Turbocharger for an internal combustion engine

Abstract

Turbocharger for an internal combustion engine, comprising a casing (2) with an air-side compressor blade (3), a shaft (1) driving the compressor blade (3), and at least one axially acting pivot bearing (5) for supporting the shaft (1), wherein the pivot bearing (5) is designed as a hydrodynamic plain bearing, wherein a stationary bearing member (6) is penetrated by the shaft (1) and forms on a first side of the bearing member (6) a first axially acting bearing against a bearing collar (7) rotating with the shaft (1), wherein the bearing member (6) forms on an opposite second side a second axially acting bearing against a sealing bush (8) rotating with the shaft (1), wherein an oil supply (9) for supplying the bearings is formed in the bearing member (6), wherein a plurality of flow surfaces (10) with a height (h) variable in the circumferential direction are formed on a surface of the bearing member (6) facing the bearing collar (7) in the axial direction, and wherein for each of the bearings an individually dimensioned throttle element (11, 12) is formed in the oil supply (9), characterized in that the throttle members (11, 12) are each designed as an annular gap between the shaft (1) and the bearing member (6), wherein in the axial direction between the annular gaps (11, 12) at least one annular groove (13) is formed, which has a larger maximum diameter than the annular gaps (11, 12).

Description

[0001] The invention relates to a turbocharger for an internal combustion engine according to the preamble of claim 1. The invention also relates to a turbocharger for an internal combustion engine according to the preamble of claim 7.

[0002] WO 2013 / 169505 A1 describes an axial bearing for a turbocharger of an internal combustion engine. A shaft passes through a bearing member in a central bore. A plurality of wedge surfaces are formed on an axially directed side of the bearing member, with a bearing collar rotating with the shaft bearing against the side of the bearing member and projecting radially beyond the wedge surfaces. To supply the bearing, oil-supply bores are provided between the wedge surfaces, located radially approximately halfway up the wedge surfaces. The bores pass completely through the bearing member and also supply oil to a rear side of the bearing member, which is axially supported against a sealing bushing.

[0003] DE 10 2013 224 413 A1 relates to an axial bearing for a high-speed shaft, which consists of two axial bearing halves connected to each other.

[0004] DE 10 2013 224 416 A1 relates to an axial bearing for the axial support of a rotor shaft in a bearing housing of an exhaust gas turbocharger. The axial bearing is characterized by the fact that it consists of two axial bearing disks that are separated from one another and lie directly against one another. A lubricant supply device is provided for supplying a lubricant into a bearing plane between the axial bearing and the rotor shaft. The lubricant supply device has supply channels for the lubricant in a space between the two axial bearing disks.

[0005] DE 32 44 893 A1 relates to a thrust bearing for a turbocharger, comprising a thrust ring, a thrust bushing, and an oil slinger. The thrust ring is equipped with an oil supply channel and has an internal oil reservoir. A small-diameter bore penetrates the wall of the thrust ring in the axial direction, providing a connection between the reservoir and the outer surface of the thrust ring, through which lubricating oil can reach the load-bearing surfaces.

[0006] CN 2 01 588 879 U relates to a thrust bearing for a pressurization device. A plurality of oil holes are arranged on a side surface of the thrust bearing and radially penetrate the inner wall of the thrust bearing to the inner ring of a central axle hole.

[0007] It is the object of the invention to provide a turbocharger for an internal combustion engine in which a hydrodynamic axial bearing has effective lubrication.

[0008] This object is achieved for a turbocharger mentioned at the outset with the characterizing features of claim 1. By individually dimensioning the throttle elements, the optimal oil quantity and the optimal oil pressure can be provided for each of the two bearings. In particular, it is taken into account that during operation both bearings alternately absorb oppositely directed axial supporting forces depending on the current requirements, but one of the bearings is subjected to more continuous loading and / or absorbs greater maximum forces. In addition, heat dissipation through the oil can be optimized in this way. The throttle elements are each designed as an annular gap between the shaft and the bearing element, with at least one annular groove being formed in the axial direction between the annular gaps, said groove having a larger maximum diameter than the annular gaps.This allows for a simple, defined distribution of the supplied oil between the first and second bearings, as the annular gaps act as flow restrictors. The annular groove can advantageously act as a reservoir upstream of the annular gaps, ensuring a constant oil flow.

