Abstract:
A self-adjusting non-contact seal for sealing the circumferential gap between a first machine component and a second machine component includes structure which undergoes wear in the event of inadvertent contact with one of the machine components in such a way as to allow a reset of its radial distance from such machine component, compared to initial installation tolerances, while minimizing leakage.

Description:
RELATED APPLICATIONS 
     This application is a continuation-in-part application of U.S. patent application Ser. No. 13/009,155 filed Jan. 19, 2011, which is a continuation-in-part application of U.S. patent application Ser. No. 11/953,099 filed Dec. 10, 2007 and now U.S. Pat. No. 7,896,352, which is a continuation-in-part application of U.S. patent application Ser. No. 11/669,454 filed Jan. 31, 2007 and now U.S. Pat. No. 7,410,173, which is a continuation-in-part application of U.S. patent application Ser. No. 11/226,836 filed Sep. 14, 2005 and now U.S. Pat. No. 7,182,345, which is a continuation of U.S. patent application Ser. No. 10/832,053 filed Apr. 26, 2004, now abandoned, which claims the benefit of U.S. Provisional Application Ser. No. 60/466,979 filed May 1, 2003, under 35 U.S.C. §119(e) for all commonly disclosed subject matter. U.S. Provisional Application Ser. No. 60/466,979 is expressly incorporated herein by reference in its entirety to form part of the present disclosure. 
    
    
     FIELD OF THE INVENTION 
     This invention relates to seals for sealing a circumferential gap between two machine components that are relatively rotatable with respect to each other, and, more particularly, to a non-contact seal having at least one shoe formed with a nozzle and two or more projections which are effective to reset the radial position of the at least on shoe relative to one of the machine components in the event of inadvertent contact and wear of such projections in order to substantially maintain a non-contact seal of the circumferential gap between such machine components. 
     BACKGROUND OF THE INVENTION 
     Turbomachinery, such as gas turbine engines employed in aircraft, currently is dependent on either labyrinth (see  FIGS. 1A-1E ), brush (see  FIGS. 2A and 2B ) or carbon seals for critical applications. Labyrinth seals provide adequate sealing but they are extremely dependent on maintaining radial tolerances at all points of engine operation. The radial clearance must take into account factors such as thermal expansion, shaft motion, tolerance stack-ups, rub tolerance, etc. Minimization of seal clearance is necessary to achieve maximum labyrinth seal effectiveness. In addition to increased leakage if clearances are not maintained, such as during a high-G maneuver, there is the potential for increases in engine vibration. Straight-thru labyrinth seals ( FIG. 1A ) are the most sensitive to clearance changes, with large clearances resulting in a carryover effect. Stepped labyrinth seals ( FIGS. 1B and 1C ) are very dependent on axial clearances, as well as radial clearances, which limits the number of teeth possible on each land. Pregrooved labyrinth seals ( FIG. 1D ) are dependent on both axial and radial clearances and must have an axial clearance less than twice the radial clearance to provide better leakage performance than stepped seals. 
     Other problems associated with labyrinth seals arise from heat generation due to knife edge to seal land rub, debris from hardcoated knife edges or seal lands being carried through engine passages, and excessive engine vibration. When seal teeth rub against seal lands, it is possible to generate large amounts of heat. This heat may result in reduced material strength and may even cause destruction of the seal if heat conducted to the rotor causes further interference. It is possible to reduce heat generation using abradable seal lands, but they must not be used in situations where rub debris will be carried by leakage air directly into critical areas such as bearing compartments or carbon seal rubbing contacts. This also holds true for hardcoats applied to knife edges to increase rub capability. Other difficulties with hardcoated knife edges include low cycle fatigue life debits, rub induced tooth-edge cracking, and the possibility of handling damage. Engine vibration is another factor to be considered when implementing labyrinth seals. As mentioned previously, this vibration can be caused by improper maintenance of radial clearances. However, it can also be affected by the spacing of labyrinth seal teeth, which can produce harmonics and result in high vibratory stresses. 
