Spiral groove face seal

Abstract: A spiral groove face seal has a stationary sealing ring (4) and a rotating sealing ring (3) with opposed radially extending sealing surfaces for restricting a fluid flow from a high pressure fluid side (H) to a low pressure fluid side (L). The rotating sealing ring (3) is in contact with the stationary sealing ring (4) at at least a part of the sealing surfaces when the rotating ring (3) is stopped, and the rotating ring (3) is brought out of contact with the stationary sealing ring (4) when the rotating sealing ring (3) is rotated. The spiral groove face seal comprises spiral grooves (31) formed on one of the sealing surfaces and extending from one circumference of the sealing surface in a radial direction and from the high pressure fluid side (H) to the low pressure fluid side (L) so as to pump fluid in the high pressure fluid side (H) towards the low pressure fluid side (L), and a recess (h t ) formed on one of the sealing surfaces and located in an area of the spiral grooves (31) or a part of the spiral grooves (31).

Description: BACKGROUND OF THE INVENTION The present invention relates to a spiral groove face seal, and more particularly to a spiral groove face seal in which an end face of a rotating sealing ring rotatable with a rotating shaft and an end face of a stationary sealing ring facing the rotating sealing ring form sealing surfaces for restricting a fluid flow from a high pressure side to a low pressure side, according to the preambles of claims 1 and 8. In general, a spiral groove face seal is used for sealing the space between a rotating shaft and housing thereof. The spiral groove face seal is incorporated in, for example, gas turbines or compressors which generate high pressure. Conventionally, the end face seal comprises two annular sealing rings facing with each other, one of which includes a spiral groove area having a plurality of spiral grooves and a dam area extending from the spiral groove area and having a flat surface, the other of which has a flat surface. The end face seal of this type is disclosed in, for example, U.S. Pat. No. 4,212,475. In U.S. Pat. No. 4,212,475 spiral grooves are formed on a rotating sealing ring as shown in Figs. 13 through 15. In Figs. 13 through 15, a rotating shaft 1 accommodated in a housing 9 is provided with a sleeve 2. The sleeve 2 is connected to a rotating sealing ring 3 through a pin. A stationary sealing ring 4 is pressed against the rotating sealing ring 3 by urging force of springs 10 interposed between a sealing ring retainer 5 and a spring retainer 6. The rotating sealing ring 3 includes a spiral groove area 35 having a plurality of spiral grooves 33 and a dam area 34 extending from the spiral groove area 35 and having a flat surface. The spiral grooves 33 serves to pump fluid (gas) inwardly, and the dam area 34 serves to restrict fluid flow pumped by the spiral grooves 33. Relative motion of the rotating sealing ring 3 and the stationary sealing ring 4 is caused by rotating the rotating shaft 1. Under dynamic conditions, the end face seal will be exposed to hydrostatic forces from fluid pressure and hydrodynamic forces from the spiral grooves 33. The spiral groove pattern pumps fluid (gas) inwardly to thereby form fluid film on the sealing surfaces, whereby the rotating sealing ring 3 is brought out of contact with the stationary sealing ring 4. The leakage of fluid is restricted by keeping the gap between the sealing surfaces small. On the other hand, another type of end seal face is disclosed in U.S. Pat. No. 3,499,653. In U.S. Pat. No. 3,499,653 the end face seal comprises a rotating sealing ring and a stationary sealing ring as well as the end face seal in Fig. 13 through 15. One of the sealing rings has spiral grooves and a tapered surface to provide the necessary wedge-shaped space at the radially outer regions thereof. In the end face seal of this type, when fluid enters into the wedge-shaped space between the rotating sealing ring and the stationary sealing ring, the sealing surfaces become noncontact by hydrostatic forces from fluid pressure. When the fluid has low pressure, the sealing surfaces do not become noncontact only by hydrostatic forces, but become noncontact by hydrodynamic forces from the spiral grooves in addition to hydrostatic forces. However, in the end face seal disclosed in U.S. Pat. No. 4,212,475, a ridge portion (on which spiral grooves are not formed) on the spiral groove area 35 is located on the same plane as the dam area 34 as shown in Figs. 13 through 15. Therefore, when starting the end face seal, the ridge portion and the dam area 34 of the rotating sealing ring 3 contact directly the surface of the stationary sealing ring 4 facing thereto. In this specification, this direct contact of two solid members is defined as "solid contact." As a result, fluid (gas) cannot be entered into the contacting surfaces of two sealing members, a large quantity of starting torque is thus required due to a large quantity of sliding resistance when starting in such a state that fluid pressure is applied to the sealing rings. The same is true in the stationary sealing ring having spiral grooves. On the other hand, the end face seal disclosed in U.S. Pat. No. 3,499,653 is a hydrostatic