Turbocharger device based on high loadable high speed rotor shafting structure

Abstract

This invention relates to the field of internal combustion engine turbocharger technology, specifically a turbocharger device based on a high-load-bearing, high-speed rotor shaft system. The device includes a bearing housing, a rotor shaft, a floating bearing, a thrust sleeve, a thrust bearing, and a shaft seal. The floating bearing is housed within the bearing housing, and the rotor shaft is rotatably mounted on it. The shaft seal and thrust sleeve are axially mounted side-by-side on the rotor shaft, with the thrust bearing located between the shaft seal and the thrust sleeve. The floating bearing contains two symmetrically arranged inner ring support surfaces. Each inner ring support surface includes an inner ring support surface one and an inner ring support surface two arranged axially side-by-side, with an included angle between them. This invention addresses the problem in conventional technologies where, when the rotor wobbles, only one or two effective oil wedge surface bearing areas are typically formed, while the remaining oil wedge surface bearing areas cannot maintain sufficient pressure support due to increased oil film clearance, leading to a decrease in overall load-bearing capacity.

Description

Technical Field

[0001] This invention relates to the field of internal combustion engine turbocharger technology, and more specifically to a turbocharger device based on a high-load-bearing, high-speed rotor shaft system structure. Background Technology

[0002] In a turbocharger, the compressor impeller and turbine impeller are connected via a rotor shaft, forming a high-speed rotating core component. Its basic working principle is to utilize the exhaust energy of the engine to drive the turbine, which in turn drives the coaxial compressor impeller to rotate, increasing the intake pressure and thus improving the engine's power density and combustion efficiency. In this device, the bearing structure, as a key support for the rotor, directly affects its operational stability, mechanical efficiency, and service life. An ideal bearing system should, across the entire operating range of the rotor, form a uniform, continuous, and sufficiently rigid oil film in the bearing clearances through lubrication, achieving stable, non-contact rotation between the rotor and the bearing surfaces. Radial bearings primarily support the radial load of the rotor, maintaining its concentricity; while thrust bearings withstand the axial thrust generated during rotor operation, limiting its axial movement.

[0003] In traditional turbochargers, such as Figure 8 As shown, the rotor shaft system 001 is typically supported radially by the floating bearing 002 and axially by the thrust bearing 003. The conventional design is based on the assumption that the rotor 004 operates under ideal alignment. However, in actual operation, the rotor 004 will produce a certain range of conical oscillations when rotating at high speed due to factors such as imbalance, oil film whirl, and external vibrations. This oscillation causes an angle between the oil wedge surface 005 of the thrust bearing 003 and the bearing surface 006, and between the journal 007 and the inner wall 008 of the floating ring. This transforms the originally designed surface or line contact into a local point contact, thereby significantly weakening the actual load-bearing capacity of the bearing under high-speed conditions.

[0004] To alleviate the aforementioned problems, existing technologies have introduced a 3D oil wedge thrust structure. By setting an inclination angle in the radial direction of the thrust bearing 003, it partially adapts to the changes in the contact surface angle caused by the rotor 004's wobble, thereby mitigating the decrease in load-bearing capacity. However, when the rotor 004 wobbles, this structure typically only forms 1-2 effective oil wedge surface 005 load-bearing areas. The remaining oil wedge surface load-bearing areas 005 cannot maintain sufficient pressure support due to the increased oil film gap, resulting in an overall load-bearing capacity that is still far below the design expectation.

[0005] With the widespread application of high-power-density, fast-response engines, the requirements for the operational stability and reliability of turbochargers are increasing. Insufficient bearing capacity of the shaft system has become a key bottleneck restricting its reliability improvement. Therefore, how to maintain the bearing system's near-ideal design bearing capacity under high-speed rotor yaw conditions has become a technical challenge that urgently needs to be overcome in this field.

