FACE SEAL WITH DOUBLE GROOVE ARRANGEMENT

Abstract: A non-contacting grooved face seal for a shaft rotating at high speed with a combination of two groove patterns (34, 44) on one of the two sealing faces of mating seal rings; one pattern being relatively deep angled grooves (34), and the other pattern being relatively shallow grooves (44). The relatively deep groove pattern (34) is optimized for hydrodynamic operation and on shaft rotation pumps the sealed fluid between the sealing faces to set the running clearance. The relatively shallow pattern (44) is designed to hydrostatically prevent a friction lock of the sealing faces when the shaft is at or near to a stationary condition.

Description: FACE SEAL WITH DOUBLE GROOVE ARRANGEMENT FIELD OF THE INVENTION This invention relates to sealing devices for rotating shafts, wherein a sealed fluid is employed to generate hydrostatic-hydrodynamic or aerostatic- aerodynamic forces between opposed interacting face- type sealing elements, one stationary and the other rotating. These forces provide for slight separation and non-contacting operation of the sealing elements, thereby minimizing face wear and friction power losses while maintaining low fluid leakage. BACKGROUND OF THE INVENTION Rotary fluid film face seals, also called gap or non-contacting face seals, are usually applied to high¬ speed and/or high-pressure rotating equipment wherein the use of ordinary mechanical face seals with face contact would result in excessive heat generation and wear. Non-contacting operation avoids this undesirable face contact at times when the shaft is rotating above a certain minimum speed, which is called a lift-off speed. There are various ways of accomplishing the above non-contacting operation. One of the more commonly used ways includes the formation of a shallow spiral groove pattern in one of the sealing faces. The sealing face opposite the grooved face is relatively flat and smooth. The face area where these two sealing faces define a sealing clearance is called the sealing interface. The above-mentioned spiral groove pattern on one of the sealing faces normally extends inward from the outer circumference and ends at a particular face diameter called the groove diameter, which is larger then the inner diameter of the seal interface. The non-grooved area between the groove diameter and the inner interface diameter serves as a restriction to fluid outflow. Fluid delivered by the spiral pattern must pass through this restriction and it can do so only if the sealing faces separate. The way this works is through pressure build-up. Should the faces remain in contact, fluid will be compressed just ahead of the restriction, thus building up pressure. The pressure causes separation force which eventually becomes larger than the forces that hold the faces together. In that moment the sealing faces separate and allow the fluid to escape. During operation of the seal, an equilibrium establishes itself between fluid inflow through spiral pumping and fluid outflow through face separation. Face separation is therefore present as long as the seal is operating, which means as long as one face is rotating in relation to the opposite face. However, spiral pumping is not the only factor that determines the amount of the separation between the sealing faces. Just as the spirals are able to drive the fluid into the non-groove portion of the sealing interface past the groove diameter, so can the pressure differential. If enough of a pressure difference exists between the grooved end of the interface and the non- grooved end, fluid will also be forced into the non- grooved portion of the interface, thereby separating the faces and forming the clearance. Both ways in which clearance can be formed between the sealing faces, one with speed of rotation, the other with pressure differential, are distinct and separate, even though the effects of both combine on the operating seal. If there is no pressure difference and the seal face separation occurs strictly due to face rotation, forces due to fluid flow are known as hydrodynamic forces if the fluid sealed is a liquid, and aerodynamic forces if the fluid sealed is a gas. On the other hand, if there is no mutual rotation between the two sealing faces and face separation is strictly the consequence of pressure differential between both ends of the sealing interface, forces due to fluid flow are called hydrostatic forces if the fluid sealed is a liquid, and aerostatics forces if the fluid sealed is a gas. In the following, the terms hydrostatic and hydrodynamic are used for both liquid and gas effects since these latter terms are more conventionally used when describing both liquid and gas seals. A typical spiral groove seal needs to provide acceptable performance in terms of leakage and the absence of face contact during all regimes of seal operation. It must do so not only at top speed and pressure, but also at standstill, at start-up, acceleration, at periods of equipment warm-up or at shutdown. At normal operating conditions, pressure and speed vary constantly, which results in continuous adjustments to the running clearance. These adjustments are automatic; one of the key properties of spiral groove seals is their self-adjustment capability. On change in speed or pressure, the face clearance adjusts automatically to a new set of conditions. Hydrostatic and hydrodynamic forces cause this adjustment. The operating envelope of speeds and pressures is usually very wide and a seal design of necessity must be a compromise. For its performance to be acceptable at near-zero speed or pressure, it is less than optimum at operating speed and pressure. This is simply due to the fact that, both in terms of pressure and speed, the seal has to be brought up to operating conditions from zero speed and zero pressure differential. Especially critical to seal operation is the start¬ up. If the seal is applied to a centrifugal gas compressor, the full suction pressure differential is often imposed onto the seal before the shaft starts turning. This presents a danger in that the sealing faces will lock together with friction. Face lock results when the hydrostatic force is insufficient to counter pressure forces that maintain the seal faces in contact. Face lock can