TURBOCHARGER

Abstract

Turbochargers (10), comprising: a wheel (31A) comprising a hub (32) designed to be rotatable about a central axis (C) and a plurality of turbine blades (33) arranged on the outside of the hub (32) in a radial direction (Dr) at intervals in a circumferential direction (Dc) about the central axis (C); and a turbine housing (40) which is arranged on the outside of the impeller (31A) in the radial direction (Dr) and forms a spiral flow path (43, 44) which directs an exhaust gas to the impeller (31A) on the inside of the radial direction (Dr) while the exhaust gas is rotated in the circumferential direction (Dc), wherein a flow path width in the circumferential direction (Dc) of at least one of a plurality of inter-blade flow path sections (R, R21, R22) formed between the plurality of turbine blades (33) differs from a flow path width of another of the plurality of inter-blade flow path sections (R, R21, R22); wherein a radial dimension (ra) of the hub (32) of a first of the plurality of inter-blade flow path sections (R21) is larger than a radial dimension (rb) of the hub (32) of a second of the plurality of inter-blade flow path sections (R22) with a larger flow path width than that of the first of the plurality of inter-blade flow path sections (R21), wherein the hub (32) comprises a disk surface (32f, 32fa, 32fb) on which the majority of turbine blades (33) are arranged, and wherein the disk surface (32f, 32fa, 32fb) of the hub (32) in the first inter-blade flow path section (R21) is located further outwards in the radial direction (Dr) than the disk surface (32f, 32fa, 32fb) of the hub (32) in the second inter-blade flow path section (R22).

Description

[Technical field]

[0001] The present disclosure relates to a turbocharger. State of the art

[0002] A rotating machine, such as a turbocharger, has been constantly improved in various ways to enhance efficiency, suppress vibrations, reduce noise, and the like.

[0003] For example, patent document 1 discloses a configuration that avoids resonance with a turbine blade by allowing one throat width of a part of a turbine guide vane to differ from the other normal throat width in one circumferential direction. [Citation list][Patent document]

[0004] [Patent document 1] Japanese patent JP 3 181 200 B2

[0005] DE 10 2015 005 121 A1 relates to a turbine wheel for a turbine of an exhaust gas turbocharger, with a blading comprising a plurality of turbine blades, each spaced apart from the others by at least one pitch angle in the circumferential direction of the turbine wheel, wherein at least two different pitch angles are provided in the circumferential direction of the turbine wheel.

[0006] US 2014 / 0 241 866 A1 reveals a turbocharger turbine wheel comprising a hub and a multitude of blades.

[0007] DE 11 2015 001 237 T5 relates to an exhaust gas turbocharger, preferably for a passenger car, comprising a compressor with a compressor wheel, a turbine with a turbine wheel and a turbine housing, and a shaft connecting the compressor wheel to the turbine wheel. [Summary of the invention][Technical problem]

[0008] Furthermore, the turbocharger incorporates a spiral section to guide exhaust gas from the engine to a turbine blade on the inside of an impeller in a radial direction, while the exhaust gas is rotated radially on the outside of the impeller in a circumferential direction. To improve turbocharger efficiency, it is preferable to minimize the gap between a tongue section of the spiral section, located on the outside of the impeller in a radial direction, and the turbine blade that forms the impeller. However, if the gap between the tongue section and the turbine blade is made too small, noise can occur due to pressure fluctuations when the turbine blade passes through a low-velocity exhaust gas flow region that occurs on the downstream side of the tongue section.

[0009] The present disclosure was made to solve the problems described above and provides a turbocharger capable of suppressing noise. [Solution to the problem]

[0010] The aforementioned disadvantages are at least partially solved by a turbocharger according to claim 1. [Advantageous effects of the invention]

[0011] According to the turbocharger described in this disclosure, it is possible to suppress noise. [Brief description of the drawings] Fig. Figure 1 is a view showing a schematic configuration of a turbocharger of an embodiment of the present disclosure. Fig. Figure 2 is a view in which a turbine of a turbocharger of a first embodiment not belonging to the invention is viewed from a central axis direction. Fig. Figure 3 is a view in which an impeller forming the turbine of the first embodiment not belonging to the invention of the present disclosure is viewed from the central axis direction. Fig. Figure 4 is a view showing a load fluctuation simulation result of the impeller of the first embodiment of the present disclosure, which does not belong to the invention. Fig. 5 is a view that shows a noise frequency distribution based on the simulation result of Fig. 4 shows. Fig. Figure 6 shows a load fluctuation simulation result of an impeller of a modified example of the first embodiment of the present disclosure, which is not part of the invention. Fig. Figure 7 is a view that shows a noise frequency distribution based on the simulation result of Fig. 6 shows. Fig. Figure 8 is a view in which an impeller forming a turbine of a second embodiment not belonging to the invention is viewed from a central axis direction. Fig. Figure 9 is a view showing a load fluctuation simulation result of the impeller of the second embodiment of the present disclosure, which does not belong to the invention. Fig. 10 is a view that shows a noise frequency distribution based on the simulation result of Fig. 9 shows. Fig. Figure 11 is a view in which an impeller forming a turbine of a third embodiment according to the invention of the present disclosure is viewed from a central axis direction. Fig. Figure 12 is a view in which an impeller forming a turbine of a fourth embodiment not belonging to the invention is viewed from a central axis direction. [Description of embodiments]<Erste Ausführungsform>

