Abstract:
The disclosed finger seals are designed to be operational under rotational velocity or a stationary condition. The contact surface of the finger seals is inclined in an axial direction. The rotational velocity of the bore does not affect the hydrodynamic lift and the finger seals can operate at any rotational speed, unlike prior finger seal where the hydrodynamic lift is generated by rotational velocity. Each finger seal is pressure balanced. The pressure chambers on the two sides of each finger seal are connected through the finger cutouts. The finger seal design is such that the fingers lift and move away from the bore surface in radial direction. In one form, each finger seal is designed with the specific required length to allow sufficient surface area for the hydrodynamic force such that the finger seals would be lifted from the bore surface at the design pressure.

Description:
RELATED APPLICATIONS 
     This application claims priority benefit of U.S. Provisional Ser. No. 61/183,846, filed Jun. 3, 2009. 
    
    
     BACKGROUND OF THE DISCLOSURE 
     a) Field of the Disclosure 
     This disclosure relates to the field of hydrodynamic bore seals, and in particular to layered finger seals of a novel design and arrangement. 
     b) Background Art 
     Some examples of dry running seals to seal against leakage of a compressed gas include finger seals, brush seals and labyrinth seals. Typically, these seals are used in turbines, or other high-temperature, high-speed applications where lubricated seals or positive seals fail. Some examples of finger seals are described in patents such as US 2008/0122183, U.S. Pat. No. 6,196,550, U.S. Pat. No. 5,108,116, and U.S. Pat. No. 6,736,401. 
     In patent application US 2008/0122183 is disclosed hydrodynamic sealing pads comprising one or two taper angles, one taper angle in the direction of shaft speed (tangential) and the other taper angle in the axial direction. The taper in the tangential direction to rotation allows for an increasing hydrodynamic lift due to increasing RPM of the shaft being sealed. The taper in the axial direction results in a desirable lifting force due only to differential pressure. The two taper angles could be combined or used separately to create a more desirable operating range of RPMs and pressures for a given application. 
     SUMMARY OF THE DISCLOSURE 
     The disclosed finger seals are designed to be operational under rotational velocity or stationary conditions. The contact surface of the finger seals is inclined in an axial direction. This incline causes a convergent leakage path between the finger foot surface and the bore. Therefore, the leakage flow passing through this gap exerts hydrodynamic lift on the finger and lifts the finger from the bore surface at design pressure. Since the slope is in the axial direction, the rotational velocity of the bore does not significantly affect the hydrodynamic lift and the finger seals can operate at any rotational speed, unlike prior finger seals where the hydrodynamic lift is generated by rotational velocity. 
     Each finger seal is pressure balanced. The pressure chambers on the two sides of each finger seal are connected through the finger cutouts. The finger seal design is such that the fingers lift and move away from the bore surface in a radial direction. Therefore, the angle between the finger seal foot and the bore is constant at any lifted distance. In each seal stack up, in one form, each finger seal is designed with the specific required length to allow sufficient surface area for the hydrodynamic force, such that the finger seals would be lifted from the bore surface at the design pressure. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a front view of a compliant seal rear plate, in one form. 
         FIG. 2  is a side view of the compliant sear rear plate shown in  FIG. 1 . 
         FIG. 3  is a detail view of the region  3  of  FIG. 2   
         FIG. 4  is a front view of a compliant seal middle plate, in one form. 
         FIG. 5  is a front view of a seal assembly, in one form. 
         FIG. 6  is a side view of the seal assembly of  FIG. 5 . 
         FIG. 7  is an isometric view of the embodiment shown in  FIG. 5 . 
         FIG. 8  is a detail view of a region  8  of  FIG. 7 . 
         FIG. 9  is a detail view of the region  9  of  FIG. 6 . 
         FIG. 10  is a front view of a compliant disk seal, in one form. 
         FIG. 11  is a side view of the compliant disk seal of  FIG. 10 . 
         FIG. 12  is a detail view of the region  12  of  FIG. 10 . 
