JPS5914647B2 - Bearing mechanism for high-speed rotating shafts - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention is directed to a turbo charger.
The present invention relates to a bearing mechanism for a high-speed rotating shaft, such as an er) shaft.

This bearing mechanism includes a sleeve bearing and a ball bearing. The sleeve bearing supports radial loads near one end of the shaft and the ball bearing supports radial loads near the other end of the shaft. The shaft carries thrust loads in either direction, and the ball bearings also support thrust loads in at least one direction. Bearings for turbocharger rotors of internal combustion engines present very difficult technical problems.

This problem stems from the following fact. That is, the rotor shaft operates at high speeds, on the order of 125,000 RPM, and the shaft is subject to radial and axial thrust loads. Thrust loads are primarily in one direction, but in some cases axial loads may be present in the opposite direction. Additionally, the turbine end of the shaft can become very hot, and this heat can damage the bearing lubricant and cause changes in the properties of the bearing material. Also,
The shaft and rotor tend to move about a center of mass, usually not a geometric center, especially at certain critical (harmonic) speeds. Therefore, as the rotor rotates on its central axis, the rotor's central axis or geometric centerline oscillates in a circular path, and this oscillation presents another difficult problem. Of course, bearings have been provided that attempt to solve these special problems. For example, US Pat. No. 3,056,634 discloses a turbocharger including a floating sleeve bearing. A film of lubricant is present between the shaft and the sleeve bearing to support rotational movement of the shaft, and another film is present between the bearing and the bearing support or housing so that the bearing "floats". Adjust the rocking. Also, another sliding thrust bearing (thrust bearing)
is provided at one end of the sleeve bearing. US Pat. No. 3,043,636 also discloses a floating sleeve bearing, but in this patent the sliding thrust bearing is integral with the sleeve bearing. Although the bearings shown in these patents and other conventional bearings are operable, they suffer from drawbacks such as the excessive force losses that occur primarily in thrust bearings because the radial loading force is generally light. Frictionless bearings, such as ball bearings and roller bearings, have lower losses than hydrodynamic sleeve bearings, but to date have not been used in turbochargers.

If such bearings were used at the turbine end,
Heat may reduce bearing life or make bearings more expensive. Furthermore, when operating with light radial loads and high speeds, the balls or rollers tend to slide rather than roll;
Such operation can cause bearings to deteriorate rapidly. It is an overall object of the present invention to provide an improved bearing mechanism with relatively long life, low cost and low force losses.

This low power loss results in high efficiency and rapid response to changing engine operation. The device according to the invention consists of a bearing arrangement that supports a high speed rotating shaft within a non-rotating housing.

A hole is formed in the housing and a generally cylindrical sleeve is operatively positioned within the hole. The sleeve is floatingly supported by being surrounded by resilient attachment means between the sleeve and the housing. The sleeve has a circular port therethrough for rotatably receiving the shaft. Hydrodynamic bearing means are provided proximate and inside the one end of the sleeve to support radial loads from the shaft at the = end of the sleeve.

A rolling bearing is provided at the other end of the sleeve. Bearing support means are provided on the sleeve to support the rolling bearing. The rolling bearing supports radial loads from the shaft at the other end. Engagement means are provided on the rolling bearing and the support means. The engagement means are capable of engaging portions of the housing to support bidirectional axial thrust loads. Means is provided for cooperating with the housing to prevent rotation of the sleeve. The sleeve may be formed with a passageway leading to the bearing support means for conveying lubricating oil to the rolling bearing.

The flow path may include means for injecting lubricating oil into the rolling bearing. The sleeve and the hydrodynamic bearing means may be integrally formed and a passageway may be formed for conveying lubricating oil to the sleeve.

The means of the sleeve cooperating with the housing may have a non-circular outer shape of the bearing support means engageable with the housing to prevent rotation of the sleeve.

The rolling bearing has outer and inner races and a plurality of balls between the races, and at least one of the races can be split.

