Patent Publication Number: US-9890788-B2

Title: Turbocharger and method

Description:
TECHNICAL FIELD 
     This patent disclosure relates generally to turbochargers and, more particularly, to turbochargers used on internal combustion engines. 
     BACKGROUND 
     Internal combustion engines are supplied with a mixture of air and fuel for combustion within the engine that generates mechanical power. To maximize the power generated by this combustion process, the engine is often equipped with a turbocharged air induction system. 
     A turbocharged air induction system includes a turbocharger having a turbine that uses exhaust from the engine to compress air flowing into the engine, thereby forcing more air into a combustion chamber of the engine than a naturally aspirated engine could otherwise draw into the combustion chamber. This increased supply of air allows for increased fuelling, resulting in an increased engine power output. 
     In conventional turbochargers, engine oil is provided to lubricate and cool bearings in the bearing housing that rotatably support a turbocharger shaft that transfers power from the turbine to the compressor. In addition to cooling and lubrication, the oil provides dampening for shaft and bearing vibrations when provided in thin films as it passes though control or bearing surfaces. Such dampening, which is sometimes referred to as squeeze film dampening, can provide vibration dampening but is often insufficient to provide sufficient dampening in cartridge style bearings. Typical designs may suspend the bearing arrangement within a bore in a bearing housing and use a pin to prevent rotation of the bearing assembly along with the turbocharger shaft relative to the bearing housing. 
     SUMMARY 
     The present disclosure is applicable to turbochargers for use in internal combustion engines. In one embodiment, a turbocharger includes a turbine, a compressor, and a bearing housing forming a bearing bore. A bearing arrangement is disposed between a shaft interconnecting the turbine and compressor wheels, and the bearing housing. The bearing arrangement includes an outer bearing race element that frictionally engages the bearing bore at both ends. At a first end, a bearing retainer that is connected to the bearing housing includes a cylindrical element that extends into the bearing bore and abuts and frictionally engages a first end of the outer bearing race element. The outer bearing race element is otherwise free to rotate within the bearing bore under a rotational motion that is dampened by viscosity of oil films present along bearing surfaces between the outer bearing race element and the bearing bore. In order for oil to gravity scavenge multiple drain holes are located in the outer race to ensure proper oil scavenging at all circumferential orientations of the bearing outer race. 
     Therefore, in one aspect, the disclosure describes a turbocharger that includes a turbine having a turbine wheel, a compressor having a compressor wheel, and a bearing housing disposed and connected between the turbine and the compressor. The bearing housing forms a bearing bore having a stop surface at one end. A shaft is rotatably disposed within the bearing housing and extends into the turbine and the compressor. The turbine wheel is connected to one end of the shaft and the compressor wheel is connected to an opposite end of the shaft such that the turbine wheel is rotatably disposed in the turbine and the compressor wheel is rotatably disposed in the compressor. A bearing arrangement is disposed between the shaft and the bearing housing, and includes an outer bearing race element disposed in the bearing bore. The outer bearing race element has a hollow cylindrical shape that engages the bearing bore frictionally along a first end and a second end, and an inner bearing race element, which engages the shaft and is rotatably disposed within the outer bearing race element. A bearing retainer has a cylindrical portion extending into the bearing bore and abutting the first end of the outer bearing race element such that, during operation, the outer bearing race element is free to rotate within the bearing bore at a rotational motion that is dampened by a viscosity of oil present at bearing surfaces disposed between the outer bearing race element and the bearing bore. 
     In another aspect, the disclosure describes a method for rotatably and sealably supporting a shaft within a bearing housing of a turbocharger. The method includes connecting a turbine wheel at one end of the shaft, forming a first roller bearing by engaging a first plurality of rolling elements in a first inner race formed in an inner bearing race element and in a first outer race formed in an outer bearing race element, forming a second roller bearing by engaging a second plurality or rolling elements in a second inner race formed in the inner bearing race element and in a second outer race formed in the outer bearing race element, and engaging the outer bearing race element between a bearing bore formed in the bearing housing and the shaft, which extends through the bearing bore, such that the inner bearing race element rotates with the shaft with respect to the outer bearing race element. The method further includes frictionally engaging the bearing bore with the outer bearing race element along a first end and a second end, and engaging the shaft with an inner bearing race element, which is rotatably disposed within the outer bearing race element. A bearing retainer having a cylindrical portion extending into the bearing bore and abutting the first end of the outer bearing race element is provided. The method also includes allowing the outer bearing race element to rotate within the bearing bore, and dampening a rotation of the outer bearing race element within the bearing bore by providing a squeeze film of oil along bearing surfaces between the outer bearing race element and the bearing bore. 