mechanical seal of the gap type so that the sealing surfaces become noncontact by hydrostatic forces from fluid pressure. Therefore, there is no problem in starting resistance at the time of starting. However, since the sealing surfaces become noncontact by hydrostatic force from fluid, fluid is leaked from the sealing surfaces having a certain gap when the machine incorporating the seal face is stopped. SUMMARY OF THE INVENTION In accordance with the present invention a spiral groove face seal set forth in claim 1 and a rotating sealing ring as set forth in claim 8 is provided. Preferred embodiments of the invention are disclosed in the dependent claims. It is an object of the present invention to provide a spiral groove face seal which can lower starting torque by reducing starting resistance when starting the machine and prevent fluid from leaking from a high pressure side to a low pressure side when the machine is stopped. According to the present invention, there is provided a spiral groove face seal having a stationary sealing ring and a rotating sealing ring with opposed radially extending sealing surfaces, the rotating sealing ring being in contact with the stationary sealing ring at at least a part of the sealing surfaces when the rotating sealing ring is stopped, the rotating ring being brought out of contact with the stationary sealing ring when the rotating sealing ring is rotated, the improvement comprising: spiral grooves formed in one of the sealing surfaces and extending from one circumference of the sealing surface in a radial direction and from a high pressure fluid side to a low pressure fluid side so as to pump fluid in the high pressure fluid side towards the low pressure fluid side; and a recess formed in one of the sealing surfaces, the recess being located in an area of the spiral grooves or a part of the spiral grooves. With the above structure, fluid can enter into the sealing surfaces through the recess in such a state that the rotating sealing ring is in contact with the stationary sealing ring at at least a part of the sealing surfaces, the starting resistance can be remarkably reduced and the starting torque required to start the machine is reduced under high pressure, thus meeting the needs of saving energy. In the case where the depth of the recess is large, there is a possibility that pumping action generated by the spiral grooves is affected by the recess to thereby impair the noncontacting function. However, adverse influences on pumping action is not caused because the depth of the recess is not large. The depth of recess is preferably not more than 3 µm in consideration of surface roughness caused by working and unevenness of surface. Further, according to the present invention, since the rotating sealing ring contacts the stationary sealing ring at a part of sealing surfaces (the dam area), leakage of fluid from the high pressure side to the low pressure side can be prevented when the machine is stopped. The above and other objects, features and advantages of the present invention will become more apparent from the following description when taken in conjunction with the accompanying drawings in which preferred embodiments of the present invention are shown by way of illustrative example. BRIEF DESCRIPTION OF THE DRAWINGS In the drawings: Fig. 1 is a cross-sectional view showing a spiral groove face seal according to an embodiment of the present invention; Fig. 2 is an end view showing spiral grooves formed on a rotating sealing ring of the spiral groove face seal according to the embodiment of the present invention; Fig. 3 is an enlarged cross-sectional view showing the rotating sealing ring and a stationary sealing ring of the spiral groove face seal according to the embodiment of the present invention; Fig. 4 is a cross-sectional view showing an operation of the spiral groove face seal according to the embodiment of the present invention; Fig. 5 is a graph showing the relationship of the dam width ratio (DW) and the load ratio (F₁/F₂) of the end face seal having the recess h t relative to the end face seal having no recess; Fig. 6 is a graph showing the relationship between the dam width ratio (DW) and the gap (h) of the sealing surfaces of the end face seal; Fig. 7 is a graph showing the relationship between the dam width ratio (DW) and the leakage rate (Q) of the end face seal; Fig. 8 is a graph showing the relationship between the dam width ratio (DW) and the film stiffness (dW/dh); Fig. 9 is a cross-sectional view showing a spiral groove face seal according to a second embodiment of the present invention; Fig. 10 is a cross-sectional view showing a spiral groove face seal according to a third embodiment of the present invention; Fig. 11 is a cross-sectional view showing a spiral groove face seal according to a fourth embodiment of the present invention; Fig. 12 is a cross-sectional view showing a spiral groove face seal according to a fifth embodiment of the present invention; Fig. 13 is a cross-sectional view showing a conventional spiral groove face seal; Fig. 14 is a cross-sectional view showing a rotating sealing ring of the conventional spiral groove face seal; and Fig. 15 is an end view showing spiral grooves of the conventional spiral groove face seal. DETAILED