[0009] A turbocharger according to the invention comprises a compressor blade driven by the shaft. On the drive side, the turbocharger can preferably have an exhaust turbine. Alternatively, other drive mechanisms for the shaft, such as an electric motor, are also possible.

[0010] The oil supply can in particular be connected to an oil circuit of the combustion engine.

[0011] An axial effect of the rotary bearing is preferably understood to mean that the bearing member supports at least a predominant part, preferably all of the forces acting on the rotating shaft in the axial direction during operation and introduces them into the housing of the turbocharger.

[0012] A bearing member according to the invention is preferably immovably mounted or fixed relative to the housing. It can be a substantially disc-shaped component, with two opposite sides of the disc serving as bearing surfaces or support surfaces. These bearing surfaces can have a corresponding shape or surface treatment to enable hydrodynamic support. Such a bearing member is preferably made of an alloy different from the housing, preferably a copper alloy.

[0013] In the context of the invention, a flow surface is understood to be a specifically structured surface, whereby the structuring achieves a targeted hydrodynamic pressure distribution of the supplied oil during operation.

[0014] In generally preferred embodiments of the invention, the oil supply for at least one of the bearings, preferably for both bearings, opens at a minimum radius of the bearing member relative to the shaft. This prevents a radially inner region of the bearing collar and / or sealing bush from being wetted with oil. In particular, the oil supply also allows the shaft itself to be wetted with oil from the oil supply, which further improves lubrication and cooling. The oil is distributed very evenly in the circumferential direction among the various flow surfaces. Each of the flow surfaces receives oil from the radially inner region and thus at the same, low temperature.

[0015] In a structurally simple and practical design, the oil supply comprises an oil channel formed in the bearing element that opens directly into the annular groove. Such an oil channel can be implemented, for example, as a bore.

[0016] With an optimized geometry of the throttle members, an axial width of at least one of the annular gaps, preferably each of the annular gaps, is at least 15%, preferably at least 20%, of a total width of the bearing member in the region of the annular gaps.

[0017] Alternatively or additionally, in an optimized geometry, it is provided that a maximum radial height of the annular groove above a surface of the shaft is at least 15%, preferably at least 20%, greater than a maximum radial height of the annular gap above the shaft.

[0018] Furthermore, alternatively or additionally, in an optimized geometry, it is provided that an axial width of the annular groove is at least 20%, preferably at least 25%, of a total width of the bearing member in the region of the annular gap.

[0019] In a second preferred embodiment of the invention, the oil supply comprises an oil channel formed in the bearing member, which branches into a first sub-channel leading to the first bearing and a second sub-channel leading to the second bearing. The throttle elements are formed by the sub-channels and a defined dimensioning of the sub-channels. This allows for precise oil distribution to be achieved in a simple manner. The sub-channels can be created, for example, by bores of a defined diameter.

[0020] In a preferred embodiment, the two sub-channels branch off from the oil channel in a Y-shape and each open at axially opposite ends of an annular gap between the shaft and the bearing member. In this way, oil can be supplied to a radially inner region, preferably with wetting of the shaft, in a simple and effective manner.

[0021] The object of the invention is achieved for a turbocharger according to the preamble of claim 7 with the characterizing features of claim 7. The radial projection of the flow surfaces allows for improved discharge of the oil from the hydrodynamic plain bearing. This allows for an increase in the amount of oil flowing through the bearing. Furthermore, a better discharge of unwanted particles from the bearing area can be achieved.

