Abstract:
A turbocharger shaft is sealed with respect to the bore of the bearing housing at the turbine wheel end by one or more shaft seal rings. Conventionally these rings are seated in annular grooves provided in the shaft behind the turbine wheel. Problems can arise with this conventional shaft seal arrangement, particularly where the turbocharged engine has an engine brake valve located downstream of the turbine, and the back pressure in the exhaust line, and thus in the turbine wheel housing, can reach 7 bar. The inventive shaft seal design avoids grooves, and makes it possible to assemble a seal with ring seals with higher wear resistance and thus to maintain the seal&#39;s effectiveness over heavy use.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
       [0001]    This application claims the benefit of U.S. Patent Application No. 62/078631 filed on Nov. 12, 2014, the disclosure of which is herein incorporated by reference in its entirety. 
     
    
     BACKGROUND OF THE INVENTION 
       [0002]    1. Field of the Invention 
         [0003]    In a turbocharger the shaft is typically sealed with respect to the bore of the bearing housing at the turbine wheel end by one or more shaft seal rings (or piston rings) which sit in annular grooves provided in the shaft behind the turbine wheel. Problems can arise with this conventional shaft seal arrangement where the turbocharged engine has an engine brake valve located downstream of the turbine. As the engine brake is activated, the back pressure in the exhaust line, and thus in the turbine wheel housing, rapidly rises and can reach  7  bar. The inventive shaft seal design makes it possible to assemble a seal with ring seals with higher wear resistance and thus to maintain the seal&#39;s effectiveness over heavy use. 
         [0004]    2. Description of the Related Art 
         [0005]    Turbochargers use the exhaust flow from the engine exhaust manifold, which enters the turbine stage of the turbocharger at a turbine housing inlet, to drive a turbine wheel, which is located in the turbine housing. The turbine wheel is affixed to one end of a shaft that is rotatably supported within a bearing housing. The shaft drives a compressor wheel mounted on the other end of the shaft. The turbine wheel, shaft and compressor wheel form a rotating assembly that is supported within the bearing housing. As such, the turbine wheel provides rotational power to drive the compressor of the turbocharger. This compressed air is then provided to the engine intake at a greater density than would be possible in a normal aspirated configuration. This allows more fuel to be combusted, thus boosting the engine&#39;s horsepower without significantly increasing engine weight. 
         [0006]    When a commercial truck, equipped with an engine compression type exhaust brake, travels down a grade with a long incline, the exhaust brake can be used to block the flow of exhaust gas downstream of the turbine wheel and provide retardation to the vehicle, independent of the vehicle&#39;s wheel brakes. The mass and inertia of the truck rolling down the incline forces rotation of the engine through the vehicle gearbox which is driven by the wheels. With no fuel being introduced into the engine, the engine acts like an air pump against the blockage of the exhaust brake, consuming energy and retarding the velocity of the truck. 
         [0007]    The turbocharger shaft is typically sealed with respect to the bore of the bearing housing at the turbine wheel end by shaft seal rings (or piston rings) which sit in an annular grooves provided in the shaft behind the turbine wheel (see, e.g., BorgWarner WO2014099289, FIG. 2), The shaft seal prevents oil from the bearing lubrication systems from leaking into the turbine housing which can cause blue smoke and oil drips from the exhaust pipe, as well as preventing exhaust gas pollution of the bearing housing which can cause overheating and adversely effect bearing life. 
         [0008]    Problems can arise with this conventional shaft seal arrangement upon activation of the engine brake valve located downstream of the turbine. As the engine brake is activated the back pressure in the exhaust line, and thus in the turbine wheel housing, rapidly rises and can reach 7 bar. As pressure behind the turbine wheel increases, the shaft seal ring(s) can be pushed inboard. This movement, together with the high rotational speed of the shaft, can generate excessive frictional heating which can cause the shaft seal ring to overheat. This in turn can cause the rings to be more susceptible to movement in the bore and induce a rapid failure of the seal. 
         [0009]    It is an object of the present invention to overcome the above problem. 
       SUMMARY 
       [0010]    The present invention is made based on the realization that, in the conventional turbocharger, where the shaft is sealed with respect to the bore of the bearing housing by one or more shaft seal rings or ring seals which sit in annular grooves provided in the shaft behind the turbine wheel, the need to expand the rings during installing so that their inner diameter can pass over the areas of greater shaft outer diameter adjacent the groove and then contract the ring for seating of the ring in the groove of the shaft dictates that the rings be made of a malleable or elastic material having high tensile yield strength. Such materials are disadvantageous in that they tend to be compromised in hardness and wear resistance. 