DESCRIPTION OF THE INVENTION A spiral groove face seal of the present invention will be described below with reference to Figs. 1 through 8. In Fig. 1, a sleeve 2 is fixedly secured to a rotating shaft 1 accommodated in a housing 9. A rotating sealing ring 3 is connected to the sleeve 2 through a pin 20. The rotating sealing ring 3 faces a stationary ring 4, and when starting, the rotating sealing ring 3 contacts the stationary ring 4 at the dam area 32 in a state of solid contact. The rotating sealing ring 3 comprises an annular body defining annular surfaces on the opposite sides thereof. Incidentally, the rotating sealing ring 3 is fixedly secured to the sleeve 2 by a sleeve 7 and a retaining nut 8. A stationary sealing ring 4 is pressed against the rotating sealing ring 3 by urging force of coiled springs 10 interposed between a sealing ring retainer 5 and a spring retainer 6. The rotating sealing ring 3 includes a spiral groove area 38 having a plurality of spiral grooves 31 which are directed from the high pressure fluid side H to the low pressure fluid side L as shown in Fig. 2. The rotating sealing ring 3 also has a dam area 32 extending from the spiral groove area 38. Fig. 3 is an enlarged cross-sectional view showing the rotating sealing ring 3 and the stationary sealing ring 4. The rotating sealing ring 3 is provided with a recess in the spiral groove area 38. The height of the ridge portion of the spiral groove area 38 is lower than that of the dam area 32 by the recess h t . That is, there is provided a gap between the ridge portion of the spiral groove area 38 of the rotating sealing ring 3 and the surface of the stationary sealing ring 4 when the machine is stopped. Reference numerals 42, 44, 46 and 48 are sealing members such as O ring in Fig. 1. With the above structure, the face seal prevents fluid having high pressure at the high pressure fluid side H from leaking out towards the low pressure fluid side L to the utmost extent. Next, the operation of the spiral groove face seal thus constructed will be described below with reference to Fig. 4. Relative motion of the rotating sealing ring 3 and the stationary sealing ring 4 is caused by rotating the rotating shaft 1. Under dynamic conditions, the face seal will be exposed to hydrostatic forces from fluid pressure and hydrodynamic forces from the spiral grooves. The spiral groove pattern pumps fluid (gas) inward to thereby form fluid film on the sealing surfaces. As a result, the rotating sealing ring 3 is brought out of contact with the stationary sealing ring 4, thus forming the gap h between the sealing surfaces of the rotating sealing ring 3 and the stationary sealing ring 4. That is, the dam area 32 of the rotating sealing ring 3 is brought out of contact with the surface 50 of the stationary sealing ring 4. The leakage of fluid is restricted by keeping the gap h between two sealing surfaces small. In the conventional end face seal, since reduction of starting torque is not considered, the recess is not provided (h t = 0 µm). Fluid pressure at the high pressure side H is represented by P₂, fluid pressure at the low pressure side L is represented by P₁. Assuming that fluid is not entered into the sealing surfaces of the solid contact (contacting portion of the surface 32 and the surface 50 at the dam area) from the surroundings. The width of the spiral groove is equal to the width of the ridge. In other words, the area of the ridge portion is equal to that of the spiral grooves. In the conventional end face seal in Figs. 13 through 15, the load applied to the solid contact portion of the sealing surfaces is expressed by the following formula. F = π(r g ² - r b ²)P₂ + π(r b ² - r₁²)P₁ + F sp + π(r₂² - r g ²)P₂/2 where F sp is urging force generated by the coiled springs 10. On the other hand, in the end face seal of the present invention in Figs. 1 and 2, the load applied to the solid contact portion of the sealing surfaces is expressed by the following formula. F = π(r g ² - r b ²)P₂ + π(r b ² - r₁²)P₁ + F sp According to the present invention, the load applied to the solid contact portion of the sealing surfaces is reduced by π(r₂² - r g ²)P₂/2 . The starting resistance is decreased by the quantity corresponding to the reduced load when starting the machine. In the case where the recess h t is large, pumping action by the spiral grooves is decreased. Therefore, there is a possibility that function of avoiding the solid contact of the sealing surfaces is impaired, it is desirable to make the recess h t small. Taking into consideration unevenness of surface condition of the sealing ring caused by working such as lapping (not more than 1 µm), deformation caused by heat or pressure, it is desirable to make the recess h t small under the condition that high pressure fluid can be sufficiently entered into the recess. That is, the recess of not more than 3 µm is preferably selected. The end face seal of the present invention is provided with the recess h t unlike the conventional end face seal for the purpose of reducing the starting resistance. The end face seal includes a sealing ring which has a step on the surface extending radially outwardly. That is, the sealing ring has two surfaces, one of which is formed with spiral grooves on the