The inner race may be split into two parts and include a sleeve that is press fit within the two parts to hold the two parts in assembled relationship.

The inner and outer races may have relatively high radial spacing to withstand relatively large axial thrust loads.

The engagement means may include solid film lubricant means between the housing and the bearing assembly.

These and other objects and advantages of the present invention will become more apparent from the following detailed description taken in conjunction with the accompanying drawings.

Referring to FIG. 1, a turbocharger including a bearing mechanism constructed in accordance with the present invention has a housing 10 formed by a compressor casing 11, a turbine casing 12, and a solid bearing support 13. .

An oil-proof plate 14 is provided between the support body 13 and the casing 11. The support 13, the retaining plate 14 and both casings are fixed together by suitable fasteners (not shown). Compressor casing 11 receives compressor wheel 18 , and similarly turbine casing 12 surrounds turpin wheel 21 . The two wheels 18 and 21 are connected to the rotation axis 2
6, these parts 18, 21 and 2
6 forms a turbine rotor 25.

Turpin wheel 21 is fixed firmly at the left end (in FIG. 1), and compressor wheel 18 is located on shaft 26 and fixed by a nut 29. The bearing support or housing 13 includes a plurality of radial struts (Stnlt) 32
It includes a generally cylindrical part 31 supported by.

A hole 37 is formed in the component 31 to receive a bearing mechanism 38 that rotatably supports the shaft 26. A lubricating oil (liquid) passage 39 is formed in one of the struts 32 of the bearing housing 13 , at its radially outer end the passage 30 forms a lubricating oil inlet opening 41 . The support 13 has a generally cylindrical shape O with an opening 41 formed therein.
It has an outer wall 33. Further, a lubricating oil outlet 40 is formed on the lower side of the outer wall 33. During operation of the engine and turbocharger, the lubricating oil of the engine lubrication system is drawn under pressure into the inlet opening 41 and the passage 39;
After flowing through the tributaries 39a, 39b and 39c of 9 and lubricating the parts of the bearing mechanism 38 as described below, the housing 13
A pipe (not shown) falling into a lower chamber 42 and connected to an outlet 40 carries lubricating oil for the oil of the lubrication system. The bearing mechanism 38 includes a sleeve (cylindrical body) 51 and a ball bearing 5.
Contains 2.

The bore 37 of the bearing housing 13 receives the cylindrical outer surface 53. The cylinder 51 is connected to the shaft 2 only at the end of the pinwheel 21.
Close to the outer surface of 6, hydrodynamic (HydrednErr
llc) forming the bearing portion 54; Except for this portion 54, the inside of the cylinder 51 forms a concave shape 56, and an interval space 57.
form. A plurality of inclined holes 58 (FIGS. 1 and 3) are formed in the cylinder 5.
1 , the outer end of the hole 58 is located within an annular groove 59 in the outer surface of the cylinder 51 . Groove 59 is aligned with passageway branch 39b as shown in FIG. Axial groove 61
are formed in the bearing 54 at the inner end of each hole 58. As a result, during operation of the turbocharger, the lubricating oil flows through the tributary 39b.
, through groove 59, hole 58, groove 61, and axially between shaft 26 and portion 54 of the bearing. As is well known to those skilled in the bearing art, this lubricant creates a lubricant film between the shaft and the bearing. A plurality of radial holes 66 and an annular groove 67 are provided on the outer surface of the radial holes 66 approximately in the middle of the entire length of the cylindrical body 51. The hole 66 allows the lubricating oil flowing into the spacing space 57 to flow into the groove 67 and into the bearing housing 13.
It exits the bearing mechanism through a hole 68 (FIG. 1) formed in the underside of the bearing mechanism. At the compressor side end of the bearing mechanism 38, a thrust bearing support portion 71 is formed in the cylindrical body 51.

