阅读说明：本技术 具有流体动力滑动轴承的涡轮增压机或流体动力滑动轴承 (Turbocharger or hydrodynamic plain bearing having a hydrodynamic plain bearing ) 是由 M·贝格尔 R·克兰施米特 F·施泰特 O·库恩 S·施密特 于 2019-11-22 设计创作，主要内容包括：本发明涉及具有流体动力滑动轴承的涡轮增压机,流体动力滑动轴承具有转子(10)和为转子(10)配备的配对轴承部件(50),其中,转子(10)的转子支承面与配对轴承部件(50)的配对面相对,以便形成流体动力滑动轴承,其中,转子支承面和/或配对面在沿着并且通过旋转轴线(R)的剖面中在剖视图中形成连续的支承轮廓,该连续的支承轮廓由至少两个轮廓区段(44.1至44.3；53.1至53.3)形成,以便在径向方向以及轴向方向上产生流体动力的承载能力,其中,转子(10)的支承面由与转子轴(11)连接的转子部件(40)形成,转子部件保持在转子轴(11)上,并且其中,转子部件(40)在转子轴(11)的支撑区段(14)的区域中相对于转子轴(11)被支撑。为了能够提供具有紧凑构造且包括流体动力滑动轴承的高效的轴承组件的这种涡轮增压机,同时能够简单地以低的部件成本安装流体动力滑动轴承,根据本发明使得支撑区段(14)和配对轴承部件(50)的轮廓区段(53.1至53.3)中的其中至少一个沿旋转轴线(R)的方向至少局部地重合。(The invention relates to a turbocharger having a hydrodynamic plain bearing having a rotor (10) and a counter bearing part (50) associated with the rotor (10), wherein a rotor bearing surface of the rotor (10) is opposite a counter surface of the counter bearing part (50) in order to form the hydrodynamic plain bearing, wherein the rotor bearing surface and/or the counter surface in a sectional view along and through a rotational axis (R) form a continuous bearing contour which is formed by at least two contour sections (44.1 to 44.3; 53.1 to 53.3) in order to generate a hydrodynamic bearing capacity in the radial direction as well as in the axial direction, wherein the bearing surface of the rotor (10) is formed by a rotor part (40) connected to the rotor shaft (11), which is held on the rotor shaft (11), and wherein, the rotor part (40) is supported relative to the rotor shaft (11) in the region of the support section (14) of the rotor shaft (11). In order to be able to provide such a turbocharger with a compact design and with an efficient bearing assembly comprising a hydrodynamic plain bearing, while at the same time being able to install the hydrodynamic plain bearing simply at low component costs, according to the invention at least one of the support section (14) and the contour section (53.1 to 53.3) of the counter bearing component (50) coincides at least in regions in the direction of the axis of rotation (R).)

1. Turbocharger having a hydrodynamic, in particular hydrodynamic, plain bearing, or a hydrodynamic, in particular hydrodynamic, plain bearing, having a rotor (10) and a counter-bearing part (50) associated with the rotor (10),

wherein a rotor bearing surface of the rotor (10) is opposed to a mating surface of the mating bearing member (50) to form a hydrodynamic plain bearing,

wherein the rotor bearing surface and/or the counter surface form, in a sectional view along and through the axis of rotation (R), a continuous bearing contour which forms at least two contour sections (44.1 to 44.3; 53.1 to 53.3) which merge into one another in order to generate a hydrodynamic bearing capacity in the radial direction as well as in the axial direction,

wherein the bearing surface of the rotor (10) is formed by a rotor part (40) connected to a rotor shaft (11), which rotor part is held on the rotor shaft (11),

and wherein the rotor part (40) is supported relative to the rotor shaft (11) in the region of a support section (14) of the rotor shaft (11),

characterized in that at least one of the support section (14) and the profile section (53.1 to 53.3) of the counter bearing component (50) at least partially coincides in the direction of the axis of rotation (R).

2. Turbocharger or hydrodynamic plain bearing according to claim 1, characterized in that the rotor part (40) has a bearing region (46) which forms a profile section (44.1 to 44.3) and the bearing region (46) is arranged such that at least one of the profile sections (44.1 to 44.3) at least partially coincides with the support section (14) in the direction of the axis of rotation (R).

3. The turbocharger or the hydrodynamic plain bearing according to claim 1 or 2, characterized in that the radial clearance between the supporting section (14) of the rotor shaft (11) and the region of the rotor part (40) which bears against the supporting section (14) is less than the radial clearance between the rotor part (40) and the counter bearing part (50), preferably less than 80%, particularly preferably less than 60%, of the radial clearance between the rotor part (40) and the counter bearing part (50).