lead to seal destruction, in which excessive break-away friction between contacting seal faces can cause heavy wear or breakage of internal seal components. First then, spiral grooves must be able to separate the sealing faces hydrodynamically for full speed non- contacting operation. This normally requires fairly short and relatively deep spiral grooves. Second, the spiral grooves must be able to unload the sealing faces hydrostaticaUy for start/stops to prevent face lock. For this, the grooves have to be extended in length. The extended grooves in turn cause more separation and leakage during full speed operation. The full speed leakage of a typical 3.75 inch shaft seal with short and relatively deep spirals may be about .9 SCFM (i.e. Standard Cubic Feet per Minute) at 1,000 psig and 10,000 rpm. However, full speed leakage for such a seal with extended grooves may reach 2.4 SCFM under the same conditions, almost triple the previous value. The constant burden of larger-than-necessary leakage represent significant operating costs and is highly undesirable. Spiral groove design practice goes back to US Patent No. 3,109,658 wherein two opposing spiral grooves pump oil against each other to develop a liquid barrier capable of sealing a gas. Such an arrangement is limited in pressure as well as speed capability, as is inherent in the use of liquid forces to seal gas. Another known arrangement is shown in US Patent No. 3,499,653. This interface design with partial spiral grooves relies heavily on hydrostatic effects. The interface gap is designed with a tapered shape which is narrower at the non-grooved end and wider at the spiral grooves. The effect of the spiral grooves and therefore the hydrodynamic forces are suppressed since spiral groove pumping becomes less effective across the wider gaps. This likewise affects the stability of the seal and limits its top pressure and speed capability. A further known arrangement is shown by US Patent No. 4,212,475. Here the spiral groove itself attempts to act both as a hydrostatic as well as a hydrodynamic pattern and is used to eliminate the need for the tapered shape of the gap so that a considerable degree of spiral groove hydrodynamic force can be applied to impart a self-aligning property to the sealing interface. The self-aligning property forces the sealing interface back towards a parallel position, regardless of whether deviations from parallel position during seal operation occur in radial or tangential directions. This resulted in improvement stability and increased performance limits in terms of pressure and speed. While the known fluid seals as briefly summarized above have attempted to provide both hydrodynamic and hydrostatic sealing properties, nevertheless the known seals have been deficient with respect to their ability to optimize the combination of these hydrostatic and hydrodynamic properties so as to provide desirable hydrostatic properties which facilitate starting and stopping of seals while effectively minimizing or avoiding direct face contact and minimizing face loading between the seals so that the assembly can be started up with minimal friction to avoid severe frictional power requirements and direct frictional wear between the faces, and at the same time provide desirable hydrodynamic properties between the relatively-rotatable seal f ces under a wide range of operating conditions particularly those involving high speed and high pressure. Accordingly, it is an object of this invention to provide an improved fluid seal of the type employing a grooved pattern on one of the opposed seal faces, which improved seal provides a more optimized combination of hydrodynamic and hydrostatic sealing characteristics so as to permit improved seal performance under a significantly greater range of operating conditions, including operating conditions ranging from start-up to conditions involving high speed and high pressure. In the improved seal arrangement of the present invention, the groove pattern (which is typically defined on only one of the seal faces) includes first and second groove arrangements which communicate with one another, one being significantly deeper than the other, whereby the deeper arrangement is particularly effective for providing the desired hydrodynamic characteristics, whereas the shallower groove arrangement is more effective for providing the desired hydrostatic characteristics. At the same time, these arrangements are positioned such that the shallower arrangement is interposed generally between the deeper groove arrangement and a non-grooved annular land or dam which effectively separates the groove pattern from the low pressure side of the seal, whereby desirable hydrostatic and hydrodynamic seal properties can both be obtained but at the same time leakage of sealing fluid (for example, a gas) across the dam to the low pressure side is minimized so as to improve the performance efficiency of the seal. In addition, this optimization of the seal properties and performance characteristics is further improved by optimizing the groove pattern or configuration relative to the surrounding lands defined on the seal face so that the fluid film which is created between the opposed seal faces provides a more uniform pressure distribution and sealing characteristics while minimizing distortion of the seal face, which in turn assists in optimizing the seal performance with minimum width of gap between the opposed seal faces while still avoiding or minimizing direct contact and frictional wear between the opposed seal faces. In the improved seal of this invention, as briefly discussed above, the groove pattern includes the deep groove arrangement which is defined by a circumfer- entially arranged series of grooves which angle circum- ferentially and radially inwardly from the surrounding high-pressure side of the seal, which angled grooves may be of spiral, circular or straight configuration. These angled grooves are relatively deep and project only partway across the seal face. The angled deep grooves, at their radially inner ends, communicate with the shallow groove arrangement which is positioned radially inwardly of the deeper groove arrangement, but which is separated from the low pressure side of the seal .by the intermediate