[0022] In particular, the parts of the flow surfaces that overlap with the bearing collar are largely responsible for the hydrodynamic properties of the bearing. The non-overlapping parts, or the radial projection, are largely responsible for improved oil discharge from the bearing. For the purposes of the invention, the projection beyond the bearing collar is generally understood to mean a projection beyond a hydrodynamically effective part of the bearing collar.

[0023] In an optimized dimensioning of the invention, it is provided that a maximum diameter of the flow surfaces is at least 2%, preferably at least 5%, larger than a maximum diameter of the bearing collar over a hydrodynamically effective area. This allows for a continuous outflow of oil over the flow surfaces.

[0024] In a generally preferred embodiment, the flow surfaces rise monotonically in the circumferential direction over a first angular section, with a locking surface of constant height being formed over a subsequent second angular section in the circumferential direction. The locking surface of constant height prevents an undefined separation of the rising flow surface and reduces wear due to abrasion.

[0025] In a preferred embodiment, the locking surface of the second section can protrude in steps by an offset height above a maximum height of the flow surface of the first section. This step allows for even greater wear on the locking surface of constant height in the event of running-in and / or operational wear, without significantly altering the hydrodynamic properties of the bearing.

[0026] To ensure a constant supply of oil, it is advantageous to provide an axial recess in the form of a pocket in the circumferential direction at least adjacent to one of the flow surfaces. The recess preferably has an opening directed radially toward the shaft. Such a pocket serves as an oil reservoir and releases the oil constantly and uninterruptedly into the area of the flow surfaces.

[0027] In a preferred embodiment, at least one of the flow surfaces has a radially inner wedge surface portion and a radially outwardly adjoining discharge portion, with the discharge portion having a height that decreases radially outward. The wedge surface portion is essentially responsible for building up the hydrodynamic pressure of the bearing. The different shape of the discharge portion allows for the radial discharge of the oil to be optimized, with the primary goal being a continuous and uninterrupted flow in the radial direction.

[0028] With an optimized geometry in this regard, the outflow section is provided with a curved profile in a radially directed section. This prevents the oil from being thrown off edges or steps. The radius of curvature of the profile is preferably between 1 and 4 times, particularly preferably between 1.5 and 2.5 times, the radial width of the wedge surface section. These dimensional relationships allow, in particular, a low-wear milling tool to be used to produce the flow surfaces of the bearing member. Overall, this makes the production of the bearing member more cost-effective.

[0029] A turbocharger according to one of claims 10 to 16 can also comprise one or more further features of claims 1 to 10. Conversely, a turbocharger according to one of claims 1 to 9 can also comprise one or more further features of claims 10 to 16. In the interest of improved lubrication, the favorable properties of the oil supply defined by the throttle elements and the favorable properties of the improved oil flow through the flow surfaces complement each other. This reduces overall wear on the bearings, as the accumulation of wear-causing particles in the bearings is reduced.

[0030] Further advantages and features of the invention emerge from the exemplary embodiments described below and from the dependent claims.

[0031] Several preferred embodiments of the invention are described below and explained in more detail with reference to the accompanying drawings. Fig. 1 shows a schematic overall view of a turbocharger according to the invention. Fig. 2 shows a plan view of a first side of a bearing member with bearing collar and sealing bush from the area Z of Fig. 1. Fig. 3 shows a sectional view through the bearing member, bearing collar and sealing bush along the line AA from Fig. 2. Fig. 4 shows the bearing link from Fig. 2 without bearing collar. Fig. 5 shows a sectional view along the line AA of Fig. 4 Fig. 6 shows a sectional view along the curved section line BB of Fig. 4 Fig. 7 shows a plan view of a second side of the bearing member from Fig. 2. Fig. 8 shows a part of a sectional view of the bearing member of Fig. 4 along the section line CC. Fig. 9 shows an enlarged section of the area Z from Fig. 6. Fig. 10 shows a plan view of a second embodiment of the invention with regard to a shape of the flow surfaces. Fig. 11 shows a sectional view through the bearing member, bearing collar and sealing bush of a third embodiment of the invention.