4. The turbocharger or the hydrodynamic plain bearing according to one of claims 1 to 3, characterized in that a press fit in the form of a transition fit is formed between the support section (14) of the rotor shaft (11) and the region of the rotor part (40) which bears against the support section (14), and the relative radial bearing play of the hydrodynamic plain bearing relative to the diameter of the support section (14) is in the range between minus 6 to plus 6 thousandths.

5. Turbocharger or hydrodynamic plain bearing according to one of claims 1 to 3, characterized in that the counter bearing component (50) is inserted into a bearing housing (60) or housing component in such a way that a preferably circumferential clearance region (57) is formed between the outer contour of the counter bearing component (50) and the bearing housing (60) or housing component for forming a trapped oil film, wherein the clearance region (57) is spatially connected with a lubricant guide channel (61) and preferably the clearance region (57) and the support section (14) thereby at least partially coincide in the direction of the axis of rotation (R).

6. Turbocharger or hydrodynamic plain bearing according to any of claims 1 to 5, characterized in that the radial clearance between the counter bearing part (50) and the bearing housing (60) or housing part is larger than the radial clearance between the rotor part (40) and the counter bearing part (50).

7. The turbocharger or the hydrodynamic plain bearing according to claim 5 or 6, characterized in that a relative gap extending in the radial direction relative to the outer diameter of the gap region (57) is formed between the counter bearing part (50) and the bearing housing (60) or the housing part in the gap region (57), which relative gap is in the range of 5 to 10 thousandths, particularly preferably in the range of 7 to 9 thousandths.

8. The turbocharger or the hydrodynamic plain bearing according to any one of claims 5 to 7, characterized in that the axial coincidence between the counter bearing component (50) and the bearing housing (60) or the housing component in the direction of the axis of rotation (R) is dimensioned in a clearance region (57) for the formation of a trapped oil film, the extension of the clearance region (57) in the direction of the axis of rotation and the radial clearance between the counter bearing component (50) and the bearing housing (60) or the housing component being such that the ratio of:

the axial extension/radial clearance of the clearance region (57) is 40 to 80,

it is particularly preferred, among others, for the ratio to be selected in the range between 45 and 70.

9. Turbocharger or hydrodynamic plain bearing according to any of claims 1 to 8, characterized in that the radial clearance between the counter bearing component (50) and the bearing housing (60) or housing component and the axial extension of the clearance region (57) are defined by the following relationship:

the axial extension in millimeters of the clearance region (57) is equal to 9 minus C times the radial clearance in millimeters between the mating bearing component (50) and the bearing housing (60) or housing component, where C is in the range of 61 to 75, preferably C is in the range of 66 to 70.

10. Turbocharger or hydrodynamic plain bearing according to one of claims 1 to 9, characterized in that the rotor component (40) has a base component (41) which is provided with at least one seal receptacle (42) into which a sealing element (43) is placed.

11. Turbocharger or hydrodynamic plain bearing according to one of claims 1 to 10, characterized in that the rotor part (40) has an appendage (45) which comprises a stop face and which stops surface-wise on a radially running face of a flange (12) of the rotor shaft (11).

12. Turbocharger or hydrodynamic plain bearing according to one of claims 1 to 11, characterized in that the rotor part (40) is tensioned between the compressor wheel (30) and the flange (12), preferably by means of a screw connection, so that it is fixed rotationally fixed in the axial direction and in the circumferential direction.

13. Turbocharger or hydrodynamic plain bearing according to one of claims 1 to 12, characterized in that the continuous bearing contour of the rotor bearing surface and/or of the counter surface, which contour is formed by contour segments (44.1 to 44.3; 53.1 to 53.3), in a sectional view along and through the axis of rotation (R), is differentiable continuously at least once.

14. Turbocharger or hydrodynamic plain bearing according to one of claims 1 to 13, characterized in that a preferably annular spacing space is formed between the rotor shaft (11) and the rotor part (40) in the direction of the axis of rotation (R) indirectly or directly beside the support section (14), preferably by an enlarged diameter section in the rotor part (40) which is connected on the bearing region (46) of the rotor part (40).

15. Bearing assembly with a hydrodynamic, in particular hydrodynamic, plain bearing according to one of claims 1 to 14.

Technical Field

The invention relates to a turbocharger having a hydrodynamic (hydrodynamic) plain bearing with a rotor and a counter-bearing part assigned to the rotor, wherein the rotor bearing surface of the rotor is opposed to the mating surface of the mating bearing member to form a hydrodynamic plain bearing, wherein the rotor bearing surface and/or the counter surface form a continuous bearing contour in a sectional view in a section along and through the axis of rotation, the continuous bearing contour is formed by at least two contour sections which merge into one another in order to generate a hydrodynamic bearing capacity in the radial direction as well as in the axial direction, wherein the bearing surface of the rotor is formed by a rotor part connected to the rotor shaft, the rotor part being held on the rotor shaft, and wherein the rotor part is supported relative to the rotor shaft in the region of the supporting section of the rotor shaft.

