。
In addition, the boundary between the end of the first inclined surface 18a in the circumferential direction and the sliding surface can be connected by a curved surface (R shape). By forming the boundary portion in an R shape, the refrigerant or the like can easily flow out from the sliding surface, and it becomes easier to further reduce the torque.
動圧溝１８の最深部１８ｄの摺動面からの深さは、リング総幅の４５％以下とすることが好ましく、３０％以下とすることが更に好ましい。なお、ここでの「深さ」は、動圧溝をリングの両側面に形成する場合には、各側面の凹部の深さを合計したものであり、この場合の片面の凹部の深さはリング総幅の２２．５％以下、好ましくは１５％以下である。リング総幅の４５％をこえる場合、環状シール部材が強度不足になり変形するおそれがある。
The depth from the sliding surface of the deepest portion 18d of the hydrodynamic groove 18 is preferably 45% or less of the total width of the ring, more preferably 30% or less. When the dynamic pressure grooves are formed on both sides of the ring, the "depth" here is the sum of the depths of the recesses on each side. In this case, the depth of the recess on one side is It is 22.5% or less of the total ring width, preferably 15% or less. If it exceeds 45% of the total width of the ring, the strength of the annular seal member may be insufficient and it may be deformed.
リング側面の内径側端部に形成される略Ｖ字状の動圧溝は、上記図３および図４の形態に限定されるものではない。例えば、図５（ａ）に示すように、動圧溝１９のリング周方向の端部に最深部１９ａを配置してもよい。可動スクロール体の旋回運動の方向（回転方向）は一方向であることから、非対称形状とすることができる。この場合、可動スクロール体の回転方向はＸ方向となる。また、動圧溝１９の底面として、上述したような第１の傾斜面および第２の傾斜面を適宜採用することができる。また、図５（ｂ）に示すように、動圧溝２０の最深部２０ａを摺動面に対して平行な平面で形成してもよい。また、最深部２０ａを曲線状にしてもよい。この場合も、動圧溝２０の底面として、上述したような第１の傾斜面および第２の傾斜面を適宜採用することができる。また、これらの例において、動圧溝の底面を構成する平面を適宜曲面で構成してもよい。
The substantially V-shaped dynamic pressure groove formed on the inner diameter side end of the ring side surface is not limited to the forms shown in FIGS. 3 and 4 above. For example, as shown in FIG. 5A, the deepest portion 19a may be arranged at the end portion of the dynamic pressure groove 19 in the ring circumferential direction. Since the orbiting direction (rotational direction) of the orbiting scroll body is unidirectional, an asymmetrical shape can be achieved. In this case, the direction of rotation of the movable scroll body is the X direction. Further, as the bottom surface of the hydrodynamic groove 19, the above-described first inclined surface and second inclined surface can be appropriately employed. Further, as shown in FIG. 5B, the deepest portion 20a of the hydrodynamic groove 20 may be formed in a plane parallel to the sliding surface. Also, the deepest portion 20a may be curved. Also in this case, the first inclined surface and the second inclined surface as described above can be appropriately employed as the bottom surface of the dynamic pressure generating groove 20 . Further, in these examples, the flat surface forming the bottom surface of the groove for hydrodynamic bearing may be appropriately curved.
図６に示すように、環状シール部材１６は、可動スクロール体の底板部４ａの背面に設けられた環状溝４ｄに装着される。図中左側が背圧室１５ａ側であり、図中右側が空間１５ｂ側である。図中の矢印が冷媒等からの圧力が加わる方向である。このシール構造により、背圧室１５ａと空間１５ｂとを仕切っている。そして、可動スクロール体の旋回運動に伴って、環状シール部材１６が連れ回りして、リング側面１７で主軸受部材１２の端面に公転摺動しながら摺動接触する。この際、連れ回りによって生じる冷媒等の流れによって、動圧溝１８に冷媒等が導入されることで動圧が発生する。この動圧によって、主軸受部材１２から離れる方向の力が環状シール部材１６の摺動面に作用するため、主軸受部材１２に対する環状シール部材１６の摺動抵抗が更に低減される。
As shown in FIG. 6, the annular seal member 16 is mounted in an annular groove 4d provided on the back surface of the bottom plate portion 4a of the movable scroll. The left side of the drawing is the back pressure chamber 15a side, and the right side of the drawing is the space 15b side. The arrows in the drawing indicate the directions in which the pressure from the refrigerant or the like is applied. This sealing structure partitions the back pressure chamber 15a and the space 15b. As the orbiting scroll body rotates, the annular seal member 16 rotates with the ring side surface 17 and comes into sliding contact with the end surface of the main bearing member 12 while revolving and sliding. At this time, a dynamic pressure is generated by introducing the refrigerant or the like into the dynamic pressure grooves 18 due to the flow of the refrigerant or the like caused by co-rotation. This dynamic pressure acts on the sliding surface of the annular seal member 16 in a direction away from the main bearing member 12 , thereby further reducing the sliding resistance of the annular seal member 16 with respect to the main bearing member 12 .
なお、環状溝は、可動スクロール体の底板部４ａ側ではなく、主軸受部材側に設けられてもよい。その場合、該環状溝に装着された環状シール部材はその環状溝内に固定される。その環状シール部材のリング側面は、旋回運動する可動スクロール体の底板部の背面に対して摺動接触する。このリング側面には、上述したような動圧溝が設けられる。なお、第２実施形態で述べる潤滑溝が設けられてもよい。
Note that the annular groove may be provided on the main bearing member side instead of the bottom plate portion 4a side of the orbiting scroll. In that case, the annular seal member mounted in the annular groove is fixed in the annular groove. The ring side surface of the annular seal member is in sliding contact with the back surface of the bottom plate portion of the orbiting orbiting scroll body. This ring side surface is provided with the hydrodynamic grooves as described above. In addition, the lubricating groove described in the second embodiment may be provided.
冷媒等は用途に応じた種類が適宜用いられる。また、冷媒等の温度は、例えば－２０℃～１４０℃程度である。可動スクロール体の旋回運動における回転数として５０００～８０００ｒｐｍ程度を主に想定している。
The type of refrigerant, etc., is appropriately used according to the application. Also, the temperature of the refrigerant or the like is, for example, about -20°C to 140°C. The rotational speed of the orbiting movement of the orbiting scroll body is mainly assumed to be about 5000 to 8000 rpm.
第１実施形態において、リング側面に設けられる動圧溝は、可動スクロール体の旋回運動によって生じる流体の流れにより該流体を導入して動圧を発生させる溝であればよく、種々の形状を採用できる。例えば、ヘリングボーン（図７（ａ）参照）、スパイラル（図７（ｂ）参照）、またはこれらを併用した形状などが挙げられる。図７は動圧溝の平面形状を示しており、図中の黒塗り部分が動圧溝である。なお、動圧溝は、環状シール部材のリング側面の内外径を連通していない溝（非連通溝）が望ましい。非連通溝では、途中で流体の流れが絞られるため、動圧が発生しやすい。
In the first embodiment, the dynamic pressure grooves provided on the side surface of the ring may be grooves that generate dynamic pressure by introducing the fluid caused by the orbiting motion of the orbiting scroll body, and adopting various shapes. can. For example, a herringbone (see FIG. 7(a)), a spiral (see FIG. 7(b)), or a shape using these together can be used. FIG. 7 shows the planar shape of the groove for dynamic pressure, and the black portion in the figure is the groove for dynamic pressure. It should be noted that the dynamic pressure groove is desirably a groove (non-communication groove) that does not communicate the inner and outer diameters of the ring side surface of the annular seal member. In non-communication grooves, dynamic pressure is likely to occur because the flow of fluid is throttled in the middle.
なお、図７（ａ）などのへリングボーンにおける溝の折り返し位置は、適宜設定できる。図７（ａ）に示す形状と可動スクロール体の回転方向では、折り返し位置が円周外側に行くほど、内径側から外径側に向かう力が大きくなる。
It should be noted that the folding position of the groove in the herringbone such as that shown in FIG. 7(a) can be set as appropriate. In the shape shown in FIG. 7A and the direction of rotation of the orbiting scroll body, the force directed from the inner diameter side to the outer diameter side increases as the folding position moves toward the outside of the circumference.