     In comparison to labyrinth seals, brush seals can offer very low leakage rates. For example, flow past a single stage brush seal is approximately equal to a four knife edge labyrinth seal at the same clearance. Brush seals are also not as dependent on radial clearances as labyrinth seals. Leakage equivalent to approximately a 2 to 3 mil gap is relatively constant over a large range of wire-rotor interferences. However, with current technology, all brush seals will eventually wear to line on line contact at the point of greatest initial interference. Great care must be taken to insure that the brush seal backing plate does not contact the rotor under any circumstances. It is possible for severing of the rotor to occur from this type of contact. In addition, undue wire wear may result in flow increases up to 800% and factors such as changes in extreme interference, temperature and pressure loads, and rubbing speeds must be taken into account when determining seal life. 
     The design for common brush seals, as seen in  FIGS. 2A and 2B , is usually an assembly of densely packed flexible wires sandwiched between a front plate and a back plate. The free ends of the wires protrude beyond the plates and contact a land or runner, with a small radial interference to form the seal. The wires are angled so that the free ends point in the same direction as the movement of the runner. Brush seals are sized to maintain a tight diametral fit throughout their useful life and to accommodate the greatest combination of axial movement of the brush relative to the rotor. 
     Brush seals may be used in a wide variety of applications. Although brush seal leakage generally decreases with exposure to repeated pressure loading, incorporating brush seals where extreme pressure loading occurs may cause a “blow over” condition resulting in permanent deformation of the seal wires. Brush seals have been used in sealing bearing compartments, however coke on the wires may result in accelerated wear and their leakage rate is higher than that of carbon seals. 
     One additional limitation of brush seals is that they are essentially uni-directional in operation, i.e., due to the angulation of the individual wires, such seals must be oriented in the direction of rotation of the moving element. Rotation of the moving element or rotor in the opposite direction, against the angulation of the wires, can result in permanent damage and/or failure of the seal. In the particular application of the seals required in the engine of a V-22 Osprey aircraft, for example, it is noted that during the blade fold wing stow operation, the engine rotates in reverse at very low rpm&#39;s. This is required to align rotor blades when stowing wings. This procedure is performed for creating a smaller aircraft footprint onboard an aircraft carrier. Reverse rotation of the engine would damage or create failure of brush seals such as those depicted in  FIGS. 2A and 2B . 
     Carbon seals are generally used to provide sealing of oil compartments and to protect oil systems from hot air and contamination. Their low leakage rates in comparison to labyrinth or brush seals are well-suited to this application but they are very sensitive to pressure balances and tolerance stack-ups. Pressure gradients at all operating conditions and especially at low power and idle conditions must be taken into account when considering the use of carbon seals. Carbon seals must be designed to have a sufficiently thick seal plate and the axial stack load path must pass through the plate as straight as possible to prevent coning of the seal. Another consideration with carbon seals is the potential for seepage, weepage or trapped oil. Provisions must be made to eliminate these conditions which may result in oil fire, rotor vibration, and severe corrosion. 
     According to the Advanced Subsonic Technology Initiative as presented at the NASA Lewis Research Center Seals Workshop, development of advanced sealing techniques to replace the current seal technologies described above will provide high returns on technology investments. These returns include reducing direct operating costs by up to 5%, reducing engine fuel burn up to 10%, reducing engine oxides of emission by over 50%, and reducing noise by 7 dB. For example, spending only a fraction of the costs needed to redesign and re-qualify complete compressor or turbine components on advanced seal development can achieve comparable performance improvements. In fact, engine studies have shown that by applying advanced seals techniques to just a few locations can result in reduction of 2.5% in SFC. 
     SUMMARY OF THE INVENTION 
     This invention is directed to a non-contact seal for sealing the circumferential gap between a first machine component such as a stator and a second machine component such as a rotor which has the capability of resetting the radial distance between itself and the rotor or stator in the event of inadvertent contact during operation. 