spiral groove area for pumping action to pump fluid of the high pressure side to the low pressure side, the other of which has a flat surface on the dam area for restricting the fluid flow from the high pressure side to the low pressure side and performing sealing action. In other words, the sealing ring has the spiral groove area for pumping action and the dam area for sealing action. As mentioned above, the end face seal of the present invention is provided with the recess h t for the purpose of reduction of the starting resistance. As a result, the gap and film stiffness are lowered compared with the conventional spiral groove face seal. Therefore, the groove depth (h₀), the balance ( BR = (r₂² - r b ²)/(r₂² - r₁²) ) and the dam width ratio ( DW = (r g - r₁)/(r₂ - r₁) ) must be properly chosen to maintain a narrow gap between two surfaces of the sealing rings without causing the solid contact by various disturbance such as vibration in an axial direction of the sealing surfaces, vibration in an inclined direction relative to the axis, or deformation. Next, characteristics of the end face seal of the present invention is shown in Figs. 5 through 8 on condition of table 1. inside diameter of the sealing surface (2r₁ mm) 138 outside diameter of the sealing surface (2r₂ mm) 180 balance diameter (2r b mm) 142.9 spiral groove angle ( α degree) 15 pressure at low pressure side (P₁ kgf/cm²) 1 pressure at high pressure side (P₂ kgf/cm²) 31 temperature of fluid at the sealing surfaces ( C) 30 rotational speed (N/min) 13000 fluid (gas) air where the spiral groove angle is defined as an angle between velocity vector at a point of the spiral groove and tangent of the point of the spiral groove. Fig. 5 shows the effect of reduction of load of the sealing surfaces by forming the recess h t . In Fig. 5, the horizontal axis represents the dam width ratio (DW), the vertical axis represents the load ratio (F₁/F₂) of the end face seal having the recess h t relative to the end face seal having no recess, wherein the former (F₁) is obtained using the formula (2) and the latter (F₂) is obtained using the formula (1). As is apparent from Fig. 5, the smaller the dam width ratio is becoming, the smaller the load ratio is becoming. Assuming that coefficients of friction at all points on the sealing surfaces are the same, the starting resistance can be greatly reduced by forming the recess h t . Further, when the balance is constant (BR=0.897), and the dam width ratio (DW) and the groove depth (h₀) are variable, the gap of the sealing surface (h), the leakage rate (Q) and the film stiffness (dW/dh) having 1 µm in thickness are shown in Figs. 6 through 8. When calculating the film stiffness (dW/dh), the gap of the sealing surfaces of 1 µm is chosen because there is a fair chance to occur the solid contact. In Figs. 6 through 8, the sealing surfaces having the recess is shown using the solid lines, the sealing surfaces having no recess is shown using broken lines. The film stiffness is remarkably decreased by forming the recess h t as shown in Fig. 8. The film stiffness is decreased, especially, at the range of shallow groove depth (h₀). On the other hand, the gap of the sealing surfaces (h) and the leakage rate (Q) are not affected so much by the recess in the range of small dam width ratio of 0.3 or thereabouts. Next, in the end face seal of the present invention, a process of determining preferable range of the groove depth (h₀), the dam width ratio (DW) and the balance (BR) will be described below on the basis of the result of Figs. 5 through 8. In the end face seal of the present invention, the starting resistance can be greatly reduced by forming the recess as shown in Fig. 5, however, the film stiffness is lowered as shown in Fig. 8. The decrease of the film stiffness is disadvantageous because there develops a tendency to cause the solid contact and to prevent normal operation. Therefore, the film stiffness must be improved. The groove depth (h₀) and the dam width ratio (DW) are closely connected with the film stiffness among the above three factors as shown in Fig. 8. To be more specific, when the groove depth h₀ is becoming smaller, the film stiffness is becoming higher. On the contrary, when the groove depth h₀ is becoming smaller, the gap of the sealing surfaces is becoming narrower as shown in Fig. 6 and the leakage rate smaller as shown in Fig. 7. However, in the case where the gap of the sealing surfaces is too small, there is a possibility to occur the solid contact of the sealing surfaces. Therefore, an extremely small groove depth h₀ cannot be chosen. Further, taking into consideration working error of the groove, the groove depth h₀ of an extremely small value cannot be chosen from this aspect too. Accordingly, in the end face seal of the present invention, the lower limit of the groove depth h₀ is preferably 3 µm. On the contrary, the upper limit of the groove depth is expanded up to 15 µm in consideration of rough sealing effect which is applicable to a relatively low pressure fluid. Next, the relationship between the dam width ratio and the film stiffness will be explained below. When the dam width ratio is becoming smaller, the film stiffness is becoming higher as shown in Fig. 8. Therefore, as far as the film