リング側面の内径側端部の少なくとも一部に非連通の動圧溝を形成する例としては、例えば上述した図２～図５の例が挙げられる。一方、リング側面の外径側端部の少なくとも一部に非連通の動圧溝を形成する例としては、例えば図８の例が挙げられる。図８に示す環状シール部材２１には、リング側面２２の外径側端部に、周方向に沿ってリングの幅方向側に凹んだ略Ｖ字状の動圧溝２３が複数設けられており、図８はその一部拡大図を示している。この動圧溝２３は、リング側面における形成位置を除いて、上述のＶ字状の動圧溝１８（図３参照）と同様の構成である。なお、動圧溝２３は、上述の動圧溝の変形例の構成を適宜採用することができる。
Examples of forming non-communicating dynamic pressure grooves in at least a part of the inner diameter side end of the ring side surface include the examples shown in FIGS. 2 to 5 described above. On the other hand, as an example of forming non-communicating hydrodynamic grooves in at least a part of the outer diameter side end of the ring side surface, for example, the example of FIG. 8 can be given. An annular seal member 21 shown in FIG. 8 is provided with a plurality of substantially V-shaped dynamic pressure grooves 23 recessed in the width direction of the ring along the circumferential direction at the outer diameter side end of the ring side surface 22 . , and FIG. 8 shows a partially enlarged view thereof. The hydrodynamic groove 23 has the same structure as the above-described V-shaped hydrodynamic groove 18 (see FIG. 3) except for the formation position on the ring side surface. It should be noted that the dynamic pressure groove 23 can appropriately adopt the configuration of the modified example of the dynamic pressure groove described above.
また、非連通の動圧溝の形成位置は、リング側面の内径側端部のみまたは外径側端部のみに限らず、リング側面の内径側端部および外径側端部の両方でもよい。この場合、例えばリング周方向に沿って、内径側端部の動圧溝と外径側端部の動圧溝とを交互に形成してもよい。また、内径側端部の動圧溝と外径側端部の動圧溝はリング径方向に重ならないように形成してもよい。
In addition, the formation position of the non-communicating dynamic pressure grooves is not limited to only the inner diameter side end portion or the outer diameter side end portion of the ring side surface, but may be both the inner diameter side end portion and the outer diameter side end portion of the ring side surface. In this case, for example, the dynamic pressure grooves on the inner diameter side end and the dynamic pressure grooves on the outer diameter side end may be alternately formed along the ring circumferential direction. Also, the dynamic pressure grooves on the inner diameter side end and the dynamic pressure grooves on the outer diameter side end may be formed so as not to overlap in the ring radial direction.
（第２実施形態）
本発明の環状シール部材の第２実施形態を図９に基づいて説明する。図９に示すように、環状シール部材３１は、断面略矩形の環状体であり、全周にわたって繋がった形状であり、合い口を有していない。図９において、環状シール部材３１には、リング側面３２に、リング外径またはリング内径のいずれか一方に開口した潤滑溝が複数設けられている。具体的には、リング側面３２の外径側端部に、リング外周面３１ａに開口した外径側潤滑溝３３が複数設けられているとともに、リング側面３２の内径側端部に、リング内周面３１ｂに開口した内径側潤滑溝３４が複数設けられている。なお、外径側潤滑溝３３および内径側潤滑溝３４は、リング外周面３１ａとリング内周面３１ｂを貫通していない。 (Second embodiment)
A second embodiment of the annular seal member of the present invention will be described with reference to FIG. As shown in FIG. 9, theannular seal member 31 is an annular body having a substantially rectangular cross section, has a shape that is continuous over the entire circumference, and does not have a joint. In FIG. 9, the ring side surface 32 of the annular seal member 31 is provided with a plurality of lubricating grooves that are open to either the ring outer diameter or the ring inner diameter. Specifically, a plurality of outer lubrication grooves 33 that open to the ring outer peripheral surface 31 a are provided at the outer diameter side end of the ring side surface 32 , and the ring inner peripheral lubrication groove 33 is provided at the inner diameter side end of the ring side surface 32 . A plurality of inner lubrication grooves 34 are provided that are open on the surface 31b. The outer lubrication groove 33 and the inner lubrication groove 34 do not penetrate the ring outer peripheral surface 31a and the ring inner peripheral surface 31b.
本発明の環状シール部材の第２実施形態を図９に基づいて説明する。図９に示すように、環状シール部材３１は、断面略矩形の環状体であり、全周にわたって繋がった形状であり、合い口を有していない。図９において、環状シール部材３１には、リング側面３２に、リング外径またはリング内径のいずれか一方に開口した潤滑溝が複数設けられている。具体的には、リング側面３２の外径側端部に、リング外周面３１ａに開口した外径側潤滑溝３３が複数設けられているとともに、リング側面３２の内径側端部に、リング内周面３１ｂに開口した内径側潤滑溝３４が複数設けられている。なお、外径側潤滑溝３３および内径側潤滑溝３４は、リング外周面３１ａとリング内周面３１ｂを貫通していない。 (Second embodiment)
A second embodiment of the annular seal member of the present invention will be described with reference to FIG. As shown in FIG. 9, the
図９において、リング外周面３１ａと両側のリング側面３２（外径側潤滑溝３３を含む）との角部や、リング内周面３１ｂと両側のリング側面３２（内径側潤滑溝３４を含む）との角部は、直線状、曲線状の面取りが設けられていてもよい。また、環状シール部材を射出成形で製造する場合、リング内周面３１ｂと両側のリング側面３２との角部に金型からの突出し部分となる段部３１ｃを設けてもよい。
In FIG. 9, the corners between the ring outer peripheral surface 31a and the ring side surfaces 32 on both sides (including the outer diameter side lubrication grooves 33), and the ring inner peripheral surface 31b and the ring side surfaces 32 on both sides (including the inner diameter side lubrication grooves 34). The corners of and may be chamfered in a straight line or a curved line. Further, when the annular seal member is manufactured by injection molding, a stepped portion 31c that protrudes from the mold may be provided at the corners of the ring inner peripheral surface 31b and the ring side surfaces 32 on both sides.
環状シール部材３１は、一方のリング側面が主軸受部材の可動スクロール体に向く端面と摺動する側の面となり、このリング側面３２に主軸受部材の端面との非接触部となる略すり鉢状の潤滑溝（内径側潤滑溝３４および外径側潤滑溝３３）が形成されている。これら潤滑溝と摺動面との境界部（角部）には、後述の図１１などに示すように傾斜面が設けられている。傾斜面を設けることで、冷媒等が主軸受部材の端面と摺動する部分に適度に流出することで、低トルク化が図れる。また、上記潤滑溝は動圧溝としても機能することができる。すなわち、冷媒等が当該潤滑溝に流入することでくさび作用が発生し、一層の低トルク化が図れ、低摩擦耐摩耗特性を向上させることができる。また、図９の構成では、潤滑溝はリング径方向に非連通の凹部であることから、冷媒等の低オイルリーク性にも繋がる。
One ring side surface of the annular seal member 31 slides against the end surface of the main bearing member facing the orbiting scroll body, and the ring side surface 32 is substantially mortar-shaped as a non-contact portion with the end surface of the main bearing member. lubrication grooves (inner diameter side lubrication grooves 34 and outer diameter side lubrication grooves 33) are formed. Boundaries (corners) between the lubricating grooves and the sliding surfaces are provided with inclined surfaces as shown in FIG. By providing the inclined surface, the coolant or the like flows appropriately to the portion that slides on the end surface of the main bearing member, so that the torque can be reduced. In addition, the lubricating grooves can also function as dynamic pressure grooves. That is, a wedge action is generated by the flow of the coolant or the like into the lubricating grooves, which further reduces the torque and improves the low friction and wear resistance characteristics. In addition, in the configuration of FIG. 9, since the lubrication groove is a concave portion that does not communicate in the ring radial direction, it also leads to low oil leakage of refrigerant or the like.
図９において、潤滑溝は少なくとも公転摺動する摺動面側のリング側面に形成すればよいが、組み付け方向の依存性がなく、重量バランスにも優れることから、反摺動面側を含めた両側のリング側面に対称に形成することが好ましい。
In FIG. 9, the lubrication grooves should be formed at least on the side of the ring on the side of the sliding surface that revolves and slides. It is preferably formed symmetrically on both sides of the ring.