[0032] The Fig. The turbocharger for an internal combustion engine shown in Figure 1 comprises a central shaft 1, which is mounted in a housing 2 and drives an air-side compressor blade 3. The shaft 1 is in turn driven by an exhaust-gas turbine or exhaust-side turbine blade 4. As an alternative to the turbine blade 4, another drive for the shaft is also possible, for example, by an electric motor.

[0033] The shaft 1 is supported by an axially acting pivot bearing 5 (see also area Z and Fig. 2, Fig. 3), which is designed as a hydrodynamic plain bearing. A stationary bearing member 6 is penetrated by the shaft. On a first side of the bearing member 6 (view Fig. 2, Fig. 4) a first axially acting bearing is formed against a bearing collar 7 rotating with the shaft 1.

[0034] The bearing member 6 also forms on an opposite second side (view Fig. 7) a second axially acting bearing against a sealing bush 8 rotating with the shaft 1.

[0035] An oil supply 9 is formed in the bearing member to supply the bearings. The oil supply 9 begins in the supply direction with a milled recess 9a in the surface of the essentially disc-shaped bearing member 6. The oil supply is further connected to an oil circuit of the internal combustion engine via channels in the housing 2.

[0036] A plurality of flow surfaces 10 are formed on a surface of the bearing member facing the bearing collar 7 in the axial direction. The flow surfaces 10 have a height h that varies in the circumferential direction, at least over part of their radial extent. The height h is defined as the axially directed distance of the flow surface from a surface plane of the bearing member. The flow surfaces are formed from the flat surface of the bearing member by a material-removing process (milling).

[0037] The bearing member 6 is preferably fixed immovably relative to the housing. It is a substantially disc-shaped component, with two opposite sides of the disc serving as bearing surfaces or support surfaces. The flow surfaces 10 represent a shaped or machined surface to enable hydrodynamic support. In this case, the bearing member is made of an alloy different from the housing, in this case a copper alloy.

[0038] According to the invention, an individually dimensioned throttle element 11, 12 is formed in the oil supply 9 for each of the two bearings. In the case of the embodiment according to Fig. 1 to 9 each comprise an annular gap 11, 12 formed between the shaft 1 and the bearing member 6. In the axial direction, an annular groove 13 is provided between the annular gaps, which has a larger maximum diameter than the annular gaps. This allows for a simple, defined distribution of the supplied oil to the first and second bearings, as the annular gaps act as flow restrictors. The annular groove 13 also acts as a reservoir upstream of the annular gaps 11, 12, thus ensuring a constant oil flow.

[0039] The oil supply 9 comprises an oil channel 9b formed as a bore, which begins in the milled portion 9a and opens directly into the annular groove 13. In the present embodiment, the oil supply 9 for supplying the two bearings comprises the milled portion 9a, the oil channel 9b, the annular groove 13, and the throttle elements or annular gaps 11, 12.

[0040] Due to the design of the throttle elements as annular gaps 11, 12, the oil supply 9 opens via the annular groove 13 for both bearings at a minimum radius of the bearing element 6 relative to the shaft 1. This prevents a radially inner area of the bearing collar and / or sealing bush from being wetted with oil. Through the oil supply 9, the shaft itself is also wetted with oil from the oil supply, which further improves lubrication and cooling. The oil is distributed very evenly in the circumferential direction among the various flow surfaces. Each of the flow surfaces receives oil from the radially inner area and thus at the same, low temperature.

[0041] In the present example, the geometry of the throttle members 11, 12 is designed such that an axial width of each of the annular gaps is approximately 27% of a total width of the bearing member 6 in the region of the annular gaps 11, 12.

[0042] A maximum radial height of the annular groove 13 above a surface of the shaft 1 is in this case approximately 66% greater than a maximum radial height of the annular gaps 11, 12 above the shaft 1.

