Turbocharger bearing and lubrication system

This invention relates generally to turbochargers, systems and methods, and particularly to ball bearings used within turbochargers. The standard turbocharger arrangement consists basically of an exhaust gas driven turbine wheel on a common shaft with a compressor wheel. The manner in which the shaft is supported is critically important. A properly constructed bearing supports the shaft in such a manner as to minimize rotational friction. Other characteristics of the bearing scheme include isolation of the gases in contact with the turbine and compressor wheels, minimization of heat transfer from the hot turbine gases and provision of adequate lubrication to combat wear.

The turbocharger of the present invention employs a ball bearing to rotationally support the shaft and reduce friction. The high rotational speed of turbochargers have heretofore tended to prohibit the use of ball bearings in the design. This is because of the extreme centrifugal fores of the orbiting balls can exert more pressure on the outer ball bearing race than the inner ball bearing race exerts on the orbiting balls. In this case, the balls will separate from contact with the inner race, allowing the shaft to wobble. At the high rotational speeds of turbochargers, mainly on the order of from 100,000-125,000 RPM, this problem is severe. The manner in which the outer races are preloaded substantially eliminates this problem in the present invention.

Another major problem involving the turbochargers is the cost and difficulty of manufacture. In building a turbocharger, a number of parts must be cast, machined and then assembled. The time and effort required to assemble the critical bearings support parts during overall turbocharger construction, in good working order, is especially great. This relationship also exists for other part groupings, such as the lubrication parts group. The ability to assemble parts into cartridge-type assemblies prior to ultimate construction of the turbocharger will save time and labor costs.

Another problem is overheating of turbocharger internals from frictionally produced as well as exhaust heat. The present invention employs an air slinger to produce a cooling air flow along the shaft, countercurrent to the heat source.

Another problem encountered in turbocharger design involves the lubrication system. Turbochargers often use the same lubrication system of the engines which they serve. Engine lubricant systems usually allow abrasive matter in the lubricant which, given the close tolerancing of turbochargers, will lead to undue scoring and wear of the bearing surfaces. Normally a separate lubrication supply system requires a mechanical connection to the engine which the turbocharger serves. The separate lubrication system of the present invention is a compact, self-continaed, electrically driven cartridge type which can, in the event of failure, be easily replaced.

Further objects and advantages of the invention will become apparent from the following description and claims, and from the accompanying drawings wherein:

FIG. 1 is a cross-sectional view of a turbocharger showing the present invention;

FIG. 2 is an enlarged cross-sectional view of a portion of the turbocharger of FIG. 1 showing the improved ball bearing cartridge of the present invention;

FIG. 3 is an isometric view of the horseshoe shaped spacer assembly of the ball bearing cartridge of FIG. 2;

FIG. 4 is an enlarged isometric view of a portion of the turbocharger shaft of FIGS. 1 and 2;

FIG. 5 is a top sectional view partially cut away of the ball bearing cartridge of FIG. 2;

FIG. 6 is an enlarged cross-section of the end of the outer races of the ball bearing cartridge of FIG. 2, illustrating the centrifugal, preload, and resultant forces acting upon the ball bearing; and

FIG. 7 is a cross section of a turbocharger illustrating the cartridge-lubrication system of the present invention.

Referring to FIG. 1, the turbocharger has a turbine wheel 11 and a compressor wheel 10 which are joined by shaft 16. A portion of shaft 16 nearest turbine blade assembly 11 contains void 19 which reduces the area available for thermal conduction from turbine blade assembly 11 through shaft 16.

Now, referring to FIG. 2 for an enlarged view of the details of the bearing assembly, shaft 16 has a shaft piston ring slot 8 near turbine wheel 11. The turbocharger body 12 surrounds this portion of shaft 16, and shaft piston ring 7 is fitted in frictional contact with body 12 such that shaft piston ring 7 fits closely within shaft piston ring slot 8 on shaft 16. The resulting assembly will aid in the prevention of the hot exhaust gases, which drive turbine wheel 11, from entering the area near the turbocharger bearing area, and thus enhance thermal isolation of the hot side of the turbocharger.

Referring to FIG. 4 as well as FIG. 2, the turbine end of shaft 16 is fitted with a series of air passage fins 26 about the periphery of shaft 16, thereby forming air passage slots 25. Air passage fins 26 are equally spaced around shaft 6, and engage inner race 20 on its inner surface as shown in FIG. 2. Immediately adjacent air passage fins 26, inner race 20 is equipped with air slinger 27 having an air slinger bore 28. The turning of shaft 16, inner race 20, and air slinger 27 create a centrifugal force on the air within air slinger bore 28, causing it to pass out of air slinger bore 28 and creating an air flow along the outer surface of shaft 16 in the annular space between shaft 16 and the inner surface of inner race 20. Air enters this passage through air inlet passage 29 near the compressor end of the turbocharger. The air flow produces a countercurrent cooling effect on inner race 20, to compress the temperature gradient created in inner race 20 by virtue of its contact with the hot exhaust gases driving turbine wheel 11.