[0012] The following describes a turbocharger of the present disclosure with reference to the drawings. (Turbocharger configuration)

[0013] As in Fig. As shown in Figure 1, a turbocharger 10 comprises a turbocharger body 11, a compressor 20, and a turbine 30. This turbocharger 10 is, for example, mounted on an automobile or the like as an engine accessory.

[0014] The turbocharger body 11 includes a bearing housing 12 and a rotating shaft 13.

[0015] The bearing housing 12 is supported by a vehicle body or the like by a bracket (not shown), the compressor 20, the turbine 30, and the like. The rotary shaft 13 is housed in a bearing housing 12. The rotary shaft 13 is supported by bearings 14A and 14B such that it can rotate about a central axis C in the bearing housing 12. Both ends of the rotary shaft 13, in the direction of the central axis C, project towards the outside of the bearing housing 12.

[0016] The compressor 20 is arranged at one end of the bearing housing 12 in the direction of the central axis C. The compressor 20 comprises a compressor wheel 21 and a compressor housing 22. The compressor wheel 21 is connected to the end section of the rotating shaft 13 outside the bearing housing 12. The compressor wheel 21 rotates about the central axis C integrally with the rotating shaft 13. The compressor housing 22 is connected to one end of the bearing housing 12 in the direction of the central axis C. The compressor housing 22 accommodates the compressor wheel 21. A flow path 22r for externally supplied air is formed within the compressor housing 22. The flow path 22r guides the externally supplied air to the compressor wheel 21 and directs the air that has passed through the compressor wheel 21 to a motor (not shown). (Turbine configuration)

[0017] The turbine 30 is located at the other end of the bearing housing 12. The turbine 30 comprises a turbine housing 40 and an impeller 31A.

[0018] The turbine housing 40 is connected to the other end of the bearing housing 12. The turbine housing 40 is arranged on the outside of the impeller 31A, centered radially Dr on the central axis C. The turbine housing 40 accommodates the impeller 31A within it.

[0019] As in Fig. As shown in Figure 2, the turbine housing 40 includes a gas inlet section (not shown) and spiral flow paths 43 and 44. The gas inlet section (not shown) opens to the outside and directs exhaust gas emitted by the engine (not shown) into the turbine housing 40.

[0020] The spiral flow paths 43 and 44 are formed within the turbine housing 40. The spiral flow paths 43 and 44 guide the exhaust gas, which drives the impeller 31A by rotating it, to gradually move inwards in the radial direction Dr while it rotates in the circumferential direction Dc.

[0021] The turbine casing 40 comprises a tubular circumferential wall 41 and an intermediate wall 42 formed within the circumferential wall 41. The circumferential wall 41 gradually extends inwards in the radial direction Dr, while overall it is continuous in a spiral shape from the gas inlet section (not shown) to the circumferential direction Dc. The intermediate wall 42 divides the inner side of the tubular circumferential wall 41 into the outer side in the radial direction Dr and the inner side centered on the central axis C. Accordingly, a spiral flow path 43, located on the outer side in the radial direction Dr with respect to the intermediate wall 42, and a spiral flow path 44, located on the inner side in the radial direction Dr with respect to the intermediate wall 42, are defined within the circumferential wall 41 of the turbine casing 40.

[0022] The circumferential wall 41 and the intermediate wall 42 include tongue sections 41s and 42s at a position adjacent to the impeller 31A on the outside of the radial direction Dr. The tongue section 41s of the circumferential wall 41 and the tongue section 42s of the intermediate wall 42 are arranged at positions that differ from each other by approximately 180° in the circumferential direction Dc, with the impeller 31A positioned between them. The spiral flow path 43 and the spiral flow path 44 include nozzle sections 43n and 44n that open towards the impeller 31A on the inside of the radial direction Dr. The nozzle section 43n and the nozzle section 44n face each other in the radial direction Dr, with the impeller 31A positioned between them. The nozzle section 43n of the spiral flow path 43 is open over approximately 180° in the circumferential direction Dc between the tongue section 41s of the circumferential wall 41 and the tongue section 42s of the intermediate wall 42.The nozzle section 44n of the spiral flow path 44 is open over approximately 180° in the circumferential direction Dc between the tongue section 42s of the intermediate wall 42 and the tongue section 41s of the circumferential wall 41.

[0023] Furthermore, as in Fig. Figure 1 shows the turbine housing 40 being provided with an exhaust section 45 formed on the inside of the radial direction Dr of the circumferential wall 41 to discharge an exhaust gas emitted by the impeller 31A to the central axis C.