         [0011]    The present inventor discovered that by re-designing the shaft and seals to eliminate grooves, such that the ring seals do not have to be expanded and contracted during installing, there is no longer a constraint to selection of sealing rings made of materials with high yield strength. It now becomes possible to select sealing rings made of higher hardness and wear resistance, regardless of yield strength. 
         [0012]    As used herein, the term wear resistant refers to materials having a high hardness, such as a hardness of greater than 50 C in a Rockwell test, or a hardness of greater than 85 C when using a smaller HR15N indenter to accommodate smaller parts. In addition, as used herein, a low yield strength refers to a yield strength of less than 500 MPa. In one example, the ring seals are formed of a sintered steel embedded with solid lubricants. In another example, the ring seals are formed of a ceramic material. In yet another example, the ring seals are formed of conventional materials coated to improve wear resistance properties. Although such coatings are known, it has not previously been possible to use them to coat ring seals of conventional turbocharger labyrinth seals since the coatings tend to become cracked or otherwise damaged when the ring seals are stretched open to accommodate the shaft outer diameter while being delivered to their corresponding grooves provided in an outer surface of the shaft. Thus, the modular seal assembly permits a greatly enlarged range of materials that can be used to form the ring seals as compared to some conventional seal assemblies. The grooveless seal assembly advantageously permits assembly of the wear resistant, low yield strength ring seals without an assembly step necessitating deformation thereof. 
         [0013]    In accordance with the invention, the turbocharger includes a shaft having at least first and second shoulders delimiting segments of different shaft diameter. The modular seal assembly further comprises at least one retainer and one retainer ring, as well as first and second seal rings. The retainer and retainer ring are separated by a spacer, integrally formed with either the retainer or retainer ring. The retainer and retainer ring when assembled form between them an annular recess adapted to receive the first seal ring. The retainer ring abuts against a first shoulder. Between the first shoulder and a second shoulder is a shaft segment having an outer diameter dimensioned to receive the second seal ring. 
         [0014]    As a result of this design, the shaft seal is produced by simply sliding onto the shaft the second shaft seal, the retainer ring until it abuts the second shoulder, the first shaft seal, and the retainer until it abuts the retainer ring. It is not necessary to expand the seal rings to pass over a shaft segment of larger diameter in order to seat the ring in a groove machined into the shaft. 
         [0015]    In one embodiment of the invention the outboard and inboard ring seals are dimensioned to interference fit in the bearing housing bore, so that they are static, while the retainer ring and retainer are interference fit or otherwise attached to the rotating shaft. In another embodiment, the retainer ring is interference fit in the bearing housing bore, while the outboard and inboard ring seals are mounted to rotate with the shaft. 
     
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         [0016]    Embodiments are illustrated by way of example and not limitation in the accompanying drawings in which like reference numbers indicate similar parts. 
           [0017]      FIG. 1  is a cross-sectional view of an exhaust gas turbocharger including a modular seal assembly at the turbine-end of the shaft. 
           [0018]      FIG. 2  is an enlarged cross sectional view of the turbine-end of the shaft illustrating a first embodiment of the modular seal assembly. 
           [0019]      FIG. 3 . is a simplified illustration of a modification of the first embodiment of the invention. 
           [0020]      FIG. 4  is an enlarged cross sectional view of the turbine-end of the shaft illustrating a first embodiment of the modular seal assembly. 
       
    
    
     DETAILED DESCRIPTION 
       [0021]    Arrangements described herein relate to sealing systems and methods for use between the dynamic rotating assembly components and the complementary static components on the turbine-end of a turbocharger. More particularly, embodiments herein are directed to forming sealing systems that can reduce turbine end blow-by leakage. Detailed embodiments are disclosed herein; however, it is to be understood that the disclosed embodiments are intended only as exemplary. Therefore, specific structural and functional details disclosed herein are not to be interpreted as limiting, but merely as a basis for the claims and as a representative basis for teaching one skilled in the art to variously employ the aspects herein in virtually any appropriately detailed structure. Further, the terms and phrases used herein are not intended to be limiting but rather to provide an understandable description of possible implementations. 