The invention also relates to such a hydrodynamic plain bearing having such a rotor and a counter bearing component.

In the context of the present invention, hydrodynamic plain bearings are to be understood as meaning, in particular, hydrodynamic plain bearings. In the present invention, such hydrodynamic plain bearings are driven by a liquid lubricant, for example oil, an oil mixture or water.

In addition, in the context of the present invention, a hydrodynamic plain bearing is to be understood as an aerodynamic plain bearing having any kind of gas as lubricant. Air, hydrogen or other gases are used as lubricants for this purpose.

Background

Rotatable mechanical elements, such as shafts, rollers, gears or pump wheels, require guides in the radial and axial directions in order to be able to transmit forces and torques. Hydraulically acting plain bearings can take over this task. The function of this bearing type is based on the physical principle of hydrodynamic pressure generation. In hydrodynamic plain bearings, a suitable lubricant is maintained between the rotor and the stator (the counter bearing component). Shear forces occur in the lubricant during the rotational movement of the rotor relative to the stator, which shear forces then cause the lubricant to be transported through the bearing at a defined speed. Hydrodynamic pressure rise is thereby obtained in the converging bearing gap. In the converging gap course, the bearing gap which is immediately followed by the convergence produces a pressure drop. If the relative speed between the rotor and the stator is sufficiently high, a sufficiently tight lubricant layer is built up by hydrodynamic pressure, which separates the two sliding partners from one another. In this operating state, friction (liquid friction) occurs in the lubricant layer. The hydrodynamic pressure thus generated, in combination with the surfaces used, remains balanced with external forces and represents the load-bearing capacity of the plain bearing. In order to generate the hydrodynamic pressure, no additional energy in the form of compression work or lubricant volume which is transported via the grooves or recesses at a defined pressure is required. The load-bearing capacity is derived from the operating parameters. The basis for the numerical calculation of the hydrodynamic pressure is described in DIN 31652 part 1(DIN sleeve book 198; plain bearings 2; Beuth Press Ltd.; Berlin, Colon 1991).

The prior art includes two basic bearing types:

1. hydrodynamic radial sliding bearing

Hydrodynamic radial plain bearings are usually in the form of cylindrical sleeves and are embodied as segment-like variants or tilting pad bearings. See (DIN 31652 part 2 and VDI criteria 2204) for this. The hydrodynamic elements (e.g. segments) of the plain bearing are arranged cylindrically and thus parallel to the axis of rotation. The converging gap course results from the eccentric position of the rotor relative to the stator.

2. Hydrodynamic axial sliding bearing

The hydrodynamic axial plain bearing is designed in the form of a starting disk with different grooves or surface modifications in the form of a blocking edge, wedge surface or spiral groove. They can also be designed as so-called tilting pad bearings (see DIN 31653 parts 1 to 3; DIN 31654 parts 1 to 3 for this purpose). For this purpose, the axial plain bearing is arranged orthogonally to the axis of rotation, with a thrust washer, which can usually rotate, as a mating running partner. The convergent gap profile required for hydrodynamic pressure generation results from the surface structure (grooves, ramps, etc.), from the inclination of the movable segments or from the angular displacement between the bearing and the thrust washer.

In one technical implementation, both of the aforementioned bearing types must be used in the presence of radial and axial loads. The axial load is then diverted by the axial plain bearing and the radial load is diverted by the radial plain bearing. The two bearing types are then each calculated and designed separately from one another, which results in correspondingly high costs in terms of construction and production.

DE 102008059598 a1 discloses a turbocharger. The turbocharger has a shaft which carries a turbine or compressor wheel on its end. The shaft is supported in the housing by means of two hydrodynamic plain bearings. The hydrodynamic plain bearing is embodied in the form of a conical bearing.

Another turbocharger is described in WO2014/105377a1 and EP1972759B 1. It is disclosed, for example, in EP1972759B1 that a rotor rotating at high speed, which is supported in bearings, is generally characterized by a relatively low load-bearing capacity, in particular in the radial direction. In this way, the hydrodynamic bearing film which is produced in such a bearing requires a relatively robust damping characteristic of the bearing section, which in turn leads to a relatively long construction of the radial bearing in order to be able to carry away the radial loads which are produced by the radial acceleration.