[0024] Exhaust gas is fed from the engine (not shown) to the gas inlet section (not shown) of such a turbine housing 40. The exhaust gas flowing from the gas inlet section (not shown) gradually flows inwards in the radial direction Dr, while rotating in the circumferential direction Dc on the outside of the radial direction Dr of the impeller 31A along the spiral flow paths 43 and 44. The exhaust gas is fed from the nozzle sections 43n and 44n to the impeller 31A on the inside of the radial direction Dr. The impeller 31A rotates about the central axis C by being struck by the exhaust gas flow. The exhaust gas that has passed through the impeller 31A is discharged to the outside from the inside of the radial direction Dr of the impeller 31A through the outlet section 45.

[0025] Due to the rotation of such impeller 31A, the rotating shaft 13 and the compressor wheel 21 rotate about the central axis C in conjunction with the impeller 31A. The compressor wheel 21 rotates about the central axis C to compress air introduced from outside into the compressor housing 22. The air compressed by the compressor 20 is supplied from the compressor housing 22 to the motor (not shown). (Configuration of the wheel)

[0026] The impeller 31A is located at the opposite end of the rotating shaft 13, in the direction of the central axis C, on the outside of the bearing housing 12. The impeller 31A rotates about the central axis C in conjunction with the rotating shaft 13. As shown in the Fig. 1 to Fig. As shown in Figure 3, the impeller 31A is fitted in one piece with a hub 32 and a turbine blade 33.

[0027] The hub 32 is attached to the other end of the rotating shaft 13 in the direction of the central axis C. The hub 32 is integrally rotatable with the rotating shaft 13 about the central axis C. The hub 32 has a disk shape when viewed from the direction of the central axis C. The hub 32 has a constant thickness in the direction of the central axis C at a hub center section 32a on the inside in the radial direction Dr. The hub 32 gradually decreases in thickness in the direction of the central axis C outwards in the radial direction Dr from the hub center section 32a. The hub 32 includes a disk surface 32f which faces the outlet section 45 in the direction of the central axis C.The disk surface 32f is formed by a concave curved surface that gradually extends from the side closer to the outlet section 45 in the direction of the central axis C to the bearing housing 12 on the side opposite the outlet section 45 in the direction of the central axis C, while extending outwards in the radial direction Dr.

[0028] A plurality of the turbine blades 33 are arranged on the disk surface 32f at intervals in the circumferential direction Dc around the central axis C. In the embodiment of the present disclosure, the impeller 31A includes, for example, ten turbine blades 33.

[0029] Each turbine blade 33 includes a leading edge 33f and a trailing edge 33r. The leading edge 33f faces the nozzle sections 43n and 44n of the spiral flow paths 43 and 44, oriented outwards in the radial direction Dr. The leading edge 33f is arranged to separate a predetermined gap in the radial direction Dr between the tongue section 41s of the circumferential wall 41 and the tongue section 42s of the intermediate wall 42, which are located on the outside of the radial direction Dr. The trailing edge 33r faces the outlet section 45, oriented in the opposite direction along the central axis C.

[0030] As in Fig. As shown in Figure 3, in such an impeller 31A, an inter-blade flow path section R is formed between the turbine blades 33, which are adjacent to each other in the circumferential direction Dc. The exhaust gas, which flows inwards in the radial direction Dr from the spiral flow paths 43 and 44 on the outside of the impeller 31A in the radial direction Dr, passes through the inter-blade flow path section R between the turbine blades 33, which are adjacent to each other in the circumferential direction Dc. The exhaust gas, which flows inwards in the radial direction Dr from the leading edge 33f of the turbine blade 33, changes the flow direction due to the curvature of the disk surface 32f and is discharged along the central axis C from the trailing edge 33r of the turbine blade 33. The impeller 31A includes a plurality of inter-blade flow path sections R in the circumferential direction Dc.

[0031] In the impeller 31A, the flow path width in the circumferential direction Dc of at least one inter-blade flow path section R1 of the plurality of inter-blade flow path sections R formed between the plurality of turbine blades 33 differs from the flow path width of the other inter-blade flow path section R2. In the embodiment of the present disclosure, the impeller 31A includes an inter-blade flow path section R1 and an inter-blade flow path section R2 with different flow path widths in the circumferential direction Dc.

[0032] For example, the flow path width can be a clamping angle on the side of the leading edge 33f of the turbine blade 33 formed between the adjacent turbine blades 33.

[0033] For example, the inter-blade flow path section R1 can be formed between the turbine blades 33, which are adjacent to each other at a clamping angle θ1.

[0034] For example, the inter-blade flow path section R2 can be formed between the turbine blades 33, which are adjacent to each other at a clamping angle θ2.

[0035] For example, in the embodiment of the present disclosure, the clamping angle θ1 can be smaller than the clamping angle θ2.