         [0022]    Referring to  FIG. 1 , an exhaust gas turbocharger  10  includes a turbine section  12 , a compressor section  14 , and a center bearing housing  23  disposed between, and connecting, the compressor section  14  to the turbine section  12 . The turbine section  12  includes a turbine housing  17  and a turbine wheel  16  disposed in the turbine housing  17 . The compressor section  14  includes a compressor housing  19  and a compressor wheel  18  disposed in the compressor housing  19 . The turbine wheel  16  is connected to the compressor wheel  18  via a shaft  2 . 
         [0023]    In the illustrated embodiment, a journaled portion  5  of the shaft  2  is supported for rotation about a rotational axis R within a bore  27  formed in the bearing housing  23  via a pair of axially-spaced radial bearings  26 . The axial spacing between the radial bearings  26  is maintained by a cylindrical bearing spacer (not shown). In addition, a thrust bearing assembly  24  is disposed in the bearing housing  23  so as to axially locate the shaft  2 . The shaft  2  is reduced in diameter on the compressor side outboard of the compressor-side journal bearing  26  relative to the journaled portion  5 , and a first shoulder  101  is formed in the shaft  2  at the transition between diameters. The compressor wheel  18  and the thrust bearing assembly  24 , including a thrust bearing  28 , a thrust washer assembly  22 , and an oil flinger, are all located on the shaft  2  in the reduced diameter portion  6 , also referred to as the “stub portion”  6 . The terminal end  3  of the stub portion of the shaft  2  extends axially beyond the compressor wheel  18  and includes an external thread. A nut  32  engages the thread, and is tightened sufficiently to clamp the compressor wheel  18 , the thrust washer assembly  22  and flinger against the first shoulder  101 . Rather than being clampled into the rotating assembly, the thrust bearing  28  is stationary, and is clamped to the bearing housing  23  by a retaining ring and the insert  34 . 
         [0024]    In use, the turbocharger  10  uses the exhaust flow from the exhaust manifold of an engine (not shown) to drive the turbine wheel  16 . Once the exhaust gas has passed through the turbine wheel  16  and the turbine wheel  16  has extracted energy from the exhaust gas, the exhaust gas exits the turbine housing  17  through the exhaust gas outlet  13  and is ducted to the vehicle downpipe and usually to after-treatment devices such as catalytic converters, particulate traps, and NO x  traps. The energy extracted by the turbine wheel  16  is translated to a rotational motion that is used to drive a compressor wheel  18 . As the compressor wheel  18  rotates, it increases the air mass flow rate, airflow density and air pressure delivered to the engine&#39;s cylinders via an outflow from the compressor section  14 , which is connected to the engine&#39;s air intake manifold (not shown). 
         [0025]    The turbocharger bearing system is lubricated by oil from the engine. The oil is fed under pressure into the bearing housing  23  via an oil supply port  36  to lubricate the thrust bearing assembly  24  and the shaft bearing surfaces within and about the radial bearings  26 . 
         [0026]    More specifically, oil passes through individual bearing supply channels  38 ,  40  for lubricating the thrust bearing  28  and the radial bearings  26 . Oil is eventually collected within a bearing housing sump chamber  29  for return circulation to the engine crankcase through an outlet port  30 . 
         [0027]    Referring also to  FIG. 2 , the shaft  2  includes a conventional compressor-side clearance or labyrinth seal  21  disposed on the between an insert  34  and the flinger to address compressor end blow-by leakage to the space  52  between insert  34  and compressor wheel  18  backwall. 
         [0028]    To address turbine end blow-by leakage, a modular seal assembly  110  is disposed between a seal portion  7  of the shaft  20  and a shaft bore portion  27   a  of the bore  27 , described in detail below. The shaft seal portion  7  of the shaft has an enlarged diameter D sp  relative to a diameter D jp  of the journaled portion  5  of the shaft  20 , and a second shoulder  102  is optionally formed in the shaft  2  at the transition between the diameters Dsp and Djp. 
         [0029]    The modular seal assembly  110  is configured to effectively minimize or prevent turbine-end blow-by leakage and blow by both at the time of initial installation and over the lifetime of the turbocharger  10 . The modular seal assembly  110  is disposed on the seal portion  7  of the shaft  2 , and provides a labyrinth seal between the bearing housing shaft bore  27   a  and the shaft  2 . The modular seal assembly  110  includes an outboard ring seal  142   b,  a retainer ring  162   a  that is urged against a shaft shoulder  105  to retain the outboard ring seal  142   b  in a recess between the retainer ring  162   a  and shaft shoulder  103 , an inboard ring seal  142   a,  and an retainer  112  that, in conjunction with the retainer ring  162   a,  axially retains the inboard ring seal  142   a  in a recess between retainer  112  and retainer ring  162   a  formed by a spacer  126 . The spacer may be formed as an axial extension of the retainer  112 , and spaces the retainer  112  and retainer ring  162   a  sufficient to allow inboard ring seal  142   a  to rotate freely with close tolerance. 