[0043] Furthermore, an axial width of the annular groove 13 in the present case amounts to approximately 46% of a total width of the bearing member 6 in the region of the annular gaps 11, 12.

[0044] Regarding the design of the oil supply 9 with two individually dimensioned throttle elements, Fig. 11 shows a second possible embodiment of the invention. Here, a milled recess 9a and an oil channel 9b are also formed in the bearing member. However, the oil channel 9b branches into a first sub-channel 14 leading to the first bearing and a second sub-channel 15 leading to the second bearing. In this embodiment, the throttle elements are formed by the sub-channels 14, 15 and a defined dimensioning of the sub-channels 14, 15. This allows precise oil distribution to be achieved in a simple manner. The sub-channels 14, 15 can, for example, be created by bores of a defined diameter.

[0045] Just as in the first embodiment, the partial channels for both bearings open at a minimum radius of the bearing member 6 with respect to the shaft 1. As a result, as in the first example, each of the two sides of the bearing member 6 is wetted with oil starting from the shaft 1.

[0046] The two sub-channels 14, 15 branch off in a Y-shape from the oil channel 9b and each open at axially opposite ends of an annular gap 18 between the shaft 1 and the bearing member 6. In this way, an oil supply to a radially inner region, including wetting of the shaft, is achieved simply and effectively.

[0047] As a further object of the invention, in the embodiments described above, the flow surfaces 10 have a radial projection 16 beyond the bearing collar 7. The radial projection 16 of the flow surfaces 10 improves the discharge or drainage of the oil from the hydrodynamic plain bearing. Furthermore, the total amount of oil flowing through the bearing is increased. Furthermore, this results in better discharge of unwanted particles from the bearing area.

[0048] The parts 17 of the flow surfaces 10 that overlap with the bearing collar are primarily responsible for the hydrodynamic properties of the bearing. The non-overlapping parts, or the radial projection 16, are primarily responsible for improving oil discharge from the bearing.

[0049] In the case of the first embodiment according to Fig. 1 to Fig. 9, a maximum diameter d1 of the flow surfaces is approximately 17% larger than a maximum diameter d2 of the bearing collar 7 over a hydrodynamically effective area. This allows for a continuous outflow of oil over the flow surfaces.

[0050] The individual flow surfaces 10 can each be considered as divided into a radially inner wedge surface portion 10a and a radially outwardly adjoining discharge portion 10b. The discharge portion 10b has a height h that decreases radially outward. The wedge surface portion 10a is essentially responsible for building up the hydrodynamic pressure of the bearing. The different shape of the discharge portion 10b allows the radial discharge of the oil to be optimized, with the primary goal being a continuous and uninterrupted flow in the radial direction.

[0051] The flow surfaces 10 increase monotonously in the circumferential direction over a first angular section W1 (in Fig. 4 counterclockwise or in the direction of rotation of the shaft 1). A locking surface 10c with a constant height h is formed over a second angular section W2 following in the circumferential direction. The locking surface 10c of constant height h prevents an undefined separation of the rising flow surface and reduces wear due to abrasion. In this case, the locking surface has a height of zero or is part of the flat surface of the bearing member 6.

[0052] In this example (see Fig. 9), the locking surface 10c of the second section W2 protrudes in steps by an offset height k above a maximum height h of the flow surface of the first section. This step 10d allows for even greater removal of the locking surface 10c at a constant height h without significantly altering the hydrodynamic properties of the bearing.

[0053] To ensure a constant supply of oil, an axial recess 17 in the form of a pocket is formed circumferentially next to each of the flow surfaces 10. The recess 17 is U-shaped in cross-section and has an opening 17a directed radially toward the shaft. Such a recess or pocket 17 serves as an oil reservoir and releases the oil constantly and uninterruptedly into the area of the flow surfaces 10.