As best shown in FIG. 3, portion 71 is radially enlarged and has a circular inner opening 72. As best shown in FIG. The bearing mechanism 38 is prevented from rotating within the bearing housing 13, and in this example:
To achieve this, a square-like non-circular (0ut-0f-ROld) contour is formed on the outside 73 of the enlarged portion 71,
The retaining plate 14 is provided with a flange 74 that engages with the outside 73. As shown in FIGS. 1 and 5, the oil-damping plate 14 has at least two axial flanges 74, which rest on the outer surface of at least two parts of the outer side 73. As shown in FIGS. Oil stop plate 14
Since is a stationary part of the turbocharger housing, flange 74 prevents rotation of the bearing mechanism. However, there is a slight spacing between the outer side γ3 and the flange R4, so that some radial movement of the bearing mechanism is possible. As mentioned above, instead of being square, the portion 71 may have another non-circular shape or may be provided with an entirely different type of device to prevent rotation, such as a pin and mating hole mechanism. As shown in FIG.
is provided within the circular opening 72 of the portion 71.

The bearing 52 includes a circular outer race 81 and an inner race 82, and a plurality of balls 83 provided within a cage. The outer race 31 is secured to the portion 71 within the opening 72 by, for example, a brace fit, and the turbine side of the outer race 81 abuts the bottom wall 84 of the opening 72 . The oil plate 14 includes another plurality of axial flanges 86 in close contact with the compressor side of the outer race 81 . The outer race 81 of the ball pairing is therefore prevented from rotating and moving axially in both directions. In the particular example illustrated, the oil plate 14 has two straight flanges 74 (FIG. 5) spaced 180 degrees apart and radially inwardly spaced apart from the flanges 74 by 90 degrees. Two arcuate flanges 86 are formed.

Preferably, the bearing assembly 38 also includes a cylindrical or annular spacer 37 that is press fit into the inner race of the bearing 52 and retains it as described below.

An annular spacer 87 is located on the shaft 26. Spacer 87
The turbine end of the turbine has a radial shoulder 88 formed on the shaft 26.
Abutting this end of the spacer is formed with a radially outwardly flared flange 89 that fits between shoulder 88 and inner race 82 . A cylinder 91 is positioned around the shaft 26 between the ball bearing 52 and the compressor wheel 18, and the cylinder 91 abuts the inner race 82. As a result, when the nut 29 is tightened with the shaft 26, the spacer 87 and the inner race 82 are firmly fixed between the shoulder 83 and the tube 91. The cylinder 91 fits snugly onto the shaft 26,
Coincides with the oil plate 14 in the radial direction.

An annular groove 93 on the outside of the cylinder 91 receives a sealing ring 94 that prevents lubricating oil from leaking. Oil drain 96 on the outside of the cylinder 91
releases the lubricating oil from the parking area. A branch 39c of the lubricating oil passage 39 is connected to supply lubricating oil to the ball bearing 52.

The tributary 39c slopes from the radial passage 39 towards the compressor end of the body 51. An annular groove 101 on the outer surface of the cylinder 51 communicates with the tributary stream 39c, and an axial passage 102 extends from the groove 101 to the bottom wall 84 of the opening 72. Insert 103 is in passage 102
Preferably, it is provided in a small diameter hole 104.
has. The hole 104 roughly corresponds to the position of the ball 83, and during operation, lubricating oil is injected from the hole 104 to the ball 83.
3 Sprayed on the soil. Also, a plurality of spaced holes 10 are provided in the wall of portion 71 for draining lubricating oil through opening 72.
6 is formed. A third branch 39a of the passageway 39 slopes radially outward and towards the turbine side of the bearing housing and jets lubricating oil against the adjacent wall 105 of the bearing support to direct the turbine side of the turbocharger. Cooling.

During operation of the turbocharger, the axial thrust load on shaft 26 is normally in the direction of compressor wheel 18, but in some cases this thrust load may change direction.