non-grooved annular land or dam. This shallow groove arrangement has a depth which is a small fraction of the deeper groove arrangement and is effective for creating a hydrostatic force between the opposed sealing faces substantially in the central region thereof as defined between the radially outer and inner boundaries of the seal interface. In a preferred embodiment, all grooves associated with the groove pattern are formed such that the sides of adjacent grooves extend generally in parallelism with one another so that the intermediate land area between adjacent grooves maintains a substantially constant width, even adjacent the radially inner ends of the grooves, to maximize squeeze film effects in the fluid which flows over these lands and thus enhance the thrust bearing support these lands provide for avoidance of seal face contact at or near the full speed rotation. The improved seal arrangement, as aforesaid, also preferably forms the shallow groove arrangement by a circumferentially-spaced series of shallow grooves which are contiguous with and project radially inwardly from the inner ends of the angled deep grooves, which shallow grooves terminate at the dam. These shallow grooves provide improved hydrostatic seal characteristics in the central seal face region, and angle radially inwardly at a smaller angle (which angle is zero in a preferred embodiment) relative to the radial direction than do the deep grooves so as to increase the land area between the adjacent shallow grooves, particularly adjacent the radially inner ends of the shallow grooves, to provide a better fluid squeeze film effect between the opposed seal faces during high speed rotation. Further improvement to the hydrostaticaUy effective relatively shallow inner groove pattern is aimed at reduction and elimination of any seal face distortions that might occur as a result of circumferential non- uniformity of hydrostatic pressure fields as these form above groove and land regions of the shallow groove pattern at conditions at or near to the zero speed of rotation. This improvement is a narrow and shallow circumferential groove interconnecting inner ends of the shallow inner groove pattern. Such a shallow circumferential groove acts to equalize pressure field non-uniformities circumferentially, as a result suppressing any face distortions and producing a uniform face separation with no or only minimal face-to-face contact even at extremely low magnitudes of separation between the faces. Other objects and purposes of the invention will be apparent to persons familiar with seals of this general type upon reading the following specification and inspecting the accompanying drawings. BRIEF DESCRIPTION OF THE DRAWINGS Figure 1 is a fragmentary central sectional view illustrating a generally conventional fluid face seal arrangement, such as a grooved face seal, associated with a rotating shaft. Figure 2 is a view taken generally along line 2-2 in Figure 1 and illustrating the groove pattern associated with a face of the rotating seal ring according to an embodiment of this invention. Figure 3 is a fragmentary enlargement of a part of Figure 2 so as to illustrate the groove pattern in greater detail. Figure 4 is a fragmentary sectional view taken substantially along line 4-4 in Figure 3. Figures 5 and 6 are views which correspond respectively to Figures 3 and 4 but illustrate a variation thereof. Figure 7 is a view similar to Figure 2 but showing a further variation of the inner groove pattern. Figure 8 is a view similar to Figure 2 but showing still a further and preferred variation of the inner groove pattern. Figure 9 is a fragmentary enlargement of a part of Figure 8 so as to illustrate the groove pattern in greater detail. Figure 10 is a fragmentary sectional view taken substantially along line 10-10 in Figure 9. Figures 11 and 12 are views which correspond respectively to Figures 9 and 10 but illustrate a variation thereof. Certain terminology will be used in the following description for convenience in reference only, and will not be limiting. For example, the words "upwardly", "downwardly", "rightwardly" and "leftwardly" will refer to directions in the drawings to which reference is made. The words "inwardly" and "outwardly" will refer to directions toward and away from, respectively, the geometric center of the assembly and designated parts thereof. Said terminology will include the words specifically mentioned, derivatives thereof, and words of similar import. DETAILED DESCRIPTION Referring to Figure 1, there is shown a typical grooved face seal assembly 10 and its environment. This environment comprises a housing 11 and a rotatable shaft 12 extending through said housing. The seal assembly 10 is applied to seal a fluid (such as a pressurized gas) within the annular space 13 and to restrict its escape into the environment at 14. Basic components of the seal assembly includes an annular, axially movable but non-rotatable sealing ring 16 having a radially extending flat face 17 in opposed sealing relationship with a radially extending flat face 18 of an annular rotatable sealing ring 19 which is non-rotatably mounted on the shaft 12. Ring 19 normally rotates in the direction of the arrow (Figure 2) . The sealing ring 16 is located within cavity 21 of housing 11 and held substantially concentric to rotatable sealing ring 19. Between housing 11 and the sealing ring 16 is a conventional anti-rotation device (not shown) for preventing rotation of ring 16, as well as a plurality of springs 22 spaced equidistantly around the cavity 21. Springs 22 urge the sealing ring 16 toward engagement with the sealing ring 19. An O-ring 23 seals the space between the sealing ring 16 and the housing 11. The sealing ring 23 is retained in the axial position by a sleeve 24 which is concentric with and locked on the shaft 12, such as by locknut 25 threaded on shaft 12 as shown. O-ring seal 26 precludes leakage between the sealing ring 19 and the shaft 12. The radially extending face 18 of the sealing ring 19 and radially extending face 17 of sealing ring 16 are in sealing relationship, and define an annular contact area 27 therebetween, this being the seal interface. This seal interface 27 is defined by a surrounding outer diameter 28 of ring 19 and an inner diameter 