[0054] In the present case, the geometry is optimized so that the outlet part 10b of a flow surface 10 in a radially directed section (see approximately Fig. 8) has a curved profile. This prevents the oil from being thrown off edges or steps. In this case, the radius of curvature r of the profile is approximately twice the radial width b of the wedge surface portion 10a. This dimensional ratio allows a low-wear milling tool to be used to produce the flow surfaces 10 of the bearing member 6.

[0055] In this case, the diameter d2 of the bearing collar 7 is equal to its hydrodynamically effective diameter, since the entire axially directed surface of the bearing collar 7 is subjected to oil pressure. The circumference of the bearing collar essentially corresponds to a circle that runs along the boundaries between the wedge surface parts 10a and the outlet parts 10b of the flow surfaces 10. In other words, the radius of the bearing collar 7 essentially corresponds to the outer radius of the wedge surface parts 10a.

[0056] Fig. Figure 7 shows the rear side, or second side, of the bearing member facing the sealing bushing. Flow surfaces 10 are also formed there, which are shaped analogously to the flow surfaces 10 on the bearing collar side, or first side. The rise of the wedge surface parts 10a thus occurs clockwise here. The only partial radial overlap by the bearing collar 7 in the first bearing arrangement corresponds to only partial overlap by the sealing bushing 8 in the second bearing arrangement. In this case, the two bearing arrangements are therefore identical in terms of the flow surfaces 10.

[0057] Fig. 10 shows a further embodiment of the invention. In contrast to the example described above according to Fig. 1 to Fig. 9, no step 10d is provided here between the highest end of the wedge surface part 10a and the locking surface 10c located in the surface of the bearing member 6. Accordingly, in the plan view shown, the outflow parts 10c of the flow surfaces taper in the ascending direction of the wedge surface parts 10a.

[0058] It is understood that the respective features of the three embodiments ( Fig. 1 to Fig. 9; Fig. 10 and Fig. 11) can be combined with each other depending on requirements. In particular, each of the two variants of the flow surfaces ( Fig. 4 or Fig. 10) with each of the two variants of oil supply ( Fig. 3 or Fig. 11). Furthermore, flow surfaces according to the invention can be combined with conventional oil feeds. Furthermore, oil feeds according to the invention can be combined with conventional flow surfaces. List of reference symbols 1 wave 2 housings 3 compressor blades 4 turbine blades 5 axial pivot bearing 6 bearing link 7 bearing bundle 8 Sealing bushing 9 Oil supply 9a Milling 9b Oil channel 10 flow surfaces 10a Wedge surface part 10b Drainage section 10c Rest area 10d level 11 first throttle element (annular gap) 12 second throttle element (annular gap) 13 Ring groove 14 first throttle element (partial channel) 15 second throttle element (partial channel) 16 (radial) overhang 17 Pocket, recess 17a Opening of the recess 18 Annular gap h Height of the flow area k Height of step 10d d1 maximum diameter of the flow surfaces d2 maximum diameter of bearing collar W1 first, monotonically increasing angle segment W2 second angular segment of constant height b radial width of the wedge surface part