The ball bearing 52 is connected to the outer race 81. inner lace 82
Alternatively, by dividing both races, it can be designed to withstand thrust loads in either direction. In the particular example shown, only the inner race 82 is split to have two portions 108 and 109. These two parts are attached to the flange 8 as shown.
9 and the cylinder 91, they are tightly clamped together. Before mounting the bearing 52 on the turbocharger, the two parts 108 and 109 are held assembled by the cylinder 87. The inner corner 111 of each section has an arcuate cutout for receiving the ball 83. Also outside lace 8
1 has an arcuate annular groove 112 for receiving the ball. By splitting the inner race 82 as shown, the ball pairing can be configured to withstand relatively high axial and radial loads in both directions.

To assemble the ball bearing 52, first attach the balls 83 and ball retainer to the outer race 81, and since there is sufficient play between the balls and the retainer, the balls fit into the grooves 112.
Press-fitted inside. And the two parts 108 of the inner lace 82
and 109 are attached to both sides of the ball. This arrangement allows grooves 111 and 112 to have relatively large and strong radial shoulders that can withstand axial loads in either direction. During operation of the turbocharger, lubricating oil flows through the passage 39.
and flowing through tributaries 89a, 89b and 39c,
Shaft 26, wheels 18 and 20, cylinder 91, spacer 87
, and the inner race of the ball bearing 52 rotate.

The lubricating oil flowing through the tributary 39b, the hole 58 and the groove 61 forms a film between the barrel bearing 51 and the shaft, and the barrel bearing portion 54 supports the radial loads at the turbine end of the shaft 26. . A bearing mechanism 38 is also resiliently installed in the support 13. Various resilient mounts (MOunts) can be provided, such as a thin layer of resilient material or a thin film of lubricating oil, although a thin lubricating oil film is preferred in this particular example. Lubricating oil is in the passage 39,
Flowing into grooves 59 and 101 through tributaries 39b and 39c, a membrane is formed between housing 51 and hole 37, and commercial body 51 floats in the bearing housing. When the rotor 25 rotating at high speed rotates near a critical speed, the rotor begins to swing around its center of mass. The oil film between the mechanism 38 and the support 13 forms an elastic or semi-rigid support and thus allows such rocking, but this film also has a viscous damping effect and the bearing mechanism Support 1
Isolate from 3. Bearing supports can be designed to prevent such rocking by holding the bearing mechanism securely, but doing so can result in high forces being applied to the component, especially at critical speeds, which can lead to failure. Become. This elastic or floating attachment therefore allows the rotor to pass through the critical speed zone without creating large forces. However, the rotation of the bearing mechanism is caused by the side T3 of the square.
and are blocked by the flange 74. Of course, lubricating oil also passes through passageway 102 to lubricate ball bearing 52, which supports radial loads at the compressor end of shaft 26.

The path of the radial load at this end passes through the spacer 87, the ball bearing 52, and the bearing support part 71 to the hole 37 near the support part 71. If the axial load is directed toward the compressor wheel 18, the load path is directed toward the shoulders or ridges 88, the flange 89 of the spacer 37, the portion 108 of the inner race, the ball 83, the outer race 81, and the flange 86. Passes through the end face.

If the axial load is in the opposite direction, the load path is from the wheel 18 to the cylinder 91. Inner lace part 109, ball 83
, outer lace 81. It passes through the bottom wall 84 of the bearing support part 71 and reaches the adjacent radial surface 113 of the bearing housing 13 . However, some axial clearance is required to obtain the radial damping effect. The direction of the thrust applied to the rotor 25 is normally to the right in FIG. 1, and this thrust load is supported by the oil stop plate 14.