また、図９に示すように、外径側潤滑溝３３および内径側潤滑溝３４はそれぞれ、リング周方向で等間隔に離間して複数個設けられている。この場合、リング側面３２から見て外径側潤滑溝３３と内径側潤滑溝３４とがリング周方向に沿って交互に設けられることが好ましい。隣り合う潤滑溝同士の間のリング側面は主軸受部材に対して摺動する部分となり、摺動面の一部を構成する。潤滑溝の面積（複数個の場合は合計の面積、以下同じ）は特に限定されないが、リング側面に対する潤滑溝の面積が小さくなりすぎるとトルク低減効果が小さくなり、大きくなりすぎると過剰面圧となり摩耗が促進されるおそれがある。このような観点から、潤滑溝の面積はリング側面の全体の面積に対して５％～７５％であることが好ましく、２０％～６０％であることがより好ましい。なお、リング側面の全体の面積とは、環状シール部材の公転摺動する摺動面側（片側）のリング側面を正面から見た平面視における面積（潤滑溝の面積も含む）であり、潤滑溝の面積は同平面視における面積である。
Further, as shown in FIG. 9, a plurality of outer diameter side lubrication grooves 33 and inner diameter side lubrication grooves 34 are provided at equal intervals in the ring circumferential direction. In this case, it is preferable that the outer lubricating grooves 33 and the inner lubricating grooves 34 are alternately provided along the ring circumferential direction when viewed from the ring side surface 32 . A ring side surface between adjacent lubricating grooves is a portion that slides on the main bearing member, and constitutes a part of the sliding surface. The area of the lubricating groove (the total area if there are multiple lubricating grooves, hereinafter the same) is not particularly limited, but if the area of the lubricating groove with respect to the ring side surface is too small, the torque reduction effect will be small, and if it is too large, excessive surface pressure will occur. Wear may be accelerated. From this point of view, the area of the lubricating groove is preferably 5% to 75%, more preferably 20% to 60%, of the entire area of the ring side surface. The total area of the ring side surface is the area (including the area of the lubrication groove) in a plan view of the ring side surface (one side) on which the annular seal member revolves and slides from the front. The area of the groove is the area in the same plan view.
潤滑溝のそれぞれのリング周方向の長さは、個数に応じて、リング円周長さの約０．５％～５％とすることが好ましい。潤滑溝のリング径方向の長さは、摺動面の径方向厚みの１０％～８０％とすることが好ましい。また、摺動特性が安定することから、潤滑溝は全て同サイズとし、略等間隔で離間して複数個（図９では内径側潤滑溝１２個、外径側潤滑溝１２個で、潤滑溝として計２４個）設けることが好ましい。なお、潤滑溝の数は、内径側と外径側で同じ数でなくてもよく、例えば、内径側潤滑溝を外径側潤滑溝よりも多くしてもよい。
The length of each lubricating groove in the ring circumferential direction is preferably about 0.5% to 5% of the ring circumferential length, depending on the number of lubrication grooves. The length of the lubricating groove in the radial direction of the ring is preferably 10% to 80% of the radial thickness of the sliding surface. In addition, since the sliding characteristics are stabilized, all the lubrication grooves are of the same size and are spaced apart at approximately equal intervals (in FIG. 9, 12 lubrication grooves on the inner diameter side and 12 lubrication grooves on the outer 24 in total) is preferably provided. The number of lubrication grooves on the inner diameter side and the outer diameter side may not be the same. For example, the inner diameter side lubrication grooves may be larger than the outer diameter side lubrication grooves.
図１０および図１１を用いて、略すり鉢状の潤滑溝について説明する。図１０は、図９のＤ部の拡大図である。図１０に示すように、外径側潤滑溝３３および内径側潤滑溝３４の平面形状は、円弧状とされている。これら潤滑溝３３、３４は、リング周方向に沿ってリングの幅方向側（軸方向側）に凹んだ略円錐状の凹部である。具体的には、外径側潤滑溝３３は、リング外径の中心点と同芯でリング外径よりも大径の円周上に中心軸を有する略円錐状の凹部である。外径側潤滑溝３３のリング周方向の略中央部の外径端部に位置する最深部から、リング周方向の両側に、かつ、リング内径側に向けて放射状に溝深さが浅くなっている。また、内径側潤滑溝３４は、リング内径の中心点と同芯でリング内径よりも小径の円周上に中心軸を有する略円錐状の凹部である。内径側潤滑溝３４のリング周方向の略中央部の内径端部に位置する最深部から、リング周方向の両側に、かつ、リング外径側に向けて放射状に溝深さが浅くなっている。
Using Figs. 10 and 11, the substantially mortar-shaped lubricating grooves will be described. FIG. 10 is an enlarged view of part D in FIG. As shown in FIG. 10, the planar shapes of the outer lubricating groove 33 and the inner lubricating groove 34 are arcuate. These lubricating grooves 33 and 34 are substantially conical recesses recessed in the width direction side (axial direction side) of the ring along the ring circumferential direction. Specifically, the outer diameter side lubrication groove 33 is a substantially conical recess having a center axis on a circumference concentric with the center point of the ring outer diameter and having a larger diameter than the ring outer diameter. From the deepest part located at the outer diameter end of the outer diameter side lubrication groove 33 in the ring circumferential direction, the groove depth becomes shallow radially on both sides in the ring circumferential direction and toward the ring inner diameter side. there is The inner lubrication groove 34 is a substantially conical recess having a central axis on a circumference concentric with the center point of the inner diameter of the ring and having a smaller diameter than the inner diameter of the ring. The depth of the lubricating groove 34 on the inner diameter side radially becomes shallower from the deepest part located at the inner diameter end part of the substantially central part in the ring circumferential direction to both sides in the ring circumferential direction and toward the ring outer diameter side. .
図１１では、内径側潤滑溝について更に説明する。なお、以下では潤滑溝として、内径側潤滑溝を用いて説明するが、外径側潤滑溝についても同様の形状などを採用できる。図１１（ａ）は潤滑溝をリング内径側から見た図であり、図１１（ｂ）はそのＥ－Ｅ線に沿って切断した切断面を表している。
In Fig. 11, the lubricating groove on the inner diameter side will be further explained. In the following description, the inner diameter side lubrication groove is used as the lubrication groove, but the outer diameter side lubrication groove may have the same shape. FIG. 11(a) is a view of the lubricating groove viewed from the inner diameter side of the ring, and FIG. 11(b) is a cross-sectional view taken along the line EE.
図１１（ａ）において、内径側潤滑溝３４は、略円錐状の中心軸をリング内径側の摺動面内径端に有する場合であり、図１０のように略円錐状の中心軸がリング内径よりも小径の円周上には無い凹部である。略円錐状の中心軸をリング内径側の摺動面内径端に有する場合、内径側潤滑溝３４の底面は、当該円錐の外周面に対応した面で形成され、中心軸に向かって直線状の傾斜面３４ａで形成されている。内径側潤滑溝３４の摺動面からの深さは、内径側潤滑溝３４のリング周方向の中央部に最深部３４ｂがあり、最深部３４ｂから径方向の放射状に向けて浅くなる。すなわち、リング径方向で摺動面に近い領域程浅くなっている（図１１（ｂ）参照）。
In FIG. 11(a), the inner lubrication groove 34 has a substantially conical central axis at the inner diameter end of the sliding surface on the inner diameter side of the ring, and as shown in FIG. It is a concave portion that is not on the circumference of a smaller diameter than . When a substantially conical central axis is provided at the inner diameter end of the sliding surface on the inner diameter side of the ring, the bottom surface of the inner lubricating groove 34 is formed with a surface corresponding to the outer peripheral surface of the cone, and extends linearly toward the central axis. It is formed by an inclined surface 34a. The depth from the sliding surface of the inner lubrication groove 34 has the deepest portion 34b at the central portion of the inner lubrication groove 34 in the ring circumferential direction, and becomes shallower radially from the deepest portion 34b. That is, the area closer to the sliding surface in the ring radial direction is shallower (see FIG. 11(b)).
図１１（ａ）に示すように、内径側潤滑溝３４は、最深部３４ｂを中心に対称形状になっている。内径側潤滑溝３４の内径縁は直線状に形成されている。また、図１１（ｂ）の軸方向断面において、内径側潤滑溝３４の底面（傾斜面３４ａ）は直線状に形成されている。なお、図１０の内径側潤滑溝３４をリング内径側から見た場合、内径縁は直線状に見えることはなく、図１４（ｂ）に近い形状に見える。
As shown in FIG. 11(a), the inner lubricating groove 34 has a symmetrical shape centering on the deepest portion 34b. An inner diameter edge of the inner diameter side lubrication groove 34 is formed in a straight line. Further, in the axial cross section of FIG. 11(b), the bottom surface (inclined surface 34a) of the inner lubricating groove 34 is formed linearly. When the inner diameter side lubrication groove 34 in FIG. 10 is viewed from the ring inner diameter side, the inner diameter edge does not appear to be straight, but appears to have a shape similar to that shown in FIG. 14(b).