[0006] In conclusion, the existing technology obviously has inconveniences and defects in practical use, so it is necessary to improve it. Summary of the Invention

[0007] To address the shortcomings of existing technologies, this invention provides a turbocharger device based on a high-load-bearing high-speed rotor shaft system structure. This addresses the problem in traditional technologies where, when the rotor wobbles, only 1-2 effective oil wedge bearing areas are typically formed, while the remaining oil wedge bearing areas cannot maintain sufficient pressure support due to increased oil film gaps, leading to a decrease in overall load-bearing capacity.

[0008] To address the above problems, the present invention provides the following technical solution: A turbocharger device based on a high-load-bearing high-speed rotor shaft system includes a bearing housing, a rotor shaft, a floating bearing, a thrust sleeve, a thrust bearing, and a shaft seal. The floating bearing is disposed in the bearing body, the rotor shaft is rotatably mounted on the floating bearing, the shaft seal and the thrust sleeve are mounted side by side on the rotor shaft along the axial direction, and the thrust bearing is located between the shaft seal and the thrust sleeve; The floating bearing has two symmetrical inner ring support surfaces inside. Each inner ring support surface includes an inner ring support surface one and an inner ring support surface two arranged side by side along the axial direction. There is an included angle between the inner ring support surface one and the inner ring support surface two. When the rotor shaft rotates horizontally, an inner oil film gap one is formed between the rotor coarse shaft of the rotor shaft and the inner ring support surface one. When the rotor shaft rotates yaw, an inner oil film gap two is formed between the rotor coarse shaft of the rotor shaft and the inner ring support surface two. The two side walls of the thrust bearing are respectively the pressure end axial force bearing surface and the vortex end axial force bearing surface. An arc-shaped gap one is provided between the pressure end axial force bearing surface and the thrust sleeve bearing surface of the thrust sleeve; an arc-shaped gap two is provided between the vortex end axial force bearing surface and the shaft seal bearing surface of the shaft seal.

[0009] As an optimized solution, the meridional contours of the pressure end axial force bearing surface one, the vortex end axial force bearing surface two, the thrust sleeve bearing surface three, and the shaft seal bearing surface four are all circular arcs, with the center of the circular arc located at the yaw center of the rotor shaft.

[0010] As an optimized solution, the yaw center is located on the rotor shaft axis at the midpoint between the two inner ring support surfaces.

[0011] As an optimized solution, the floating bearing is provided with a number of radial oil supply holes 1 around the middle position of the inner ring support surface 1, and the floating bearing is provided with a number of radial oil supply holes 2 around the middle position of the inner ring support surface 2, and the radial oil supply holes 1 and radial oil supply holes 2 are arranged alternately.

[0012] As an optimized solution, the bearing body is provided with a main oil supply hole, an axial oil supply hole and a floating oil supply hole that are interconnected. The outlet end of the floating oil supply hole is located at the position corresponding to the radial oil supply hole one and the radial oil supply hole two on the surface hole formed in the bearing body. The axial length of the surface hole is not less than the sum of the axial lengths of the radial oil supply hole one and the radial oil supply hole two.

[0013] As an optimized solution, the thrust bearing is provided with a central oil supply hole. One end of the central oil supply hole is connected to the axial oil supply hole, and the other end penetrates the inner ring surface of the thrust bearing and is connected to the thrust oil supply radial clearance between the inner ring surface of the thrust bearing and the outer wall surface of the small diameter of the thrust sleeve.

[0014] As an optimized solution, several sets of oil wedge surfaces, oil wedge flat sections and radial oil grooves are evenly distributed along the circumference on the pressure end axial force bearing surface and the vortex end axial force bearing surface. The oil wedge surfaces have a height difference of 15-40μm at both ends in the circumferential direction.

[0015] As an optimized solution, the floating bearing is provided with two symmetrical outer ring support surfaces, and an outer oil film gap is formed between the outer ring support surfaces and the bearing body.

[0016] As an optimized solution, one end of the floating bearing is circumferentially positioned inside the bearing body by a floating bearing locating pin, and the other end of the floating bearing is axially positioned inside the bearing body by a thick shaft shoulder.