29 of ring 16, these being the diameters exposed to the high and low pressure fluid respectively in the illustrated embodiment. In operation, a very narrow clearance is maintained between the seal faces 17-18, due to a fluid film as generated by a groove pattern (as described below) formed in the sealing face 18 of the sealing ring 19. Alternately, the groove pattern can be formed in the sealing face 17 of the sealing ring 16 and still be effective. Said narrow clearance is maintained by the fluid between the seal faces which prevents generation of friction heat and wear, but the narrow clearance limits outflow of the sealed fluid from the space 13 into the region 14. Referring now to Figure 2, there is illustrated the sealing face 18 of the sealing ring 19, which face has a groove arrangement 31 formed therein. This groove arrangement 31 includes a first groove pattern 32 which is positioned primarily on the radially outer portion of the face 18. This groove pattern 32 normally provides both hydrodynamic and hydrostatic force in the seal interface 27, although it is the primary source for generating hydrodynamic force and hence will herein often be referred to as the hydrodynamic region. The groove arrangement 31 also includes a second groove pattern 33 which is disposed generally radially inwardly of the groove pattern 32 and is positioned generally within the center radial region of the face 18, that is the region which is spaced radially from both of the interface diameters 28 and 29. This latter groove pattern or region 33 functions primarily to provide a hydrostatic force between the opposed seal faces 17-18 at conditions of near zero rotational speeds. The groove patterns 32 and 33 may be formed in the face 18 using conventional fabrications techniques. Considering first the hydrodynamic groove pattern 32, it is defined by a plurality of angled grooves 34 which are formed in the face 18 in substantially uniformly angularly spaced relationship therearound. These grooves 34 are all angled such that they open radially inwardly from the outer diameter 28 in such fashion that the grooves simultaneously project circumferentially and radially inwardly, and have an angled relationship with respect to both the circumferential and radial directions of the seal face. The angled groove 34, as represented by the centerline 36 thereof where the groove intersects the outer diameter 28, normally opens inwardly of the outer diameter 28 at an acute angle relative to a tangent to the outer diameter, which acute angle may be in the neighborhood of 15 degrees. Each angled groove 34 is defined by a pair of side or edge walls 37 and 38. The inner ends of grooves 34 terminate generally at shoulders or abutments 39 which are generally rather abrupt and are defined about a radius designated R4 as generated about the center point 0 of the face ring, this radius R4 defining the groove diameter for the grooves 34 of the outer groove pattern 32. The opposed side walls 37-38 defining each of the grooves 34 generally and preferably slightly converge relative to one another as the groove angles radially inwardly. These side walls 37-38 may assume different configurations including straight lines, circular arcs or spiral profiles. When the sides 37-38 are defined as circular arcs or spirals, then the side wall 37 is of a convex configuration, and the opposed wall 38 is of a concave configuration. In the illustrated and preferred embodiment, the opposed sides 37-38 are of circular configuration, but are preferably generated about different radii having different centerpoints. For example, and referring to Figure 3, the concave side 38' of groove 34' is generated about a radius designated R5 having a first centerpoint Cl, and the convex side 37 of the adjacent groove 34 is generated about a radius R6 which is swung about the same centerpoint Cl, whereby the radius R6 exceeds the radius R5 by the perpendicular distance which separates the edges 37 and 38' of the adjacent pair of grooves 34 and 34*. This results in the flat or land 41 as defined between the edges 37 and 38' being of constant transverse width as it angles radially inwardly toward the center of the ring. In similar fashion, the concave edge 38 of groove 34 is also generated about the radius R5, which radius is now generated about a second centerpoint C2 spaced from the first centerpoint, and similarly the convex edge 37" of the next groove 34" is generated about the radius R6 which is also swung about the second centerpoint C2, whereby the land 41 between the edges 38 and 37" again has a constant transverse dimension therebetween as this land angles inwardly toward the center of the ring. The two centerpoints themselves are located on a circle which is concentric about the center O, and all of the grooves 34 are generated in a similar fashion. Each of the grooves 34 is of substantial depth relative to the groove pattern 33, which depth is illustrated by the generally flat bottom wall 42 of the groove 34 as illustrated by Figure 4. The groove depth in a preferred embodiment as illustrated by solid line 42 is substantially uniform throughout the length of the groove 34. However, the groove 34 can be of a tapered configuration throughout its length so that the depth varies throughout the length, such being diagrammatically illustrated by the variations indicated by dotted lines designated at 42a and 42b in Figure 4. As to the groove bottom wall designated at 42a, this groove has its maximum depth at the radially outer end, and its minimum depth at the radially inner end, although the depth at the radially inner end is still sufficient so as to result in a significant shoulder or step 39 at the radially inner end thereof. Further, with this variation designated at 42a, the average depth of the groove substantially midway throughout the length thereof preferably substantially corresponds to the uniform depth of the groove as indicated by the bottom wall 42. In this tapered variation designated at 42a, the groove depth at the radially outer end is sufficiently deep as to minimize the hydrodynamic force effect. This latter effect is more pronounced adjacent the radial inner end of the groove 34 in the region of the face ring which is more centrally located, and is believed more