DE 202016105071U 1 describes a turbocharger which supports a rotor in a bearing housing. For this purpose, the rotor has a rotor shaft which is held rotatably in a counter bearing. Hydrodynamic plain bearings are provided between the rotor and the counter bearing on opposite sides of the counter bearing. For this purpose, the rotor and the counter bearing form a bearing contour (Lagerkontur). The bearing contour is composed of contour segments. In a section along and through the axis of rotation, the bearing profile forms a continuous geometry in the sectional view. In this case, the continuous bearing contour is geometrically designed such that it is continuously differentiable in a sectional view and along the axis of rotation. In this way, a high-performance hydrodynamic plain bearing is achieved, which is suitable for carrying axial and radial loads in a minimum space. This results in a compact design but high support forces which must be reliably removed. In the turbocharger known from DE 202016105071U 1, a sealing sleeve is used in the region of the compressor wheel to reduce the component costs. The gland is configured in the form of a rotor component which is pushed onto the rotor shaft. In hydrodynamic plain bearings, the rotor parts form the bearing contour of the rotor. At the same time, the rotor part has a seal receptacle, for example in the form of one or more piston ring grooves, which for example receive piston rings and can thus seal the interior of the bearing housing from the compressor housing of the turbocharger. The rotor parts are pushed onto the rotor shaft with a precise fit, so that the smallest possible radial play is achieved.

Disclosure of Invention

The object of the invention is to provide a turbocharger having a compactly constructed and effective bearing assembly with hydrodynamic, in particular hydrodynamic, plain bearings, wherein the hydrodynamic plain bearings can be mounted simply and at low component costs.

It is also an object of the present invention to provide such a hydrodynamic plain bearing.

The object of the invention with regard to a turbocharger or a hydrodynamic plain bearing is achieved in that the support section and at least one of the contour sections of the counter bearing part at least partially coincide in the direction of the axis of rotation.

During operational use, high bearing forces occur in the region of the hydrodynamic plain bearing, as already mentioned above in the discussion of the prior art. The inventors have now found that it is necessary to reliably introduce loads, in particular high radial loads, from the hydrodynamic plain bearing into the rotor shaft in order to ensure a reliable manner of operation. To this end, it is proposed according to the invention that the support section of the rotor part relative to the rotor shaft support coincides in the direction of the axis of rotation with at least one of the contour sections of the counter bearing part, preferably with the bearing region supporting the majority of the radial load. In this way, the forces are conducted away directly in the radial direction and directly from the profile section via the support section into the rotor shaft. The rotor part can be mounted in a simple manner, since it can be pushed onto the rotor shaft, for example, wherein the support sections are then assigned to the respective bearing regions of the rotor part. In the mounted state, the rotor part is held in particular against tilting, which ensures that the lubricant gap in the hydrodynamic plain bearing is reliably maintained during operational use even with load changes.

According to a preferred variant of the invention, it is possible that the rotor part has a bearing region which forms the profile section and is arranged such that at least one of the profile sections at least partially coincides with the support section in the direction of the axis of rotation. By directly forming the rotor component as a profile section for the bearing area, the component costs are significantly reduced. Since the number of components in the bearing region of the plain bearing, which adjoins the support section in the radial direction, is small, the sum of the manufacturing tolerances in this direction is also minimized, which results in a dimensionally stable and simply reproducible manufacture of the bearing assembly, wherein the bearing gap can be maintained precisely in the hydrodynamic plain bearing.

According to a preferred variant of the invention, the radial gap between the supporting section of the rotor shaft and the region of the rotor part resting on the supporting section can be smaller than the radial gap between the rotor part and the counter bearing part, preferably smaller than 80%, particularly preferably smaller than 60%, of the radial gap between the rotor part and the counter bearing part.

It has been found that the press fit between the supporting section of the rotor shaft and the bearing region of the rotor part then produces a slide bearing which can operate reliably, i.e. a press fit is provided between the supporting section of the rotor shaft and the region of the rotor part which bears against the supporting section, in the range between minus 6 to plus 6 thousandths of a relative radial bearing play between the rotor and the stator of the hydrodynamic slide bearing relative to the diameter of the supporting section. When the radial play in this press fit is formed in the ISO base tolerance classes IT3 to IT8, a good compromise between unbalanced behavior of the rotor and simple mounting of the rotor components is then achieved.

Components designed according to one or more of the aforementioned dimensioning rules are particularly suitable for this purpose in turbochargers for passenger cars. In particular, only a small imbalance arises in such an assembly. In addition, here, a sufficient lubrication gap is always ensured even when the hydrodynamic plain bearing is implemented as a hydrodynamic plain bearing. The smallest possible lubrication gap thickness is dimensioned in this context in such a way that a sufficient lubricant flow is achieved in the special bearing type according to the invention. In this case, too, the lubricant flow is such that, in particular, no turbulence occurs in the lubricant gap of the hydrodynamic plain bearing, which turbulence can lead to interfering, higher-order effects. In particular in hydrodynamic plain bearings of this type, there are no self-excited eddy currents in the lubricant.

A particularly preferred variant of the invention provides for the counter bearing part to be inserted into the bearing housing or into the housing part in such a way that a preferably circumferential gap region is formed between the outer contour of the counter bearing part and the bearing housing or housing part, wherein the gap region is spatially connected to the lubricant supply channel. Preferably, the gap region and the support section can at least partially coincide here in the direction of the axis of rotation. A trapped oil film may be created in the clearance area. This is possible because the gap region is connected to a lubricant supply, for example with a pressure pump. In this way, the pressure is generated by the pressing of the lubricant and a loadable oil trapping film is generated in the gap region.