[0036] For example, the clamping angle θ1 can be set to 31°.

[0037] For example, the clamping angle θ2 can be set to 41°.

[0038] For example, the inter-blade flow path section R1 and the inter-blade flow path section R2 can be arranged alternately in the circumferential direction Dc with different flow path widths.

[0039] If the thickness of the turbine blade 33 in the circumferential direction Dc is not negligible with respect to the clamping angle, the clamping angle can be an angle formed by the facing surfaces of the adjacent turbine blades 33. At this point, for example, the clamping angle θ1 can be set to an angle obtained by subtracting an angle corresponding to the thickness of the turbine blade 33 in the circumferential direction Dc from 31°, and the clamping angle θ2 can be set to an angle obtained by subtracting an angle corresponding to the thickness of the turbine blade 33 in the circumferential direction Dc from 41°.

[0040] However, the following describes a case that is different from the case of Fig. 3 includes, in which the thickness of the turbine blade 33 in the circumferential direction Dc is negligible with respect to the clamping angle.

[0041] In this way, the impeller 31A comprises a plurality of turbine blades 33 arranged at non-uniform intervals in the circumferential direction Dc. The impeller 31A has a total of five sets of two types of inter-blade flow path sections R1 and R2 with different flow path widths (clamping angles) in the circumferential direction Dc.

[0042] Here, two types of interblade flow path sections R1 and R2 with different flow path widths (clamping angles) in the circumferential direction Dc include a turbine blade group G1 (a first group of turbine blades) and a turbine blade group G2 (a second group of turbine blades), which comprises five turbine blades 33. Each of the turbine blade groups G1 and G2 comprises five turbine blades 33 arranged at equal intervals of 72° in the circumferential direction Dc. Each turbine blade 33 of turbine blade group G2 is arranged with a clockwise phase shift in the circumferential direction Dc relative to each turbine blade 33 of turbine blade group G1.

[0043] Therefore, it is possible to easily configure the majority of turbine blades 33, which are arranged at unequal intervals in the circumferential direction Dc, as a whole by combining five elements of each of the turbine blade groups G1 and G2 and forming them with a phase shift.

[0044] Thus, it is possible to easily manufacture the turbocharger 10.

[0045] When a simulation of load fluctuations (pressure fluctuations) applied to the tongue sections 41s and 42s is performed using such a runner 31A, the runner 31A rotates about the central axis C and the tongue sections 41s and 42s and the majority of turbine blades 33 move closer together, so that load fluctuations with unequal spacing, such as a pressure fluctuation waveform W1, which is in Fig. As shown in section 4, a similar simulation was performed for a case in which the majority of turbine blades were arranged at equal intervals in the circumferential direction Dc for comparison, and a pressure fluctuation waveform W2 exhibited load fluctuations at equal intervals.

[0046] Furthermore, when a frequency analysis is performed on the pressure fluctuation waveforms W1 and W2, load fluctuations are concentrated in a specific frequency range f if the majority of turbine blades are arranged at equal intervals in the circumferential direction Dc, as in Fig. 5 shown. In contrast, when the majority of turbine blades 33 were arranged at unequal intervals and the flow path width of the inter-blade flow path section R1 differed from the flow path width of the inter-blade flow path section R2 in the circumferential direction Dc, load fluctuations were distributed over several frequency ranges.

[0047] According to the turbocharger 10 of the embodiment described above, the circumferential flow path width Dc of a portion of the inter-blade flow path sections R1 differs from the flow path width of another inter-blade flow path section R2. Consequently, the pressure fluctuations occurring in the gaps between the majority of turbine blades 33 and the tongue sections 41s and 42s are unequal. Therefore, it is possible to distribute the frequency of the noise generated in the gaps without widening the gaps between the majority of turbine blades 33 and the tongue sections 41s and 42s. Accordingly, it is possible to reduce noise at a specific frequency. Thus, it is possible to suppress noise.

[0048] Furthermore, in the turbocharger 10 described above, each turbine blade 33 of the second group of turbine blades G2 is arranged with a phase shift in the circumferential direction Dc with respect to each turbine blade 33 of the first group of turbine blades G1.

[0049] Accordingly, it is possible to easily configure the majority of turbine blades 33, which are arranged at unequal intervals in the circumferential direction Dc, as a whole by combining each of the first group of turbine blades G1 and the second group of turbine blades G2 and arranging them with a phase shift.

[0050] Thus, it is possible to easily manufacture the turbocharger 10.

[0051] In particular, if the number of turbine blades 33 in the impeller 31A is an even number as in the turbocharger 10 described above, the impeller 31A can be configured such that the number of turbine blades 33 of the first group of turbine blades G1 and the number of turbine blades 33 of the second group of turbine blades G2 are the same.