[0017] As an optimized solution, the included angle between the inner ring support surface one and the inner ring support surface two is 0.09-0.11°.

[0018] As an optimized solution, the gap size between the inner oil film gap one and the inner oil film gap two is 0.003-0.005mm, and the axial width is 2-6mm.

[0019] As an optimized solution, the gap size between the first arc-shaped gap and the second arc-shaped gap is 10-40 μm.

[0020] As an optimized solution, the radial clearance of the thrust oil supply is 0.2-0.5 mm.

[0021] Compared with the prior art, the beneficial effects of the present invention are: The core of this invention is to enable the bearing system to have the ability to adapt to rotor shaft runout: the floating bearing is divided into two sections with a small included angle, inner ring support surface one and inner ring support surface two, to ensure that at least one side of the rotor can maintain a stable and uniform inner oil film gap when it runs out of control, thus avoiding local point contact. Meanwhile, the bearing surfaces of the thrust bearing corresponding to the thrust sleeve and the shaft seal both adopt an arc profile with the rotor's theoretical deflection center as the center. This ensures that even under deflection conditions, the mating surfaces can still maintain uniform arc-shaped gap one and arc-shaped gap two. This key design ensures that the lubricating oil wedge can continuously and effectively build up the design pressure throughout the entire circumference. Thus, when the rotor deflects, the axial and radial bearings can still maintain a near-ideal alignment state with multi-region uniform bearing, significantly improving the overall stiffness and load-bearing capacity of the shaft system under high speed and variable working conditions. The staggered radial oil supply holes 1 and 2 on the floating bearing precisely correspond to the floating oil supply holes on the bearing body, achieving dynamic and sufficient lubrication of the gap between the inner and outer oil films. The axial and circumferential positioning structures of the floating bearing enhance the position retention of the entire bearing system under complex stress. These measures together ensure that the lubricating oil film is evenly distributed and has sufficient pressure under complex rotor dynamics, effectively reducing the risk of failure caused by local temperature rise, wear or instantaneous oil shortage, and further consolidating operational stability. Therefore, this invention significantly improves the reliability and service life of turbochargers under high speed and dynamic load. Its ingenious and compact structural design, without significantly increasing manufacturing costs and complexity, enables the bearing system to have excellent runout tolerance through a series of mutually cooperating geometric configurations and oil circuit optimizations. It is particularly suitable for high power density engines with extremely high requirements for reliability and response speed, and has important engineering application value and promotion prospects. Attached Figure Description

[0022] To more clearly illustrate the specific embodiments of the present invention or the technical solutions in the prior art, the accompanying drawings used in the description of the specific embodiments or the prior art will be briefly introduced below. In all the drawings, similar elements or parts are generally identified by similar reference numerals. In the drawings, the elements or parts are not necessarily drawn to scale.

[0023] Figure 1 This is a schematic diagram of the structure of the present invention; Figure 2 for Figure 1 An enlarged schematic diagram of part A in the middle; Figure 3 for Figure 2 Enlarged schematic diagram of part B; Figure 4 This is a schematic diagram of the thrust bearing of the present invention; Figure 5 for Figure 4 An enlarged schematic diagram of section C; Figure 6 This is a schematic diagram of the structure of the floating bearing of the present invention; Figure 7 This is a schematic diagram of the bearing body of the present invention; Figure 8 This is a schematic diagram of the existing technology.