The bearing types used in the invention, which have a continuous and continuously differentiable bearing contour and which have different contour sections, are characterized by a particularly quiet and low-noise operating mode. The gap region can therefore be dimensioned such that a relatively soft damping action of the trapped oil film results. A bearing design with soft damping properties is thereby obtained in the region of the trapped oil film. This ultimately also results in the need for a smaller load-bearing capacity of the hydrodynamic bearing as the damping is softer, which further reduces the structural dimensions of the hydrodynamic plain bearing and its frictional properties.

It should also be taken into account when designing oil trapping films with gentle damping characteristics for reducing the frictional behavior, which lead to an increased deflection of the rotor, which leads to an increased contour clearance between the turbine or compressor wheel and the turbine or compressor wheel housing.

For this purpose and for setting the damping characteristics, it is possible in the invention for a relative play, which extends in the radial direction relative to the outer diameter of the clearance region, to be formed in the clearance region between the mating bearing component and the bearing housing or housing component, which relative play is in the range from 5 to 10 thousandths of an inch. In particular for use in turbochargers for passenger vehicles, a relative play in the range of 7 to 9 per thousand with respect to the outer diameter of the clearance region is suitable. This results in an advantageous compromise between as little deflection as possible and as gentle a damping behavior as possible, in order to optimize the overall efficiency of the turbocharger, which is composed of the compressor efficiency and the turbine efficiency and the efficiency of the bearings.

The gap region generally has the shape of a hollow cylinder. In principle, it is also conceivable for the gap region to have other geometries, in particular the shape of a hollow cone. In the case of gap regions having other geometries, the outer diameter considered is the average diameter.

According to a conceivable variant of the invention, the radial clearance between the mating bearing part and the bearing housing or housing part can be made larger than the radial clearance between the rotor part and the mating bearing part.

According to a conceivable variant of the invention, the axial extension of the radial gap and the gap region between the mating bearing component and the bearing housing (or housing component) can be defined by the following relationship:

the axial extension (in mm) of the clearance region is equal to 9 minus C times the radial clearance (in mm) between the mating bearing component and the bearing housing, where C is in the range of 61 to 75. Preferably C is selected in the range between 66 and 70.

In this way, a relatively soft damping behavior in the trapped oil film can be achieved. Low radial bearing forces can thereby be achieved. This enables the required axial bearing length to be reduced with concomitant reduction in friction performance, but without incurring significant losses in terms of thermodynamic efficiency of the turbine and compressor. This is particularly important in high-speed applications, for example in modern turbochargers with a speed of more than 200000 rpm.

The lower load-bearing capacity of the hydrodynamic plain bearing, which is required for a suitable damping, can be achieved in this context in particular by the axial overlap in the direction of the axis of rotation between the counter-bearing part and the bearing housing or housing part being dimensioned in the clearance region for the formation of a trapped oil film in such a way that the ratio of the extent of the clearance region in the direction of the axis of rotation to the radial clearance between the counter-bearing part and the bearing housing or housing part is:

the axial extension/radial clearance of the clearance region in the direction of the axis of rotation is 40 to 80.

Particularly preferably, the ratio can be selected in the range of 45 to 70.

In order to reduce the component costs, the rotor component can have a base component which is provided with at least one seal receptacle into which the encircling sealing element is placed.

In order to be able to ensure a precise correspondence of the rotor part to the mating bearing part, a variant according to the invention provides that the rotor part has an appendage with a stop surface, and the stop surface is in surface contact with a radially running surface of the flange of the rotor shaft.

The aforementioned surface-mounted contact also makes it possible to introduce axial tensioning forces into the rotor part, so that the rotor part is tensioned between the compressor wheel and the flange, preferably by means of a screw connection, is secured axially and is held in a rotationally fixed manner in the circumferential direction. This measure enables a simple assembly with low component costs.

As already mentioned, the measures mentioned above are particularly suitable in connection with the particularly preferred bearing type according to the invention, in which the continuous bearing contour of the rotor bearing surface and/or counter surface, which contour is formed by two or more contour segments, is continuously differentiable in a sectional view along and through the axis of rotation. In this bearing type, hydrodynamic, in particular hydrodynamic, load-bearing capacity can be generated in the axial direction as well as in the radial direction on the contour section and preferably on the entire bearing contour. Hydrodynamic, in particular hydrodynamic, plain bearings can be designed as multi-faced plain bearings with two or more lubricating wedges.

