Annular seal member for scroll compressor

Abstract

To provide an annular seal member of a compressor which can exert stable low-torque performance without causing the deterioration of durability and the lowering of a seal function.SOLUTION: In a scroll compressor, an annular seal member 16 comprises a fixed scroll body 3, a movable scroll body 4, a shaft 7, a main bearing 9 for rotatably supporting the shaft 7, and a main bearing member 12 for fixing the main bearing 9 thereto. The scroll compressor compresses fluid in a compression chamber 5 by revolving the movable scroll body 4 around an axial line of the fixed scroll body 3 by the rotation of the shaft 7, and the fluid is supplied to a backpressure chamber 15 of the movable scroll body 4 at a back face side. A seal member is attached to at least one annular groove 4d which is formed at a back face of a bottom plate part 4a of the movable scroll body 4, and seals the backpressure chamber 15a, and a dynamic pressure groove is formed at a slide face which at least revolutionarily slides at a ring side face.SELECTED DRAWING: Figure 1

Description

[Technical Field]

[0001] The present invention relates to an annular seal member that is attached to a bottom plate portion or the like of a movable scroll that constitutes a scroll compressor. [Background technology]

[0002] A scroll compressor has a scroll-type compression mechanism consisting of a fixed scroll and a movable scroll that orbits relative to the fixed scroll. The fixed scroll and movable scroll each have a bottom plate and a spiral wall extending from the surface of the bottom plate. They are meshed with each other at the spiral walls to form a compression chamber between them. This compression chamber moves toward the center of the spiral due to the action of the movable scroll revolving around the axis of the fixed scroll, thereby compressing a refrigerant or the like.

[0003] An annular seal 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 due to a compression reaction force. This thrust load increases the friction between the annular seal provided on the back side of the movable scroll and the main bearing member that slides against it, which can cause wear of the annular seal.

[0004] As a countermeasure against such frictional wear of annular seal members, a method of reducing frictional wear by using a lubricant such as oil is known (see Patent Document 1).

[0005] Another known method is to connect the discharge pressure region and the back pressure chamber through a pressure introducing hole in order to reduce the thrust load acting unilaterally from the movable scroll to the main bearing member. Furthermore, in this method, it is also known to provide a radially communicating groove on the side surface of the annular seal member to connect the back pressure chamber and the suction pressure region (see Patent Document 2). However, Patent Document 2 is a technology that makes it easy to set the intended back pressure in the back pressure chamber when the atmospheres in the back pressure chamber and the suction pressure region are not uniform.

[0006] Furthermore, as another method, a means for reducing the load is also known in which a thrust receiving member is interposed as a member separate from the above-mentioned annular sealing member to receive the thrust force from the bottom plate side of the movable scroll body to the main bearing member side (see Patent Document 3). [Prior art documents] [Patent documents]

[0007] [Patent Document 1] Japanese Patent Application Publication No. 8-121366 [Patent Document 2] Japanese Patent Application Laid-Open No. 2007-211702 [Patent Document 3] Japanese Patent Application Laid-Open No. 2012-17656 Summary of the Invention [Problem to be solved by the invention]

[0008] While the use of lubricants, for example, can reduce frictional wear on the annular seal, this method requires that the sliding surfaces be kept in a constant state of good lubrication. Therefore, if the lubricant runs out locally, torque becomes unstable, raising concerns about maintaining stable performance of the compressor itself. On the other hand, if a thrust bearing member is installed separately from the annular seal, the number of parts increases, which may lead to increased costs for the entire unit.

[0009] The present invention has been made in consideration of the above circumstances, and aims to provide an annular seal member for a compressor that can exhibit stable low torque performance without compromising durability or sealing function. [Means for solving the problem]

[0010] The annular seal member of the present invention is provided in a scroll compressor comprising: a fixed scroll body having a bottom plate portion and a spiral wall erected on its surface; a movable scroll body having the bottom plate portion and a spiral wall erected on its surface; a shaft; a main bearing rotatably supporting the shaft; and a main bearing member for fixing the main bearing, wherein rotation of the shaft causes the movable scroll body to revolve around the axis of the fixed scroll body to compress a fluid in a compression chamber and supply the fluid to a back pressure chamber on the back surface side of the movable scroll body.The annular seal member of the present invention is attached to at least one annular groove formed on either the back surface of the bottom plate portion of the movable scroll body or the end face of the main bearing member facing the movable scroll body, and seals the back pressure chamber, and is characterized in that the annular seal member has dynamic pressure grooves formed on at least the sliding surface that revolves and slides on the ring side.