[0024] In the diagram: 1-Centrifugal compressor; 101-Compressor casing; 102-Impeller; 103-Locking nut; 2-Radial turbine; 201-Vortex casing; 202-Turbine; 203-Heat shield; 3-Bearing system; 4-Rotor shaft; 401-Rotor coarse shaft; 402-Coarse shaft shoulder; 403-Rotor fine shaft; 404-Fine shaft shoulder; 5-Floating bearing; 501-Inner ring support surface; 502-Inner ring support surface one; 5 03-Inner ring support surface two; 504-Inner oil film gap one; 505-Inner oil film gap two; 506-Outer ring support surface; 507-Outer oil film gap; 508-Radial oil supply hole one; 509-Radial oil supply hole two; 510-Floating bearing locating pin; 511-Surface hole; 6-Bearing body; 601-Bearing seat; 602-Main oil supply hole; 603-Axial oil supply hole; 604-Floating oil supply hole; 605-Thrust bearing 7-Thrust sleeve; 701-Thrust sleeve minor diameter outer wall surface; 702-Thrust sleeve inner hole; 703-Thrust sleeve bearing surface; 704-Meridian profile three; 705-Arc-shaped clearance one; 706-Thrust sleeve non-bearing surface; 8-Thrust bearing; 801-Central oil supply hole; 802-Thrust bearing inner ring surface; 803-Thrust oil supply radial clearance; 804-Pressure end axial force bearing surface; 805-Scroll end axial force bearing surface. Cross-section; 806-Meridian profile one; 807-Oblique center; 808-Meridian profile two; 809-Oil wedge surface; 810-Oil wedge flat section; 811-Radial oil groove; 9-Compressor sealing assembly; 901-Impeller sleeve; 902-Oil baffle; 903-Shaft seal; 904-Sealing ring; 9041-Shaft seal inner hole; 9042-Shaft seal bearing surface; 9043-Meridian profile four; 9044-Arc-shaped gap two. Detailed Implementation

[0025] The embodiments of the technical solution of the present invention will now be described in detail with reference to the accompanying drawings. These embodiments are merely illustrative of the technical solution of the present invention and are therefore intended to limit the scope of protection of the present invention.

[0026] like Figures 1 to 7As shown, a turbocharger device based on a high-load-bearing, high-speed rotor shaft system includes: a centrifugal compressor 1, a radial turbine 2, and a bearing system 3. The bearing system 3 includes a rotor shaft 4, a floating bearing 5, a bearing housing 6, a thrust sleeve 7, a thrust bearing 8, and a compressor sealing assembly 9.

[0027] The floating bearing 5 has two symmetrical inner ring support surfaces 501 inside. Each inner ring support surface 501 is divided into inner ring support surface one 502 and inner ring support surface two 503.

[0028] The inner ring support surface 502 and the inner ring support surface 503 form an angle of 0.09-0.11°. When the rotor is running horizontally, an inner oil film gap 504 is formed between the inner ring support surface 502 and the rotor coarse shaft 401 of the rotor shaft 4. The gap size of the inner oil film gap 504 is controlled within the range of 0.003-0.005mm, and the width of the inner oil film gap 504 is controlled within the range of 2-6mm.

[0029] When the rotor shaft 4 is wobbling, due to the tilt of the rotor shaft 4, an inner oil film gap 2 505 is formed between the inner ring support surface 2 503 and the rotor coarse shaft 401. The gap size of the inner oil film gap 2 505 is controlled within the range of 0.003-0.005mm, and the width of the inner oil film gap 2 505 is controlled within the range of 2-6mm.

[0030] The floating bearing 5 is provided with two outer ring support surfaces 506. The outer ring support surfaces 506 and the bearing housing 601 form an outer oil film gap 507. The gap size of the outer oil film gap 507 is controlled within the range of 0.03-0.05mm.

[0031] At the axial midpoint of the inner ring support surface 502 corresponding to the floating bearing 5, radial oil supply holes 508 are evenly distributed around the circumference. At the axial midpoint of the inner ring support surface 503, radial oil supply holes 509 are evenly distributed around the circumference. The diameter of radial oil supply holes 508 and radial oil supply holes 509 is set in the range of 1-3mm, and the number is set to 3-4, which are staggered in the circumferential direction.

[0032] The floating bearing 5 is circumferentially positioned inside the bearing body by the floating bearing positioning pin 510, and axially positioned by the end face of the floating bearing positioning pin 510 and the shoulder 402 of the coarse shaft.