By means of the continuous and variable cross-section bearing contour, in particular in the region of the convergent play of hydrodynamic, in particular hydrodynamic, plain bearings, a pressure region can be generated which conducts axial and radial loads away. This results in a three-dimensional hydrodynamic, in particular hydrodynamic, load-bearing capacity in the hydrodynamic, in particular hydrodynamic, plain bearing. The invention makes use of the physical effect according to which the locally generated hydrodynamic, in particular hydrodynamic, pressure acts orthogonally on the surface. From which the local load-bearing capacity is derived. Since in the present invention the surface of the bearing contour can be designed three-dimensionally, a local force component with a corresponding direction is thereby obtained. From the integrated sum of the individual force components, the bearing capacity component of the bearing and the three-dimensional bearing capacity can be calculated and designed for the desired application.

In this case, the hydrodynamic, in particular hydrodynamic, plain bearing can be designed as a multi-faced plain bearing with two or more lubricating wedges, as described above. The friction reduction is achieved by segmenting the bearing in the region of the profile section. In addition, the axial load-bearing capacity is also improved in this bearing, since a continuous and continuously differentiable transition is realized between the individual profile sections. This results in a higher load capacity overall while maintaining the same frictional properties. The segmentation of the bearing assembly also results in a further reduction of sound emission.

A possible variant according to the invention can provide that a preferably annular gap space is formed between the rotor shaft and the rotor part in the direction of the axis of rotation, indirectly or directly next to the support section. This achieves a defined support of the rotor component on the support section. Furthermore, the installation is simplified due to the smaller guide length of the rotor components. In this case, it is particularly preferred if the diameter enlargement in the rotor part, which is connected to the bearing region of the rotor part, also forms a clearance space. Thereby simplifying manufacture. Furthermore, the rotor shaft can remain unaffected in this region, which leads to greater stability.

Drawings

The invention is described in detail below on the basis of embodiments shown in the drawings. In which is shown:

fig. 1 shows a sectional view of a turbocharger; and

fig. 2 and 3 show an enlarged detail of fig. 1.

Detailed Description

Fig. 1 shows a side view and a sectional view of a turbocharger. The turbocharger has a rotor 10 comprising a rotor shaft 11. The rotor shaft 11 has a middle section, which may have a narrowing. The middle section has a stop 13 at its end facing the compressor. A circumferential collar 12 can thus be formed between the stop 13 and the narrowing of the intermediate section. The stop 13 is visible in fig. 2. As shown schematically, the stop 13 can preferably be embodied as a shaft shoulder with a radially oriented surface which surrounds in a ring shape. The rotor shaft 11 has a support section 14 on the compressor side next to the flange 12. The support section can be configured in the form of a machined, circumferential surface. The support section 14 merges into a shaft section 15, which ends in a threaded section 16.

On the side of the rotor shaft 11 opposite the threaded section 16, a bearing section 17 may preferably be provided. The bearing section 17 may be formed from the rotor shaft by machining the rotor shaft 11. As can be seen from the schematic illustration according to fig. 3, the bearing section 17 of the rotor shaft 11 has a circumferential bearing contour. The bearing contour has a plurality of contour sections 17.1 to 17.3 and is preferably constructed in one piece with the rotor shaft 11. The profile section 17.1 which carries away the axial or radial and/or axial forces can be embodied, for example, as a truncated cone to absorb the radial forces. The profile section can also be of convex or concave design. The contour section 17.3 can be cylindrical in shape. The two profile sections 17.1 and 17.3 are connected to each other via the profile section 17.2. In this case, the matching is carried out in such a way that the profile sections 17.1 to 17.3 merge continuously into one another and thus form a continuous bearing profile. In a section through the rotational axis R of the rotor shaft 11, as shown in fig. 3, the bearing profile is continuously differentiable along the rotational axis R. It is also conceivable for the profile sections 17.1 to 17.3 to be formed by functions which can be differentiated several times in succession and thus form a bearing profile with a continuous curvature.

The rotor shaft 11 can have a deflector 18 in the form of a slinger, for example in the form of an enlarged diameter section, next to the profile section 17.1. In this embodiment, the diameter enlargement is configured in the form of a circumferential flange. The diverter 18 may have other suitable profiles that effectively prevent or at least reduce oil leakage through the shaft penetration in the bearing housing.

The rotor shaft 11 can also have at least one seal receptacle 19, as can be seen in fig. 3. In this exemplary embodiment, two seal receptacles 19 are used, which are, for example, in the form of piston ring grooves and are arranged axially spaced apart from one another. A piston ring is inserted into the seal receiving portion 19. A turbine wheel 20 is arranged on the end of the rotor shaft 11 opposite the compressor wheel. The turbine 20 is typically materially connected to the rotor shaft 11.

The rotor 10 has a rotor part 40 on the side facing away from the turbine 20. Rotor member 40 is shown enlarged in figure 2. As schematically shown, the rotor component 40 has a base component 41. In the base part 41, for example, a circumferential seal receptacle 42 in the form of a piston ring groove can be provided. In this exemplary embodiment, two circumferential seal receptacles 42 are used. An annular sealing element 43 in the form of a piston ring is inserted into the seal receptacle 42.