[0011] In the present invention, the dynamic pressure grooves are grooves that generate dynamic pressure by introducing a fluid due to a fluid flow generated by the orbital motion of the movable scroll, and the cross-sectional area of ​​the grooves decreases in the direction of fluid flow so as to throttle the fluid. The direction of decrease may be the depth direction of the groove (FIG. 3), the width direction of the groove (FIG. 7), or both.

[0012] The area of ​​the dynamic pressure grooves is characterized by being 5% to 75% of the total area of ​​the ring side surface.

[0013] The shape of the dynamic pressure groove is a generally V-shape recessed toward the width direction of the ring along the circumferential direction of the ring, and the depth of the dynamic pressure groove from the sliding surface decreases from the deepest part toward both ends in the circumferential direction of the ring and is constant in the radial direction of the ring.

[0014] The dynamic pressure groove is not flush with the sliding surface from the sliding surface to the deepest part, but has a first inclined surface connected to the sliding surface and a second inclined surface connected to the deepest part and forming a smaller inclination angle with respect to the sliding surface than the first inclined surface.

[0015] The inclination angle of the first inclined surface relative to the sliding surface is 50° to 80°, and the inclination angle of the second inclined surface relative to the sliding surface is 0.1° to 15°.

[0016] In the dynamic pressure groove, the boundary between the first inclined surface and the second inclined surface is connected by a curved surface.

[0017] The hydrodynamic grooves are provided in a plurality at intervals in the circumferential direction of the ring, and the ring side surface between adjacent hydrodynamic grooves constitutes part of the sliding surface.

[0018] 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. [Effects of the Invention]

[0019] The annular seal of the present invention is a seal member that is attached to an annular groove formed on either the back surface of the bottom plate of the movable scroll or the end face of the main bearing member facing the movable scroll in a scroll compressor, sealing the back pressure chamber. The ring side has dynamic pressure grooves on at least the sliding surface that revolves and slides, thereby reducing the sliding area. This reduces the sliding torque due to the dependency of the friction coefficient on surface pressure. Furthermore, the dynamic pressure grooves allow fluid to flow into the dynamic pressure grooves, making it easier for a wedge action to occur, further reducing torque. This improves friction and wear characteristics, allowing for stable low torque performance without the need for a thrust bearing member, without compromising durability or sealing function.

[0020] The area of ​​the hydrodynamic grooves is 5% to 75% of the total area of ​​the ring side surface, so that the torque reduction effect is ensured while the promotion of wear is suppressed.

[0021] The shape of the hydrodynamic groove is roughly V-shaped, recessed toward the width direction of the ring along the circumferential direction of the ring, and the depth of the hydrodynamic groove from the sliding surface becomes shallower from the deepest part to both ends in the circumferential direction of the ring and is constant in the radial direction of the ring, making it easier to introduce fluid into the hydrodynamic groove and to generate a wedge action.

[0022] The dynamic pressure groove is not flush with the sliding surface from the sliding surface to the deepest part, but has a first inclined surface connected to the sliding surface and a second inclined surface connected to the deepest part and forming a smaller inclination angle with respect to the sliding surface than the first inclined surface. The first inclined surface is formed at a steeper inclination with respect to the sliding surface than the second inclined surface. Therefore, even if the sliding surface wears, the decrease in the opening area of ​​the dynamic pressure groove is small (i.e., the increase in the sliding area is small), making it difficult for a change in torque to occur. In particular, since the inclination angle of the first inclined surface with respect to the sliding surface is 50° to 80°, it is possible to effectively generate a wedge action while suppressing an increase in the sliding area. Furthermore, since the inclination angle of the second inclined surface with respect to the sliding surface is 0.1° to 15°, it is possible to effectively generate a wedge action by the inflowing fluid.