     In one presently preferred embodiment, the seal comprises at least one shoe extending along one of the rotor and stator in a position to create a non-contact seal therewith. At least one spring element is connected between one of the rotor and stator and the at least one shoe. The spring element(s) is flexible in the radial direction, but axially stiff so that it can function to assist in preventing roll over of the shoes with respect to the rotor or stator where it is located, thus maintaining an effective seal under pressure load. Preferably, stops are provided to limit the extent of radial motion of the shoe with respect to the rotor or stator. The spring elements deflect and move with the shoe(s) in response to the application of fluid pressure to the shoe(s) to create a primary seal, within design tolerances, along the gap between the machine components. 
     The shoe(s) is formed with a slot that receives at least two secondary sealing elements which are oriented side-by-side and are connected to one of the first and second machine components. The secondary sealing elements radially deflect and move with the shoe(s) in response to the application of fluid pressure applied to the shoe(s) to assist in the creation of a secondary seal along the gap between the machine components. Preferably, each of the secondary sealing elements comprises an annular plate, which, in alternative embodiments described below, may be formed with structure to enhance the radial deflection thereof The secondary sealing elements may be formed of sheet metal of varying thickness, or other suitable heat-resistant, flexible material. 
     In a further embodiment of this invention, the shoe(s) is formed with structure which undergoes wear in the event of inadvertent contact with one of the machine components in such a way as to allow the shoe(s) to reset its/their radial distance from such machine component compared to initial installation tolerances while minimizing leakage. 
    
    
     
       DESCRIPTION OF THE DRAWINGS 
       The structure, operation and advantages of this invention will become further apparent upon consideration of the following description, taken in conjunction with the accompanying drawings, wherein: 
         FIGS. 1A-1E  are schematic views of a number of prior art labyrinth seals; 
         FIGS. 2A and 2B  depict views of a prior art brush seal; 
         FIG. 3  is a perspective view of the hybrid seal of this invention; 
         FIG. 4  is an enlarged perspective view of a portion of the seal depicted in  FIG. 3 , with the sealing elements removed; 
         FIG. 5  is a cross sectional view of one of the spring elements and shoes; 
         FIG. 6  is a cross sectional view of the seal shown in  FIGS. 3 and 4  with the sealing elements inserted; 
         FIG. 7  is an elevational view of a portion of a sealing element; 
         FIG. 8  is an elevational view of an alternative embodiment of the sealing elements of this invention; 
         FIG. 9  is an elevational view of a further embodiment of the sealing elements herein; 
         FIG. 9A  is an enlarged view of the encircled portion of  FIG. 9 ; 
         FIG. 10  is an elevational view of still another embodiment of the sealing elements of this invention; 
         FIG. 10A  is an enlarged view of the encircled portion of  FIG. 10 ; 
         FIG. 11  is a view similar to  FIG. 4  but of a further embodiment of this invention; 
         FIG. 12  is a cross sectional view of the seal depicted in  FIG. 11  in position relative to a machine element; and 
         FIG. 12A  is an enlarged view of a portion of the shoe shown in  FIG. 12 . 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Referring now to  FIGS. 3-6 , one embodiment of a seal  10  according to this invention is illustrated which creates a non-contact seal of the circumferential gap  11  between two relatively rotating components, namely, a fixed stator  12  and a rotating rotor  14 . The seal  10  includes at least one, but preferably a number of circumferentially spaced shoes  16  which are located in a non-contact position along the exterior surface of the rotor  14 . Each shoe  16  is formed with a sealing surface  20  and a slot  22  extending radially inwardly toward the sealing surface  20 . For purposes of the present discussion, the term “axial” or “axially spaced” refers to a direction along the longitudinal axis of the stator  12  and rotor  14 , e.g. axis  18  shown in  FIGS. 3 and 6 , whereas “radial” refers to a direction perpendicular to the longitudinal axis  18 . 