         [0030]    In the embodiment of the invention shown in  FIG. 2 , the inboard and outboard ring seals  142   a,    142   b  are dimensioned to interference fit in the bearing housing bore, so that they are static, while the retainer ring and retainer are interference fit or otherwise attached to the rotating shaft. In another embodiment, the retainer ring is interference fit in the bearing housing bore, while the first and second shaft seals are mounted to rotate with the shaft. 
         [0031]    In  FIG. 2  the retainer  112  is a cylinder with a bore matched to the shaft, having a radially outward-facing surface  114  that faces the ring bore portion  27   a,  and a radially inward-facing surface  116  that faces an outer surface of the shaft  2 . The retainer  112  includes a planar, axially inward-facing surface  118 , and an opposed, planar, axially outward facing surface  120 . In addition, the retainer  112  optionally includes a circumferentially-extending groove  124  formed in the radially outward-facing surface  114  that serves as an oil flinger. Separation of the retainer  112  from the shaft due to thermal expansion and high rotational speeds is avoided by careful selection of materials used to form the retainer  112 . 
         [0032]    The retainer ring  162  is annular and has in the illustrated embodiment has a rectangular cross section. The retainer ring  162  has a radially outward-facing edge  164  that faces the bearing housing shaft bore portion  27   a,  and a radially inward-facing edge  166  that is press fit or shrink fit on an outer surface of the shaft  2  in the embodiment shown in  FIG. 2 . In addition, the retainer ring  162  includes a planar, axially inboard-facing surface and an opposed, planar, axially outboard facing surface. 
         [0033]    The outboard ring seal  142   b  and the inboard ring seal  142   a  are substantially similar, and thus common reference numbers will be used to refer to common parts. The outboard ring seal  142   b  and the inboard ring seal  142   a  are each annular and have a rectangular cross section. The outboard ring seal  142   b  and the inboard ring seal  142   a  each have a radially outward-facing edge  164  that faces the bearing housing shaft bore portion  27   a,  and a radially inward-facing edge  146  that encircles the shaft  2 . In addition, the outboard ring seal  142   b  and the inboard ring seal  142   a  each include a planar, axially inboard-facing surface, and an opposed, planar, axially outboard facing surface  180 . 
         [0034]    The outboard ring seal  142   b  and the inboard ring seal  142   a  are formed of a highly wear resistant material. For example, in some embodiments, the surface of the ring seals  142   a,    142   b  may have a surface hardness of Rockwell C 50 and above (or a hardness of greater than 85 C when using a smaller HR15N indenter to accommodate smaller parts) or a Knoop hardness of 1000 and above. In addition, the material used to form the outboard ring seal  142   b  and the inboard ring seal  142   a  may have a low yield strength material relative to some conventional ring seals used in turbocharger shaft sealing mechanisms. For example, a conventional ring seal used in this application, such as a ring seal formed of a hardened tool steel such as M2 Tool steel, has a yield strength of 2400 MPa in order to accommodate a ring opening stress of about 2300 MPa. In contrast, the material used to form the outboard ring seal  142   b  and the inboard ring seal  142   a  in the present invention may be a highly wear resistant, low yield strength material, such as a ceramic material or a sintered tool steel embedded with solid lubricants. Such highly wear resistant, low yield strength materials may have a yield strength of zero (in the case of some ceramic materials) to about 400 MPa (in the case of some types of lubricant-embedded sintered tool steels). In other embodiments, the material used to form the outboard ring seal  142   b  and the inboard ring seal  142   a  may be a conventional material that includes a highly wear resistant coating. Such coatings may include, but are not limited to, titanium or chromium nitride coatings and diamond-like carbon (DLC) coatings. Such alternative materials can be employed to form the ring seals  142   a,    142   b  in the modular seal assembly  110  since, due to the modular nature of the assembly, the ring seals  142   a,    142  are assembled without stretching to pass over shaft diameters greater than the ring seal inner diameter before being slid onto the gap on the shaft  2 . The modular nature of the seal assembly  110  removes risk of cracking a surface coating during assembly. 