effective for applying greater pressure against the central portion of the face ring so as to resist the typical thermal distortion (i.e. crowning) which occurs in operation. As to the other tapered variation of the groove 34 as illustrated by the bottom 42b in Figure 4, in this variation the groove 34 is shallowest at its radially outer end and deepest at its radially inner end adjacent the shoulder or step 39. The shallowness of the groove at the radially outer end is such as to effectively starve this region of the groove of fluid, and again minimizes the hydrodynamic effect in this region so that greater pressure is developed closer to the center of the face ring so as to tend to provide increased pressure resistance against the distortion of the ring which normally occurs during operation. Considering now the hydrostatic groove region or pattern 33, this groove pattern is disposed generally radially inwardly of the hydrodynamic groove pattern 32 and is generally of significantly shallower depth so as to prevent it from having any significant hydrodynamic effect. This hydrostatic groove pattern 33 also includes a plurality of angled grooves 44 which are formed in the central radial region of the seal face 18, with these grooves 44 being uniformly angularly disposed around the seal face. The grooves 44 are contiguous with and project radially inwardly from the radially inner ends of the angled grooves 34, with grooves 44 being angled such that they simultaneously project circumferentially and radially inwardly from the diameter which defines the steps 39. The grooves 44 thus have an angled relationship with respect to both the circumferential and radial directions of the seal face 18. The angled grooves 44, in the embodiment illustrated by Figures 2 and 3, are angled in the reverse circumferential direction from the outer grooves 34, whereby a centerline 46 of the groove 44 intersects a radial line 45 at an acute angle α which, in the embodiment illustrated by Figure 3, is about 45°. Each groove 44 is defined between opposed edge or side walls 47-48, with the radially inner ends of grooves 44 terminating at abrupt shoulders or abutments 49, the latter being defined generally on a radius R3 generated about the centerpoint 0, this latter radius defining the inner groove diameter. The side walls of adjacent grooves, such as the adjacent side walls 48 and 47", define therebetween a flat land 51 which is an extension of the flat land 41 defined between the adjacent grooves 34 and 34". This land 51 projects radially inwardly and connects to a further annular flat land 53, the latter being defined between the inner face diameter 29 (i.e., radius R2) and the radius R3. This land 53 is free of grooves and functions as a dam to significantly restrict flow of sealing fluid thereacross into the low pressure region defined at the diameter 29. The inclined orientation (i.e. angle α) of the grooves 44 relative to the radial direction 45 is selected so that the grooves have a significant radially-directed flow component and hence these grooves 44 have a less steep angle relative to radial direction 45 than do the grooves 34. More specifically, the inclination angle α is preferably selected so as to be within the range of about ± 45° relative to the radial direction 45. This maximizes the area of the land 51 as measured transversely between the side walls (for example the side walls 48 and 47") of adjacent grooves, thereby permitting creation of a more effective land 51 for trapping pressure fluid therebetween so as to create a thrust bearing effect at times of operation at relatively high speeds of rotation. That is, a squeeze film effect is created at the lands 51 which is effective for resisting changes in gap width due to high speed vibrations or oscillations. In fact, in the preferred embodiment, the directly adjacent sides of adjacent grooves 44 and 44", such as the sides 48 and 47", preferably extend in parallel relationship to one another. Similarly, the adjacent sides 47 and 48' of the next adjacent pair of grooves 44 and 44' also preferably extend in parallel relationship with one another. This necessarily results in the opposed sides 47-48 of each groove being of a slightly converging relationship as they project radially inwardly, and results in the transverse width of the land 51 between each adjacent pair of grooves 44 being substantially constant and hence of maximum width as the land project radially inwardly, and maximizes the width of land 51 at the mouth thereof where the land meets the groove diameter defined by the radius R3. The pair of side walls 47-48 which cooperate to define each groove 44 may be straight for manufacturing convenience, or may be generated with spiral or circular profiles, which circular profiles will preferably be generated in a manner similar to the circular profiles of the side walls 37-38 for the grooves 34 as explained above. As to the depth of the grooves 34 and 44, the grooves 34 have a depth which is several times greater than the depth of grooves 44 and which is preferably in the range of about five to about ten times the depth of the grooves 44. More specifically, the deep grooves 34 will normally have an average depth of from about .0001 inch to about .001 inch although a more practical maximum depth is believed to be about .0005 inch with a depth of from about .0001 inch to about .0003 inch being preferred, and the shallow grooves 44 will normally have a depth of from about .00001 inch to about .00008 inch with a depth of about .00002 inch to about .00005 inch being preferred. As to the radial positional relationships between the deep grooves 34, the shallow grooves 44 and the land 53, these relationships are determined relative to the radial width ΔR of the seal interface 27 as measured between the high pressure radius 28 (radius Rl) and the low pressure radius 29 (radius R2) . The hydrodynamic groove pattern 32 will normally occupy about the radially outer one-third of the radial dimension ΔR, the hydrostatic groove pattern 33 will normally occupy about the middle one-third of the radial distance ΔR, and the dam 53 will normally occupy about the radially inner one-third of the distance ΔR. However, the shallow groove pattern 33 can be either radially narrowed or widened