Claims

[1] Turbocharger for an internal combustion engine, comprising a casing (2) with an air-side compressor blade (3), a shaft (1) driving the compressor blade (3), and at least one axially acting pivot bearing (5) for supporting the shaft (1), wherein the pivot bearing (5) is designed as a hydrodynamic plain bearing, wherein a stationary bearing member (6) is penetrated by the shaft (1) and forms on a first side of the bearing member (6) a first axially acting bearing against a bearing collar (7) rotating with the shaft (1), wherein the bearing member (6) forms on an opposite second side a second axially acting bearing against a sealing bush (8) rotating with the shaft (1), wherein an oil supply (9) for supplying the bearings is formed in the bearing member (6), wherein a plurality of flow surfaces (10) with a height (h) variable in the circumferential direction are formed on a surface of the bearing member (6) facing the bearing collar (7) in the axial direction, and wherein for each of the bearings an individually dimensioned throttle element (11, 12) is formed in the oil supply (9), characterized by that the throttle members (11, 12) are each designed as an annular gap between the shaft (1) and the bearing member (6), wherein in the axial direction between the annular gaps (11, 12) at least one annular groove (13) is formed, which has a larger maximum diameter than the annular gaps (11, 12). [2] Turbocharger according to claim 1, characterized by that the oil supply (9) for at least one of the bearings opens at a minimum radius of the bearing member (6) with respect to the shaft (1). [3] Turbocharger according to claim 1 or claim 2, characterized bythat the oil supply (9) comprises an oil channel (9a) formed in the bearing member (6) which opens directly into the annular groove (13). [4] Turbocharger according to any one of the preceding claims, characterized by that an axial width of at least one of the annular gaps (11, 12) is at least 15% of a total width of the bearing member (6) in the region of the annular gaps (11, 12). [5] Turbocharger according to any one of the preceding claims, characterized by that a maximum radial height of the annular groove (13) above a surface of the shaft (1) is at least 15% greater than a maximum radial height of the annular gaps (11, 12) above the shaft (1). [6] Turbocharger according to any one of the preceding claims, characterized by that an axial width of the annular groove (13) is at least 20% of a total width of the bearing member (6) in the region of the annular gaps (11, 12). [7] Turbocharger for an internal combustion engine, comprising a casing (2) with an air-side compressor blade (3), a shaft (1) driving the compressor blade (3), and at least one axially acting pivot bearing (5) for supporting the shaft (1), wherein the pivot bearing (5) is designed as a hydrodynamic plain bearing, wherein a stationary bearing member (6) is penetrated by the shaft (1) and forms at least on a first side of the bearing member (6) a first axially acting bearing against a bearing collar (7) rotating with the shaft (1), wherein an oil supply (9) for supplying at least the first bearing is formed in the bearing member (6), wherein a plurality of flow surfaces (10) with a height (h) variable in the circumferential direction are formed on a surface of the bearing member (6) facing the bearing collar (7) in the axial direction, characterized by , that the flow surfaces (10) have a projection (16) over the bearing collar (7) in the radial direction. [8] Turbocharger according to claim 7, characterized by that the bearing member (6) forms on an opposite second side a second axially acting bearing against a sealing bush (8) rotating with the shaft (1). [9] Turbocharger according to claim 7 or claim 8, characterized by that a maximum diameter (d1) of the flow surfaces (10) is at least 2% larger than a maximum diameter (d2) of the bearing collar (7) over a hydrodynamically effective area. [10] Turbocharger according to any one of claims 7 to 9, characterized by that the flow surfaces (10) increase monotonously in the circumferential direction over a first angular section (W1), wherein a locking surface (10c) with a constant height (h) is formed over a second angular section (W2) following in the circumferential direction. [11] Turbocharger according to claim 10, characterized bythat the locking surface (10c) of the second angular section (W2) projects in steps by an offset height (k) over a maximum height of the flow surface (10a, 10b) of the first angular section (W1). [12] Turbocharger according to one of claims 7 to 11, characterized by that in the circumferential direction at least next to one of the flow surfaces (10) an axial recess (17) in the manner of a pocket is formed, wherein the recess (17) has an opening (17a) directed radially towards the shaft (1). [13] Turbocharger according to one of claims 7 to 12, characterized by that at least one of the flow surfaces (10) has a radially inner wedge surface part (10a) and a radially outwardly adjoining outlet part (10b), wherein the outlet part (10b) has a height (h) decreasing radially outwards. [14] Turbocharger according to claim 13, characterized bythat the outlet part (10b) has a curved profile in a radially directed section, wherein a radius of curvature (r) of the profile is between 1 and 4 times a radial width (b) of the wedge surface part (10a). [15] Turbocharger according to one of claims 7 to 14, comprising the features according to one of claims 1 to 6.

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