     In yet another aspect, the disclosure describes an internal combustion engine having a plurality of combustion chambers formed in a cylinder block, an intake manifold disposed to provide air or a mixture of air with exhaust gas to the combustion chambers, and an exhaust manifold disposed to receive exhaust gas from the combustion chambers. The engine further includes a turbine having a turbine housing surrounding a turbine wheel, the turbine housing being fluidly connected to the exhaust manifold and disposed to receive exhaust gas therefrom to drive the turbine wheel, a compressor having a compressor housing that surrounds a compressor wheel, the compressor housing being fluidly connected to the intake manifold and disposed to provide air thereto, and a bearing housing disposed and connected between the turbine and the compressor. The bearing housing forms a bearing bore therethrough that accommodates a shaft interconnecting the turbine wheel and the compressor wheel to transfer power therebetween. The shaft is rotatably mounted within the bearing housing and extends into the turbine and the compressor such that the turbine wheel is connected to one end of the shaft and the compressor wheel is connected to an opposite end of the shaft. A bearing arrangement is disposed between the shaft and the bearing housing. The bearing arrangement includes first and second bearings, each of the first and second bearings formed by a respective first and second plurality of roller elements engaged between a respective first and second inner race and a respective first and second outer race. 
     In one embodiment, an outer bearing race element is disposed within the bearing bore and forms the respective first and second outer races, and an inner bearing race element forms the respective first and second inner races. The outer bearing race element is disposed in the bearing bore and has a hollow cylindrical shape that engages the bearing bore frictionally along a first end and a second end. A bearing retainer having a cylindrical portion extends into the bearing bore and abuts the first end of the outer bearing race element. The outer bearing race element, during operation, is free to rotate within the bearing bore at a rotational motion that is dampened by a viscosity of oil present at bearing surfaces disposed between the outer bearing race element and the bearing bore. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a block diagram of an internal combustion engine in accordance with the disclosure. 
         FIG. 2  is an outline view from a side perspective of a turbocharger in accordance with the disclosure. 
         FIG. 3  is a fragmented view through a center of the turbocharger shown in  FIG. 2 . 
         FIG. 4  is an enlarged detail view of the turbocharger bearings shown in  FIG. 3 . 
         FIGS. 5 and 6  are enlarged detailed views of seals at both ends of the shaft of the turbocharger shown in  FIG. 3 . 
         FIG. 7  is an illustration of the fragmented view of  FIG. 3  showing flow paths of oil through the bearing housing of the turbocharger shown in  FIG. 2 . 
         FIG. 8  is an enlarged detail of  FIG. 7 . 
         FIG. 9  is a fragmented view of two turbocharger bearings in accordance with the disclosure. 
         FIGS. 10 and 11  are graphical representations of roto-dynamics for a turbocharger in accordance with the disclosure. 
         FIGS. 12-15  are illustrations of a bearing housing assembly process in accordance with the disclosure. 
     
    
    
     DETAILED DESCRIPTION 
     This disclosure relates to an improved turbocharger used in conjunction with an internal combustion engine to promote the engine&#39;s efficient operation and also the robust and reliable operation of the turbocharger. A simplified block diagram of an engine  100  is shown in  FIG. 1 . The engine  100  includes a cylinder case  104  that houses a plurality of combustion cylinders  106 . In the illustrated embodiment, six combustion cylinders are shown in an inline or “I” configuration, but any other number of cylinders arranged in a different configuration, such as a “V” configuration, may be used. The plurality of combustion cylinders  106  is fluidly connected via exhaust valves (not shown) to first exhaust conduit  108  and the second exhaust conduit  110 . Each of the first exhaust conduit  108  and the second exhaust conduit  110  is connected to a turbine  120  of a turbocharger  119 . In the illustrated embodiment, the turbine  120  includes a housing  122  having a gas inlet  124 , which is fluidly connected to the first exhaust conduit  108  and the second exhaust conduit  110  and arranged to receive exhaust gas therefrom. Exhaust gas provided to the turbine  120  causes a turbine wheel (not shown here) connected to a shaft  126  to rotate. Exhaust gas exits the housing  122  of the turbine  120  through an outlet  128 . The exhaust gas at the outlet  128  is optionally passed through other exhaust after-treatment components and systems such as an after-treatment device  130  that mechanically and chemically removes combustion byproducts from the exhaust gas stream, and/or a muffler  132  that dampens engine noise, before being expelled to the environment through a stack or tail pipe  134 . 
     Rotation of the shaft  126  causes a wheel (not shown here) of a compressor  136  to rotate. As shown, the compressor  136  can be an axial, radial or mixed flow compressor configured to receive a flow of fresh, filtered air from an air filter  138  through a compressor inlet  140 . Pressurized air at an outlet  142  of the compressor  136  is routed via a charge air conduit  144  to a charge air cooler  146  before being provided to an intake manifold  148  of the engine  100 . In the illustrated embodiment, air from the intake manifold  148  is routed to the combustion cylinders  106  where it is mixed with fuel and combusted to produce engine power. 