as desired so that it will occupy anywhere from the middle one-quarter to about the middle one-half of the width ΔR so as to maximize the fluid pressures in this central region of the face ring so as to provide increased resistance against the conventional distortion and crowning which normally occurs in operation, such as due to thermal effects. In operation, the high pressure fluid surrounding the outer diameter 28 enters into the deep grooves 34 and the shallow grooves 44, but is then restricted from further radial inward flow by the land or dam 53. This pressure fluid within the grooves creates sufficient hydrostatic pressure to effect significant unloading of force or a small separation between the opposed seal faces 17-18 throughout the interface area 27, there thus being created a hydrostatic force between the opposed seal faces. A small but controlled amount of the sealing fluid will pass over the dam 53 to the low pressure side 29 of the seal. The presence of this hydrostatic force, however, greatly minimizes frictional contact between the opposed sealing faces, and greatly facilitates start-up of the seal both by reducing the stresses imposed on seal structural elements that transmit the seal face friction to the seal housing 11 or shaft 12, and by significantly reducing or eliminating direct frictional contact between the opposed relatively rotatable seal faces 17-18 as rotation is initiated. As the seal arrangement operates at higher rotational speed, the high pressure fluid enters the deep grooves 34, and is effectively pumped out over the shallow groove region 33 and the lands 41 to create and increase the dimension of the gap or clearance between the opposed faces 17-18 so as to permit relative high speed rotation between the faces while effectively avoiding or greatly minimizing any direct frictional contact therebetween. The fluid pressure profile (i.e. hydrodynamic force) created between these opposed faces under this later condition, however, is subject to its highest pressure in the vicinity of the steps 39 disposed circumferentially between the adjacent groove regions 32 and 33. For this pressure fluid to escape to the lower pressure side 29 of the seal, it must first flow over the shallow groove region 33 which creates significant flow resistance, and in addition must also flow across the relatively wide dam or land 53. This significant radial extent as defined by the land 53 and the shallow groove region 33 severely impedes the escape of the sealing fluid to the low pressure side of the system, and permits the development of a desirable hydrodynamic force while at the same time providing for controlled and acceptable rates of sealing fluid leakage to the low pressure side. Referencing now Figures 5 and 6, there is illustrated a variation of the invention. In this variation, the seal face is again provided with a groove pattern which incorporates a radially outer series of angled deep grooves 34 contiguous with a radially intermediate series of angled shallow grooves 44 constructed and positioned in a manner substantially identical to that illustrated by Figures 3 and 4. In this variation of Figures 5-6, however, the hydrostatic groove pattern 33' additionally includes a shallow annular groove 61 formed in the seal face 18 in concentric relationship to the centerpoint O, which annular groove 61 is formed at and continuously connects the radially inner ends of the shallow grooves 44. This annular groove 61 has an inner annular wall 62 which effectively defines the radius R3 which is the radially inner groove diameter, whereby the non-grooved land 53 projects radially inwardly from this boundary wall 62. The groove 61 is generally of uniform depth circumferentially throughout, which depth preferably substantially identically corresponds to the depth of the shallow grooves 44. The groove 61 is preferably of rather narrow radial width, which radial width as defined between the radially inner boundary wall 62 and the radially outer boundary wall 63 will typically be in the neighborhood of about 1/16 inch or less. When the hydrostatic groove pattern 33* includes therein the shallow annular groove 61 as shown in Figures 5-6, this effectively equalizes pressures circumferentially in the vicinity of annular groove 61. Thus, the fluid film created between adjacent grooves in the presence of the lands 41 and 51 can be maintained at a substantially uniform magnitude circumferentially. Since the pressure fluid occupies not only the grooves 34, 44 but also the annular groove 61, this minimizes distortions of both sealing faces in circumferential directions and permits therefore smaller hydrostatic face separation with smaller leakage while avoiding or minimizing face contact when at or near zero rotational speed. Since the high pressure fluid exists continuously throughout the annular groove 61 in a hydrostatic condition, the pressure drop of the fluid as it escapes radially across the land 53 to the low pressure side 29 creates uniform pressure gradients which extend circumferentially of the seal ring, thereby also minimizing distortion circumferentially of the seal ring in the area of the land 53, and hence minimizing the tendency of the seal ring to deform into a wavy circumferentially-extending configuration. However, under a hydrodynamic condition, the entire shallow groove region 33 effectively acts as an extension of the land 53 to provide for controlled and minimal leakage of sealing fluid thereacross during operation near to or at full speed. While the grooves 44 illustrated by Figure 3 and as described above are reversely angled relative to the grooves 34, the grooves 44 can also be angled in the same circumferential direction as the grooves 34 as illustrated by Figure 7. In this latter variation, the inner grooves 44 still preferably have the centerlines 46 thereof intersecting the radial direction 45 at an angle α which is preferably no greater than about 45°, with the inclination of the centerlines of grooves 44 preferably being positioned so as to lie within the extremes illustrated by the positions of Figures 3 and 7. In the positional relationship wherein the grooves 44 angle in the same circumferential direction as the grooves 34, such as illustrated by Figure 7, the inner shallow