Referring to FIGS. 2 and 5, the inner race 20 supports the ball bearings generally designated as 30 for the compressor side of the turbocharger and 31 for the turbine side of the turbocharger. The ball bearings 30 and 31 ar e located within inner race groove 62 and inner race groove 63 respectively. Ball bearings 30 and 31 are separated peripherally about inner race 20 by ball bearing separators 60 and 61 respectively.

FIG. 6, which shows an enlarged view of compressor side outer race 14, the turbine side of inner race 20, and ball bearing 30, illustrates the forces at work under preloading. Compressor side outer race 14 has a shoulder 64, on its inner surface, to engage ball bearings 30, and outer race 15 has a shoulder 65, on its inner surface, to engage ball bearings 31. Under a preloaded, (non-operated) condition, compressor side outer race 14 is urged to the left, parallel to and relative to inner race 20. The inward sloping outer race shoulder 64 of outer race 14 engages ball bearings 30 against inner race goove 62 of inner race 20. Compressor side outer race 14 and inner race 20 are then connected by forces in ball bearings 30 which exist from a point labeled F P on compressor side outer race 14 to a point also labeled F P shown on inner race 20. Absent preloading, the centrifugal orbit of bal bearing 30 against compressor side outer race 14 would produce a centrifugal force when running labeled F C . The F C force, which is perpendicular to the axis of the shaft, would tend to press the inner surface of the outer race centrifugally outward. If any two races and ball bearings were designed without preloading, and even if the tolerances between the ball bearings and both of the other surfaces were set to zero, the centrifugal forces of the ball bearings 30 would push the compressor side outer race 14 radially outward, thus creating a gap between ball bearings 30 and inner race 20. Such a gap would allow the shaft to "wobble" and nutate at high speeds. In the case of a turbocharger where extremely high speeds are the norm, the destruction of the turbocharger would occur. Under conditions of preloading, ball bearings 30 at rest exhibits force F P against compressor side outer rae 14. As the bearings begin to turn, the ball bearings 30 attempt to exert an increasing centrifugal force against the internal surface of compressor side outer race 14 represented by FC. The combination of these two forces results in a resultant force F R shown on FIG. 6 as being angularly between F P and F C . As the speed increases, FC becomes larger with respect to F P , and the resultant force F R shifts counterclockwise between F P and F C . The leftward force on compressor side outer race 14 must always be great enough that the resultant force F R will never be close enough to F C to allow ball bearings 30 to break contact with inner race 20, or the compressor side outer race 14. In this manner, regardless of the wear on inner race 20, compressor side outer race 14, or ball bearings 30, the shaft is always kept in perfect alignment. Similarly, regardless of wear on the turbine side of inner race 20, turbine side outer race 15 or ball bearings 31, the shaft is always kept in perfect alignment.

To urge compressor side outer race 14 and turbine side outer race 15 away from each other, horseshoe spacer assembly 35 fits therebetween. Referring to FIG. 3, the horseshoe spacer assembly 35 is provided with preload spring pockets 36 containing preload springs 37 which act to force compressor side outer race 14 and turbine side outer race 15 axially away from each other. The length of horseshoe spacer assembly 35 is important because once fitted in place between compressor side outer race 14 and turbine side outer race 15, the total clearance between horseshoe spacer assembly 35 and both the compressor side outer race 14 and the turbine side outer race 15 will determine the amount of axial float allowed for the shaft 16. This clearance is chosen such that regardless of how much the outer races 14 and 15 shift axially, the multiple coil springs 37 are always enabled to exert the aforementioned sufficient amount of force on the compressor side outer race 14 and turbine side outer race 15 in order to keep the ball bearings 30 and 31 in full contact with inner race 20. The horseshoe spacer assembly 35 is held in place within body 12 by locating pin 9 which is fit through aperture 5 of body 12. Horseshoe spacer assembly 35 is equipped with a threaded bore 38 into which the locating pin threads 6 of locating pin 9 screwedly fit.

Referring again to FIG. 2, the compressor side of inner race 20 is equipped with slots 21 and 22. Piston ring 23 and piston ring 24 frictionally fit within the inner surface of body bore 13, and are adapted to fit within piston ring slot 21 and piston ring slot 22 respectively to aid in isolating the inside of the turbocharger of the present invention from the compressed fluid with which compressor wheel 10 is exposed. Oil drainage vent 39 is provided within body 12 between piston ring 23 and piston ring 24 in order to allow the escape of any of the compressor side fluid, if the fluid is able to creep past piston ring 23, before it is able to creep past piston ring 24.