     An EGR system  102 , which is optional, includes an EGR cooler  150 , which is also optional, that is fluidly connected to an EGR gas supply port  152  of the first exhaust conduit  108 . A flow of exhaust gas from the first exhaust conduit  108  can pass through the EGR cooler  150  where it is cooled before being supplied to an EGR valve  154  via an EGR conduit  156 . The EGR valve  154  may be electronically controlled and configured to meter or control the flow rate of the gas passing through the EGR conduit  156 . An outlet of the EGR valve  154  is fluidly connected to the intake manifold  148  such that exhaust gas from the EGR conduit  156  may mix with compressed air from the charge air cooler  146  within the intake manifold  148  of the engine  100 . 
     The pressure of exhaust gas at the first exhaust conduit  108 , which is commonly referred to as back pressure, is higher than ambient pressure, in part, because of the flow restriction presented by the turbine  120 . For the same reason, a positive back pressure is present in the second exhaust conduit  110 . The pressure of the air or the air/EGR gas mixture in the intake manifold  148 , which is commonly referred to as boost pressure, is also higher than ambient because of the compression provided by the compressor  136 . In large part, the pressure difference between back pressure and boost pressure, coupled with the flow restriction and flow area of the components of the EGR system  102 , determine the maximum flow rate of EGR gas that may be achieved at various engine operating conditions. 
     An outline view of the turbocharger  119  is shown in  FIG. 2 , and a fragmented view is shown in  FIG. 3 . In reference to these figures, and in the description that follows, structures and features that are the same or similar to corresponding structures and features already described may be, at times, denoted by the same reference numerals as previously used for simplicity. As shown, the turbine  120  is connected to a bearing housing  202 . The bearing housing  202  surrounds a portion of the shaft  126  and includes bearings  242  and  243  disposed within a lubrication cavity  206  formed within the bearing housing  202 . The lubrication cavity  206  includes a lubricant inlet port  203  and a lubricant outlet opening  205  that accommodate a flow of lubrication fluid, for example, engine oil, therethrough to lubricate the bearings  242  and  243  as the shaft  126  rotates during engine operation. 
     The shaft  126  is connected to a turbine wheel  212  at one end and to a compressor wheel  213  at another end. The turbine wheel  212  is configured to rotate within a turbine housing  215  that is connected to the bearing housing  202 . The compressor wheel  213  is disposed to rotate within a compressor housing  217 . The turbine wheel  212  includes a plurality of blades  214  radially arranged around a hub  216 . The hub  216  is connected to an end of the shaft  126 . In the illustrated embodiment, the turbine wheel  212  is connected at the end of the shaft  126  by welding, but other methods, such as by use of a fastener, may be used to connect the turbine wheel to the shaft. The turbine wheel  212  is rotatably disposed between an exhaust turbine nozzle  230  defined within the turbine housing  215 . The exhaust turbine nozzle  230  provides exhaust gas to the turbine wheel  212  in a generally radially inward and axial direction relative to the shaft  126  and the blades  214  such that the turbine  120  is a mixed flow turbine, meaning, exhaust gas is provided to the turbine wheel in both radial and axial directions. Exhaust gas passing over the turbine wheel  212  exits the turbine housing  215  via an outlet bore  234  that is formed in the housing. The outlet bore  234  is fluidly connected to the outlet  128  ( FIG. 1 ). The exhaust turbine nozzle  230  is fluidly connected to an inlet gas passage  236  having a scrolled shape and formed in the turbine housing  215 . The inlet gas passage  236  fluidly interconnects the exhaust turbine nozzle  230  with the gas inlet  124  (also see  FIG. 1 ). It is noted that a single, inlet gas passage  236  is shown formed in the turbine housing  215  in  FIG. 3 , but in alternative embodiments separated passages may be formed in a single turbine housing. 
     In the embodiment shown in  FIG. 3 , the inlet gas passage  236  wraps around the area of the turbine wheel  212  and outlet bore  234  and is open to the exhaust turbine nozzle  230  around the entire periphery of the turbine wheel  212 . A cross sectional flow area of the inlet gas passage  236  decreases along a flow path of gas entering the turbine  120  via the gas inlet  124  and being provided to the turbine wheel  212  through the exhaust turbine nozzle  230 . 