         FIG. 13  is a detail view of the region  13  of  FIG. 11 . 
         FIG. 14  is a detail view of the region  14  of  FIG. 13   
         FIG. 15  is a front view of a compliant disk seal, in one form. 
         FIG. 16  is a side view of the compliant disk seal of  FIG. 15 . 
         FIG. 17  is a detail view of the region  17  of  FIG. 16 . 
         FIG. 18  is a detail view of the region  18  of  FIG. 17 . 
         FIG. 19  is a front view of a compliant disk seal, in another form. 
         FIG. 20  is a detail view of the region  20  of  FIG. 19   
     
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     The attached figures illustrate the design of an improved finger seal  20  and a preferred assembly embodiment. The seal  20 , in this embodiment, is configured to seal inside of a bore, however, a similarly designed seal could be designed to seal on a shaft by inverting the features. 
     To ease in understanding, an alphanumeric number system will be used comprising numeric references to groups, such as the number  24  regarding a pad, and an alpha suffix referring to particular elements with the group, such as individual pads  24   a  and  24   b.    
     The seal  20  comprises one or more sealing rings  26 , as shown in  FIG. 9 , that rest up against one or more backing plates (seal holder)  46 . Each of the sealing rings  26  is designed to have an “L” shape profile and a taper angle  22  on the sealing diameter  36  (see  FIG. 12 ) of the sealing ring  26 . There are a number of “pads”  24  on the outside diameter of each sealing ring  26  shown in the figures. The taper angle  24  is provided on the outside diameter of the seal pads  24  and each pad  24  tapers in the direction of the axis of rotation, or central axis of the circular seal. The principle of operation of the finger pads  24  is as follows: each pad  24  is supported by a spring structure  28 , comprising a plurality of spring members, which can be more easily understood by looking to  FIG. 12 . The seal pads, in one form, are connected to each other and cannot move independent of the adjacent pads. One key advantage of this arrangement is that the spring understructure is designed such that the seal pad moves radially (linearly) and its angle between the bore or shaft surface stays constant as the sealing pad moves (lifts off from the bore or shaft surface). This spring under structure in this particular embodiment has a symmetric spring support structure (unlike finger seals in the patents listed previously which are spiral beams in shape) using a simple beam spring  52  design. 
     Under a differential pressure, the seal experiences a leakage over each pad  24  and through the gaps  50  between the sides of the pads, which are ideally laser-cut with a gap thickness that is very small, such as 2 or 3 thousandths of an inch. In another embodiment, the gap is minimized as much as possible. Most of the leakage typically occurs between the bore  30  and the tapered pad faces  32  rather than between the pads  24 . A non-linear pressure drop occurs on the outside diameter of the pads. The difference in pressure between the outside  40  of the pads and the inside  42  of the pads  24  (see  FIG. 12 ) creates a net lifting force away from the bore  30  so as to preferably lift the pads  24  and reduce or prevent sliding friction under certain differential pressures, in this particular embodiment. In one form, the length of the L-lip  44  ( FIG. 9 ) is designed so as to provide the desired net lifting-force away from the bore  30 —the longer the lip, the higher the net lift force will be. There is a designed spring pre-load interference between the pads  24  and the bore  30 . The lifting force must come close to or overcome the preloading force in order to reduce or eliminate friction wear of the pads  24  on the bore  30 . The spring force, in one form, shall be sufficient to overcome the axial friction of the seal and bore. The taper  22  on the outside of the pads  24  is used in order to better maximize the pressure along the outside of the pad  24 . It can be shown that the taper produces a desirable differential pressure for lift, in contrast to prior, no taper designs. 