[0023] In the dynamic pressure groove, 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, which makes it easier for fluid to flow out to the sliding surface between adjacent dynamic pressure grooves, for example, thereby achieving further torque reduction. [Brief explanation of the drawings]

[0024] [Figure 1] 1 is a partial cross-sectional view showing an example of a scroll compressor equipped with an annular seal member of the present invention. [Figure 2] 1 is a perspective view showing an example of an annular seal member of the present invention. [Figure 3] FIG. 3 is an enlarged view of part A in FIG. 2. [Figure 4] 10 is a view of the dynamic pressure groove of the annular seal member as viewed from the inner diameter side of the ring. FIG. [Figure 5] 10A and 10B are diagrams showing another example of a substantially V-shaped dynamic pressure groove in the annular seal member. [Figure 6] 3 is a cross-sectional view showing a state in which the annular seal member of FIG. 2 is fitted into an annular groove. [Figure 7] 10A and 10B are diagrams showing other examples of dynamic pressure grooves in the annular seal member. [Figure 8] 10A and 10B are diagrams showing other examples of dynamic pressure grooves in the annular seal member. [Figure 9] FIG. 1 is a schematic diagram of a thrust test. [Figure 10] FIG. 10 is a diagram showing test results of a thrust test. DETAILED DESCRIPTION OF THE INVENTION

[0025] An example of a scroll compressor equipped with the annular seal member of the present invention will be described with reference to Fig. 1. Fig. 1 is a partial cross-sectional view of the scroll compressor. This scroll compressor compresses a fluid such as a refrigerant such as carbon dioxide gas, a refrigeration oil such as polyalkylene glycol oil (PAG oil), or a mixture of these (hereinafter collectively referred to as a refrigerant, etc.).

[0026] In FIG. 1, compressor 1 has a compression mechanism and a motor mechanism inside housing 2 and is connected to the outside via an intake port (not shown) and a discharge port (not shown). The compression mechanism compresses refrigerant or the like drawn in through the intake port and discharges it from a discharge port, and is composed of a fixed scroll body 3 and a movable scroll body 4. Fixed scroll body 3 has a bottom plate portion 3a and a spiral wall 3b extending vertically from bottom plate portion 3a, with an opening 3c at its center. Moving scroll body 4 has a bottom plate portion 4a and a spiral wall 4b extending vertically from bottom plate portion 4a. Fixed scroll body 3 and movable scroll body 4 are eccentrically intermeshed, and a compression chamber 5 is formed between the spiral walls 3b, 4b of each scroll body.

[0027] Although not shown, a spiral seal member (tip seal) is attached to the axial end face of the spiral walls 3b and 4b of each scroll body, thereby preventing leakage of refrigerant, etc., from the compression chamber.

[0028] The motor mechanism provides the orbiting driving force to the movable scroll 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 rotor 6b form an electric motor, and when current is applied to the stator 6a, the rotor 6b and shaft 7 rotate together. The shaft 7 is rotatably supported via a main bearing 9 and an auxiliary bearing 10. An eccentric shaft 7a is formed integrally with one end of the shaft 7, and a balance weight 8 is supported on this. The shaft 7 and balance weight 8 form a rotating member.

[0029] A boss 4c is provided at approximately the center of the back surface side of the bottom plate 4a of the movable scroll 4 so as to protrude vertically, and an orbiting bearing 11 is press-fitted into this boss 4c. The orbiting bearing 11 supports the eccentric shaft 7a, and the movable scroll 4 is configured to orbit by the orbiting bearing 11.

[0030] The main bearing 9 is fixed to a bearing support portion formed on the center side of the main bearing member 12. The main bearing member 12 is fixed inside the housing, and the fixed scroll body 3 is connected to the main bearing member 12 with bolts or the like. A shaft seal 13 is installed on the side of the main bearing 9, between the outer circumferential surface of the shaft 7 and the main bearing member 12. This shaft seal 13 blocks communication between the motor chamber 14 and the back pressure chamber 15a.

[0031] An annular seal member 16 is provided between the main bearing member 12 and the back surface of the bottom plate 4a of the movable scroll 4. In FIG. 1, the annular seal member 16 is fitted in an annular groove 4d formed in the back surface of the bottom plate 4a of the movable scroll 4. In this structure, the annular seal member 16 revolves and slides against the end face of the main bearing member 12 facing the movable scroll. The back pressure chamber 15a is sealed by the annular seal member 16 and the shaft seal 13, and a sealed space is formed between these seals, the main bearing member 12, and the bottom plate 4a of the movable scroll 4.