     A radial nozzle ring  238 , which also forms a shroud for the turbine wheel  212 , is disposed substantially around the entire periphery of the turbine wheel  212 . As will be discussed in more detail in the paragraphs that follow, the radial nozzle ring  238  is disposed in fluid communication with the inlet gas passage  236  and defines the exhaust turbine nozzle  230  around the turbine wheel  212 . As shown in  FIG. 3 , the radial nozzle ring forms a plurality of vanes  246 , which are fixed and which are symmetrically disposed around the radial nozzle ring  238  and operate to direct exhaust gas form the inlet gas passage  236  towards the turbine wheel  212 . The shape and configuration of the plurality of vanes  246  can vary. Flow channels  250  having an inclined shape are defined between adjacent vanes in the first plurality of vanes  246 . A flow momentum of gas passing through the flow channels  250  is directed generally tangentially and radially inward towards an inner diameter of the turbine wheel  212  such that wheel rotation may be augmented. Although the vanes  246  further have a generally curved airfoil shape to minimize flow losses of gas passing over and between the vanes, thus providing respectively uniform inflow conditions to the turbine wheel, they also provide structural support to a shroud portion of the radial nozzle ring  238 . The radial nozzle ring  238 , which includes the shroud portion, is connected to the turbine via a plurality of fasteners  252 , but other methods can be used. The fasteners  252  engage a heat shield  254 , which is connected to a turbine flange  256  formed on the bearing housing  202  with an interference fit and stakes  258 . 
     The bearing housing  202  encloses a portion of the shaft  126 , which is rotationally mounted in a bearing bore  260  formed in the bearing housing by bearings  242  and  243 . Each of the bearings  242  and  243  includes an outer race  261  that engages an inner diameter surface of the bearing bore  260 , rollers, and an inner race  262  that has a generally tubular shape and extends around the shaft  126  along its length. Oil from the lubricant inlet port  203  is provided by an external oil pump to the bearings  242  and  243  during operation via passages  264 , from where it washes over the bearings to cool and lubricate them before collecting in the lubrication cavity  206  and draining out of the bearing housing through the lubricant outlet opening  205 . 
     The bearings  242  and  243  are axially retained within the bearing bore  260  by a bearing retainer  266  disposed between a compressor mounting plate  268  formed on the bearing housing  202  and the compressor wheel  213 . The bearing retainer  266  forms a central opening  270  having an inner diameter that is smaller than an inner diameter of the bearing bore  260  such that, when the bearing retainer  266  is connected to the bearing housing  202 , the bearings  242  and  243  are retained within the bearing bore  260 . The bearing retainer  266  is fastened to the compressor mounting plate  268  by fasteners  272 , but other fastening or retention structures may be used, and has a cylindrical portion that engages an outer race  261  axially. The engagement between the cylindrical portion (adjacent  288  as shown in  FIG. 3 ) can create, at times, a frictional engagement between the bearing housing and the outer race via the retaining plate that removes the necessity of using structures to otherwise prevent rotation of the outer race such as by use of a radially extending pin or similar structures. To prevent material wear at this interface, a nitride surface treatment may be applied to the annular end face of the bearing retainer  266  that abuts the outer race  261 , or a similar anti-wear treatment that can increase material hardness can be used. In the illustrated embodiment, the outer race  261  may in fact rotate relative to the bearing housing, at a rotation rate that is much slower than the shaft, during operation. Such rotation during operation may be dampened by the viscosity of the oil film that is present in the bearing surfaces B 1 , B 2 , B 3  and B 4  (see  FIG. 9 ). A component of this frictional engagement is also provided on a turbine side of the outer race, which abuts a stop surface at the end of the bearing bore  260 , as shown in  FIG. 6  (stop surface adjacent end of B 4 ), when the turbine is not operating. During operation, a gap of a few thousandths of an inch may form between the outer race and the stop surface that permits oil to pass therethrough. 
     The compressor  136  includes a compressor vane ring  274  that forms vanes  276  disposed radially around the compressor wheel  213 . The vanes  276  fluidly connect a compressor inlet bore  278 , which contains the compressor wheel  213 , with a compressor scroll passage  280  that is formed in the compressor housing  217  and that terminates to a compressor outlet opening  282 . Bolts  284  and circular plate segments  286  connect the turbine housing  215  to the turbine flange  256  and the compressor housing  217  to the compressor mounting plate  268 . A nut  288  engaged on the shaft  126  retains the shaft  126  within the bearings  242  and  243 . 
     An enlarged detailed view of the bearings  242  and  243  is shown in  FIG. 4 . In this illustration, and in the other illustrations that follow, structures that are the same or similar to structures previously described herein will be denoted by the same reference numerals previously used for simplicity. Accordingly, the first bearing  242 , which can also be referred to as the compressor-side bearing, is formed by a plurality of roller elements  302  that are confined in rolling or sliding motion between an outer race groove  304 , which is formed in the outer race  261 , and an inner race groove  306 , which is formed close to the compressor-side end of the inner race  262 . Similarly, the second bearing  243 , which can also be referred to as the turbine-side bearing, is formed by a plurality of roller elements  308  that are confined in rolling or sliding motion between a corresponding outer race groove  310  and inner race groove  312 . 