[0052] Therefore, it is easy to balance the center of the impeller 31A when the first group of turbine blades G1 and the second group of turbine blades G2 are combined. (Modified example of the first embodiment)

[0053] In addition, in the embodiment described above, the clamping angle θ1 of the inter-blade flow path section R1 was set to 31° and the clamping angle θ2 of the inter-blade flow path section R2 was set to 41°. However, the present disclosure is not limited to this. The clamping angle θ1 of the inter-blade flow path section R1 and the clamping angle θ2 of the inter-blade flow path section R2 can be suitably modified.

[0054] For example, the clamping angle θ1 of the inter-blade flow path section R1 can be set to 26° and the clamping angle θ2 of the inter-blade flow path section R2 can be set to 46°. In this case, as shown in Fig. Figure 6 shows that load fluctuations with unequal spacing, such as a pressure fluctuation waveform W1', also occur in the impeller 31A, where the clamping angles θ1 and θ2 differ. The pressure fluctuation waveform W1' exhibits a greater deviation than the pressure fluctuation waveform W2 when the turbine blades 33 are installed at equal intervals. As shown in Fig. As shown in 7, if a frequency analysis is performed on the pressure fluctuation waveform W1' of Fig. 6 is carried out, the load fluctuations in other frequency ranges are more strongly distributed and the load peak (noise peak) in the specific frequency range f was greatly reduced by arranging the majority of turbine blades 33 at unequal intervals and adjusting the flow path width of the inter-blade flow path section R1 in the circumferential direction Dc so that it differs from the flow path width of the inter-blade flow path section R2. <Zweite Ausführungsform>

[0055] Next, a turbocharger of a second embodiment of the present disclosure will be described.

[0056] In the second embodiment, the same reference numerals are given for the same components as in the first embodiment, and detailed descriptions thereof are omitted. The turbocharger of the second embodiment differs from that of the first embodiment in the arrangement of the plurality of turbine blades 33.

[0057] As in Fig. Figure 8 shows an impeller 31B forming the turbine 30 of the turbocharger 10 of the second embodiment, for example ten turbine blades 33 arranged at intervals in the circumferential direction Dc similar to the first embodiment described above.

[0058] In the impeller 31B of the embodiment of the present disclosure, the plurality of inter-blade flow path sections R formed between the plurality of turbine blades 33 comprise five types of inter-blade flow path sections R11 to R15 with different flow path widths in the circumferential direction Dc. In five types of inter-blade flow path sections R11 to R15, the clamping angles θ11 to θ15 of the adjacent turbine blades 33 are different from one another. The inter-blade flow path section R11 is formed between the turbine blades 33 that are adjacent to each other at the clamping angle θ11. In the embodiment of the present disclosure, the clamping angle θ11 is, for example, set to 42°. The inter-blade flow path section R12 is formed between the turbine blades 33 that are adjacent to each other at the clamping angle θ12. The clamping angle θ12 is set to 39°, for example.The inter-blade flow path section R13 is formed between the turbine blades 33 that are adjacent to each other at the clamping angle θ13. The clamping angle θ13 is set to 33°, for example. The inter-blade flow path section R14 is formed between the turbine blades 33 that are adjacent to each other at the clamping angle θ14. The clamping angle θ14 is set to 30°, for example. The inter-blade flow path section R15 is formed between the turbine blades 33 that are adjacent to each other at the clamping angle θ15. The clamping angle θ15 is set to 36°, for example.

[0059] In this way, the impeller 31B incorporates the majority of turbine blades 33, which are arranged at unequal intervals in the circumferential direction Dc. The impeller 31B contains a total of two sets of inter-blade flow path sections R11 to R15, which include five types of clamping angles θ11 to θ15, arranged in the circumferential direction Dc.

[0060] Here, five types of interblade flow path sections R11 to R15, in which the flow path widths (clamping angles) differ in the circumferential direction Dc, include turbine blade groups G1 to G5 (a first group of turbine blades up to a fifth group of turbine blades), each containing two turbine blades 33. Each of the turbine blade groups G1 to G5 contains two turbine blades 33 arranged at equal intervals of 180° in the circumferential direction Dc. Each turbine blade 33 of the turbine blade groups G1 to G5 is arranged with a phase shift in the circumferential direction Dc.

[0061] Therefore, it is possible to easily configure the majority of turbine blades 33, which are arranged at unequal intervals in the circumferential direction Dc, as a whole by combining two elements of each of the turbine blade groups G1 to G5 and arranging them with a phase shift.

[0062] Thus, it is possible to easily manufacture the turbocharger 10.

[0063] When a simulation of load fluctuations (pressure fluctuations) applied to the tongue sections 41s and 42s is performed using such a runner 31B, the runner 31B rotates about the central axis C and the majority of turbine blades 33 and the tongue sections 41s and 42s approach each other, so that load fluctuations with an unequal spacing, such as the pressure fluctuation waveform W11, which is shown in Fig. 9 shows that load fluctuations occur when a frequency analysis is performed on the pressure fluctuation waveform W11, compared with the pressure fluctuation waveform W1 of the first embodiment, which is shown in Fig. 5 and Fig. Figure 7 shows a more pronounced distribution when the majority of turbine blades 33 are arranged at unequal intervals and the flow path widths of the inter-blade flow path sections R11 to R15 differ from each other in the circumferential direction Dc, as shown in Fig. 10 shown.