     Under some operating conditions, particularly at higher pressures, it is desirable to limit the extent of radial movement of the shoes  16  with respect to the rotor  14  to maintain tolerances, e.g. the spacing between the shoes  16  and the facing surface of the rotor  14 . The seal  10  preferably includes a number of circumferentially spaced spring elements  24 , the details of one of which are best seen in  FIGS. 3 and 4 . Each spring element  24  is formed with an inner band  26 , and an outer band  28  radially outwardly spaced from the inner band  26 . One end of each of the bands  26  and  28  is mounted to or integrally formed with the stator  12  and the opposite end thereof is connected to a first stop  30 . The first stop  30  includes a leg  32  which is connected to or integrally formed with a shoe  16 , and has an arm  34  opposite the shoe  16  which may be received within a recess  36  formed in the stator  12 . The recess  36  has a shoulder  38  positioned in alignment with the arm  34  of the first stop  30 . 
     A second stop  40  is connected to or integrally formed with the shoe  16 . The second stop  40  is circumferentially spaced from the first stop  30  in a position near the point at which the inner and outer bands  26  and  28  connect to the stator  12 . The second stop  40  is formed with an arm  42  which may be received within a recess  44  in the stator  12 . The recess  44  has a shoulder  46  positioned in alignment with the arm  42  of second stop  40 . 
     Particularly when the seal  10  of this invention is used in applications such as gas turbine engines, aerodynamic forces are developed which apply a fluid pressure to the shoe  16  causing it to move radially inwardly toward the rotor  14 . The spring elements  24  deflect and move with the shoe  16  to create a primary seal of the circumferential gap  11  between the rotor  14  and stator  12 . The purpose of first and second stops  30  and  40  is to limit the extent of radially inward and outward movement of the shoe  16  with respect to the rotor  14 . A gap is provided between the arm  34  of first stop  30  and the shoulder  38 , and between the arm  42  of second stop  40  and shoulder  46 , such that the shoe  16  can move radially inwardly relative to the rotor  14 . Such inward motion is limited by engagement of the arms  34 ,  42  with shoulders  38  and  46 , respectively, to prevent the shoe  16  from contacting the rotor  14  or exceeding design tolerances for the gap between the two. The arms  34  and  42  also contact the stator  12  in the event the shoe  16  moves radially outwardly relative to the rotor  14 , to limit movement of the shoe  16  in that direction. 
     In one presently preferred embodiment illustrated in  FIGS. 6 and 7 , the seal  10  is also provided with a secondary seal comprising a stack of at least two secondary sealing elements  48  and  50 . Each of the secondary sealing elements  48  and  50  comprises an outer ring  52  formed with a number of circumferentially spaced openings  54 , a spring member  56  mounted within each opening  56  and a number of inner ring segments  58  each connected to at least one of the spring members  56 . The spring member  56  is depicted in  FIG. 7  as a series of connected loops, but it should be understood that spring member  56  could take essentially any other form, including parallel bands as in the spring elements  24 . The secondary sealing elements  48  and  50  are oriented side-by-side and positioned so that the inner ring segments  58  extend into the slot  22  formed in the shoe  16 . The spring members  56  deflect with the radial inward and outward movement of the shoe  16 , in response to the application of fluid pressure as noted above, and create a secondary seal of the gap  11  between the rotor  14  and stator  12 . As such, the secondary sealing elements  58  and  50  assist the spring elements  24  in maintaining the shoe  16  within design clearances relative to the rotor  14 . 
     In the presently preferred embodiment, the secondary sealing elements  48  and  50  are formed of sheet metal or other suitable flexible, heat-resistant material. The secondary sealing elements  48  and  50  may be affixed to one another, such as by welding, a mechanical connection or the like, or they may merely placed side-by-side within the slot  22  with no connection between them. In order to prevent fluid from passing through the openings  54  in the outer ring  52  of each secondary sealing element  48  and  50 , adjacent sealing elements are arranged so that the outer ring  52  of one secondary sealing element  48  covers the openings  54  in the adjacent secondary sealing element  50 . Although not required, a front plate  60  may be positioned between the spring element  24  and the secondary sealing element  48 , and a back plate  62  may be located adjacent to the secondary sealing element  50  for the purpose of assisting in supporting the secondary sealing elements  48 ,  50  in position within the shoe  16 . See  FIG. 5 . 