stiffness is concerned, it is desirable to make the dam width ratio small. On the other hand, when making the dam width ratio small, the leakage rate increases as shown in Fig. 7. Further, in the case where the dam width ratio is extremely small and there are flaws or chips on the sealing ring at the dam area 32, the end face seal does not function as seal due to leakage of fluid therefrom. Accordingly, in the end face seal of the present invention, the lower limit of the dam width ratio is preferably 0.2. The upper limit of the dam width ratio is preferably 0.4 so that the film stiffness does not become excessively small. Concerning the balance, taking into consideration the fact that the solid contact is kept at the dam area of the sealing surfaces when the machine is stopped and leakage of fluid is minimized from the high pressure side to the low pressure side, it is desirable to select the balance of between 0.8 and 0.9. As is apparent from the foregoing description, the preferable range of the groove depth (h₀), the dam width ratio ( DW = (r g - r₁)/(r₂ - r₁) ) and the balance ( BR = (r₂² - r b ²)/(r₂² - r₁²) ) are shown in table 2. recess (h t µm) not more than 3 the groove depth (h₀ µm) 3 to 15 the dam width ratio (DW) 0.2 to 0.4 the balance (BR) 0.8 to 0.9 Next, other embodiments of the present invention will be described below with reference to Figs. 9 through 12. Those parts and components in Figs. 9 through 12 which are structurally and functionally identical to those shown in Figs. 1 through 4 are denoted using the same reference numerals. Fig. 9 is a cross-sectional view showing a spiral groove face seal according to a second embodiment of the present invention. According to this embodiment, the recess h t is provided on the stationary ring 4. The area where the recess h t is formed is the same area as the spiral groove area 38. The second embodiment has the same function and effect as the first embodiment in Fig.s 1 through 4. Fig. 10 is a cross-sectional view showing a spiral groove face seal according to a third embodiment of the present invention. According to this embodiment, the recess h t extends in the dam area 32. The recess h t is formed on the rotating sealing ring 3 in Fig. 10(a), and the recess h t is formed on the stationary sealing ring 4 in Fig. 10(b). The load applied to the solid contact portion of the sealing surfaces is expressed by the following formula. F = π(r t ² - r b ²)P₂ + π(r b ² - r₁²)P₁ + F sp (r t > r b ) where 2r t is inside diameter of the recess h t . Therefore, the load of this embodiment is smaller than that of the first embodiment in Fig.s 1 through 4. Fig. 11 is a cross-sectional view showing a spiral groove face seal according to a fourth embodiment of the present invention. According to this embodiment, the recess h t is located inside the spiral groove area 38. The recess h t is formed on the rotating sealing ring 3 in Fig. 11(a), and the recess h t is formed on the stationary sealing ring 4 in Fig. 11(b). The load applied to the solid contact portion of the sealing surfaces is expressed by the following formula. F = π(r g ² - r b ²)P₂ + π(r b ² - r₁²)P₁ + F sp + π(r t ² - r g ²)P₂/2 where 2r t is inside diameter of the recess h t . Therefore, the load of this embodiment is larger than that of the first embodiment in Figs. 1 through 4. In the embodiments in Figs. 1 through 11, the outside of the spiral groove area is high pressure side and the inside of the spiral groove is low pressure side, the fluid flows inwardly at the sealing surfaces. On the contrary, in a fifth embodiment in Fig. 12, the outside of the spiral groove area is low pressure side and the inside of the spiral groove area is high pressure side, the fluid flows outwardly at the sealing surfaces. The fifth embodiment has the same function and effect as the embodiments in Fig.s 1 through 11. As is apparent from the foregoing description, according to the spiral groove face seal of the present invention, since fluid can enter into the sealing surfaces through the recess in such a state that the rotating sealing ring is in contact with the stationary sealing ring at at least a part of the sealing surfaces, the starting resistance can be remarkably reduced and the starting torque required to start the machine are reduced under high pressure, thus avoiding a failure in starting the machine and preventing the sealing surfaces from being damaged. Further, according to the present invention, the recess is provided on one of the sealing rings, so that pumping action generated by the spiral grooves is affected by the recess. However, a desirable performance is obtained by selecting the range of the groove depth (h₀), the dam width ratio (DW) and the balance (BR), thereby preventing noncontacting function of the sealing surfaces from being impaired. Furthermore, according to the present invention, since the rotating sealing ring contacts the stationary sealing ring at the dam area, leakage of fluid from the high pressure side to the low pressure side can be prevented when the machine is stopped. Although certain preferred embodiments have been shown and described, it should be understood that many changes and modifications may be made therein without departing from the scope of the appended claims.