傾斜面３４ａの摺動面に対する傾斜角度γ（図１１（ａ）、図１１（ｂ）参照）は特に限定されないが、０．１°～１５°の範囲であることが好ましく、１°～１０°の範囲であることがより好ましい。これにより、主軸受部材の端面と摺動する部分に適度に冷媒等が流出しやすくなり、また、流入してきた流体によるくさび作用を効果的に発揮しやすくなる。傾斜角度γが０．１°未満であると流入した流体が摺動面に向かって流れにくくなり、また傾斜角度γが１５°を超えると内径側潤滑溝３４の最深部３４ｂが深くなり、該内径側潤滑溝３４の容積が増加し、圧力が分散することで、くさび作用が薄れるおそれがある。傾斜角度γは、内径側潤滑溝３４と摺動面との境界部の全域（図１０における潤滑溝の円弧全周）において、同じ角度でもよく、異なる角度でもよい。
The inclination angle γ (see FIGS. 11(a) and 11(b)) of the inclined surface 34a with respect to the sliding surface is not particularly limited, but is preferably in the range of 0.1° to 15°, more preferably 1° to 10°. ° range is more preferred. As a result, the coolant or the like tends to flow moderately into the portion that slides on the end surface of the main bearing member, and the wedging effect of the inflowing fluid can be effectively exhibited. If the inclination angle γ is less than 0.1°, the inflowing fluid becomes difficult to flow toward the sliding surface. As the volume of the inner lubricating groove 34 increases and the pressure disperses, the wedge action may be weakened. The angle of inclination γ may be the same or different over the entire boundary portion between the inner diameter side lubrication groove 34 and the sliding surface (the entire circumference of the arc of the lubrication groove in FIG. 10).
なお、潤滑溝において、潤滑溝と摺動面との境界部は傾斜面で構成されていればよく、図１１に示すように傾斜面だけで潤滑溝が構成されていてもよく、図１２に示すように傾斜面と、摺動面と平行な底面で構成されていてもよい。
また、傾斜面は、図１１に示すように直線状で摺動面に接続してもよく、図１４（ｂ）に示すように曲線状（Ｒ状）で接続してもよい。例えば、図１１（ｂ）に示す内径側潤滑溝３４の傾斜面３４ａは、リング側面に対して直角の断面から見た形状が直線であるため、流体が安定して流入することで動圧効果も安定し、更なる低トルク化を図ることができる。一方で、Ｒ状に形成することで、冷媒等が摺動面により流出しやすくなり、更なる低トルク化を図りやすくなる。 In addition, in the lubrication groove, the boundary portion between the lubrication groove and the sliding surface may be constituted by an inclined surface, and as shown in FIG. As shown, it may be composed of an inclined surface and a bottom surface parallel to the sliding surface.
The inclined surface may be connected to the sliding surface in a straight line as shown in FIG. 11, or may be connected in a curved line (R shape) as shown in FIG. 14(b). For example, since the inclined surface 34a of theinner lubricating groove 34 shown in FIG. is also stabilized, and a further reduction in torque can be achieved. On the other hand, the rounded shape makes it easier for the refrigerant or the like to flow out from the sliding surface, making it easier to further reduce the torque.
また、傾斜面は、図１１に示すように直線状で摺動面に接続してもよく、図１４（ｂ）に示すように曲線状（Ｒ状）で接続してもよい。例えば、図１１（ｂ）に示す内径側潤滑溝３４の傾斜面３４ａは、リング側面に対して直角の断面から見た形状が直線であるため、流体が安定して流入することで動圧効果も安定し、更なる低トルク化を図ることができる。一方で、Ｒ状に形成することで、冷媒等が摺動面により流出しやすくなり、更なる低トルク化を図りやすくなる。 In addition, in the lubrication groove, the boundary portion between the lubrication groove and the sliding surface may be constituted by an inclined surface, and as shown in FIG. As shown, it may be composed of an inclined surface and a bottom surface parallel to the sliding surface.
The inclined surface may be connected to the sliding surface in a straight line as shown in FIG. 11, or may be connected in a curved line (R shape) as shown in FIG. 14(b). For example, since the inclined surface 34a of the
内径側潤滑溝３４の最深部３４ｂの摺動面からの深さは、リング総幅の４５％以下とすることが好ましく、３０％以下とすることが更に好ましい。なお、ここでの「深さ」は、潤滑溝をリングの両側面に形成する場合には、各側面の凹部の深さを合計したものであり、この場合の片面の凹部の深さはリング総幅の２２．５％以下、好ましくは１５％以下である。リング総幅の４５％をこえる場合、環状シール部材が強度不足になり変形するおそれがある。
The depth from the sliding surface of the deepest portion 34b of the inner lubricating groove 34 is preferably 45% or less, more preferably 30% or less, of the total width of the ring. It should be noted that the "depth" here is the sum of the depths of the recesses on each side when lubricating grooves are formed on both sides of the ring. It is 22.5% or less of the total width, preferably 15% or less. If it exceeds 45% of the total width of the ring, the strength of the annular seal member may be insufficient and it may be deformed.
略すり鉢状の潤滑溝は、図１０および図１１の形態に限定されるものではない。例えば、図１２に示すように、内径側潤滑溝３５を略円錐台状の凹部として、内径側潤滑溝３５の最深部３５ｂを摺動面に対して平行な平面で形成してもよい。この場合も、最深部３５ｂから、リング周方向の両側に、かつ、リング外径側に向けて放射状に溝深さが浅くなる。またこの場合、最深部３５ｂを曲面で形成してもよい。
The substantially mortar-shaped lubrication grooves are not limited to the forms shown in FIGS. 10 and 11. For example, as shown in FIG. 12, the inner diameter lubrication groove 35 may be formed as a substantially truncated conical recess, and the deepest portion 35b of the inner diameter lubrication groove 35 may be formed in a plane parallel to the sliding surface. Also in this case, the depth of the groove becomes shallow radially from the deepest portion 35b toward both sides in the ring circumferential direction and toward the ring outer diameter side. Further, in this case, the deepest portion 35b may be formed with a curved surface.
なお、上述した第２実施形態において、潤滑溝の底面を構成する平面を適宜曲面で構成してもよい。
In addition, in the above-described second embodiment, the flat surface forming the bottom surface of the lubricating groove may be appropriately curved.
図１３に示すように、環状シール部材３１は、可動スクロール体の底板部４ａの背面に設けられた環状溝４ｄに装着される。このシール構造により、背圧室１５ａと空間１５ｂとを仕切っている。そして、可動スクロール体の旋回運動に伴って、環状シール部材３１が連れ回りして、リング側面３２で主軸受部材１２の端面に公転摺動しながら摺動接触する。この際、連れ回りによって生じる冷媒等の流れによって、内径側潤滑溝３４に冷媒等が導入されることで動圧が発生する。この動圧によって、主軸受部材１２から離れる方向の力が環状シール部材３１の摺動面に作用するため、主軸受部材１２に対する環状シール部材３１の摺動抵抗が更に低減される。
As shown in FIG. 13, the annular seal member 31 is mounted in an annular groove 4d provided on the back surface of the bottom plate portion 4a of the movable scroll. This sealing structure partitions the back pressure chamber 15a and the space 15b. As the orbiting scroll member rotates, the annular seal member 31 rotates with the ring side surface 32 and comes into sliding contact with the end surface of the main bearing member 12 while revolving and sliding. At this time, a dynamic pressure is generated by introducing the coolant or the like into the inner diameter side lubrication groove 34 due to the flow of the coolant or the like caused by the co-rotation. This dynamic pressure acts on the sliding surface of the annular seal member 31 in a direction away from the main bearing member 12 , thereby further reducing the sliding resistance of the annular seal member 31 with respect to the main bearing member 12 .