         [0035]    The ring seals require a radial gap to allow some tolerance variation as each ring is fit into the bearing housing shaft bore  27   a.  This is not a problem because the installed stress is low, much lower than the traditional opening stress to assemble, and within the strength of the low wear materials. The installed stress also decreases at operating temperatures because the low-wear materials tend to have very low thermal expansion compared to the bearing housing. 
         [0036]      FIG. 3  shows an alternative design of the spacer component of the modular seal assembly. Rather than being formed as a component of the retainer  112 , it is formed as a component of the retainer ring  162   b.  The advantage of this design is that it allows one, two, three, or more retainer rings with spacers to be installed, alternating with ring seals  142   a,  so that any number of ring seals can be provided on the turbine end of the shaft. 
         [0037]    The term “labyrinth seal” as used herein is intended to include the “tortuous path” seal as disclosed in WO2013106303, a “sequential chamber” (sequentially decreasing pressure, with cylincrical shaft of constant diameter) seal as disclosed in U.S. Pat. No. 6,575,693, or a combination of both. The basic idea is that as outboard pressure passes a small outer gap the pressure drops, within the volume in the space between rings there is expansion to again lower the pressure, and as gas passes the next small gap inboard pressure drops further. Small gaps effectively reduce blow-by gas leakage. A labyrinth seal may be composed of many stationary rings and rotating grooves which interdigitate to produce the long characteristic path which slows leakage so that the fluid has to pass through a long and difficult path to escape. For labyrinth seals on a rotating shaft, a very small clearance must exist between the tips of the labyrinth threads and the running surface. Labyrinth seals on rotating shafts provide non-contact sealing action by controlling the passage of fluid through a variety of chambers by centrifugal motion, as well as by the formation of controlled fluid vortices. At higher speeds, centrifugal motion forces the liquid towards the outside and therefore away from any passages. Similarly, if the labyrinth chambers are correctly designed, any liquid that has escaped the main chamber becomes entrapped in a labyrinth chamber, where it is forced into a vortex-like motion. This acts to prevent its escape, and also acts to repel any other fluid. So long as these labyrinth seals are non-contacting as designed, they do not wear out. 
         [0038]    Labyrinth sequential chamber seals can also be formed with one cylindrical surface of constant diameter (e.g., rotating shaft) and one surface forming a series of chambers. Leaking fluid is forced through a series of chambers separated by teeth, creating a decrease in both pressure and flowrate. 
         [0039]    Referring again to  FIG. 2 , the seal portion  7  of the shaft  2  includes surface features which accommodate and cooperate with the elements of the modular seal assembly  110  to form an effective labyrinth seal. In particular, the seal portion  7  includes a first reduced diameter portion  7   a  having a diameter D 1  at a location axially spaced apart from the turbine wheel-end  4  of the shaft  2 , whereby a third shoulder  103  is formed in the shaft  2  at the transition between the diameters D sp  and D 1 . The first reduced diameter portion  7   a  has an axial length that corresponds to the axial dimension of ring seal  142   b.  When assembled, the retainer ring  162  is disposed in the step with reduced diameter D 3  and is urged against shoulder  105 . The radially inward-facing surface of the retainer ring  162  may be dimensioned to provide a press fit, a slip fit or a rotational fit with respect to the shaft step  7   b.    
         [0040]    When the pressure in the turbine housing is greater than the pressure in the bearing housing, and in particular when an engine brake valve located downstream of the turbine is activated, and the back pressure in the exhaust line, and thus in the turbine wheel housing, rapidly rises and can reach 7 bar, the outboard ring seal may be urged axially inboard, against outboard surface  126   a  of retaining ring  162   a.  In this condition the present invention provides a significant advance over the state of the art, since the friction of the two high speed rubbing surfaces is reduced by the provision of an anti-friction component or coating on the ring seal, or mitigated by the use of a highly wear resistant material for the ring seal. 
         [0041]    Retainer ring his held in place against shaft shoulder  105  by spacer  126  extending from retainer  112 . The retainer and retainer ring when assembled form between them an annular recess adapted to receive the first seal ring. 