[0032] When the compressor 1 starts operating, the rotation of the rotor 6b causes the movable scroll 4 to begin orbiting. Refrigerant, etc., entering the compression mechanism through the suction port is compressed as it moves from the outer periphery to the center of the orbiting scroll wall and is then discharged to the outside through the opening 3c of the fixed scroll 3. Meanwhile, pressurized fluid from within the compression mechanism is supplied to the back pressure chamber 15a through a pressure introduction hole (not shown) provided in the bottom plate 4a of the movable scroll 4. By introducing pressurized fluid into the back pressure chamber 15a, the pressure within the back pressure chamber acts on the movable scroll 4 so as to reduce the thrust load (a force that presses the movable scroll 4 toward the main bearing member) acting on the movable scroll 4 due to the compression reaction force, or to press the movable scroll 4 toward the fixed scroll 4.

[0033] The annular seal member 16 separates the inner back pressure chamber 15a from the outer space 15b. While the space 15b has a pressure value close to the suction pressure, the back pressure chamber 15a is higher in pressure than the space 15b because compressed refrigerant or the like is introduced into the back pressure chamber 15a. As a result, one ring side of the annular seal member 16 revolves and makes sliding contact with the end surface of the main bearing member 12. The annular seal member 16 is primarily made of resin, while the main bearing member 12 is made of metal (steel or aluminum die-cast). Therefore, wear of the annular seal member 16 due to sliding contact is a concern. In particular, as the compression pressure of the fluid increases, the thrust load acting on the movable scroll body 4 also increases, making the annular seal member 16 more susceptible to wear.

[0034] In the present invention, the sliding area is reduced and torque is reduced by providing dynamic pressure grooves on the ring side surface of the annular seal member 16. Furthermore, torque can be further reduced by the wedge effect.

[0035] The annular seal member of the present invention will be described below.

[0036] An example of the annular seal member of the present invention will be described with reference to Fig. 2. Fig. 2 shows a perspective view of the annular seal member. From the viewpoint of ensuring the discharge volume of the compressor, the outer diameter φ of the annular seal member 16 is, for example, 50 mm or more, and preferably in the range of about 50 mm to 100 mm.

[0037] As shown in Fig. 2, the annular seal member 16 is an annular body with a generally rectangular cross section, is continuous around the entire circumference, and has no joints. In Fig. 2, the annular seal member 16 is provided with a plurality of V-shaped dynamic pressure grooves 18 recessed circumferentially toward the width direction of the ring at the inner diameter side end of the ring side surface 17. In addition, the corners between the inner peripheral surface 16b and both ring side surfaces 17 (including the dynamic pressure grooves 18) may be chamfered in a linear or curved manner, and when the annular seal member is manufactured by injection molding, these parts may be provided with steps 16c that protrude from the mold.

[0038] As shown in Fig. 2, one ring side of the annular seal member 16 is the surface that slides against the end face of the main bearing member facing the movable scroll, and V-shaped dynamic pressure grooves 18 are formed on this ring side so that they do not come into contact with the end face of the main bearing member. The provision of these dynamic pressure grooves 18 allows the refrigerant, etc. to flow into the dynamic pressure grooves, creating a wedge effect, and also allows the refrigerant, etc. to flow appropriately out to the part that slides against the end face of the main bearing member, thereby reducing torque and improving wear and friction resistance. Furthermore, 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 inner diameter side of the ring, which also contributes to low oil leakage of the refrigerant, etc.

[0039] In FIG. 2, it is sufficient to form the dynamic pressure grooves on at least one of the ring side surfaces, which is the sliding surface that revolves and slides, but it is preferable to form them symmetrically on both ring side surfaces, as this is not dependent on the assembly direction and provides excellent weight balance.

[0040] As shown in FIG. 2, it is preferable to provide a plurality of dynamic pressure grooves 18 spaced apart in the circumferential direction of the ring. The ring side surface between adjacent dynamic pressure grooves is the portion that slides against the main bearing member and constitutes part of the sliding surface. The area of ​​the dynamic pressure grooves (total area if there are multiple dynamic pressure grooves; the same applies below) is not particularly limited, but if the area of ​​the dynamic pressure grooves relative to the ring side surface is too small, the torque reduction effect will be reduced, and if it is too large, excessive surface pressure will occur, which may promote wear. From this perspective, the area of ​​the dynamic pressure grooves is preferably 5% to 75% of the total area of ​​the ring side surface, and more preferably 20% to 60%. The total area of ​​the ring side surface refers to the sliding area including the dynamic pressure grooves in a plan view seen from the front of the ring side surface on the sliding surface side that revolves and slides of the annular seal member, and the area of ​​the dynamic pressure grooves is the area in that plan view.