第２実施形態において、リング側面に設けられる潤滑溝は、可動スクロール体の旋回運動によって生じる流体の流れにより主軸受部材の端面と摺動する部分に適度に流体を流出させる溝であればよく、種々の形状を採用できる。例えば、図１４（ａ）～（ｃ）に示すように、略すり鉢状の内径側潤滑溝３６の底面が傾斜曲面３６ａで形成されてもよい（略球欠状）。また、図１５に示すように、内径側潤滑溝３７の平面形状が、略三角形状で形成されてもよい。なお、潤滑溝は、環状シール部材のリング側面の内外径を連通していない溝（非連通溝）が望ましい。非連通溝では、途中で流体の流れが絞られるため、動圧が発生しやすい。
In the second embodiment, the lubricating groove provided on the side surface of the ring may be a groove that allows the fluid to flow appropriately to the portion that slides on the end surface of the main bearing member due to the flow of fluid generated by the orbiting motion of the orbiting scroll body. Various shapes can be employed. For example, as shown in FIGS. 14(a) to 14(c), the bottom surface of the substantially mortar-shaped inner lubrication groove 36 may be formed with an inclined curved surface 36a (substantially spherical segment shape). Further, as shown in FIG. 15, the planar shape of the inner diameter side lubrication groove 37 may be formed in a substantially triangular shape. The lubricating groove is preferably a groove (non-communicating groove) that does not communicate with the inner and outer diameters of the ring side surface of the annular seal member. In non-communication grooves, dynamic pressure is likely to occur because the flow of fluid is throttled in the middle.
リング側面の内径側端部および外径側端部の少なくとも一部に非連通の潤滑溝を形成する例としては、例えば上述した図の例が挙げられる。なお、上述した潤滑溝を、摺動面となるリング側面の内径側端部のみに形成してもよく、外径側端部のみに形成してもよい。
Examples of forming non-communicating lubricating grooves in at least a part of the inner diameter side end and the outer diameter side end of the ring side surface include the examples shown in the above figures. The lubricating grooves described above may be formed only on the inner diameter side end of the ring side surface serving as the sliding surface, or may be formed only on the outer diameter side end.
本発明の環状シール部材の材質は特に限定されないが、合成樹脂の成形体とすることが好ましい。使用できる合成樹脂としては、例えば、熱硬化性ポリイミド樹脂、熱可塑性ポリイミド樹脂、ポリエーテルケトンエーテルケトンケトン樹脂、ポリエーテルケトン樹脂、ＰＥＥＫ樹脂、全芳香族ポリエステル樹脂、ポリテトラフルオロエチレン（以下、ＰＴＦＥと記す）樹脂等のフッ素樹脂、ＰＰＳ樹脂、ポリアミドイミド樹脂、ポリアミド樹脂などが挙げられる。なお、これらの樹脂は単独で使用しても、２種類以上混合したポリマーアロイとしてもよい。
Although the material of the annular seal member of the present invention is not particularly limited, it is preferably made of synthetic resin. Synthetic resins that can be used include, for example, thermosetting polyimide resin, thermoplastic polyimide resin, polyether ketone ether ketone ketone resin, polyether ketone resin, PEEK resin, wholly aromatic polyester resin, polytetrafluoroethylene (hereinafter referred to as PTFE Fluororesins such as resins, PPS resins, polyamideimide resins, and polyamide resins. These resins may be used singly or as a polymer alloy in which two or more kinds are mixed.
また、環状シール部材は、合成樹脂を射出成形してなる射出成形体にすることが好ましい。このため、合成樹脂としては、射出成形が可能である熱可塑性樹脂を用いることが好ましい。その中でも特に、摩擦摩耗特性、曲げ弾性率、耐熱性、摺動性などに優れることから、ＰＥＥＫ樹脂またはＰＰＳ樹脂を用いることが好ましい。これらの樹脂は高い弾性率を有し、シールする冷媒等の温度が高くなる場合でも使用でき、また、ソルベントクラックの心配もない。
Also, the annular seal member is preferably an injection molded body made by injection molding a synthetic resin. Therefore, it is preferable to use a thermoplastic resin that can be injection molded as the synthetic resin. Among them, it is preferable to use PEEK resin or PPS resin because they are particularly excellent in friction and abrasion properties, flexural modulus, heat resistance, slidability, and the like. These resins have a high modulus of elasticity, can be used even when the temperature of the sealing coolant or the like is high, and are free from solvent cracks.
また、必要に応じて上記合成樹脂に、炭素繊維、ガラス繊維、アラミド繊維などの繊維状補強材、球状シリカや球状炭素などの球状充填材、マイカやタルクなどの鱗状補強材、チタン酸カリウムウィスカなどの微小繊維補強材を配合できる。また、ＰＴＦＥ樹脂、グラファイト、二硫化モリブデンなどの固体潤滑剤、リン酸カルシウム、硫酸カルシウムなどの摺動補強材、カーボンブラック、酸化チタンなどの顔料も配合できる。これらは単独で配合することも、組み合せて配合することもできる。特に、ＰＥＥＫ樹脂またはＰＰＳ樹脂に、繊維状補強材である炭素繊維と、固体潤滑剤であるＰＴＦＥ樹脂とを含むものが、本発明の環状シール部材に要求される特性を得やすいので好ましい。炭素繊維を配合することで、曲げ弾性率等の機械的強度の向上が図れ、ＰＴＦＥ樹脂の配合により摺動特性の向上が図れる。
In addition, if necessary, the above synthetic resin may be added with fibrous reinforcing materials such as carbon fiber, glass fiber, and aramid fiber, spherical fillers such as spherical silica and spherical carbon, scaly reinforcing materials such as mica and talc, and potassium titanate whiskers. Microfiber reinforcing materials such as can be blended. Further, solid lubricants such as PTFE resin, graphite and molybdenum disulfide, sliding reinforcing materials such as calcium phosphate and calcium sulfate, and pigments such as carbon black and titanium oxide can also be blended. These may be blended singly or in combination. In particular, PEEK resin or PPS resin containing carbon fiber as a fibrous reinforcing material and PTFE resin as a solid lubricant is preferable because it facilitates obtaining the properties required for the annular seal member of the present invention. By blending carbon fiber, mechanical strength such as flexural modulus can be improved, and by blending PTFE resin, sliding property can be improved.
合成樹脂製とする場合には、以上の諸原材料を溶融混練して成形用ペレットとし、これを用いて公知の射出成形法等により所定形状に成形する。射出成形により製造する場合、そのゲート位置は特に限定されないが、シール性の確保の観点および後加工が不要になることからリング内周面に設けることが好ましい。さらに、ゲート位置は、周方向に等間隔に配置した多点ゲート（例えば３点～６点）として、ゲート位置と動圧溝の位置とがリング径方向において重ならないことがより好ましい。この場合、環状シール部材は、リング径方向において、動圧溝と重ならない位置の内周面にゲート痕を有することが好ましい。
When using a synthetic resin, the above raw materials are melted and kneaded to form pellets for molding, which are then molded into a predetermined shape by a known injection molding method or the like. When the gate is manufactured by injection molding, the position of the gate is not particularly limited, but it is preferably provided on the inner peripheral surface of the ring from the viewpoint of ensuring sealing performance and because post-processing is not required. Further, it is more preferable that the gate positions are multi-point gates (for example, 3 to 6 points) arranged at equal intervals in the circumferential direction, and that the gate positions and the positions of the dynamic pressure generating grooves do not overlap in the ring radial direction. In this case, it is preferable that the annular seal member has a gate mark on the inner peripheral surface at a position that does not overlap with the dynamic pressure groove in the ring radial direction.
摺動面積の違いによる、動摩擦係数の面圧依存性を確認する目的として、摺動面積を固定して、荷重を３水準に分けてスラスト試験を行った。
For the purpose of confirming the surface pressure dependence of the dynamic friction coefficient due to the difference in the sliding area, the sliding area was fixed and the load was divided into three levels and a thrust test was conducted.
実施例Ａおよび比較例Ａ
ＰＰＳ樹脂を主材料とし、ＰＴＦＥ樹脂および炭素繊維を配合した樹脂組成物（ＮＴＮ社製：ベアリーＡＳ５３０２）を用い、実施例Ａおよび比較例Ａの環状の試験片を射出成形により製造した。
比較例Ａの試験片は、外径寸法φ２１ｍｍ、内径寸法φ１７ｍｍ、径方向長さ２ｍｍ、軸方向長さ１．６ｍｍであり、リング側面に動圧溝が設けられていない。一方、実施例Ａの試験片は、外径寸法φ２１ｍｍ、内径寸法φ１７ｍｍ、径方向長さ２ｍｍ、軸方向長さ１．６ｍｍであり、リング側面の内径側端部に、図３に示すようなリング周方向に沿った略Ｖ字状の動圧溝が４個設けられている。動圧溝の最深部の溝深さは、０．１ｍｍであり、第１の傾斜面の摺動面に対する傾斜角度は約６５°であり、第２の傾斜面の摺動面に対する傾斜角度は約３°であった。なお、動圧溝の面積はリング側面の全体の面積に対して４０％であった。 Example A and Comparative Example A
Circular test pieces of Example A and Comparative Example A were produced by injection molding using a resin composition (BEAREE AS5302 manufactured by NTN) containing PPS resin as the main material and PTFE resin and carbon fiber.