[0064] According to the turbocharger 10 of the embodiment described above, the flow path widths of the majority of the inter-blade flow path sections R11 to R15 differ from one another. Accordingly, it is possible to distribute the frequency of the noise generated in the gaps without widening the gaps between the majority of turbine blades 33 and the tongue sections 41s and 42s. Therefore, it is possible to reduce noise in the specific frequency range f. Thus, it is possible to suppress noise. <Dritte Ausführungsform>

[0065] Next, a turbocharger of a third embodiment of the present disclosure will be described.

[0066] In the third embodiment, the same reference numerals are given for the same components as in the first and second embodiments, and detailed descriptions thereof are omitted. The turbocharger of the third embodiment differs from that of the first embodiment in the structure of the hub 32.

[0067] As in Fig. Figure 11 shows an impeller 31C, which forms the turbine 30 of the turbocharger 10 of the third embodiment, for example ten turbine blades 33, which are arranged at intervals in the circumferential direction Dc similarly to the first embodiment described above.

[0068] Similar to the first embodiment described above, the impeller 31C of the embodiment of the present disclosure comprises a first inter-blade flow path section R21 and a second inter-blade flow path section R22, which have different flow path widths in the circumferential direction Dc. The first inter-blade flow path section R21 is formed between the turbine blades 33, which are adjacent to each other at a clamping angle θ21 (for example, 31°). The second inter-blade flow path section R22 is formed between the turbine blades 33, which are adjacent to each other at a clamping angle θ22 (for example, 41°). The impeller 31C has a total of five sets of two types of first inter-blade flow path sections R21 and second inter-blade flow path sections R22 with different flow path widths (clamping angles) in the circumferential direction Dc.

[0069] In such an impeller 31C, two types of first inter-blade flow path section R21 and second inter-blade flow path section R22, with different flow path widths (clamping angles) in the circumferential direction Dc, have different radial dimensions of the hub 32. The radial dimension ra of the hub 32 of the first inter-blade flow path section R21 is larger than the radial dimension rb of the hub 32 of the second inter-blade flow path section R22, which has a larger flow path width than the first inter-blade flow path section R21. Since the radial dimension ra of the hub 32 is large in the first inter-blade flow path section R21, which has a narrow flow path width, a disk surface 32fa of the hub 32 is located on the outside of the radial direction Dr with respect to a disk surface 32fb of the hub 32 in the second inter-blade flow path section R22.Accordingly, a gap in the circumferential direction Dc between roots 33k of the turbine blades 33, which are adjacent to each other in the first inter-blade flow path section R21, is essentially widened.

[0070] According to the turbocharger 10 of the embodiment described above, the radial dimension ra of the hub 32 is large in the first inter-blade flow path section R21 with a narrow flow path width, and the radial dimension rb of the hub 32 is small in the second inter-blade flow path section R22 with a wide flow path width. Accordingly, it is also possible in the first inter-blade flow path section R21 with a narrow flow path width to easily machine or otherwise modify the curved surface of section 33j, which connects the root 33k of the turbine blade 33 and the disk surface 32fa of the hub 32. Furthermore, it becomes possible to increase the radius of curvature of section 33j, which connects the root 33k of the turbine blade 33 and the disk surface 32fa of the hub 32, and thus suppress the stress concentration at the root 33k of the turbine blade 33.

[0071] Furthermore, similar to the embodiments described above, the flow path widths of the first inter-blade flow path section R21 and the second inter-blade flow path section R22 were set to differ from each other. Accordingly, it is possible to reduce noise at the specific frequency. Thus, it is possible to suppress noise. <Vierte Ausführungsform>

[0072] Next, a turbocharger of a fourth embodiment of the present disclosure will be described.

[0073] In the fourth embodiment, the same reference numerals are given for the same components as in the first to third embodiments, and detailed descriptions thereof are omitted. The turbocharger of the fourth embodiment differs from those of the first to third embodiments in the arrangement of a plurality of turbine blades 33G on the side of the trailing edge 33r on the downstream side of the exhaust gas flow direction.

[0074] As in Fig. Figure 12 shows an impeller 31D forming the turbine 30 of the turbocharger 10 of the fourth embodiment, for example ten turbine blades 33G arranged at intervals in the circumferential direction Dc similar to the first embodiment described above.

[0075] The majority of turbine blades 33G of the embodiment of the present disclosure are arranged such that the trailing edge 33r on the downstream side of the exhaust gas flow direction has the same minimum flow path area between the blades that are adjacent to each other in the circumferential direction Dc.

