WO2023286559A1 - Élément d'étanchéité annulaire pour compresseur à spirale - Google Patents

Élément d'étanchéité annulaire pour compresseur à spirale Download PDF

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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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WIPO (PCT)
Prior art keywords
ring
groove
seal member
annular seal
dynamic pressure
Prior art date
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PCT/JP2022/025119
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English (en)
Japanese (ja)
Inventor
洋志 柳川
Original Assignee
Ntn株式会社
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Filing date
Publication date
Priority claimed from JP2021115971A external-priority patent/JP2023012367A/ja
Application filed by Ntn株式会社 filed Critical Ntn株式会社
Publication of WO2023286559A1 publication Critical patent/WO2023286559A1/fr

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    • 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.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Rotary Pumps (AREA)

Abstract

L'invention concerne un élément d'étanchéité annulaire pour un compresseur qui peut présenter une propriété de faible couple fiable sans nuire à la durabilité ni provoquer une détérioration d'une fonction d'étanchéité. Dans un compresseur à spirale comprenant un corps fixe (3), un corps mobile (4), un arbre (7), un palier principal (9) qui supporte en rotation l'arbre (7), et un élément (12) de palier principal qui fixe le palier principal (9), conjointement avec un fluide comprimé dans une chambre de compression (5) par la rotation de l'arbre (7) amenant le corps de spirale mobile (4) à tourner autour de l'axe du corps de spirale fixe (3), le fluide est acheminé vers une chambre de contre-pression (15a) sur le côté de surface arrière du corps de spirale mobile (4), cet élément d'étanchéité annulaire (16) est monté sur au moins une rainure annulaire (4d) formée dans la surface arrière d'une partie de plaque inférieure (4a) de la spirale mobile (4), et scelle hermétiquement la chambre de contre-pression (15a). Dans une surface latérale annulaire de l'élément d'étanchéité annulaire (16), une rainure de pression dynamique est ménagée dans au moins une surface coulissante qui tourne et coulisse.
PCT/JP2022/025119 2021-07-13 2022-06-23 Élément d'étanchéité annulaire pour compresseur à spirale WO2023286559A1 (fr)

Applications Claiming Priority (4)

Application Number Priority Date Filing Date Title
JP2021-115971 2021-07-13
JP2021115971A JP2023012367A (ja) 2021-07-13 2021-07-13 スクロールコンプレッサの環状シール部材
JP2021-160193 2021-09-29
JP2021160193 2021-09-29

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Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2024075740A1 (fr) * 2022-10-04 2024-04-11 Nok株式会社 Bague d'étanchéité

Citations (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPH11336676A (ja) * 1998-03-25 1999-12-07 Tokico Ltd スクロール式流体機械
JP2008215090A (ja) * 2007-02-28 2008-09-18 Denso Corp スクロール型圧縮機およびその製造方法
WO2021125201A1 (fr) * 2019-12-17 2021-06-24 イーグル工業株式会社 Composant coulissant

Patent Citations (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPH11336676A (ja) * 1998-03-25 1999-12-07 Tokico Ltd スクロール式流体機械
JP2008215090A (ja) * 2007-02-28 2008-09-18 Denso Corp スクロール型圧縮機およびその製造方法
WO2021125201A1 (fr) * 2019-12-17 2021-06-24 イーグル工業株式会社 Composant coulissant

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2024075740A1 (fr) * 2022-10-04 2024-04-11 Nok株式会社 Bague d'étanchéité

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