The test piece of Comparative Example A had an outer diameter of φ21 mm, an inner diameter of φ17 mm, a radial length of 2 mm, and an axial length of 1.6 mm, and no dynamic pressure groove was provided on the ring side surface. On the other hand, the test piece of Example A had an outer diameter of φ21 mm, an inner diameter of φ17 mm, a radial length of 2 mm, and an axial length of 1.6 mm. Four substantially V-shaped dynamic pressure grooves are provided along the ring circumferential direction. The deepest groove depth of the dynamic pressure groove is 0.1 mm, the inclination angle of the first inclined surface with respect to the sliding surface is approximately 65°, and the inclination angle of the second inclined surface with respect to the sliding surface is was about 3°. The area of the hydrodynamic grooves was 40% of the total area of the side surface of the ring.
ＰＰＳ樹脂を主材料とし、ＰＴＦＥ樹脂および炭素繊維を配合した樹脂組成物（ＮＴＮ社製：ベアリーＡＳ５３０２）を用い、実施例Ａおよび比較例Ａの環状の試験片を射出成形により製造した。
比較例Ａの試験片は、外径寸法φ２１ｍｍ、内径寸法φ１７ｍｍ、径方向長さ２ｍｍ、軸方向長さ１．６ｍｍであり、リング側面に動圧溝が設けられていない。一方、実施例Ａの試験片は、外径寸法φ２１ｍｍ、内径寸法φ１７ｍｍ、径方向長さ２ｍｍ、軸方向長さ１．６ｍｍであり、リング側面の内径側端部に、図３に示すようなリング周方向に沿った略Ｖ字状の動圧溝が４個設けられている。動圧溝の最深部の溝深さは、０．１ｍｍであり、第１の傾斜面の摺動面に対する傾斜角度は約６５°であり、第２の傾斜面の摺動面に対する傾斜角度は約３°であった。なお、動圧溝の面積はリング側面の全体の面積に対して４０％であった。 Example A and Comparative Example A
Circular test pieces of Example A and Comparative Example A were produced by injection molding using a resin composition (BEAREE AS5302 manufactured by NTN) containing PPS resin as the main material and PTFE resin and carbon fiber.
The test piece of Comparative Example A had an outer diameter of φ21 mm, an inner diameter of φ17 mm, a radial length of 2 mm, and an axial length of 1.6 mm, and no dynamic pressure groove was provided on the ring side surface. On the other hand, the test piece of Example A had an outer diameter of φ21 mm, an inner diameter of φ17 mm, a radial length of 2 mm, and an axial length of 1.6 mm. Four substantially V-shaped dynamic pressure grooves are provided along the ring circumferential direction. The deepest groove depth of the dynamic pressure groove is 0.1 mm, the inclination angle of the first inclined surface with respect to the sliding surface is approximately 65°, and the inclination angle of the second inclined surface with respect to the sliding surface is was about 3°. The area of the hydrodynamic grooves was 40% of the total area of the side surface of the ring.
スラスト試験機の概略図を図１６に示す。負荷軸４１の先端に試験片４３を取り付け、回転軸４５に取り付けられた相手材４４（ＡＤＣ１２、外径寸法φ３３ｍｍ、厚さ１０ｍｍ、試験片との摺動面は平面研磨によりＲａ０．８μｍ程度とした）に、所定の荷重Ｆで押し付け、オイル４２中で下記の条件にてスラスト試験を行った。各試験において、試験終了直前の動摩擦係数を測定した。面圧と動摩擦係数の関係を図１７に示す。
A schematic diagram of the thrust tester is shown in Fig. 16. A test piece 43 is attached to the tip of a load shaft 41, and a mating member 44 (ADC 12, outer diameter φ33 mm, thickness 10 mm, sliding surface with the test piece is polished to a Ra of about 0.8 μm by plane polishing. ) was pressed with a predetermined load F, and a thrust test was performed in the oil 42 under the following conditions. In each test, the dynamic friction coefficient was measured just before the end of the test. FIG. 17 shows the relationship between surface pressure and dynamic friction coefficient.
＜試験条件＞
速度 ：２ｍ／ｓｅｃ
面圧 ：１ＭＰａ、２ＭＰａ、３ＭＰａ
雰囲気温度：室温
潤滑 ：油中（ＰＡＧ油、出光ダフニーハーメチックオイルＰＳ）
試験時間 ：各面圧３０ｍｉｎ
試験数 ：ｎ＝１ <Test conditions>
Speed: 2m/sec
Surface pressure: 1 MPa, 2 MPa, 3 MPa
Ambient temperature: Room temperature Lubrication: In oil (PAG oil, Idemitsu Daphne hermetic oil PS)
Test time: each surface pressure 30min
Number of tests: n = 1
速度 ：２ｍ／ｓｅｃ
面圧 ：１ＭＰａ、２ＭＰａ、３ＭＰａ
雰囲気温度：室温
潤滑 ：油中（ＰＡＧ油、出光ダフニーハーメチックオイルＰＳ）
試験時間 ：各面圧３０ｍｉｎ
試験数 ：ｎ＝１ <Test conditions>
Speed: 2m/sec
Surface pressure: 1 MPa, 2 MPa, 3 MPa
Ambient temperature: Room temperature Lubrication: In oil (PAG oil, Idemitsu Daphne hermetic oil PS)
Test time: each surface pressure 30min
Number of tests: n = 1
図１７に示すように、動摩擦係数は、面圧（荷重）の増加とともに低下する傾向であることから、面圧依存性があり、摺動面の面積を減らすことで、動摩擦係数（トルク）が減少した。これにより、実施例Ａのように動圧溝を形成することで、低トルク化を図ることができる。
As shown in FIG. 17, the coefficient of dynamic friction tends to decrease as the surface pressure (load) increases. Diminished. Accordingly, by forming the hydrodynamic grooves as in the embodiment A, torque reduction can be achieved.
実施例Ｂおよび比較例Ｂ
ＰＰＳ樹脂を主材料とし、ＰＴＦＥ樹脂および炭素繊維を配合した樹脂組成物（ＮＴＮ社製：ベアリーＡＳ５３０２）を用い、実施例Ｂおよび比較例Ｂの環状の試験片を射出成形により製造した。
比較例Ｂの試験片は、外径寸法φ２１ｍｍ、内径寸法φ１７ｍｍ、径方向長さ２ｍｍ、軸方向長さ１．６ｍｍであり、リング側面に潤滑溝が設けられていない。一方、実施例Ｂの試験片は、外径寸法φ２１ｍｍ、内径寸法φ１７ｍｍ、径方向長さ２ｍｍ、軸方向長さ１．６ｍｍであり、リング側面の内径側端部、外径側端部に、図１０に示すようなリング径方向に対して放射状の潤滑溝が各４個設けられている。潤滑溝の最深部の溝深さは、０．１ｍｍであり、傾斜面の摺動面に対する傾斜角度は約３°であった。なお、潤滑溝の面積はリング側面の全体の面積に対して４０％であった。 Example B and Comparative Example B
Circular test pieces of Example B and Comparative Example B were produced by injection molding using a resin composition (manufactured by NTN: BEAREE AS5302) containing PPS resin as the main material and PTFE resin and carbon fiber.
The test piece of Comparative Example B had an outer diameter of φ21 mm, an inner diameter of φ17 mm, a radial length of 2 mm, and an axial length of 1.6 mm, and no lubricating groove was provided on the ring side surface. On the other hand, the test piece of Example B had an outer diameter of φ21 mm, an inner diameter of φ17 mm, a radial length of 2 mm, and an axial length of 1.6 mm. As shown in FIG. 10, four radial lubricating grooves are provided in the radial direction of the ring. The deepest groove depth of the lubricating groove was 0.1 mm, and the inclination angle of the inclined surface with respect to the sliding surface was about 3°. The area of the lubricating groove was 40% of the total area of the ring side surface.