The external surfaces of outer races 14 and 15 may be beveled to a slight frusto conical shape (FIG. 2), the frusto conical section at its smaller diameter further away from horseshoe shaped spacer 35, or they may be cylindrical. The slight frusto conical taper creates annular space 54 between the external surface of outer race 14 and the internal surface of body bore 13. Annular space 55 is created between the external surface of outer race 15 and the internal surface of body bore 13. This space, when occupied with a fluid of sufficient viscosity, preferably lubricating oil, creates a squeeze film damper which resists radial displacement of the bearing assembly within body bore 13. Since energy is required to "squeeze" the liquid film from one side of its annular space to the other, any exerted energy of the rotating shaft, which tends to angularly displace the axis of shaft 16 and its associated bearing parts, will be damped by the energy requirement of the shifting fluid of the squeeze film damper.

Compressor side outer race 14 contains outer race oil inlet aperture 40 and turbine side outer race 15 contains oil inlet aperture 41. Oil inlet aperture 40 is in fluid communication with body oil inlet aperture 42 and oil inlet aperture 41 is in fluid communication with body oil inlet aperture 43. Both body oil inlet apertures 42 and 43 are connected to oil supply channel 44. Oil supply channel 44 is connected tool supply channel 45 (also shown on FIG. 1).

Referring to FIG. 1, oil for lubrication and cooling is electrically pumped through oil supply channel 45, and then through oil supply channel 44, to be supplied through body oil inlet apertures 42 and 43, and oil inlet apertures 40 and 41, and then to flow into the annular space between inner race 20 and compressor side outer race 14 and turbine side outer race 15, to lubricate same. At its destination, the oil lubricates ball bearings 30 and 31, absorbs heat from inner race 20, and exits to drain area 48 from various points within the bearing structure.

The lubricant supply system is also illustrated in FIG. 1. Surrounded by body 12, oil supply channel 45 continues into a cylindrically shaped oil filter 47. Oil filter 47 is situated concentrically within oil cartridge 50 at the bottom of oil cartridge 50. Oil cartridge 50 is connected to body 12 by means of oil cartridge threads 51. The oil level shown at 49 is higher outside the periphery of filter 47 than inside the periphery of filter 47. Oil returning from lubrication and cooling surface within the turbocharger arrives outside of filter 47 and due to the higher oil level 49 outside of filter 47 then within it, the oil sets up a gravity-inducted height gradient within the walls of the filter 47 to provide the force for the transport of the oil through filter 47. Electrically driven pump 46, located concentrically within filter 47, pumps the filtered oil through oil supply channel 45 and back to the internals of the turbocharger for lubrication and temperature reduction service. The pump 46 is powered by any electrical source, which is usually the engine starting battery. While being filtered, the oil contained within oil cartridge 50 is cooled by heat conduction through the surface of oil cartridge 50, including ribs 58. FIG. 7 illustrates the cartridge-type lubrication system attached to a conventional exhaust driven turbocharger not having a ball bearing cartridge.

In the operation of the turbocharger of the present invention, hot exhaust gases are introduced into turbine housing 3 which drive turbine wheel 11. The action of preload springs 37 on compressor side outer race 14 and turbine side outer race 15 maintains full contact of the outer races 14 and 15 on ball bearings 30 and 31 and therefore full contact of ball bearings 30 and 31 on the inner race 20. This condition will hold, even when the diameter of the ball bearings 31 and 32 are reduced by wear. While the turbocharger is operating, the action of the dampening effect of squeezable clearance 54 and squeezable clearance 55 act to maintain the shaft 16 in alignment as the turbine wheel 11 is driven by the hot gases. Shaft 16 then turns compressor wheel 10 within compressor housing 4 to compress any fluid.

During operation, electrically driven pump 46 pumps oil, which has been filtered by filter 47, through oil supply channel 45, oil supply channel 44, body oil inlet apertures 42 and 43, oil inlet apertures 40 and 41, into the annular space between the turbine and compressor side outer races 14 and 15 respectively, and inner race 20 for the lubrication of ball bearings 30 and 31 or in the case of FIG. 7, to the conventional floating sleeve bearing.

While specific embodiments of this turbocharger with cartridge bearing assembly and cartridge lubrication system have been disclosed in the foregoing description, it is intended that many modifications and adaptations should and are intended to be comprehended within the meaning and range of this invention, without any such modifications and adaptations causing a departure from the spirit and scope of the invention.