[0041] The length of each dynamic 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 each dynamic pressure groove in the ring radial direction is preferably 10% to 80% of the total ring thickness. Furthermore, in order to stabilize sliding characteristics, it is preferable that all dynamic pressure grooves be the same size and that multiple grooves (13 grooves on one side in Figure 2) be provided spaced at approximately equal intervals.

[0042] The V-shaped dynamic pressure groove will be described using FIGS. 3 and 4. FIG. 3 is a perspective view of portion A in FIG. 2, FIG. 4(a) is a view of the dynamic pressure groove as seen from the inner diameter side of the ring, FIG. 4(b) is an enlarged view of portion B, and FIG. 4(c) is an enlarged view of portion C. As shown in FIGS. 3 and 4, the dynamic pressure groove 18 is V-shaped, recessed toward the width direction of the ring along the circumferential direction of the ring. As shown in FIG. 4(a), the depth of the dynamic pressure groove 18 from the sliding surface is deepest at the center of the dynamic pressure groove 18 in the circumferential direction of the ring, and becomes shallower from the deepest portion 18d toward both ends in the circumferential direction of the ring. In other words, the closer to the sliding surface in the circumferential direction of the ring, the shallower the depth becomes. Furthermore, the depth of the dynamic pressure groove 18 from the sliding surface is constant in the radial direction of the ring. In FIG. 3, the deepest portion 18d is formed linearly.

[0043] 4(a), the dynamic pressure groove 18 has a symmetrical shape with respect to the deepest part 18d, and the bottom surface of the dynamic pressure groove 18 has a pair of first inclined surfaces 18a, 18a and a pair of second inclined surfaces 18b, 18b. Specifically, the surface is not flush from the sliding surface to the deepest part 18d, but has a first inclined surface 18a connected to the sliding surface and a second inclined surface 18b connected to the deepest part 18d and forming a smaller inclination angle with respect to the sliding surface than the first inclined surface 18a.

[0044] As shown in FIG. 4(c), the first inclined surface 18a is formed at a steeper inclination with respect to the sliding surface than the second inclined surface 18b. As a result, even if the sliding surface wears, the reduction in the opening area of ​​the dynamic pressure groove 18 is small, and torque fluctuations are unlikely to occur. The inclination angle θ1 of the first inclined surface 18a with respect to the sliding surface is not particularly limited, but is preferably 50° to 80°, and more preferably 50° to 70°. If the inclination angle θ1 is less than 50°, the increase in sliding area when the sliding surface wears will be significant, raising concerns about torque fluctuations. Furthermore, if the inclination angle θ1 exceeds 80°, the wedge action may be weakened.

[0045] On the other hand, the inclination angle θ2 (see FIG. 4(b)) of the second inclined surface 18b relative to the sliding surface is not particularly limited, but is preferably an acute angle of 0.1° to 15°, and more preferably 1° to 10°. This allows the inflowing fluid to effectively exert a wedge action. On the other hand, if the inclination angle θ2 is less than 0.1°, the inflowing fluid will have difficulty flowing to the first inclined surface 18a, and if it exceeds 15°, the deepest part 18d of the dynamic pressure groove 18 will become deeper, the volume of the dynamic pressure groove 18 will increase, and the pressure will be dispersed, which may weaken the wedge action.

[0046] The configuration 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 has a constant width in the circumferential direction of the ring, and the radius of curvature of the R-surface 18c is, for example, 0.1 to 0.3. As shown in FIG. 4(c), forming the boundary between the first inclined surface 18a and the second inclined surface 18b in a R-shape facilitates the outflow of fluid to the sliding surface between adjacent hydrodynamic grooves, thereby facilitating further torque reduction.

[0047] Furthermore, the boundary between the circumferential end of the first inclined surface 18a and the sliding surface can be connected by a curved surface (R-shaped). By forming the boundary in a R-shape, the refrigerant and the like can flow more easily onto the sliding surface, which makes it easier to achieve further torque reduction.

[0048] The depth of the deepest part 18d of the hydrodynamic groove 18 from the sliding surface is preferably 45% or less of the total width of the ring, and more preferably 30% or less. Note that, when the hydrodynamic grooves are formed on both sides of the ring, the "depth" here refers to the sum of the depths of the recesses on each side. In this case, the depth of the recesses on one side is 22.5% or less of the total width of the ring, and preferably 15% or less. If the depth exceeds 45% of the total width of the ring, the annular seal member may lack strength and may be deformed or damaged.