ＰＰＳ樹脂を主材料とし、ＰＴＦＥ樹脂および炭素繊維を配合した樹脂組成物（ＮＴＮ社製：ベアリーＡＳ５３０２）を用い、実施例Ｂおよび比較例Ｂの環状の試験片を射出成形により製造した。
比較例Ｂの試験片は、外径寸法φ２１ｍｍ、内径寸法φ１７ｍｍ、径方向長さ２ｍｍ、軸方向長さ１．６ｍｍであり、リング側面に潤滑溝が設けられていない。一方、実施例Ｂの試験片は、外径寸法φ２１ｍｍ、内径寸法φ１７ｍｍ、径方向長さ２ｍｍ、軸方向長さ１．６ｍｍであり、リング側面の内径側端部、外径側端部に、図１０に示すようなリング径方向に対して放射状の潤滑溝が各４個設けられている。潤滑溝の最深部の溝深さは、０．１ｍｍであり、傾斜面の摺動面に対する傾斜角度は約３°であった。なお、潤滑溝の面積はリング側面の全体の面積に対して４０％であった。 Example B and Comparative Example B
Circular test pieces of Example B and Comparative Example B were produced by injection molding using a resin composition (manufactured by NTN: BEAREE AS5302) containing PPS resin as the main material and PTFE resin and carbon fiber.
The test piece of Comparative Example B had an outer diameter of φ21 mm, an inner diameter of φ17 mm, a radial length of 2 mm, and an axial length of 1.6 mm, and no lubricating groove was provided on the ring side surface. On the other hand, the test piece of Example B had an outer diameter of φ21 mm, an inner diameter of φ17 mm, a radial length of 2 mm, and an axial length of 1.6 mm. As shown in FIG. 10, four radial lubricating grooves are provided in the radial direction of the ring. The deepest groove depth of the lubricating groove was 0.1 mm, and the inclination angle of the inclined surface with respect to the sliding surface was about 3°. The area of the lubricating groove was 40% of the total area of the ring side surface.
これらの試験片について、実施例Ａなどと同様の条件で、スラスト試験を実施した。その結果、図１７に示したように、動摩擦係数が、面圧（荷重）の増加とともに低下する傾向がみられ、面圧依存性があり、摺動面の面積を減らすことで、動摩擦係数（トルク）が減少するという結果が得られた。これにより、実施例Ｂのように潤滑溝を形成することで、低トルク化を図ることができる。
A thrust test was performed on these test pieces under the same conditions as in Example A. As a result, as shown in FIG. 17, the coefficient of dynamic friction tends to decrease as the surface pressure (load) increases. The result was that the torque) decreased. Accordingly, by forming the lubricating grooves as in the embodiment B, the torque can be reduced.
本発明の環状シール部材は、耐久性やシール機能の低下を損なうことなく、安定した低トルク性を発揮できるので、スクロールコンプレッサの環状シール部材として広く利用できる。また、スラスト受け部材を除くことが可能となる。
The annular seal member of the present invention can exhibit stable low-torque performance without impairing the durability and sealing function, so it can be widely used as an annular seal member for scroll compressors. Also, it becomes possible to eliminate the thrust receiving member.
１ コンプレッサ
２ ハウジング
３ 固定スクロール体
３ａ 底板部
３ｂ 渦巻壁
３ｃ 開口部
４ 可動スクロール体
４ａ 底板部
４ｂ 渦巻壁
５ 圧縮室
６ａ ステータ
６ｂ ロータ
７ シャフト
８ バランスウェイト
９ 主軸受
１０ 副軸受
１１ 旋回軸受
１２ 主軸受部材
１３ シャフトシール
１４ モータ室
１５ａ 背圧室
１５ｂ 空間
１６ 環状シール部材
１７ リング側面
１８ 動圧溝
１８ａ 第１の傾斜面
１８ｂ 第２の傾斜面
１８ｃ Ｒ面
１８ｄ 最深部
１９ 動圧溝
１９ａ 最深部
２０ 動圧溝
２０ａ 最深部
２１ 環状シール部材
２２ リング側面
２３ 動圧溝
３１ 環状シール部材
３２ リング側面
３３ 外径側潤滑溝（潤滑溝）
３４ 内径側潤滑溝（潤滑溝）
３４ａ 傾斜面
３４ｂ 最深部
３５ 内径側潤滑溝（潤滑溝）
３５ａ 傾斜面
３５ｂ 最深部
３６ 内径側潤滑溝（潤滑溝）
３６ａ 傾斜面
３７ 内径側潤滑溝（潤滑溝）
４１ 負荷軸
４２ オイル
４３ 試験片
４４ 相手材
４５ 回転軸 1Compressor 2 Housing 3 Fixed Scroll 3a Bottom Plate 3b Spiral Wall 3c Opening 4 Movable Scroll 4a Bottom Plate 4b Spiral Wall 5 Compression Chamber 6a Stator 6b Rotor 7 Shaft 8 Balance Weight 9 Main Bearing 10 Sub-Bearing 11 Orbiting Bearing 12 Main Bearing member 13 Shaft seal 14 Motor chamber 15a Back pressure chamber 15b Space 16 Annular seal member 17 Ring side face 18 Dynamic pressure groove 18a First inclined surface 18b Second inclined surface 18c R surface 18d Deepest part 19 Dynamic pressure groove 19a Deepest part 20 dynamic pressure groove 20a deepest portion 21 annular seal member 22 ring side surface 23 dynamic pressure groove 31 annular seal member 32 ring side surface 33 outer diameter side lubrication groove (lubrication groove)
34 Inner diameter side lubrication groove (lubrication groove)
34aInclined surface 34b Deepest part 35 Inner diameter side lubrication groove (lubrication groove)
35aInclined surface 35b Deepest part 36 Inner diameter side lubrication groove (lubrication groove)
36aInclined surface 37 Inner diameter side lubrication groove (lubrication groove)
41Load shaft 42 Oil 43 Test piece 44 Mating material 45 Rotating shaft
２ ハウジング
３ 固定スクロール体
３ａ 底板部
３ｂ 渦巻壁
３ｃ 開口部
４ 可動スクロール体
４ａ 底板部
４ｂ 渦巻壁
５ 圧縮室
６ａ ステータ
６ｂ ロータ
７ シャフト
８ バランスウェイト
９ 主軸受
１０ 副軸受
１１ 旋回軸受
１２ 主軸受部材
１３ シャフトシール
１４ モータ室
１５ａ 背圧室
１５ｂ 空間
１６ 環状シール部材
１７ リング側面
１８ 動圧溝
１８ａ 第１の傾斜面
１８ｂ 第２の傾斜面
１８ｃ Ｒ面
１８ｄ 最深部
１９ 動圧溝
１９ａ 最深部
２０ 動圧溝
２０ａ 最深部
２１ 環状シール部材
２２ リング側面
２３ 動圧溝
３１ 環状シール部材
３２ リング側面
３３ 外径側潤滑溝（潤滑溝）
３４ 内径側潤滑溝（潤滑溝）
３４ａ 傾斜面
３４ｂ 最深部
３５ 内径側潤滑溝（潤滑溝）
３５ａ 傾斜面
３５ｂ 最深部
３６ 内径側潤滑溝（潤滑溝）
３６ａ 傾斜面
３７ 内径側潤滑溝（潤滑溝）
４１ 負荷軸
４２ オイル
４３ 試験片
４４ 相手材
４５ 回転軸 1
34 Inner diameter side lubrication groove (lubrication groove)
34a
35a
36a
41
Claims (14)
- 底板部とその表面に立設する渦巻壁を有する固定スクロール体と、底板部とその表面に立設する渦巻壁を有する可動スクロール体と、シャフトと、該シャフトを回転可能に支持する主軸受と、該主軸受を固定する主軸受部材とを備え、
前記シャフトの回転により、前記可動スクロール体を前記固定スクロール体の軸線の周りで公転させて流体を圧縮室にて圧縮するとともに、前記流体が前記可動スクロール体の背面側の背圧室に供給されるスクロールコンプレッサにおいて、
前記可動スクロール体の前記底板部の背面と、前記主軸受部材の前記可動スクロール体に向く端面のいずれか一方の面に形成された少なくとも１個の環状溝に装着され、前記背圧室をシールする環状シール部材であって、
環状シール部材は、リング側面において少なくとも公転摺動する摺動面に動圧溝が設けられていることを特徴とする環状シール部材。 A fixed scroll body having a bottom plate portion and a spiral wall erected on its surface, a movable scroll body having a bottom plate portion and a spiral wall erected on its surface, a shaft, and a main bearing rotatably supporting the shaft , and a main bearing member for fixing the main bearing,
The rotation of the shaft causes the orbiting scroll to revolve around the axis of the fixed scroll, compressing the fluid in the compression chamber, and supplying the fluid to the back pressure chamber on the back side of the orbiting scroll. In a scroll compressor with
mounted in at least one annular groove formed in either the back surface of the bottom plate portion of the orbiting scroll body or the end surface of the main bearing member facing the orbiting scroll body to seal the back pressure chamber; An annular seal member for
An annular seal member, wherein dynamic pressure grooves are provided on at least a sliding surface on which a ring side surface revolves and slides. - 前記動圧溝の面積は前記リング側面の全体の面積に対して５％～７５％であることを特徴とする請求項１記載の環状シール部材。 The annular seal member according to claim 1, wherein the area of the dynamic pressure groove is 5% to 75% of the total area of the ring side surface.