     In other words, the angled foot surface develops a converging channel for the leakage flow. The flow passing between the finger surface and bore or shaft pushes the finger away from the bore or shaft surface. This force is caused by the leakage passing underneath the finger and is independent of rotation speed. Therefore, the fingers will not touch the bore or shaft as there is a pressure differential between the high-pressure side  58  and low pressure side  56  of the seal. The fingers are I-shaped in one form to increase the surface area of the finger foot where the hydrodynamic lift is applied. In one form each finger seal is pressure balanced in the axial direction, the cutouts in the finger seal connect the two sides of the finger seals for pressure balancing. A lip feature  44  is designed at the inner angle of the L-shaped finger to seal the pressure balance chamber from the lower pressure side of the seal. 
     The cutouts in the finger seal connect the two sides of the finger seals for pressure balancing. As the fluid passes through the finger seals, the pressure decreases. Therefore, the pressure (hydrodynamic lift) applied to each finger foot is different than the other. Hence, each finger is designed with the required length to provide enough surface area for the lift force to lift the finger from the bore or shaft surface at the design pressure. 
       FIG. 19  shows a similar arrangement to that of  FIG. 10 , with the under spring structure substantially reversed in that a single brace  62 ′ extends between the main structure and the beam springs  52 ′ adjacent each pad  24 ′, and a plurality of braces  62 ′ extends between the beam springs  52 ′ and each pad  24 ′ on the ends thereof. 
     The embodiment shown in  FIG. 20  also utilizes gaps  50   b ′, which do not extend through the entire pad  24 ′ but rather function as a living hinge type structure to maintain relative position therebetween. 
     In operation, there is a net axial force that acts on the pads  24  due to differential pressure. The hydrodynamic force should also overcome this axial sliding friction between pads  24  and backup plates  38 , in order to allow the pads  24  to comply to the bore  30  if there is changing eccentricity of the bore  30  with respect to the seal  26 . To minimize this axial frictional force, the L-shaped lip  44  is designed to be as thin as reasonably possible for strength and machineability, and the gap between the lips and the outside diameter of the backup plates  38  is minimized. This design results in a net pressure-area that is minimized. Each finger seal is pressure balanced. The gaps between the seal ring and the spacer are connected through the spring cutouts in the seal structure. Therefore, the two sides of the seal are at the same pressure and the seal is pressure balanced. To achieve this, the thickness of the finger seal is smaller than the space between the two adjacent spacers, such that the finger has some clearance, gaps  53  and  54 , from the two spacers, plates  38   a  and  38   b  when the seal ring  26   b  is sandwiched in between. Fluid fills these gaps and since the gaps on the two sides are connected through the cutouts in the finger seal structure, they are the same pressure. Therefore, little axial load is applied on the finger seal structure. The upstream gap  53  and downstream gap  54  are shown in  FIG. 9 . To prevent leakage from the downstream gap  54  to the low-pressure region  56  from the high-pressure region  58 , a lip  60  is provided on the finger seal that rests on the backup plate  38   a  and effectively seals the gap  54  chamber from the downstream low-pressure region  56 . The compliant spring loaded sealing rings are pressure balanced axially to allow the pads  24  to slide in and out more freely, in order to allow them to be more compliant under eccentricities (such as due to thermal expansion, centrifugal expansion of the bore, startup and shutdowns, vibration, or out of tolerance parts rotating, to list a few examples). The underlying spring structure  28  under the pads  24  is designed to be stiff enough so as to be able to overcome the axial frictional forces, so as to allow for the compliance under eccentricities, but the underling spring structure  28  is also designed to be relatively thick and strong (or rather not very fragile) so as to be robust under high vibration applications, such as in automotive engines, for example. This design is therefore possibly less fragile than prior “finger seal” designs. 
     This particular embodiment has a taper angle  22  in the axial direction in order to create a hydrodynamic lift due only to leakage flow due to differential pressure. In another embodiment, another taper angle could be added in the circumferential direction so as to create a compound angle in two planar directions. This could be designed to result in hydrodynamic lift from both differential pressure in the axial direction, as well as hydrodynamic lift caused from rotation of the bore with respect to the seal pads, as disclosed in US patent application 2008/0122183. That is, the lifting force could be designed for specific differential pressures and specific ranges of revolutions per minute (RPMs) of the bore  30  being sealed. Particularly with sealing against liquids, this rotational hydrodynamic lift is not negligible. Another embodiment could be such that the seal pad  24  is designed to have only a circumferential hydrodynamic lifting force and zero lifting force due to pressure differential/axial leakage flow. Typically, this design would be good for lower differential pressures and higher rotational speeds. 