Claims: A spiral groove face seal having a stationary sealing ring (4) and a rotating sealing ring (3) with opposed radially extending sealing surfaces, said rotating sealing ring (3) being in contact with said stationary sealing ring (4) at at least a part of said sealing surfaces when said rotating ring (3) is stopped, said rotating sealing ring (3) being brought out of contact with said stationary sealing ring (4) when said rotating sealing ring (3) is rotated, and spiral grooves (31) formed in one of said sealing surfaces and extending from one circumference of said sealing surface in a radial direction and from a high pressure fluid side (H) to a low pressure fluid side (L) so as to pump fluid in said high pressure fluid side towards said low pressure fluid side; characterized in that a recess is formed in one of said sealing surfaces, said recess being located in an area of said spiral grooves (31) or a part of said spiral grooves (31). The spiral groove face seal according to claim 1, further comprising said recess extending beyond said spiral grooved area (38) to a flat surface of the ungrooved area. The spiral groove face seal according to claim 1, wherein said recess is formed on said sealing surface of said rotating sealing ring (3). The spiral groove face seal according to claim 1, wherein said recess is formed on said sealing surface of said stationary sealing ring (4). The spiral groove face seal according to claim 1, wherein said spiral grooves (31) extend from the outside circumference of said sealing surface towards the center of said sealing surface. The spiral groove face seal according to claim 1, wherein said spiral grooves (31) extend from the inside circumference of said sealing surface towards the center of said sealing surface. The spiral groove face seal according to claim 1, wherein said spiral grooves (31) have a depth of between approximately 0.003 and 0.015 mm; said face seal has a dam width ratio of between approximately 0.2 and 0.4 according to one of the following formulas Dam Width Ratio = (r g - r₁)/(r₂ - r₁) Dam Width Ratio = (r₂ - r g )/(r₂ - r₁) where groove diameter 2r g is the diameter of a circle defined by the boundary of the grooved area (38) and the ungrooved area in said sealing surface, 2r₁ is the inside diameter and 2r₂ is the outside diameter; and said face seal has a balance of between 0.8 and 0.9 according to one of the following formulas Balance = (r₂² - r b ²)/(r₂² - r₁²) Balance = (r b ² - r₁²)/(r₂² - r₁²) where 2r₂ is the outside diameter of the sealing interface, 2r₁ is the inside diameter and 2r b is the balance diameter. A rotating sealing ring (3) comprising: an annular body defining annular surfaces on the opposite sides thereof, said annular body having spiral grooves (31) formed in one of said surfaces, said spiral grooves (31) extending from a circumferential edge of said annular body in a radial direction, characterized in that said annular body having a recess formed in said one of said surfaces, said recess being located in an area (38) of said spiral grooves (31) or a part of said spiral grooves (31). The rotating sealing ring according to claim 8, further comprising said recess extending beyond said spiral grooved area (38) to a flat surface of the ungrooved area.

Inventors: KIMURA YOSHIKAZU C O EBARA CO, TSUBOI HIDEO C O EBARA CO LTD, KIMURA, YOSHIKAZU, C/O EBARA CO.,LTD., TSUBOI, HIDEO, C/O EBARA CO.,LTD.
Applicants: EBARA CORP, EBARA CORPORATION
Application year: 1991
Main classification: F16J15
IPC codes: F16J15