[0049] The substantially V-shaped dynamic pressure groove formed at the inner diameter end of the ring side surface is not limited to the configuration shown in FIGS. 3 and 4. For example, as shown in FIG. 5(a), the deepest portion 19a may be located at the end of the dynamic pressure groove 19 in the circumferential direction of the ring. Since the orbiting motion (rotational direction) of the movable scroll is unidirectional, an asymmetric shape is possible. In this case, the rotational direction of the movable scroll is the X direction. The first inclined surface and the second inclined surface described above can be appropriately used as the bottom surface of the dynamic pressure groove 19. As shown in FIG. 5(b), the deepest portion 20a of the dynamic pressure groove 20 may be formed as a flat surface parallel to the sliding surface. The deepest portion 20a may also be curved. In this case, the first inclined surface and the second inclined surface described above can be appropriately used as the bottom surface of the dynamic pressure groove 20. In these examples, the flat surface constituting the bottom surface of the dynamic pressure groove may be formed as an appropriate curved surface.

[0050] As shown in FIG. 6, the annular seal 16 is mounted in an annular groove 4d provided on the back surface of the bottom plate 4a of the movable scroll. The left side of the figure corresponds to the back pressure chamber 15a, and the right side corresponds to the space 15b. The arrow in the figure indicates the direction of pressure from the refrigerant, etc. The outer peripheral surface 16a of the annular seal 16 is pressed against and in contact with the side wall surface of the annular groove 4d on the space 15b side. This seal structure separates the back pressure chamber 15a from the space 15b. As the movable scroll rotates, the annular seal 16 rotates, and its ring side surface 17 revolves and slides against the end surface of the main bearing member 12. During this operation, the flow of refrigerant, etc., generated by the rotation, introduces the refrigerant, etc., into the dynamic pressure grooves 18, generating dynamic pressure. This dynamic pressure acts on the end surface of the annular seal 16 in a direction away from the main bearing member 12, further reducing the sliding resistance of the annular seal 16 against the main bearing member 12.

[0051] The annular groove may be provided on the main bearing member side, rather than on the bottom plate 4a side of the movable scroll. In this case, an annular seal member is attached to the annular groove and fixed within the annular groove. The ring side of the annular seal member is in sliding contact with the back surface of the bottom plate of the orbiting movable scroll. The dynamic pressure grooves described above are provided on the ring side.

[0052] The type of refrigerant used is appropriate depending on the application. The temperature of the refrigerant is, for example, about -20° C. to 140° C. The rotation speed of the orbiting scroll is mainly assumed to be about 5000 to 8000 rpm.

[0053] In the present invention, the dynamic pressure grooves provided on the ring side surface may have various shapes, as long as they are grooves that introduce a fluid and generate dynamic pressure by the fluid flow caused by the orbital motion of the movable scroll. Examples include herringbone (see FIG. 7(a)), spiral (see FIG. 7(b)), and shapes that combine these. FIG. 7 shows the planar shape of the dynamic pressure grooves, with the blackened areas in the figure representing the dynamic pressure grooves. It is desirable for the dynamic pressure grooves to be grooves that do not communicate between the inner and outer diameters of the ring side surface of the annular seal member (non-communicating grooves). Non-communicating grooves inhibit the flow of fluid midway, making it easy for dynamic pressure to be generated.

[0054] The return position of the grooves in the herringbone grooves shown in Figure 7(a) can be set as appropriate. With the shape and rotational direction of the movable scroll shown in Figure 7(a), the force acting from the inner diameter side to the outer diameter side increases as the return position moves further outward on the circumference.

[0055] 2 to 5 are examples of forming non-communicating dynamic pressure grooves in at least a portion of the inner diameter side end of the ring side surface. On the other hand, the example of FIG. 8 is an example of forming non-communicating dynamic pressure grooves in at least a portion of the outer diameter side end of the ring side surface. The annular seal member 21 shown in FIG. 8 has 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 an enlarged partial view of the same. These dynamic pressure grooves 23 have the same configuration as the V-shaped dynamic pressure groove 18 (see FIG. 3) described above, except for their positions on the ring side surface. Note that the dynamic pressure grooves 23 can adopt the configuration of the above-mentioned modified dynamic pressure grooves as appropriate.