     The outer race  261  forms various features that facilitate operation of the turbocharger  119  and also promote a desirable flow of lubrication oil through the bearing housing  202 . More specifically, the outer race  261  has a generally hollow cylindrical shape that forms an outer wall or outer casing  314 . The outer casing  314  forms the outer race grooves  304  and  310  at its ends, and encloses a cylindrical space  316  that surrounds the shaft  126  and inner race  262  during operation. Close to either end, the outer casing  314  forms two oil collection grooves or oil feed galleys  318 , each of which is axially aligned with the passages  264  formed in the bearing housing  202  such that, during operation, oil flowing through the passages  264  collects and fills each of the two oil collection grooves or oil feed galleys  318 . Lubrication passages  320  extend through the outer casing  314  and fluidly connect each respective oil feed galley  318  with the cylindrical space  316  in an area close to the inner race grooves  306  and  312 , and also the outer race grooves  304  and  310 , to lubricate and cool the bearings  242  and  243  during operation. The outer casing  314  further forms drainage openings  322  that fluidly connect the cylindrical space  316  with the lubrication cavity  206  to drain out any oil collecting within the outer race  261 . 
     The outer race  261  contacts the bearing bore  260  along four cylindrical bearing surfaces, each of which has a diameter and axial length along a shaft centerline, C/L, that has been designed and selected for optimal bearing and dampening performance during operation. Accordingly, beginning from the compressor side of the outer race  261 , a first bearing surface B 1  has an outer diameter D 1  (see  FIG. 9 ) and extends along an axial length L 1 . A second bearing surface B 2  has a diameter D 2  ( FIG. 9 ) and an axial length L 2 . A third bearing surface B 3  has a diameter D 3  ( FIG. 9 ) and extends along an axial length L 3 . Finally, a fourth bearing surface B 4  has a diameter D 4  ( FIG. 9 ) and extends along an axial length L 4 . The bearing surfaces are also illustrated in  FIG. 9 . 
     Each of the four bearing surfaces B 1 , B 2 , B 3  and B 4  permits a thin film or a squeeze film diameter of oil therein having a thickness equal to a difference between the inner diameter D of the bearing bore  260  and the outer diameters D 1 , D 2 , D 3  and D 4 . As shown, the two bearing surfaces B 1  and B 2  that straddle the compressor-side oil feed gallery  318  have the same squeeze film diameter (SFD) and are considered together in terms of axial length (L 1 +L 2 ). Similarly, the two turbine-side bearing surfaces B 3  and B 4  have the same SFD and are considered together in terms of axial length (L 3 +L 4 ). As used herein, SFD is used to refer to those hollow cylindrical areas between each bearing surface and the bearing bore through which oil passes during operation. The thickness of the cylindrical areas or gaps are referred to as SFD clearance, while the length of each cylindrical area (the “height” of the cylindrical area) along the centerline of the shaft is referred to as SFD length. 
     For the compressor side bearing surfaces, B 1  and B 2 , a ratio of the SFD clearance over the diameter, which can be expressed as (Dx−D)/D, is equal to about 0.0021, where “x” is 1 or 2 and denotes D 1  or D 2 . For the same bearing surfaces, the SFD length over the diameter, which can be expressed as (L 1  or L 2 )/D, is equal to about 0.300. For the turbine side bearing surfaces B 3  and B 4 , a ratio of the SFD clearance over the diameter, which can be expressed as (Dx−D)/D, is equal to about 0.0031, where “x” is 3 or 4 and denotes D 3  or D 4 . For the same bearing surfaces, the SFD length over the diameter, which can be expressed as (L 3  or L 4 )/D, is equal to about 0.200. In other words, in the illustrated embodiment, the cylindrical areas through which oil flows during operation, which can act to dampen shaft vibrations and other excitations, are thinner and longer on the compressor side than on the turbine side, where they are thicker and shorter, thus providing different dampening characteristics. 
     During operation, oil provided through the passages fills and, to a certain extent, pressurizes the oil feed galleys  318 . Oil from the oil feed galleys  318  is pushed or passes into the SFDs of the bearing surfaces B 1 , B 2 , B 3  and B 4 , such that oil flows out from each oil feed galley  318  towards the compressor on one side, the turbine on an opposite side, and towards the center of the bearing housing on both sides. To promote oil flow through the inner bearing surfaces B 2  and B 3 , the oil flowing towards the center of the bearing housing  202  is collected by drainage grooves  324  (also see  FIG. 8 ), which are formed on an external surface of the outer race  261 , and which direct the oil into the lubrication cavity  206 . 
     The outer race  261  surrounds the inner race  262 , which in turn surrounds a portion of the shaft  126 . The inner race  262  forms two end portions  326  having a reduced diameter portion that engages the ends of the shaft  126 . The shaft  126  includes a slender portion  328  having a reduced outer diameter  330 , which is smaller than an increased outer diameter  332  at the ends of shaft  126 . The slender portion  328  extends over an axial length  334 . The increased outer diameter  332  of the shaft  126  mates at its ends with a reduced inner diameter  336  of the two end portions  326  of the inner race  262 . 