     The type of material used for this compliant seal depends on the pressures, temperatures, and expected frictional forces on the pad faces, as well as pressure loads acting on the backup plates. Metals, such as stainless steel, spring steel etc., could be utilized. Surface coatings could be applied to the materials in order to reduce friction, and material hardness or composition could be changed to provide desirable temperature, strength or friction properties. 
     In prior art finger seals listed previously, the seals generally comprise a supporting spring structure that is spiral in shape. As the spring deflects, the angle that the finger&#39;s pad makes with the sealing surface tilts and changes, causing a change (increase) in the leakage gaps under the finger pads  24 . For the present invention where there is not necessarily a designed tilt in the tangential direction for conditions where we do not desire a hydrodynamic lift force due to rotation but only due to differential pressure, this tilting of the pad  24  due to the spring deflection under preload or lift force is not desirable and can increase leakage. 
     In the present invention, a sturdy pad and spring design is disclosed that is robust under severe operating conditions. This configuration is unique to finger seal designs as this configuration utilizes large pads  24  and a strong spring structure  28 , among other properties. A robust, strong design using the prior finger seal approach would require the finger beams to be short and the pads long, accentuating the tilting effect due to lifting of the pads, therefore creating larger gaps under the pads and higher leakage. As a result in the disclosed embodiments, a symmetric design is utilized to substantially eliminate the tilting effect. In the disclosed embodiments, there is no taper angle in the transverse direction of the pads. All leakage should therefore be through the side gaps, under the pads from hydrodynamic lift, or surface inconsistencies under the pads and bore, not due to tilting effects of the pads or taper in the transverse direction. 
     In the assembly  FIG. 9  of the disclosed embodiments, shown are two sealing pad rings  26  and two backup support plates  38 . The pads  24  are staggered in order to create a circuitous path and maximize pressure drop from leakage that occurs between the pads  24 . Layering the pads as shown also reduces the forces acting on each sealing ring  26 , so any number of sealing rings  26  and backup plates  38  could be stacked together. The benefits of this arrangement could be to reduce the forces or requirement for robustness of each layer (sealing ring/backup plate combination), or increasing the number of layers can decrease leakage by creating a more circuitous path for the leakage, for example. In one form, shown in  FIG. 8 , adjacent pads  24   a  and  24   b  having offset gaps such that adjacent gaps  50   a  and  50   b  defining a pad  24   a  do not overlap the gaps  50   c  or  50   d  in the adjacent pad  24   b  of adjacent sealing rings. 
     In one form, all plates and sealing rings are attached to the seal holder  46 . Bolts may be passed through voids  48 . For example, voids  48   a  (in the outer backing plate  38   a  as shown in  FIG. 8 ),  48   b  in each compliant disk seal (sealing ring  26  of  FIG. 10 ),  48   c  (in the middle plate  38   b  of  FIG. 4 ),  48   d  (in the rear plate  46  of  FIG. 1 ) etc. could be aligned when assembled. 
     In one form, the outer backing plate  38   a  is thicker than the middle backing plate  38   b  resulting from analysis that a higher differential pressure occurs on the last seal and therefore there is a larger axial force acting on the outer backing plate  38   a.    
     While the present invention is illustrated by description of several embodiments and while the illustrative embodiments are described in detail, it is not the intention of the applicants to restrict or in any way limit the scope of the appended claims to such detail. Additional advantages and modifications within the scope of the appended claims will readily appear to those sufficed in the art. The invention in its broader aspects is therefore not limited to the specific details, representative apparatus and methods, and illustrative examples shown and described. Accordingly, departures may be made from such details without departing from the spirit or scope of applicants&#39; general concept.