     Referring now to  FIGS. 8-10A , alternative embodiments of secondary sealing elements according to this invention are illustrated. Considering initially the embodiment shown in  FIG. 8 , a secondary sealing element  70  is shown which comprises an annular plate  72  having an inner edge  74  and an outer edge  76  that is spaced from the inner edge  74 . A slit  78  extends from the inner edge  74  to the outer edge  76  thus forming two ends  80  and  82  of the annular plate  72  which abut one another. 
     An alternative embodiment of a secondary sealing element  84  is depicted in  FIGS. 9 and 9A . In this embodiment, the secondary sealing element  84  comprises an annular plate  86  formed of the same material as annular plate  72 . The annular plate  86  has the same inner and outer edges  74 ,  76 , slit  78  and ends  80 ,  82  described above in connection with a discussion of  FIG. 8 , but with the addition of three cut-outs  88 ,  90  and  92 . The cut-outs  88  and  92  are preferably spaced about 90° from cut-out  90 , and about 90° from the slit  78 . As best seen in  FIG. 9A , the cut-out  88  comprises an elongated slot  94  that extends part way along and is radially inwardly spaced from the outer edge  76  of the annular plate  86 . A break line  96  is formed between the inner edge  74  of the annular plate  86  and the slot  94  defining opposed ends  98 ,  100  which abut one another. The break line  96  is preferably substantially perpendicular to the slot  94 . All of the cut-outs  88 - 92  are identical, and therefore cut-outs  90  and  92  have the same construction as described above with reference to cut-out  88 . 
     Referring now to  FIGS. 10 and 10A , a still further embodiment of a secondary sealing element  102  is shown. The secondary sealing element  102  comprises an annular plate  104  having an inner edge  106  and an outer edge  108  spaced from the inner edge  106 . Preferably, four deflection structure  110 ,  112 ,  114  and  116  are formed in the annular plate  102  at approximately 90° intervals about its circumference. As best seen in  FIG. 10A , each of the deflection structures  110 - 116  includes a number of circumferentially spaced inner recesses  118  that extend from the inner edge  106  toward the outer edge  108 , and a number of circumferentially spaces outer recesses  120  that extend from the outer edge  108  toward the inner edge  106 . The inner and outer recesses  118 ,  120  are circumferentially offset from one another such that each inner recess  118  is located in between two outer recesses  120 . 
     Each of the annular plates  72 ,  86  and  104  is preferably formed of sheet metal or other suitable flexible and heat-resistant material. Two or more sealing elements  70 ,  84  or  102  are preferably employed to assist in the formation of a secondary seal of the gap  11  between the rotor  14  and stator  12 . The secondary sealing elements  70 ,  84  or  102  are oriented side-by-side and positioned within the slot  22  formed in the shoe  16 , in the same manner as secondary sealing elements  48  and  50  depicted in  FIG. 6 . The secondary sealing elements  70 ,  84  and  102  may be affixed to one another, such as by welding, a mechanical connection or the like, or they may merely be placed within the slot  22  with no connection between them. The secondary sealing elements  70 ,  84  or  102  may be connected to one of the rotor  14  and stator  12 , and they may be positioned between a front plate  60  and back plate  62  as shown in  FIG. 6 . The slit  78  in secondary sealing element  70 , the slit  78  and cut-outs  88 - 92  in secondary sealing element  84  and the deflection structures  110 - 116  of the secondary sealing element  102  all function to enhance the radially flexibility of the respective secondary sealing elements  70 ,  84  and  102 , compared to a continuous annular plate, so that they move inwardly and outwardly with the shoe  16  in response to the application of fluid pressure thereto and assist in the creation of a secondary seal of the gap  11  between the rotor  14  and stator  12 . 