grooves 44 will be angled radially inwardly more sharply than the outer deep grooves 34, whereby the side walls 37-38 where they join to the side walls 47-48 effectively define a discontinuity in curvature. That is, the abutting side walls 37, 47 and 38, 48 do not define a continuous curvature or straight line, although any discontinuity can obviously be rounded to facilitate the merger of the side walls. By providing the shallow grooves 44 with a greater radially-inwardly directed inclination, the grooves 44 permit the formation of more effective land areas 51 therebetween so as to provide for an improved squeeze film effect during high speed rotation, and at the same time the retained circumferential angularity of the grooves 44 is believed to permit at least some minimal hydrodynamic force generation in the gap between the opposed seal faces 17-18 when seal rotation occurs at low speed, such as during start-up, thereby improving the fluid seal in the central radial region between the seal faces 17-18 prior to the gap being widened due to the full effectiveness of the hydrodynamic force generated by the outer grooves 34, which latter grooves become fully effective at higher speeds. More specifically, by providing the shallow grooves 44 and deep grooves 34 with different angularities as discussed above, this enables the width A of the land 51, as measured perpendicularly between the side edges of adjacent grooves 44, to be maximized, and made greater than the width B of the land 41 as measured perpendicularly between the side edges of the adjacent grooves 34. The angularities are preferably selected so that the land width A is equal to or greater than about 1.3 times the land width B. Further, by inclining the shallow grooves 44 in the reverse circumferential direction as indicated by the embodiment of Figure 3, such is believed to provide some hydrodynamic force generation in the small gap between the opposed seal faces when low speed reverse rotation occurs, such as when accidental back pressure upon shut down causes reversal of rotation in a compressor upon shut down. Since such reversal of rotation in most use applications occurs only for a relatively short time and normally involves only lower rotational speeds, the reverse angled orientation of the shallow inner grooves 44 is believed to provide generation of at least minimal hydrodynamic force to prevent or at least minimize any significant direct contact between the opposed seal faces. This Figure 7 variation is also preferably provided with the annular groove 61 in the same manner as shown in Figures 5-6. While Figures 2 and 7 illustrate the shallow grooves 44 respectively reversely and forwardly angled relative to the deep grooves 34, reference is now made to Figures 8-10 wherein there is illustrated the preferred variation of the shallow grooves. In this variation (see Figures 8 and 9) , the shallow grooves 44 project directly radially inwardly from the radially inner ends of the angled grooves 34, with each groove 44 being positioned such that it has a substantially straight centerline 46 which extends lengthwise of the groove and which projects inwardly in intersecting relationship to the centerpoint O so as to constitute a radial line. Each groove 44 is also defined between opposed edge or side walls 47-48, both of which are preferably straight. The side walls of the adjacent grooves 44, such as the adjacent side walls 48 and 47" (Figure 9) , define therebetween the flat land 51 which is an extension of the flat land 41 defined between the adjacent grooves 34 and 34". This land 51 projects radially inwardly and connects to the further annular flat land 53. The radial orientation of the grooves 44 is highly desirable since this maximizes the area of the land 51 as measured transversely between the side walls (for example the side walls 48 and 47") of adjacent grooves, thereby permitting creation of a more effective land (since the minimum transverse dimension across these lands is relatively large) for trapping pressure fluid therebetween so as to create a thrust bearing effect at times of operation at relatively high speeds of rotation. That is, a squeeze film effect is created at the lands 51 which is effective for resisting changes in gap width due to high speed induced oscillations and vibrations. In fact, in this preferred variation of Figures 8-10, the directly adjacent sides of adjacent grooves 44 and 44", such as sides 48 and 47", preferably extend in parallel relationship to one another. Similarly the adjacent sides 47 and 48' of the next adjacent pair of grooves 44 and 44' also preferably extend in parallel relationship with one another. This necessarily results in the opposed sides 47-48 of each groove being of a slightly converging relationship as they project radially inwardly, and results in the transverse width of the land 51 between each adjacent pair of grooves 44 being substantially constant as the land 51 projects radially inwardly, and maximizes the width of land 51 at the mouth thereof, that is where the land meets the groove diameter defined by the radius R2. This parallel relationship between the sides of adjacent grooves 44, and the radial direction of these grooves 44, hence maximizes the area of land 51 to provide significantly improved squeeze film characteristics which are particularly important at or near high speed rotation. Referencing now Figures 11 and 12, there is illustrated a preferred variation of Figures 8-10 modified to include the shallow annular groove 61 for continuously connecting the radially inner ends of the shallow grooves 44. This annular groove 61 functions in the same manner as described above relative to Figures 5 and 6. While the invention illustrated and described herein has the high pressure region located at the outer diameter, which is the most commonly encountered -use condition, it will be appreciated that the groove pattern can extend radially from an inner diameter if the latter is the high pressure region. Although a particular preferred embodiment of the invention has been disclosed in detail for illustrative purposes, it will be recognized that variations or modifications of the disclosed apparatus, including the rearrangement of parts, lie within the scope of the present invention.