     To provide torsional and bending rigidity to the shaft  126 , the inner race  262  is advantageously flared along a middle portion thereof to form an increased inner diameter  338 . The increased inner diameter  338  overlaps in an axial direction with the slender portion  328  to increase the bending stiffness of the combined structure of the shaft  126  and inner race  262  without considerably increasing the overall mass of the system. In the illustrated embodiment, to facilitate assembly, the inner race  262  is formed by two components, a compressor-side cup  340  and a turbine-side cup  342 . One of the cups, in this case the turbine-side cup  342 , forms a ledge and a wall that accepts therein the free, annular face of the compressor-side cup  340 . Together, the compressor-side cup  340  and the turbine-side cup  342  form the inner race  262  that has a central, flared portion  344  and two transition portions  346  connecting the flared portion  344  with the two end portions  326 . Smooth or chamfered transitions  350 , which avoid stress concentration, are provided between the end portions, the transition portions  346 , and the flared portion  344 , as shown in the enlarged detail of  FIG. 8 . In the illustrated embodiment, each chamfered transition  350 , which can be convex or concave, is formed at the same radius, but different radii can be used. 
     An enlarged detail view of an interface between the compressor wheel  213  and the shaft  126  is shown in  FIG. 5 . In this figure, a diagnostic passage  402  formed in the bearing housing  202  can be seen. The diagnostic passage  402  is plugged with a plug  404 , which can be removed during service provide access, for example, to the interior of the bearing housing for installation of instrumentation and/or access to the interior of the bearing housing. 
     As can also be seen in  FIG. 5 , a ring seal  406  is disposed to provide a sliding seal between an internal, working chamber of the compressor and the oil cavity of the bearing housing. More specifically, the ring seal  406  is disposed in an open channel  408  that, together with an annular surface  410  on the inner side of the back of the compressor wheel  213 , forms a U-shape. The open channel  408  is formed at the end of an extension of the inner race  262  that is disposed on a compressor-side of the bearing  242 . The ring seal  406  slidably and sealably engages an inner bore  412  of the bearing retainer  266  such that a sliding seal is provided between the inner race  262  and the bearing retainer  266  that provides sealing against leakage of oil from the bearing housing  202  into the compressor housing  217 . In addition, the ring seal  406  provides sealing against pressurized gas from entering the interior of the bearing housing. A bearing retainer seal  414  is disposed between an outer portion of the bearing retainer  266  and the compressor mounting plate  268 . It is noted that an interior  348  ( FIG. 4 ) of the inner race  262  is expected to be generally free of oil as no entry openings for oil are provided except, perhaps, the interface between the compressor-side cup  340  and the turbine-side cup  342 . In the event of turbocharger failure, in a condition when the shaft  126  may be pulled towards the turbine housing, the retention nut  288  may be pulled towards and sealably engage a seat  424 , to keep the piston rings engaged and retain the turbine wheel and shaft assembly within the bearing housing. 
     In the illustrated embodiment, a tortuous path is also provided to discourage oil flow towards the ring seal  406 . As shown, the end of the inner race  262  forms a radially outward extending portion  416  that slopes away from the shaft  126 . The outward extending portion forms an outer tip portion  418  that is shaped as a cylindrical wall extending towards the compressor. The bearing retainer  266  forms an inwardly facing cylindrical wall  420  that is axially aligned with the outer tip portion  418  and disposed radially inward therefrom such that a meandering or tortuous path  422  is formed therebetween leading up to the ring seal  406 . 
     An enlarged detail view of an interface between the turbine wheel  212  and the bearing housing  202  is shown in  FIG. 5 . In this figure, a drainage groove  502  is formed towards an end  504  of the shaft  126  to facilitate drainage of oil passing through the innermost bearing surface B 4  into the scavenge oil gallery. To seal against leakage of oil, and to provide sealing against pressurized gas from entering the interior of the bearing housing, two ring seals are provided between the shaft  126  and an inner bore  506  of the turbine flange  256 . More specifically, a first ring seal  508  is disposed in a channel  510  formed in the shaft  126 , and a second ring seal  512  is disposed in a channel  514 , which is also formed in the shaft  126 . 