The rotor part 40 has a bearing section 44 next to the base part 41. The bearing section 44 forms a circumferential bearing contour, which can be similar to or identical in construction to the bearing contour having the contour sections 17.1 to 17.3, wherein the bearing section 44 has the contour sections 44.1 to 44.3 forming the bearing contour. The profile section 44.1 for receiving the bearing load is preferably designed as a truncated cone, but may also be convexly or concavely curved. While the cylindrical profile section 44.3 is also used to carry radial loads. The two contour sections 44.1 and 44.3 are connected to one another by means of the contour section 44.2 or transition into one another by means of the contour section 44.2. The contour section 44.2 can be configured like the contour section 17.2 in a crescent-shaped concave shape. A continuous bearing contour is formed by means of the contour sections 44.1 to 44.3. In the sectional view through the axis of rotation R according to fig. 2, the bearing contour forms a continuous and continuously differentiable contour in the sectional view. This is clearly visible in the figures, in which the profile sections 44.1 to 44.3 merge continuously into one another without a continuous abrupt change. As in the turbine-side bearing profiles 17.1 to 17.3, it is also conceivable here for the profile sections 44.1 to 44.3 to be formed by a function which can be differentiated several times in succession and thus form bearing profiles with a continuous curvature.

Rotor member 40 has an appendage 45 at its end facing flange 12. The attachment is preferably formed by a bearing section 44. The end face of the appendage 45 is oriented in a radial direction. In this way, the attachment 45 bears with its end face against the stop 13 of the flange 12. In order to ensure the planar contact, the attachment 45 is chamfered on the inside. For this purpose, the relief is furthermore turned into the rotor shaft 11 next to the flange 12, as shown in fig. 2.

The rotor part 40 rests on the support section 14 of the rotor shaft 11 at a bearing section 44 forming a bearing region 46. In this case, a press fit, preferably in the form of a transition fit, is formed between the rotor part 40 and the rotor shaft 10. Preferably, a transition match in ISO base tolerance classes IT3 to IT8 is achieved. The bore which is introduced into the bearing region 46 and is intended to bear against the support section 14 has, next to the bearing region 46, an enlarged diameter portion which forms a setback 47. An annular free space is formed between rotor part 40 and the outer circumference of rotor shaft 11 by means of a recess 47.

The base part 41 has an annular, radially extending contact surface 48. The bearing surface 48 is therefore parallel to the bearing surface of the bearing section 44, which bears against the flange 12.

The compressor wheel 30 is pushed onto the rotor shaft 11 in the region of the shaft section 15. The compressor wheel 30 bears with a radially extending bearing surface against the bearing surface 48 of the rotor part 40. To secure the rotor component 40 and the compressor wheel 30, the nut 31 is screwed onto the threaded section 16. Thus, the nut 31 tensions the compressor wheel 30 relative to the rotor member 40 and tensions the rotor member 40 relative to the stop 13. In this way, the compressor wheel 30 and the rotor part 40 are fixed axially on the rotor shaft 11 and are held non-rotatably thereon in the circumferential direction.

As can be seen in fig. 1, the turbocharger has a counter bearing part 50 which is inserted into a bearing housing 60 of the turbocharger. The counter bearing part has an intermediate part 51. The intermediate part 51 is connected on both sides with attachments 53. The two attachments 53 each have a circumferential bearing contour. The circumferential bearing contour is complementary to the bearing contour formed by the rotor part 14 or the bearing section 17 of the rotor shaft 11. The support contour therefore has contour sections 53.1 to 53.3, which can be embodied as circumferential. The profile section 53.1 which is subjected to axial forces is, for example, made in the shape of a truncated cone, and the profile section 53.3 is, for example, made in the shape of a cylinder. The two profile sections 53.1 and 53.3 can be continuously differentiated and continuously transitioned into one another at least once via the profile section 53.2.

In order to install the assembly according to fig. 1 for a turbocharger, the counter bearing part 50 is first inserted into a correspondingly formed receptacle of the bearing housing 60. In order to fix the predetermined position of the counter bearing part 50 shown in fig. 1 in the bearing housing 60, a fastening element 70 is used. The fastening element 70 has a retaining section 72. The retaining section 72 engages into the fastening receptacle 52 of the counter bearing part 50. For mounting the fastening element 70, the fastening element can be introduced through the lubricant guide channel 61 of the bearing housing 60. In order to hold the fastening element 70 in a manner that prevents it from being lost, the fastening element can be screwed, pressed into the bearing housing 60 or fixed by means of a holding element.