     Referring now to  FIGS. 11-12A , a further embodiment of a seal  130  according to this invention is illustrated. As noted above, the task of maintaining adequate radial tolerances in turbomachinery, and particularly gas turbine engines employed in aircraft, is complicated by a number of factors such as thermal expansion, shaft motion, tolerance stack-ups, rub tolerance, the presence of debris and the like. It is desirable to provide a non-contact seal between rotor  14  and the seal  10  described above, while limiting leakage, but it is not always possible to prevent contact between the two under all operating conditions. The seal  130  of  FIGS. 11-12A  includes one or more shoes  132 , one of which is shown in such Figs., each having an inner surface  134  and an outer surface  136 . As discussed below, the shoes  132  are specifically designed to “reset” their radial position following unintended contact with rotor  14 . 
     The seal  130  may include essentially the same spring elements  24  and stops  30 ,  40 , having arms  34 ,  42 , as described above in connection with a discussion of 
       FIGS. 3-5 . Additionally, any one of the secondary sealing elements  48 ,  50 ,  70 ,  84  or  102  described above with reference to  FIGS. 6-10A  may be employed in the seal  130 . In one embodiment, depicted in  FIG. 12 , secondary sealing elements  48  and  50  are shown positioned within a slot  22  formed in the outer surface  136  of a shoe  132  between front plate  60  and back plate  62 . It is contemplated that the position of secondary sealing elements  48 ,  50 ,  70 ,  84  or  102  along the outer surface  136  of shoe  132  could be varied, and may be located, for example, in a position overlying tooth elements extending from the inner surface  134  of shoe  132  as discussed in detail below. 
     For purposes of the present discussion, a stream of fluid is presumed to be flowing over both the outer surface  136  of shoe  132  and in between the inner surface  134  of shoe  132  and the rotor  14 . See the arrow  138  in  FIG. 12 . The fluid stream  138  is considered to be at “high” pressure on the left-hand or upstream side  140  of shoe  132  and comparatively lower pressure on the downstream side  142 . The shoe  132  includes three longitudinally spaced labyrinth-type tooth elements  144 ,  146  and  148 , all located upstream from a nozzle  150 . 
     The tooth elements  144 ,  146  and  148  each project from the inner surface  134  of shoe  132  and extend in a direction toward the rotor  14 . Tooth element  144  has a substantially vertical upstream surface  152 , oriented at about 90° from inner surface  134 , a tip  154  and a downstream surface  156  which is disposed at an angle relative to inner surface  134  of greater than 90°. The thickness of tooth element  144 , as measured between its upstream and downstream surfaces  152 ,  156 , therefore decreases from the inner surface  134  of shoe  132  toward the rotor  14 . The other two tooth elements  146  and  148  have essentially the same shape as tooth element  144 , but different lengths. The middle tooth element  146  has a tip  158  connected between upstream and downstream surfaces  160 ,  162 , and the tip  164  of the third tooth element  148  is connected to its upstream and downstream surfaces  166 ,  168 . Preferably, the middle tooth element  146  has the greatest length, as measured between the inner surface  134  of shoe  132  and its tip  158 , while the tooth element  144  is shortest in length and the tooth element  148  has a length in between that of the tooth elements  144  and  146 . By way of example, the length of the middle tooth element may be such that its tip  158  is spaced about 0.010 inches from rotor  14 , whereas the tip  154  of tooth element  144  is spaced about 0.029 inches from rotor  14  and the tip  164  of tooth element  148  is spaced about 0.019 inches from rotor  14 . It should be understood that the lengths of tooth elements  144 ,  146  and  148 , and, their spacing from rotor  14 , may be varied depending upon a particular application and the desired clearance relative to rotor  14 . 