     During operation, oil from within the bearing housing  202  is discouraged from leakage into the working chamber of the turbine by the sliding and sealing contact of the first ring seal  508  and the second ring seal  512  with the shaft  126  and the inner bore  506  of the turbine flange  256 . It is noted that, in the event of a failure in the turbocharger during which the shaft  126  may displace towards the turbine, at least the first ring seal  508  can axially displace within the inner bore  506  for a predetermined distance while still maintaining contact therewith to provide a seal even under a failure mode, to avoid leakage of oil into the turbine housing. The same sliding tolerance is provided in the even the shaft  126  displaces towards the compressor, in which case the second ring seal  512  can displace within the inner bore  506  while still maintaining its sealing function. The ring seals shown herein are advantageously made of a hardened material such as M2 Steel having a yield stress of about 3,247 MPa (471,000 ksi) and can withstand temperature differences between the ring and surrounding components of about 450 deg. F. In each instance, the rings have a rectangular cross section, but other cross sections can be used, and have a C-shape that can be installed in a channel formed in a shaft to provide a spring-load against a sealably sliding surface that the ring engages. 
     A simplified oil flow diagram is shown in  FIG. 7 , where the structures shown in  FIG. 4  are used for illustration of the flow paths. In one embodiment, a main oil flow  519  is provided at the lubricant inlet port  203 . At a point A, the supply pressure and flow of oil splits into the passages  264  to reach the oil feed galleys  318 . Point B is taken to describe oil pressure in the oil feed galley  318  disposed on the compressor side (left side of the figure), and point C is taken to describe oil pressure in the oil feed galley  318  disposed on the turbine side (right side of the figure). Oil from the oil feed galleys  318  passes through the bearing surfaces as previously described, and drains into the lubrication cavity  206 . For purpose of description, point E is taken in bearing B 1 , and point F is taken in bearing B 4 . Table 1 below illustrates oil flow rates in gallons per minute (GPM) at different operating pressures (low, medium and high, depending on engine speed), and temperatures (cold and hot oil), which are representative of typical engine operating conditions: 
                     TABLE 1                  Oil Flow Data (GPM)                                         Hot Oil   Hot Oil   Cold Oil           Point   Low Pressure   Medium Pressure   High Pressure                       A   0.9   1.6   0.040           B   0.2   0.3   0.003           C   0.2   0.3   0.004           D   0.1   0.2   0.001           E   0.8   0.2   0.001                        
As can be seen from the above table, the larger gap at point E accounts for more flow of oil towards the turbine, which promotes more effective cooling. In the above table, hot oil can be anywhere within a normal oil temperature operating range for an engine such as between 190 and 230 deg. F., and cold oil can be anywhere in a cold start engine operating range such as between −30 and 0 deg. F. Similarly, low pressure can be between 20 and 40 PSI, medium pressure can be between 50 and 75 PSI, and high pressure can be between 90 and 120 PSI.
 
     As discussed above, oil passing through the bearing surfaces B 1  and B 2  on the compressor side, and bearing surfaces B 3  and B 4  on the turbine side (see  FIG. 9 ), help dampen vibrations and imbalances during operation. Such imbalances are advantageously controlled by selecting different oil film thicknesses on both sides of the shaft, which control the shaft dynamics to have natural vibration frequencies beyond the operating range of the engine. For example, for an engine operating at higher speeds and loads, the natural vibration frequencies or at least their prevalent harmonics are configured to occur above or below the expected range of engine operation. In the present embodiment, the difference between D 1  and D 2  with D 3  and D 4  in the bearing surfaces B 1 , B 2 , B 3  and B 4  produce the desired characteristics. 
       FIGS. 10 and 11  show graphical representations of the vibration characteristics of a turbocharger in accordance with the present disclosure, which was operated to sweep shaft rotation speeds using both hot oil, for example, oil at a normal operating temperature, and cold oil. As can be seen from the above table, the amount of oil flowing through the bearing areas, and also its viscosity, will change with temperature thus yielding different dampening characteristics against vibration. The vibration characteristics can be quantified from many different aspects, including a shaft displacement as a percentage of the displacement measured, observed or expected with respect to the bearing diameter at the bearing areas, averaged over the four bearing areas. 
     The results of a shaft speed sweep on shaft displacement using hot oil are shown in  FIG. 10 , where shaft speed  516 , as a percentage of maximum speed, is plotted along the horizontal axis, and percentage displacement  518 , expressed in (%), of a displacement distance with respect to the bearing diameter, is plotted along the vertical axis. Two curves are shown, the dashed lines representing a compressor response curve  520  and the solid line representing a turbine response curve  522 . The compressor response curve  520  represents a collection of points showing the percentage displacement  518  of each test point and the corresponding shaft speed  516  over a range of shaft speeds taken at the compressor wheel (e.g., compressor wheel  213 ,  FIG. 3 ). Similarly, the turbine response curve  522  represents a collection of points showing the percentage displacement  518  of each test point and the corresponding shaft speed  516  over a range of shaft speeds taken at the turbine wheel (e.g., turbine wheel  212 ,  FIG. 3 ). The same curves plotted against the same parameters, but for cold oil, are shown in  FIG. 11   
     As can be seen from the graphs in  FIGS. 10 and 11 , when the lubricating oil is warm, a peak vibration of just over 2% can occur at the compressor wheel speed below 10% of the maximum speed, as denoted by point  524  on the graph, and at about that same shaft speed, a vibration with a much lower displacement percentage of about 0.5% can occur at the turbine wheel, as denoted by point  526 . As can be seen by the compressor response curve  520  in  FIG. 10 , the percent displacement over a range of shaft speeds between 10% and about 85% of maximum speed, which accounts for most of the engine&#39;s operating range, remains constant at less than 1% for the compressor wheel. The turbine response curve  522  shows even better vibration profiles of a relatively constant peak displacement of less than 0.5% over a speed range between 10% and 100% of the maximum speed. 