[0033] A main oil supply hole 602, an axial oil supply hole 603, and a floating oil supply hole 604 are provided on the bearing housing 6. The floating oil supply hole 604 forms a surface hole 511 on the surface of the bearing seat 601. The surface hole 511 is located at the corresponding position of the first radial oil supply hole 508 and the second radial oil supply hole 509, and the axial length of the surface hole 511 is not less than the sum of the axial lengths of the first radial oil supply hole 508 and the second radial oil supply hole 509.

[0034] A thrust bearing positioning surface 605 is provided at the end face of the axial oil supply hole 603 of the bearing body 6. The impeller sleeve 901, the oil baffle 902 and the thrust bearing 8 are pressed together on the thrust bearing positioning surface 605 by clamping screws or retaining ring structure.

[0035] The thrust bearing 8 has a central oil supply hole 801. One end of the central oil supply hole 801 is connected to the axial oil supply hole 603, and the other end is connected to the inner ring surface 802 of the thrust bearing. The inner ring surface 802 of the thrust bearing and the outer wall surface 701 of the small diameter of the thrust sleeve form a thrust oil supply radial clearance 803. The thrust oil supply radial clearance 803 is controlled within the range of 0.2-0.5mm.

[0036] The thrust bearing 8 has a pressure end axial force bearing surface 804 and a volute end axial force bearing surface 805 on its two sides, respectively.

[0037] The meridional profile 806 of the axial force bearing surface 804 at the pressure end is an arc. The center of this arc is located on the yaw center 807 of the rotor shaft 4, which is the axis of the rotor shaft 4, and is located in the middle of the two inner ring support surfaces 502.

[0038] The radius of the meridian profile 806 is any point from the sway center 807 to the edge of the pressure end axial force bearing surface 804.

[0039] The meridional profile 808 of the axial force bearing surface 805 at the vortex end is an arc, and the center of this arc is also the yaw center 807. The radius of the meridional profile 808 is any point from the yaw center 807 to the edge of the axial force bearing surface 805 at the vortex end.

[0040] Four to eight sets of oil wedge surfaces 809, oil wedge flat sections 810 and radial oil grooves 811 are evenly distributed on the pressure end axial force bearing surface 804 and the vortex end axial force bearing surface 805. The oil wedge surface 809 has a height difference h in the circumferential direction, and the height difference h is controlled within the range of 15-40μm.

[0041] The thrust sleeve 7 is stepped and rotary. The inner hole 702 of the thrust sleeve and the rotor shaft 403 are in clearance fit or transition fit. The meridional profile 704 of the bearing surface 703 of the thrust sleeve is an arc. A uniform arc-shaped gap 705 is formed between the meridional profile 704 and the meridional profile 806. The gap size of the arc-shaped gap 705 is controlled between 10-40μm.

[0042] The non-load-bearing surface 706 of the thrust sleeve is an annular plane with a meridional profile perpendicular to the rotor shaft 4.

[0043] The compressor sealing assembly 9 includes an impeller sleeve 901, an oil baffle 902, a shaft seal 903, and a sealing ring 904.

[0044] The shaft seal 903 is stepped and rotary. The inner hole 9041 of the shaft seal and the rotor thin shaft 403 are clearance fit or transition fit. The meridional profile 9043 of the bearing surface 9042 of the shaft seal is an arc. A uniform arc-shaped gap 9044 is formed between the meridional profile 9043 and the meridional profile 808. The gap size of the arc-shaped gap 9044 is controlled within the range of 10-40μm.

[0045] The centrifugal compressor 1 includes a compressor housing 101, an impeller 102, and a locking nut 103.

[0046] The impeller 102, shaft seal 903 and thrust sleeve 7 are locked together on the thin shaft shoulder 404 by the locking nut 103.

[0047] The radial turbine 2 includes a vortex casing 201, a turbine 202, and a heat shield 203.

[0048] The turbine 202 is fixed to the rotor shaft 4 by friction welding or electron beam welding.

[0049] The vortex housing 201 and the pressure housing 101 are fixedly connected by bolts, pressure plate structure or clamp structure, and together they form a turbocharger device.