     The nozzle  150  is preferably formed with a tip  170  connected between an upstream surface  172  and a downstream surface  174 , both of which are oriented at an angle of greater than 90° relative to the inner surface  134  of shoe  132 . As best seen in  FIGS. 12 and 12A , the upstream surface  172  of nozzle  150  converges toward the rotor  14  whereas its downstream surface  174  diverges away from the surface of the rotor  14 . 
     The seal  130  is initially mounted to the stator  12  in a position to create a non-contact seal of the circumferential gap  11  between the stator  12  and rotor  14 , e.g. with a radial spacing from the rotor  14  within tolerances of at least about 0.010 inches, in the example given above, so that the middle tooth element  146  does not contact the rotor  14 . As the fluid stream  138  passes between the inner surface  134  of shoe  132  and rotor  16 , the tolerances are such that upon initial installation of the seal  130  the fluid flow is restricted to some extent. In turn, the velocity of the fluid stream  138  between the shoe  132  and rotor  14  increases causing a corresponding drop in the pressure along the inner surface  134  of the shoe  132 . In response to reduced pressure acting on the inner surface  134  of the shoe  132 , the force exerted by spring elements  24  and secondary sealing elements  48 ,  50 ,  70 ,  84  or  102  on the opposite, outer surface  136  of the shoe  132  urge it radially inwardly toward the rotor  14  to a position at which an initial seal is created. A force balance is obtained between the pressure exerted by the fluid stream  138  both on its inner and outer surface  134 ,  136 , in combination with the force exerted by springs  24  and secondary sealing elements  48 ,  50 ,  70 ,  84  or  102 , such that the radial spacing between the longest tooth element  146  and the rotor  14 , e.g. the “initial” seal, is within desired tolerances. 
     Conditions within turbomachinery, and especially gas turbine engines for aircraft, are dynamic and the seal  130  is designed to dynamically respond to such conditions. In the event of inadvertent contact between the seal  130  and rotor  14  at the position where the initial seal was created, the middle tooth element  146  is first to engage the surface of the rotor  14 , because it is the longest, and it begins to wear. As the middle tooth element  146  wears, the next longest tooth element  148  moves closer to the rotor  14  causing the fluid pressure beneath the shoe  132 , along its inner surface  134 , to increase. In turn, a force is exerted against the shoe  132  urging it in a radial direction away from the rotor  14  and against the force exerted by the fluid stream  138  on the outer surface  136  of the shoe  132  as well as the forces applied by springs  24  and the secondary sealing elements  48 ,  50 ,  70 ,  84  or  102  to such outer surface  136 . If the radially outwardly directed force on the shoe  132  does not balance the forces exerted in the radially inward direction, the tooth element  146  will continue to wear, and tooth element  148  may begin to wear, thus moving the nozzle  150  closer to rotor  14 . Such wear conditions create an increasingly greater force acting in the radially outward direction on the inner surface  134  of shoe  132 , until such time as the seal  130  moves to a new or “reset” seal position wherein the force exerted by the fluid pressure on the inner surface  134  of the shoe  132  balances the force applied to the shoe  132  in the opposite direction by the fluid steam  138 , the spring elements  24  and the secondary sealing elements  48 ,  50 ,  70 ,  84  or  102 . The seal  130  therefore dynamically resets its non-contact position with respect to the rotor  14 , albeit at a somewhat greater radial spacing from the rotor  14  compared to the tolerances of the initial seal at installation, but nevertheless with limited leakage within the circumferential gap  11  between the stator  12  and rotor  14 . 
     While the invention has been described with reference to a preferred embodiment, it should be understood by those skilled in the art that various changes may be made and equivalents may be substituted for elements thereof without departing from the scope of the invention. In addition, many modifications may be made to adapt a particular situation or material to the teachings of the invention without departing from the essential scope thereof. Therefore, it is intended that the invention not be limited to the particular embodiments disclosed as the best mode contemplated for carrying out the invention, but that the invention will include all embodiments falling within the scope of the appended claims.