     When the lubricating oil is cold, as shown in  FIG. 11 , a peak vibration of about 7% can occur at the turbine wheel at around 50%, as denoted by point  532  on the graph, and at about that same shaft speed, a vibration with a much lower displacement percentage of about 4.4% can occur at the compressor wheel, as denoted by point  530 . At a speed of about 5%, similar peaks as those seen in the hot oil condition ( FIG. 10 ) can be seen, with the compressor wheel having a peak displacement percentage of about 3.5%, as denoted by point  534 , and the turbine wheel having a peak displacement percentage of about 1%, as denoted by point  526 . In both cases, the peak displacement at the 5% speed with cold oil is about double that of hot oil. 
     As the shaft speed increases, still using cold oil ( FIG. 11 ), the percent displacement over a range of shaft speeds between 55% and about 115%, which accounts for most of the engine&#39;s operating range, remains constant at less than 1% for the turbine wheel. The compressor response curve  520  shows even better vibration profiles of a relatively constant peak displacement of about than 0.5% over a range between 55% and 115%. With these vibration profiles, shaft roto-dynamics is acceptable until the oil warms up, and then settles to a low peak displacement of less than 1% over the expected engine operating range. It is noted that, on the graphs of  FIGS. 10 and 11 , idle engine speed may be about 10% of the ranges shown in the chart. 
     When assembling a turbocharger in accordance with the disclosure, and especially when putting together an assembly of the bearing housing  202 , certain process steps may be carried out using a fixture, as shown in  FIGS. 12-15 . In  FIG. 12 , an assembly of the turbine wheel  212  welded to an end of the shaft  126  is mounted on a fixture  602  in a vertical position with the turbine wheel at the bottom. After the first ring seal  508  and the second ring seal  512  ( FIG. 6 ) are installed on the shaft, the bearing housing  202 , which has the heat shield  254  already installed, is inserted around the shaft  126  until the turbine flange  256  rests on a second fixture  604 , thus setting a proper distance between the turbine flange  256  and the turbine wheel  212 , as shown in  FIG. 13 . 
     Various components including the outer race  261 , inner race  263  and bearings  242  and  243  are inserted into the bearing bore  260  around the shaft  126  and, after various seals are installed, the bearing retainer  266  is assembled to close the bearing housing  202  and set a proper concentricity between the shaft  126  and the bearing bore  260 , as shown in  FIG. 14 . The compressor wheel  213  is then installed on the free end of the shaft  126 , as shown in  FIG. 15 . In the illustrated assembly sequence, the subassembly of the turbine wheel  212  onto the end of the shaft  126  may be rotationally balanced before assembly of the turbine is undertaken such that the shaft can determine the concentricity of the remaining components assembled thereafter, including the compressor wheel  213 , to maintain a balanced assembly. As an optional step, the entire assembly may be trim balanced after assembly to reduce imbalances, especially those imbalances that may be present when operating with cold oil. Trim balancing may be accomplished by removing material from the compressor wheel at the central hub and/or at the tips of the compressor blades. To determine the amount of material to be removed and the location for such removal, the entire assembly may be placed on a rotation balancing machine. It is further noted that the engagement of the radial seal within the inner bore of the bearing retainer, which helps place the shaft concentrically into the bearing bore, also reduces the amount of material that must be removed to balance the assembly when compared to turbochargers having a different sealing arrangement than what is shown herein. 
     INDUSTRIAL APPLICABILITY 
     It will be appreciated that the foregoing description provides examples of the disclosed system and technique. However, it is contemplated that other implementations of the disclosure may differ in detail from the foregoing examples. All references to the disclosure or examples thereof are intended to reference the particular example being discussed at that point and are not intended to imply any limitation as to the scope of the disclosure more generally. All language of distinction and disparagement with respect to certain features is intended to indicate a lack of preference for those features, but not to exclude such from the scope of the disclosure entirely unless otherwise indicated. 
     Recitation of ranges of values herein are merely intended to serve as a shorthand method of referring individually to each separate value falling within the range, unless otherwise indicated herein, and each separate value is incorporated into the specification as if it were individually recited herein. All methods described herein can be performed in any suitable order unless otherwise indicated herein or otherwise clearly contradicted by context.