[0076] On the other hand, in the majority of turbine blades 33G, the leading edges 33f are arranged at unequal intervals in the circumferential direction Dc on the upstream side of the exhaust gas flow direction.

[0077] Similar to the first embodiment described above, the impeller 31D of the embodiment of the present disclosure comprises a first inter-blade flow path section R31 and a second inter-blade flow path section R32, which have different flow path widths in the circumferential direction Dc. The first inter-blade flow path section R31 is formed between the turbine blades 33G, which are adjacent to each other at a clamping angle θ31 (for example, 31°) on the side of the leading edge 33f. The second inter-blade flow path section R32 is formed between the turbine blades 33G, which are adjacent to each other at a clamping angle θ32 (for example, 41°) on the side of the leading edge 33f.

[0078] On the other hand, the first inter-blade flow path section R31 and the second inter-blade flow path section R32 are arranged such that they have the same minimum flow path area between the adjacent blades by each having the same clamping angle θ41 on the trailing edge 33r side.

[0079] In the turbocharger 10 of the embodiment described above, the flow path widths of the first inter-blade flow path section R31 and the second inter-blade flow path section R32 on the side of the leading edge 33f on the upstream side of the exhaust gas flow direction were made different from each other. Accordingly, it is possible to suppress noise.

[0080] On the other hand, since the majority of turbine blades 33G are arranged such that they have the same minimum flow path area between the blades adjacent to each other in the circumferential direction Dc on the downstream side of the exhaust gas flow direction, the load applied to the majority of turbine blades 33G on the downstream side of the exhaust gas flow direction can be made uniform.

[0081] Additionally, in this embodiment, since the plurality of turbine blades 33G with the same thickness in the circumferential direction Dc of each turbine blade 33 are arranged at equal intervals in the circumferential direction Dc on the downstream side of the exhaust gas flow direction, the plurality of turbine blades are arranged such that they have the same minimum flow path area between the blades that are adjacent to each other in the circumferential direction Dc. However, any configuration can be used if the plurality of turbine blades are arranged such that they have the same minimum flow path area.

[0082] As a modified example, even if the majority of turbine blades 33G are arranged at unequal intervals in the circumferential direction Dc on the downstream side of the exhaust flow direction, the majority of turbine blades 33G can be arranged such that they have the same minimum flow path area between the blades that are adjacent to each other in the circumferential direction Dc by adjusting the thickness of each turbine blade 33 in the circumferential direction Dc to be different. (deleted)

[0083] Furthermore, in the embodiments described above, ten turbine blades 33 are provided, but the number of turbine blades 33 is not limited at all and can be set to a different number if necessary. As a modified example, the number of turbine blades 33 can be an odd number.

[0084] Furthermore, in the embodiments described above, the turbine housing 40 includes two spiral flow paths 43 and 44, but the present disclosure is not limited thereto.

[0085] As a modified example, the turbine housing 40 can only contain a spiral flow path.

[0086] As another modified example, the turbine housing 40 can have a configuration in which two spiral flow paths are arranged side by side in the direction of the central axis C. <anhang>

[0087] For example, the turbocharger 10 of each embodiment is to be understood as follows.

[0088] (1) The turbocharger 10 according to a first aspect comprises: the impeller 31A to 31D, which includes a hub 32 provided to be rotatable about a central axis C, and a plurality of turbine blades 33 and 33G, which are arranged on the outside of the hub 32 in a radial direction Dr at intervals in a circumferential direction Dc about the central axis C;and a turbine casing 40, which is arranged on the outside of the impeller 31A to 31D in the radial direction Dr and forms a spiral flow path 43 and 44, which directs an exhaust gas to the impeller 31A to 31D on the inside of the radial direction Dr, while the exhaust gas is rotated in the circumferential direction Dc, wherein a flow path width in the circumferential direction Dc of at least one R1, R11, R21 and R31 of a plurality of inter-blade flow path sections R formed between the plurality of turbine blades 33 and 33G differs from a flow path width of another R2, R12, R22 and R32 of the plurality of inter-blade flow path sections.

[0089] In this turbocharger 10, the majority of turbine blades 33 and 33G are arranged at unequal intervals. Accordingly, it is possible to distribute the frequency of the noise generated in the gaps without widening the gap between the majority of turbine blades 33 and 33G and the tongue sections 41s and 42s, which are arranged on the outside of the radial direction Dr of the turbine blades 33 and 33G in the turbine housing 40. Therefore, it is possible to reduce noise at a specific frequency. Thus, it is possible to suppress noise.

[0090] (2) The turbocharger 10 according to a second aspect is the turbocharger 10 of (1), wherein the plurality of turbine blades 33 and 33G comprises a first group of turbine blades G1 arranged at equal intervals in the circumferential direction Dc, and a second group of turbine blades G2 arranged at equal intervals in the circumferential direction, and wherein each turbine blade 33 and 33G of the second group of turbine blades G2 is arranged with a phase shift in the circumferential direction Dc with respect to each turbine blade 33 and 33G of the first group of turbine blades G1.