[0056] Furthermore, the locations where the non-communicating dynamic pressure grooves are formed are not limited to only the inner diameter end or only the outer diameter end of the ring side surface, but may be both the inner diameter end and the outer diameter end of the ring side surface. In this case, for example, the dynamic pressure grooves at the inner diameter end and the dynamic pressure grooves at the outer diameter end may be formed alternately along the circumferential direction of the ring. Furthermore, the dynamic pressure grooves at the inner diameter end and the dynamic pressure grooves at the outer diameter end may be formed so as not to overlap in the radial direction of the ring.

[0057] The material of the annular sealing member of the present invention is not particularly limited, but it is preferably a molded product of a synthetic resin. Examples of synthetic resins that can be used include thermosetting polyimide resins, thermoplastic polyimide resins, polyetherketoneetherketoneketone resins, polyetherketone resins, PEEK resins, wholly aromatic polyester resins, fluororesins such as polytetrafluoroethylene (hereinafter referred to as PTFE) resins, PPS resins, polyamideimide resins, and polyamide resins. These resins may be used alone or in combination to form a polymer alloy.

[0058] The annular seal member is preferably an injection-molded article 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 these, it is particularly preferable to use a PEEK resin or a PPS resin because of their excellent friction and wear characteristics, flexural modulus, heat resistance, and sliding properties. These resins have a high modulus of elasticity, and can be used even when the temperature of the refrigerant to be sealed is high, and there is no risk of solvent cracking.

[0059] If necessary, the synthetic resin can be blended with fibrous reinforcing materials such as carbon fiber, glass fiber, and aramid fiber; spherical fillers such as spherical silica and spherical carbon; scale-like reinforcing materials such as mica and talc; and microfiber reinforcing materials such as potassium titanate whiskers. Also blended are 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. These can be blended alone 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 preferred, as it easily achieves the properties required for the annular seal member of the present invention. The inclusion of carbon fiber improves mechanical strength, such as flexural modulus, while the inclusion of PTFE resin improves sliding properties.

[0060] When using a synthetic resin, the above-mentioned raw materials are melted and kneaded to form molding pellets, which are then molded into a predetermined shape using a known injection molding method or the like. When manufactured by injection molding, the gate position is not particularly limited, but it is preferable to provide it on the inner peripheral surface of the ring from the viewpoint of ensuring sealing performance and eliminating the need for post-processing. Furthermore, it is more preferable that the gate positions be multi-point gates (e.g., 3 to 6 points) arranged at equal intervals around the circumferential direction, and that the gate positions and the positions of the dynamic pressure grooves do not overlap in the radial direction of the ring. In this case, the annular seal member will have gate marks on the inner peripheral surface at positions that do not overlap with the dynamic pressure grooves in the radial direction of the ring. [Example]

[0061] In order to confirm the dependency of the dynamic friction coefficient on the surface pressure due to differences in the sliding area, thrust tests were conducted with the sliding area fixed and the load divided into three levels.

[0062] Examples and Comparative Examples Annular test pieces for the examples and comparative examples were manufactured by injection molding using a resin composition (Bearee AS5302, manufactured by NTN Corporation) containing PPS resin as the main material and blended with PTFE resin and carbon fiber. The test piece of the comparative example had an outer diameter of 21 mm, an inner diameter of 17 mm, a radial length of 4 mm, and an axial length of 1.6 mm, and no dynamic pressure grooves were provided on the ring side surface. On the other hand, the test piece of the example had an outer diameter of 21 mm, an inner diameter of 17 mm, a radial length of 4 mm, and an axial length of 1.6 mm, and four roughly V-shaped dynamic pressure grooves were provided along the circumferential direction of the ring at the inner diameter end of the ring side surface, as shown in FIG. 3. The depth of the deepest part of the dynamic pressure grooves was 0.1 mm, the inclination angle of the first inclined surface relative to the sliding surface was approximately 65°, and the inclination angle of the second inclined surface relative to the sliding surface was approximately 3°. The area of ​​the dynamic pressure grooves was 40% of the total area of ​​the ring side surface.

[0063] A schematic diagram of the thrust testing machine is shown in Figure 9. A test piece 33 was attached to the tip of a load shaft 31 and pressed against a mating material 34 (ADC12, outer diameter φ33 mm, thickness 10 mm, sliding surface with the test piece was flat-polished to approximately Ra 0.8 μm) attached to a rotating shaft 35 with a predetermined load F, and a thrust test was performed in oil 32 under the following conditions. For each test, the dynamic friction coefficient was measured immediately before the end of the test. The relationship between surface pressure and dynamic friction coefficient is shown in Figure 10.