The load path is from the rotor 25 to the bearing mechanism and to the arcuate flange 86. Therefore, the right side surface of the outer race 81 of the ball bearing in FIG.
Press it against the left end surface of 6. As previously mentioned, the floating arrangement of the bearing mechanism allows it to oscillate about its geometric center, and this motion, in combination with the thrust load, causes wear (
Galling and fretting may occur. Failures due to such wear and tear are caused by
This can be prevented by increasing the contact area between bearing 52 and flange 86 or by selecting materials for these parts that will withstand fraying and digging under these conditions. The modified embodiment of the invention shown in FIGS. 6 and 7 includes means to prevent fraying and digging as described above. The device shown in FIGS. 6 and 7 includes a solid bearing support 120 and an oil stop plate 121.
, a compressor wheel 122, a rotor shaft 123, and a barrel 124, which parts are the same as the pair of core parts 13, 14, 18, 26 and 91 shown in FIGS. 1-5. Furthermore, a bearing mechanism 126 is provided, which includes a barrel 127 and a ball bearing 128, the barrel 127 being resiliently mounted in the bore 129 of the support 120 by means of a lubricating oil film, as previously described. A cylinder 127 and a ball bearing 128 support the shaft 123, and lubricating oil flows through these parts similar to the arrangement shown in FIGS. 1-5. Means are provided for preventing rotation of the bearing arrangement, which means take the form of a non-circular outer surface 131 of the bearing arrangement 126 and a flange 132 on the oil stop plate 121.

The axial thrust load of the rotor 125 is normally to the right in FIG.
7, and they support thrust loads. To prevent fraying and digging of the adjacent thrust bearing surfaces of the bearing mechanism and flange 138, a solid film lubricant 136 is provided between the adjacent surfaces. The lubricant film 136 includes a thin film or sheet of Teflon infused with glass fibers. It preferably forms part of the bearing arrangement 126 and substantially covers the outer race 134 and the right end face of the barrel 127. As previously mentioned, in the particular example of the invention shown and described, as shown in FIG. 7, the outer surface 131 of the barrel 127 is square and the outer shape of the membrane 136 is also square. The outer edge of membrane 136 is bent axially, as indicated at 137 in FIG. 6, and extends between outer surface 131 of the barrel and flange 132. Membrane 136 is formed with a central circular aperture 138 that extends to the outer circumference of the ball of bearing 128 and thus also onto outer race 134 of bearing 128 . membrane 13
6 can be fixed, for example, to a bearing mechanism or to an oil holder plate by means of a suitable adhesive. The right end surfaces of the cylinder 127 and the outer race 134 are preferably coplanar to increase the bearing surface area for thrust loads. Membrane 136 is particularly advantageous because it is somewhat compressible so that thrust loads can be distributed to the end faces of the barrel and outer race. During operation of the turbocharger, lubricating fluid flows around membrane 136 from passages formed in bearing support 120 and bearing mechanism 126.

The membrane 136 is attached to the rotor 1 between the bearing mechanism and the flange 133.
Compressed by the axial thrust load on 25, the rotor also moves radially, but the bearing mechanism and flanges 132 and 13
3 and the solid film of lubricant 136 prevents the adjoining metal surfaces from wearing out and digging into each other. A membrane similar to membrane 136 is provided in conventional turbochargers, but without the bearing mechanism and seal flange as disclosed herein. Membrane 136 extends across the junction of barrel 127 and outer race 134 and is compressible to evenly distribute thrust loads and increase thrust loading area. Furthermore, the edge 137 of the membrane 136 extends between the surface 131 of the cylinder 12r and the flange 132 to prevent fraying and digging into these surfaces. As can be seen from the foregoing, an improved bearing mechanism is provided.

Thrust loads, which are the source of most bearing friction losses in conventional turbochargers, are supported by low-loss rolling bearings, which also support radial loads at the compressor end of the shaft. Ball bearings are away from the hot turbine, preventing lubrication problems. In addition, lubricating fluid is continuously injected into the ball bearing to cool and lubricate it. Because ball bearings are kept relatively cool, bearings containing cheaper materials and heat treatments can be used and have a longer life. The axial load is also
Prevents ball slippage during low radial loads. A solid film of lubricant may be provided between the metal parts to ensure that the metal parts do not wear and dig due to rocking combined with axial loads. By using split races in ball bearings, it is possible to construct ball bearings that can withstand relatively large thrust loads in either direction. Due to the low force losses of the ball bearings, the turbocharger's efficiency is high and it can respond quickly to increased airflow demands. The bearing mechanism may be part of the turbocharger's original equipment, or the mechanism may be used as a replacement for the turbocharger's conventional bearings. Accurate bearing alignment is achieved by providing hydrodynamic enclosure bearings and hall bearings as part of a single mechanism.
), and the proper spacing required for bearings designed for high-speed, high-speed operation. The axial thrust load passes through the rolling bearing and applies tension to the balls between the inner and outer races, preventing the bearing from slipping and extending the life of the bearing.