- 前記動圧溝の形状は、リング周方向に沿ってリングの幅方向側に凹んだ略Ｖ字状であり、前記動圧溝の前記摺動面からの深さは、最深部からリング周方向の両端部に向けて浅くなり、リング径方向には一定であることを特徴とする請求項１記載の環状シール部材。 The shape of the dynamic pressure groove is a substantially V shape recessed in the width direction of the ring along the ring circumferential direction, and the depth of the dynamic pressure groove from the sliding surface is 2. The annular seal member according to claim 1, wherein the depth becomes shallower toward both ends of the ring and is constant in the radial direction of the ring.
- 前記動圧溝は、前記摺動面から前記最深部に至るまで同一平面ではなく、前記摺動面に接続される第１の傾斜面と、前記最深部に接続され、前記摺動面に対して、前記第１の傾斜面よりも小さな傾斜角度をなす第２の傾斜面とを有することを特徴とする請求項３記載の環状シール部材。 The dynamic pressure groove is not flush from the sliding surface to the deepest portion, but has a first inclined surface connected to the sliding surface and a first inclined surface connected to the deepest portion, and 4. The annular sealing member according to claim 3, further comprising a second inclined surface having an inclination angle smaller than that of said first inclined surface.
- 前記摺動面に対する前記第１の傾斜面の傾斜角度が５０°～８０°であり、前記摺動面に対する前記第２の傾斜面の傾斜角度が０．１°～１５°であることを特徴とする請求項４記載の環状シール部材。 The inclination angle of the first inclined surface with respect to the sliding surface is 50° to 80°, and the inclination angle of the second inclined surface with respect to the sliding surface is 0.1° to 15°. 5. The annular seal member according to claim 4, wherein
- 前記動圧溝において、前記第１の傾斜面と前記第２の傾斜面の境界部が曲面で接続されていることを特徴とする請求項４記載の環状シール部材。 5. The annular seal member according to claim 4, wherein in said dynamic pressure groove, a boundary portion between said first inclined surface and said second inclined surface is connected by a curved surface.
- 前記動圧溝がリング周方向で離間して複数個設けられ、隣り合う動圧溝同士の間のリング側面が前記摺動面の一部を構成することを特徴とする請求項１記載の環状シール部材。 2. The annular ring according to claim 1, wherein a plurality of said dynamic pressure grooves are provided spaced apart in the ring circumferential direction, and the ring side surface between adjacent dynamic pressure grooves constitutes a part of said sliding surface. sealing material.
- 前記環状シール部材は合成樹脂製であり、該合成樹脂がポリフェニレンサルファイド樹脂またはポリエーテルエーテルケトン樹脂であることを特徴とする請求項１記載の環状シール部材。 The annular sealing member according to claim 1, wherein the annular sealing member is made of synthetic resin, and the synthetic resin is polyphenylene sulfide resin or polyether ether ketone resin.
- 底板部とその表面に立設する渦巻壁を有する固定スクロール体と、底板部とその表面に立設する渦巻壁を有する可動スクロール体と、シャフトと、該シャフトを回転可能に支持する主軸受と、該主軸受を固定する主軸受部材とを備え、
前記シャフトの回転により、前記可動スクロール体を前記固定スクロール体の軸線の周りで公転させて流体を圧縮室にて圧縮するとともに、前記流体が前記可動スクロール体の背面側の背圧室に供給されるスクロールコンプレッサにおいて、
前記可動スクロール体の前記底板部の背面と、前記主軸受部材の前記可動スクロール体に向く端面のいずれか一方の面に形成された少なくとも１個の環状溝に装着され、前記背圧室をシールする環状シール部材であって、
前記環状シール部材は、リング側面において少なくとも公転摺動する摺動面に、リング外径およびリング内径のいずれか一方に開口した潤滑溝が複数設けられており、前記潤滑溝と前記摺動面との境界部には傾斜面が設けられていることを特徴とする環状シール部材。 A fixed scroll body having a bottom plate portion and a spiral wall erected on its surface, a movable scroll body having a bottom plate portion and a spiral wall erected on its surface, a shaft, and a main bearing rotatably supporting the shaft , and a main bearing member for fixing the main bearing,
The rotation of the shaft causes the orbiting scroll to revolve around the axis of the fixed scroll, compressing the fluid in the compression chamber, and supplying the fluid to the back pressure chamber on the back side of the orbiting scroll. In a scroll compressor with
mounted in at least one annular groove formed in either the back surface of the bottom plate portion of the orbiting scroll body or the end surface of the main bearing member facing the orbiting scroll body to seal the back pressure chamber; An annular seal member for
The annular seal member has a plurality of lubricating grooves that are open to either the ring outer diameter or the ring inner diameter on at least the sliding surface that revolves and slides on the ring side surface, and the lubricating groove and the sliding surface are provided. An annular seal member characterized in that a slanted surface is provided at the boundary of the . - 前記潤滑溝の面積は前記リング側面の全体の面積に対して５％～７５％であることを特徴とする請求項９記載の環状シール部材。 The annular seal member according to claim 9, wherein the area of said lubrication groove is 5% to 75% of the total area of said ring side surface.
- 前記潤滑溝は略すり鉢状であることを特徴とする請求項９記載の環状シール部材。 The annular seal member according to claim 9, wherein the lubricating groove is substantially mortar-shaped.
- 前記環状シール部材は、前記潤滑溝として、リング外径に開口した外径側潤滑溝とリング内径に開口した内径側潤滑溝とを有することを特徴とする請求項９記載の環状シール部材。 The annular seal member according to claim 9, wherein the annular seal member has, as the lubrication grooves, an outer diameter side lubrication groove that opens to the ring outer diameter and an inner diameter side lubrication groove that opens to the ring inner diameter.
- 前記外径側潤滑溝および前記内径側潤滑溝はそれぞれ、リング周方向で離間して複数個設けられ、前記リング側面から見て前記外径側潤滑溝と前記内径側潤滑溝とがリング周方向に交互に設けられていることを特徴とする請求項１２記載の環状シール部材。 A plurality of the outer-diameter-side lubrication grooves and the inner-diameter-side lubrication grooves are provided separately in the ring circumferential direction. 13. The annular sealing member according to claim 12, wherein the annular sealing member is alternately provided with the .
- 前記傾斜面は前記摺動面に対する角度が０．１°～１５°であることを特徴とする請求項９記載の環状シール部材。 The annular sealing member according to claim 9, wherein the inclined surface has an angle of 0.1° to 15° with respect to the sliding surface.
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| JP2021-115971 | 2021-07-13 | ||
| JP2021115971A JP2023012367A (en) | 2021-07-13 | 2021-07-13 | Annular seal member of scroll compressor |
| JP2021-160193 | 2021-09-29 | ||
| JP2021160193 | 2021-09-29 |
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| WO2023286559A1 true WO2023286559A1 (en) | 2023-01-19 |
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Cited By (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| WO2024075740A1 (en) * | 2022-10-04 | 2024-04-11 | Nok株式会社 | Seal ring |
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| JPH11336676A (en) * | 1998-03-25 | 1999-12-07 | Tokico Ltd | Scroll type fluid machine |
| JP2008215090A (en) * | 2007-02-28 | 2008-09-18 | Denso Corp | Scroll compressor and its manufacturing method |
| WO2021125201A1 (en) * | 2019-12-17 | 2021-06-24 | イーグル工業株式会社 | Sliding component |
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2022
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| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPH11336676A (en) * | 1998-03-25 | 1999-12-07 | Tokico Ltd | Scroll type fluid machine |
| JP2008215090A (en) * | 2007-02-28 | 2008-09-18 | Denso Corp | Scroll compressor and its manufacturing method |
| WO2021125201A1 (en) * | 2019-12-17 | 2021-06-24 | イーグル工業株式会社 | Sliding component |
Cited By (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| WO2024075740A1 (en) * | 2022-10-04 | 2024-04-11 | Nok株式会社 | Seal ring |
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