After the counter bearing part 50 is mounted in the bearing housing 60, the rotor 10 may be assembled. For this purpose, the rotor shaft 11 is pushed from the turbine-side bearing housing side into the bore of the bearing housing 60. Here, the rotor shaft 11 is engaged through the counter bearing part 50, as shown in fig. 1. The thrust movement of the rotor 10 is limited by the contour section 17.1 of the rotor shaft 11, which abuts against a corresponding contour section 53.1 of the counter bearing (see fig. 3). The sealing element which has been inserted into the circumferential seal receptacle 19 rests in the installed position against a corresponding annular sealing surface of the bearing housing 60 (see fig. 3).

The rotor part 40 can now be pushed into the bearing housing 60 from the opposite side. In this case, the rotor part 40 is pushed with its bearing region 46 onto the rotor shaft 11. This can be done simply, since the rotor part 40 is guided on the rotor shaft 11 only in the bearing region 46 with a precise fit. In addition, the retraction 47 does not hinder the pushing-in movement. In the mounting position according to fig. 2, the rotor part 40 rests on the flange 12. The compressor wheel 30 is then pulled onto the rotor shaft 11 and the nut 31 is tightened (see description above). In the mounted state, the rotor 10 is arranged with its contour sections 53.1 to 53.3 opposite the contour sections 17.1 to 17.3 or 44.1 to 44.3 on the two attachments 53. The respective bearing gap is formed in such a way that a hydrodynamic film is guided in each case in the bearing gap to produce two hydrodynamic plain bearings. The relative radial and relative axial bearing play of each hydrodynamic plain bearing is preferably in the range between 1 and 5 per thousandth of the diameter of the profile section 17.3 or 44.3 transmitting the radial forces.

As shown in fig. 1, a surrounding clearance area 57 is provided between one, preferably both, appendages 53 of the mating bearing component 50 and the bearing housing 60. The clearance region has a relative radial clearance (absolute radial clearance/diameter of the counter bearing part 50 in the clearance region 57) extending in the radial direction, which is in the range between 5 and 10 thousandths, particularly preferably in the range between 7 and 9 thousandths.

The gap region 57 is preferably arranged such that at least one of the contour sections 44.1 to 44.3 or 17.1 to 17.3, which mainly contribute to the load-bearing capacity of the hydrodynamic plain bearing in the radial direction, coincides with the gap region 57 in the direction of the axis of rotation R, particularly preferably at least in the region of the contour sections 17.3 and 44.3. The axial overlap in the gap region 57 between the counter bearing component 50 and the bearing housing 60 in the direction of the axis of rotation R is preferably dimensioned such that the extension of the gap region 57 in the direction of the axis of rotation R and the radial gap between the counter bearing component 50 and the bearing housing 60 or housing component are in the ratio:

the axial extension/radial clearance of the clearance region 57 along the axis of rotation R is 40 to 80;

particularly preferably, the ratio is in the range of 45 to 70.

A trapped oil film is provided in the clearance region 57 by one or more of the aforementioned sizing specifications. With the hydrodynamic plain bearing according to the invention, the trapping membrane has a high axial load-bearing capacity for a conventional turbocharger. Trapping oil films designed with suitable damping are particularly suitable for reducing the bearing forces during operation, in particular caused by imbalances and load transfer processes. In this way, a turbocharger is provided which can be operated with particularly low noise on the one hand, and which has improved friction properties on the other hand.

The two gap regions 57 are spatially connected to the lubricant guide channel 61. Lubricant under pressure can be fed in via the lubricant guide channel 61. The lubricant reaches the chamber 64 through the passage 71 of the fastening element 70. Lubricant is pressed from the chamber 64 into the clearance area 57. In this way, a solution is created in the annularly encircling gap region 57, in which an adjustable damping by means of an oil trapping film is achieved. Lubricant is also delivered to both hydrodynamic plain bearings from the same chamber 64. Accordingly, the lubricant reaches the region of the hydrodynamic gap formed between the profile sections 44.1 to 44.3 of the rotor 10 and the corresponding profile sections 53.1 to 53.3 of the counter bearing component (on the one hand) and the profile sections 17.1 to 17.3 and the corresponding profile sections 53.1 to 53.3 (on the other hand). During rotation of the rotor 10, the lubricant is guided via hydrodynamic clearances of the hydrodynamic plain bearing to generate hydrodynamic pressure. Next to the hydrodynamic gap, the lubricant reaches the separation space 62. For example, the gap region 57 can also open into the separation space 62. The lubricant is collected in the collection region 63 of the bearing housing 60, conveyed back to the lubricant circuit and conveyed again to the lubricant guide channel 61.

According to fig. 2, the rotor part 40 is supported relative to the rotor shaft 11 in the region of the support section 14 of the rotor shaft 11, as previously described. The assignment is made in such a way that the support section 14 and at least one of the contour sections 53.1 to 53.3 of the counter bearing part 50 at least partially coincide in the direction of the axis of rotation R. The overlap is preferably in the region of the contour section 17.3 or 44.3.