Conventional bearing structures have a sleep pairing at each end of the rotating shaft and a thrust pairing for thrust in two directions, but the present invention uses an ordinary type of rolling bearing with a high shoulder. The cost of the device was reduced by reducing the number of bearings to two.

Rolling bearings inherently have very low friction losses, and since the bearing supports radial loads at the end of the bearing arrangement and axial loads in both directions, the bearing arrangement has low force losses.

[Brief explanation of the drawing]

1 is a sectional view of a turbocharger including a bearing mechanism constructed in accordance with the present invention; FIG. 2 is a sectional view of a bearing support of the turbocharger; and FIG. 3 is a sectional view of the bearing mechanism. 4 is an end view taken along line 4--4 of FIG. 3; FIG. 5 is an elevational view of another portion of the turbocharger; FIG. 6 shows a modification of the invention. FIG. 7 is a partial cross-sectional view taken along line 7--7 of FIG. 10...Housing, 11...Compressor casing, 12...Turbine casing, 1
3...Bearing support.

Claims (1)

Claims: 1. A bearing mechanism 38 for supporting a high-speed rotating shaft 26 in a non-rotating housing 10, in which a hole 37 is formed and a generally cylindrical sleeve 51 is operatively disposed within the hole 37. 26, the sleeve is floatingly supported by being surrounded by resilient mounting means between the sleeve and the housing, the sleeve having a circular opening therethrough for rotatably receiving the shaft 26; A rolling bearing 62 is provided at the other end of the sleeve 51 in an arrangement in which hydrodynamic bearing means 54 are provided proximate and inside the one end of the sleeve to support radial loads from the shaft at one end of the sleeve. is provided, the rolling bearing has an outer race 81 and an inner race 82, and bearing support means 71 are provided on the sleeve 51 for supporting the outer race of the rolling bearing;
The rolling bearing 52 supports radial loads from the shaft at the other end, the inner race 82 of the rolling bearing 52 engages the shaft 26 and the outer race 81 and the bearing support means 7
1 is a high-speed rotating shaft characterized in that means 73, 74 are provided which engage the housing to support axial thrust loads in both directions and cooperate with the housing to prevent rotation of the sleeve 51; Bearing mechanism for. 2. In the bearing mechanism according to claim 1, the sleeve 51 is formed with flow passages 101 and 102 that communicate with the bearing support means 71 in order to convey lubricating oil to the rolling bearing 52. 3. In the bearing mechanism according to claim 2, the flow passages 101, 102 include means 103, 104 for injecting lubricating oil into the rolling bearing 52. 4. In the bearing mechanism according to claim 1,
The sleeve 51 and the hydrodynamic bearing means 54 are integrally formed and the sleeve includes a passage 61 for carrying lubricating oil. 5. In the bearing mechanism according to claim 1,
The rolling bearing 52 has outer and inner races 81, 8
2 and a plurality of balls 83 between the races, and at least one of the races is divided. 6. In the bearing mechanism according to claim 5, the inner race 82 is split into two parts 108, 109, which are press-fitted into the two parts to bring the two parts into assembled relationship. A retaining sleeve 87 is included in the inner race. 7 In the bearing mechanism according to claim 5,
The races 81, 82 have relatively high radial shoulders to withstand relatively large axial thrust loads. 8. In the bearing arrangement according to claim 1, the engagement means includes solid film lubricant means 136 between the housing and the bearing arrangement.