[0050] The working principle of this device is as follows: The core of this invention is to enable the bearing system 3 to have the adaptive capability to the yaw of the rotor shaft 4: the floating bearing 5 is internally divided into two sections with a small included angle, an inner ring support surface 502 and an inner ring support surface 503, to ensure that at least one side of the rotor can maintain a stable and uniform inner oil film gap when yawing, thus avoiding local point contact. Meanwhile, the bearing surfaces of the thrust bearing 8, the thrust sleeve 7, and the shaft seal 903 are all designed with an arc profile centered on the rotor's theoretical deflection center 807. This ensures that even under deflection conditions, the mating surfaces can maintain uniform arc gap 705 and arc gap 9044. This key design ensures that the lubricating oil wedge can continuously and effectively build up the design pressure throughout the entire circumference. As a result, when the rotor deflects, the axial and radial bearings can still maintain a near-ideal alignment state with uniform multi-area bearing, significantly improving the overall stiffness and bearing capacity of the shaft system under high speed and variable working conditions. The radial oil supply holes 508 and 509 arranged alternately on the floating bearing 5 correspond precisely to the floating oil supply hole 604 on the bearing body 6, achieving dynamic and sufficient lubrication of the gap between the inner and outer oil films. The axial and circumferential positioning structures of the floating bearing 5 enhance the position retention of the entire bearing system 3 under complex stress. These measures together ensure that the lubricating oil film can be evenly distributed and have sufficient pressure under complex rotor dynamics, effectively reducing the risk of failure caused by local temperature rise, wear or instantaneous oil shortage, and further consolidating the operational stability. Therefore, this invention significantly improves the reliability and service life of turbochargers under high speed and dynamic load. Its structure is ingenious and compact. Without significantly increasing manufacturing costs and complexity, through a series of mutually cooperating geometric configurations and oil circuit optimizations, the bearing system 3 has excellent yaw tolerance. It is particularly suitable for high power density engines with extremely high requirements for reliability and response speed, and has important engineering application value and promotion prospects.

[0051] Finally, it should be noted that the above embodiments are only used to illustrate the technical solutions of the present invention, and not to limit them. Although the present invention has been described in detail with reference to the foregoing embodiments, those skilled in the art should understand that modifications can still be made to the technical solutions described in the foregoing embodiments, or equivalent substitutions can be made to some or all of the technical features therein. Such modifications or substitutions do not cause the essence of the corresponding technical solutions to deviate from the scope of the technical solutions of the embodiments of the present invention, and they should all be covered within the scope of the claims and specification of the present invention.

Claims

1. A turbocharger device based on a high-load-bearing high-speed rotor shaft system structure, comprising a bearing housing (6), a rotor shaft (4), a floating bearing (5), a thrust sleeve (7), a thrust bearing (8), and a shaft seal (903); the floating bearing (5) is disposed within the bearing housing (6), the rotor shaft (4) is rotatably mounted on the floating bearing (5), the shaft seal (903) and the thrust sleeve (7) are axially mounted side-by-side on the rotor shaft (4), and the thrust bearing (8) is located between the shaft seal (903) and the thrust sleeve (7); characterized in that: The floating bearing (5) is symmetrically provided with two inner ring support surfaces (501). Each inner ring support surface (501) includes an inner ring support surface one (502) and an inner ring support surface two (503) arranged side by side along the axial direction. An included angle is provided between the inner ring support surface one (502) and the inner ring support surface two (503). When the rotor shaft (4) is running horizontally, an inner oil film gap one (504) is formed between the rotor coarse shaft (401) of the rotor shaft (4) and the inner ring support surface one (502). When the rotor shaft (4) is running yaw, an inner oil film gap two (505) is formed between the rotor coarse shaft (401) of the rotor shaft (4) and the inner ring support surface two (503). The two side walls of the thrust bearing (8) are the pressure end axial force bearing surface (804) and the vortex end axial force bearing surface (805), respectively. An arc-shaped gap one (705) is provided between the pressure end axial force bearing surface (804) and the thrust sleeve bearing surface (703) of the thrust sleeve (7); an arc-shaped gap two (9044) is provided between the vortex end axial force bearing surface (805) and the shaft seal bearing surface (9042) of the shaft seal (903).