Claims: C l a i m s : The embodiments of the invention in which an exclusive property or privilege is claimed are defined as follows: 1. A fluid seal device cooperating between a housing and a rotatable shaft for creating a fluid seal between high and low pressure regions, said device comprising: a first seal ring mounted on the shaft for rotation therewith and a second seal ring disposed adjacent the first seal ring and being non-rotatably mounted relative to the housing; said first and second seal rings respectively defining thereon opposed first and second flat annular seal faces adapted to substantially axially abut to define an annular seal interface which extends radially between and is defined by radially outer and inner diameters which respectively communicate with said high and low pressure regions, one of said seal rings being axially movable and normally urged axially toward the other seal ring; a groove pattern formed in one of said seal faces for causing a thin film of pressurized fluid to be interposed between said seal faces to create a small clearance therebetween; said groove pattern including first groove means formed in said one seal face for creating a hydrodynamic fluid seal between the opposed seal faces when the first and second seal rings relatively rotate at high speed; said groove pattern including second groove means formed in said one seal face for creating a hydrostatic fluid seal between said opposed seal faces when said first and second seal rings are substantially stationary relative to one another; said first groove means including a plurality of first grooves disposed in generally uniformly angularly spaced relationship around said one seal face, said first grooves being angled so as to project circumferentially and radially inwardly from said high pressure diameter; said second groove means including a plurality of second grooves which are disposed in substantially uniformly angularly spaced relationship around said one seal face, each said second groove being contiguous with and projecting radially inwardly from a radially inner end of a respective one of said first grooves, said second grooves projecting radially inwardly from the radially inner ends of said first grooves, said second grooves being radially inwardly angled at a smaller angle relative to a radial direction than said first grooves, and said second grooves projecting radially inwardly so as to terminate at a groove diameter which is spaced radially outwardly of said low pressure diameter; said one seal face defining thereon an annular non- grooved flat land extending radially between said low pressure diameter and said groove diameter; and said first grooves having an average longitudinally- extending depth which is several times greater than the depth of said second grooves, and the inner end of said first grooves defining abrupt damlike steps where said first grooves connect to said second grooves. 2. A seal device according to Claim 1, wherein said second grooves are reversely angled in the circumferential direction relative to said first grooves. 3. A seal device according to Claim 2, wherein said second groove has a longitudinally extending centerline which intersects a radially projecting line of said one seal face at an angle of no greater than about 45°. 4. A seal device according to Claim 3, wherein said second groove means includes an annular groove concentrically formed in said one seal face and defining said groove diameter, said annular groove joining together and communicating with the radially inner ends of said plurality of second grooves, said annular groove having a depth of a magnitude similar to the depth of said second grooves. 5. A seal device according to Claim 1, said second groove having a longitudinally extending centerline which intersects a radially projecting line of said one seal face at an angle of no greater than about 45°. 6. A seal device according to Claim 5, wherein said second groove means includes an annular groove concentrically formed in said one seal face and defining said groove diameter, said annular groove joining together and communicating with the radially inner ends of said plurality of second grooves, said annular groove having a depth of a magnitude similar to the depth of said second grooves. 7. A seal device according to Claim 1, wherein said second groove means includes an annular groove concentrically formed in said one seal face and defining said groove diameter, said annular groove joining together and communicating with the radially inner ends of said plurality of second grooves, said annular groove having a depth of a magnitude similar to the depth of said second grooves. 8. A seal device according to Claim 7, wherein each said second groove is defined between first and second side edges which converge toward one another as said second groove projects radially inwardly, and wherein the first side edge of one said second groove and an adjacent second side edge of an adjacent said second groove extend in parallel relationship to define therebetween a flat land which is of constant width as it projects radially inwardly between adjacent said second grooves. 9. A seal device according to Claim 8, wherein each of said first grooves is defined between first and second side edges which converge relative to one another as they project circumferentially and radially inwardly from said high pressure diameter, and wherein the first side edge of one said first groove extends in parallelism with an adjacent said second side edge of an adjacent said first groove to define therebetween a flat land which is of constant width throughout the longitudinal extent of the first grooves. 10. A seal device according to Claim 1, wherein each said second groove is defined between first and second side edges which converge toward one another as said second groove projects radially inwardly, and wherein the first side edge of one said second groove and an adjacent second side edge of an adjacent said second groove extend in parallel relationship to define therebetween a flat land which is of constant width as it projects radially inwardly between adjacent said second grooves. 11. A seal device according to Claim 1, wherein each of said first grooves is defined between first and second side edges which converge relative to one another as they project circumferentially and radially inwardly from said high pressure diameter, and wherein the first side edge of one said first groove extends in parallelism with an adjacent said second side edge of an adjacent said first groove to define therebetween a flat land which is of constant width throughout the longitudinal extent of the first grooves. 12. A seal device according to Claim 1, wherein the sealing fluid is a gas. 13. A seal device according to Claim 1, wherein the first groove has an average depth in the range of about .0001 inch to about .0003 inch, and wherein the depth of said first groove is in the range of at least about five to about ten times the depth of the second groove. 14. A seal device according to Claim 13, wherein each of said first and second grooves are of uniform depth throughout substantially the respective longitudinal extent thereof. 15. A seal device according to Claim 13, wherein said second groove means includes an annular groove concentrically formed in said one seal face and defining said groove diameter, said annular groove joining together and communicating with the radially inner ends of said plurality of second grooves, said annular groove having a depth of a magnitude similar to the depth of said second grooves. 16. A seal device according to Claim 1, wherein said angle is zero and said second grooves project solely radially inwardly from the radially inner ends of said first grooves. 17. A seal device according to Claim 16, wherein the first groove has an average depth in the range of about .0001 inch to about .0003 inch, and wherein the depth of said first groove is in the range of at least about five to about ten times the depth of the second groove. 18. A seal device according to Claim 16, wherein each said second groove is defined between first and second straight side edges which converge toward one another as said second groove projects radially inwardly, and wherein the first side edge of one said second groove and an adjacent second side edge of an adjacent said second groove extend in parallel relationship to define therebetween a flat land which is of constant width as the land projects radially inwardly between adjacent said second grooves. 19. A seal device according to Claim 16, wherein said second groove means includes an annular groove concentrically formed in said one seal face, said annular groove joining together and communicating with the radially inner ends of said plurality of second grooves, said annular groove having a depth of a magnitude similar to the depth of said second grooves.

Inventors: SEDY JOSEF, SEDY, JOSEF
Applicants: DURAMETALLIC CORP, DURAMETALLIC CORPORATION
Application year: 1993
Main classification: F16J15
IPC codes: F16J15