[0064] <Test conditions> Speed: 2m / sec Surface pressure: 1MPa, 2MPa, 3MPa Ambient temperature: Room temperature (as usual) Lubrication: Oil (PAG oil, Idemitsu Daphne Hermetic Oil PS) Test time: 30 mins at each pressure Number of tests: n=1

[0065] As shown in Figure 10, the dynamic friction coefficient tends to decrease as the surface pressure (load) increases, indicating that it is dependent on the surface pressure, and that the dynamic friction coefficient (torque) decreased by reducing the area of ​​the sliding surface. Therefore, by forming dynamic pressure grooves as in the example, it is possible to reduce torque. [Industrial Applicability]

[0066] The annular seal member of the present invention can exhibit stable low torque performance without compromising durability or sealing function, and can therefore be widely used as an annular seal member for scroll compressors. In addition, it is possible to eliminate the need for a thrust bearing member. [Explanation of symbols]

[0067] 1 Compressor 2. Housing 3 Fixed scroll body 3a Bottom plate part 3b Spiral wall 3c opening 4. Movable scroll 4a Bottom plate part 4b Spiral wall 5 Compression Chamber 6a Stator 6b rotor 7 shaft 8 balance weights 9 Main bearing 10 Sub bearing 11 Slewing bearing 12 Main bearing member 13 Shaft seal 14 Motor Room 15a Back pressure chamber 15b Space 16 Annular seal member 17 Ring side 18 Hydrodynamic groove 18a First inclined surface 18b Second inclined surface 18c R side 18d Deepest 19 Hydrodynamic groove 19a Deepest 20 Hydrodynamic groove 20a deepest part 21 Annular seal member 22 Ring side 23 Hydrodynamic groove 31 Load axis 32 Oil 33 Test specimens 34 Counterpart 35 Rotation axis

Claims

1. The scroll assembly comprises a fixed scroll body having a bottom plate portion and a spiral wall erected on the surface thereof, a movable scroll body having a bottom plate portion and a spiral wall erected on the surface thereof, a shaft, a main bearing for rotatably supporting the shaft, and a main bearing member for fixing the main bearing; In a scroll compressor, the movable scroll body is revolved around the axis of the fixed scroll body by rotation of the shaft, compressing a fluid in a compression chamber, and the fluid is supplied to a back pressure chamber on the back side of the movable scroll body, an annular seal member that is fitted into at least one annular groove formed in either a back surface of the bottom plate portion of the movable scroll body or an end face of the main bearing member facing the movable scroll body, and that seals the back pressure chamber, The annular seal member has a dynamic pressure groove formed at least on a sliding surface that revolves and slides on a ring side surface, the hydrodynamic grooves have a generally V-shaped configuration that is recessed toward the width direction of the ring along the circumferential direction of the ring, and the depth of the hydrodynamic grooves from the sliding surface decreases from the deepest part toward both ends in the circumferential direction of the ring and is constant in the radial direction of the ring.

2. 2. The annular seal member according to claim 1, wherein the area of ​​the dynamic pressure grooves is 5% to 75% of the total area of ​​the ring side surface.

3. 3. The annular seal member according to claim 1, wherein the dynamic pressure groove is not flush with the sliding surface from the sliding surface to the deepest portion, but has a first inclined surface connected to the sliding surface and a second inclined surface connected to the deepest portion and forming an inclination angle with respect to the sliding surface smaller than that of the first inclined surface.

4. The annular seal member according to claim 3, characterized in that the inclination angle of the first inclined surface relative to the sliding surface is 50° to 80°, and the inclination angle of the second inclined surface relative to the sliding surface is 0.1° to 15°.

5. 5. The annular seal member according to claim 3, wherein the boundary between the first inclined surface and the second inclined surface in the dynamic pressure groove is connected by a curved surface.

6. 6. The annular seal member according to claim 1, wherein a plurality of the dynamic pressure grooves are provided at intervals in the circumferential direction of the ring, and a ring side surface between adjacent dynamic pressure grooves constitutes a part of the sliding surface.

7. 7. The annular seal member according to claim 1, wherein the annular seal member is made of a synthetic resin, and the synthetic resin is a polyphenylene sulfide resin or a polyether ether ketone resin.

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