2. The turbocharger device based on a high-load-bearing high-speed rotor shaft system structure according to claim 1, characterized in that: The meridional contour one (806) of the pressure end axial force bearing surface (804), the meridional contour two (808) of the vortex end axial force bearing surface (805), the meridional contour three (704) of the thrust sleeve bearing surface (703) and the meridional contour four (9043) of the shaft seal bearing surface (9042) are all circular arcs, and the center of the circular arc is located at the yaw center (807) of the rotor shaft (4). The yaw center (807) is located on the axis of the rotor shaft (4) at the middle position of the two inner ring support surfaces (502).

3. The turbocharger device based on a high-load-bearing high-speed rotor shaft system structure according to claim 1, characterized in that: The floating bearing (5) is located at the middle position of the inner ring support surface (502) and has a number of radial oil supply holes (508) arranged in the circumferential direction. The floating bearing (5) is located at the middle position of the inner ring support surface (503) and has a number of radial oil supply holes (509) arranged in the circumferential direction. The radial oil supply holes (508) and radial oil supply holes (509) are arranged alternately.

4. The turbocharger device based on a high-load-bearing high-speed rotor shaft system structure according to claim 3, characterized in that: The bearing body (6) has a main oil supply hole (602), an axial oil supply hole (603) and a floating oil supply hole (604) that are interconnected. The outlet end of the floating oil supply hole (604) forms a surface hole (511) on the bearing body (6). The surface hole (511) is located at the position corresponding to the first radial oil supply hole (508) and the second radial oil supply hole (509). The axial length of the surface hole (511) is not less than the sum of the axial lengths of the first radial oil supply hole (508) and the second radial oil supply hole (509).

5. The turbocharger device based on a high-load-bearing high-speed rotor shaft system structure according to claim 4, characterized in that: The thrust bearing (8) is provided with a central oil supply hole (801). One end of the central oil supply hole (801) is connected to the axial oil supply hole (603), and the other end passes through the inner ring surface (802) of the thrust bearing and is connected to the thrust oil supply radial clearance (803) between the inner ring surface (802) of the thrust bearing and the outer wall surface (701) of the small diameter of the thrust sleeve.

6. The turbocharger device based on a high-load-bearing high-speed rotor shaft system structure according to claim 2, characterized in that: Several sets of oil wedge surfaces (809), oil wedge flat sections (810) and radial oil grooves (811) are evenly distributed along the circumference on the pressure end axial force bearing surface (804) and the vortex end axial force bearing surface (805). The oil wedge surface (809) has a height difference of 15-40μm at both ends in the circumferential direction.

7. The turbocharger device based on a high-load-bearing high-speed rotor shaft system structure according to claim 1, characterized in that: The floating bearing (5) has two symmetrical outer ring support surfaces (506) on its outside, and an outer oil film gap (507) is formed between the outer ring support surface (506) and the bearing body (6).

8. The turbocharger device based on a high-load-bearing high-speed rotor shaft system structure according to claim 3, characterized in that: The included angle between the inner ring support surface one (502) and the inner ring support surface two (503) is 0.09-0.11°.

9. The turbocharger device based on a high-load-bearing high-speed rotor shaft system structure according to claim 1, characterized in that: The gap size between the inner oil film gap one (504) and the inner oil film gap two (505) is 0.003-0.005mm, and the axial width is 2-6mm.

10. The turbocharger device based on a high-load-bearing high-speed rotor shaft system structure according to claim 5, characterized in that: The gap size between the first arc-shaped gap (705) and the second arc-shaped gap (9044) is 10-40 μm; The clearance dimension of the thrust oil supply radial clearance (803) is 0.2-0.5mm.

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