         [0042]    As a result of this design, the shaft seal arrangement is produced by simply sliding onto the shaft the outboard shaft seal, the retainer ring until it abuts the second shoulder, the inboard shaft seal, and the retainer until it abuts the retainer ring. It is not necessary to expand the seal rings to pass over a shaft segment of larger diameter in order to seat the ring in a groove machined into the shaft. 
         [0043]    In the embodiment shown in  FIG. 2  the outboard and inboard ring seals are dimensioned to interference fit in the bearing housing bore, so that they are static, while the retainer ring  162   a  and retainer  112  are interference fit or otherwise attached to the rotating shaft. In the embodiment shown in  FIG. 4 , the retainer ring  162  is interference fit in the bearing housing bore, while the outboard and inboard ring seals are mounted to rotate with the shaft. 
         [0044]      FIGS. 2 and 4  show a shoulder  104  between shaft diameter D 2  and D 3 . In the embodiment shown in  FIG. 2  the reduced shaft diameter D 2  and shoulder  104  are optional. In  FIG. 4  the area of reduced shaft diameter D 2  is externally threaded and allows internally threaded retainer  212  to be screwed onto and secured to shaft  2 . 
         [0045]    The modular seal assembly  110  is assembled on the shaft  2  by sequentially axially sliding each element of the modular seal assembly  110  on shaft seal portion  7  as follows: First, the outboard ring seal  142   b  is slid along the shaft  2  until it resides in the first reduced diameter portion  7   a  adjacent the third shoulder  103 . Then, the retainer ring  162 ,  162   a,    162   b  is slid along the shaft  2  until it rests against shoulder  105 . The inboard ring seal  142   a  is then slid along the shaft  2  from the stub portion  6  to the seal portion  7  until it rests against the retainer ring  162 ,  162   a,    162   b.  Finally, the retainer  112  is slid along the shaft  2  from the stub portion  6  to the seal portion  7  until it resides in the second reduced diameter portion  7   c  with the spacer abutting the retainer ring  162 ,  162   a,    162   b.  The retainer  112  is press fit to the second reduced diameter portion  7   c,  whereby the inboard ring seal  142   a  is retained in the desired axial location between the retainer  112  and the retainer ring  162 . 
         [0046]    In  FIG. 2 , the inboard and outboard ring seals  142   a,    142   b  are installed under radial compression, and when in place, the inboard and outboard ring seals  142   a,    142   b  are interference fit to the ring bore  27   a.  As a result, the inboard and outboard ring seals  142   a ,  142   b  remain stationary during turbocharger operation. In addition, the labyrinth or clearance seal is formed between a first, stationary labyrinth surface defined by the ring bore  27   a  and the inboard and outboard ring seals  142   a,    142   b,  and a second, rotating labyrinth surface defined by the retainer  112  and retainer ring  162  and the above-described surface features of the shaft seal portion  7 . 
         [0047]    The retainer  212  of the modular seal assembly shown in  FIG. 4  is similar to the modular seal assembly  112  of  FIG. 2 , differs in that the radially inward facing surface  216  of the retainer  212  includes threads that are configured to engage corresponding threads formed on the outer surface of the second reduced diameter portion  7   b.  Thus, the retainer  212  is secured to the shaft  2  by the cooperative engagement of the respective threaded portions. 
         [0048]    Introduction of the shaft, with modular seal components, into the bearing housing is facilitated by the beveled opening of the shaft bore as can be seen in  FIGS. 1, 2 and 4 . 
         [0049]    While the shaft  2  has been described herein as being supported for rotation by a pair of radial bearings, it is contemplated, however, that the radial bearings used to rotatably support the shaft  2  can include, and are not limited to journal bearings and rolling element bearings (REBs) such as angular contact bearings, etc. When certain types of rolling element bearings are used, the thrust bearing assembly  24  can be omitted. 
         [0050]    Although the modular seal assembly  110 ,  210  is described with respect to addressing turbine-end blow-by leakage, the seal assembly  110 ,  210  can be easily adapted for use on the compressor-end of the shaft  2  in order to address compressor-end blow-by leakage. 
         [0051]    Aspects described herein can be embodied in other forms and combinations without departing from the spirit or essential attributes thereof. For instance, while embodiments described herein are directed to compressor end blow-by leakage, it will be appreciated that such sealing systems and methods can be applied to minimize turbine end oil discharge (i.e., the passage of oil from the bearing housing to the turbine stage). Thus, it will of course be understood that embodiments are not limited to the specific details described herein, which are given by way of example only, and that various modifications and alterations are possible within the scope of the following claims.