[0091] Accordingly, it is possible to easily configure the majority of turbine blades 33 and 33G, which are arranged at unequal intervals in the circumferential direction Dc, as a whole by combining each of the first group of turbine blades G1 and the second group of turbine blades G2 and arranging them with a phase shift.

[0092] Thus, it is possible to easily manufacture the turbocharger 10.

[0093] (3) The turbocharger 10 according to a third aspect is the turbocharger 10 of (1) or (2), wherein a radial dimension ra of the hub 32 of a first of the plurality of inter-blade flow path sections R21 is larger than a radial dimension rb of the hub 32 of a second of the plurality of inter-blade flow path sections R22 having a larger flow path width than that of the first of the plurality of inter-blade flow path sections R21.

[0094] Accordingly, the radial dimension ra of the hub 32 is large in the first of the plurality of interblade flow path sections R21 with a narrow flow path width, and the radial dimension rb of the hub 32 becomes smaller in the second of the plurality of interblade flow path sections R22 with a wide flow path width. In the first of the plurality of interblade flow path sections R21 with a narrow flow path width, it becomes possible to easily machine the curved surface of section 33j, which connects the root 33k of the turbine blade 33 and the disk surface 32f of the hub 32. Furthermore, it becomes possible to increase the radius of curvature of section 33j, which connects the root 33k of the turbine blade 33 and the disk surface 32f of the hub 32, and thus the stress concentration at the root 33k of the turbine blade 33 is suppressed.

[0095] (4) The turbocharger 10 of a fourth aspect is the turbocharger 10 according to any one of (1) to (3), wherein the plurality of turbine blades 33G are arranged such that they have the same minimum flow path area between the blades adjacent to each other in the circumferential direction Dc on a downstream side of an exhaust flow direction and are arranged at unequal intervals in the circumferential direction Dc on an upstream side of the exhaust flow direction.

[0096] Accordingly, the majority of turbine blades 33 are arranged at unequal intervals in the circumferential direction Dc on the upstream side of the exhaust gas flow direction. Consequently, the flow path widths of the inter-blade flow path sections R1 and R2 differ from each other in the circumferential direction Dc only in the section where the majority of turbine blades 33G are adjacent to the tongue sections 41s and 42s. Even with this configuration, it is possible to distribute the frequency of the noise generated in the gaps without widening the gaps between the turbine blade 33 and the tongue sections 41s and 42s. Therefore, it is possible to reduce noise at the specific frequency.

[0097] On the other hand, since the majority of turbine blades 33G are arranged at equal intervals in the circumferential direction Dc on the downstream side of the exhaust gas flow direction, the load applied to the majority of turbine blades 33G on the downstream side of the exhaust gas flow direction can be made uniform.

[0098] (5) The turbocharger 10 according to a fifth aspect is the turbocharger 10 according to one of (1) to (4), wherein the turbine housing 40 comprises a plurality of spiral flow paths 43 and 44, each of which supplies the exhaust gas to the impeller 31A to 31D from a plurality of different positions of the impeller 31A to 31D in the circumferential direction Dc.

[0099] Accordingly, it is possible to suppress noise in the turbocharger 10, which includes an exhaust gas inlet to the impeller 31A to 31D at various positions in the circumferential direction Dc. [Commercial Applicability]

[0100] According to the turbocharger described above, it is possible to suppress noise. [List of reference symbols] 10 turbochargers 11 Turbocharger bodies 12 bearing housings 13 Rotary shaft 14A, 14B bearings 20 compressors 21 compressor wheel 22 compressor housings 22r Flow path 30 Turbine 31A to 31D wheel 32 hub 32a Hub center section 32f, 32fa, 32fb disk area 33, 33G turbine blade 33f Front edge 33r trailing edge Section 33j 33k root 40 turbine housings 41 Perimeter wall 41s tongue section 42 Partition wall 42s tongue section 43, 44 Spiral flow path 43n, 44n nozzle section 45 Outlet section C Central axis DC circumferential direction Dr radial direction G1 Turbine blade group (first group of turbine blades) G2 Turbine blade group (second group of turbine blades) G3 to G5 turbine blade group (third group of turbine blades to fifth group of turbine blades) R, R1, R2, R11 to R15 Inter-blade flow path section R21, R31 first inter-blade flow path section R22, R32 second inter-blade flow path section W1, W1', W11, W2 Pressure fluctuation waveform f specific frequency range ra radial dimension rb radial dimension θ1, θ2, θ11 to θ15, θ21, θ22, θ31, θ32, θ41 clamping angle< / anhang>

Claims

[1] Turbocharger (10), comprising: a wheel (31A) comprising a hub (32) designed to be rotatable about a central axis (C) and a plurality of turbine blades (33) arranged on the outside of the hub (32) in a radial direction (Dr) at intervals in a circumferential direction (Dc) about the central axis (C); and a turbine housing (40) which is arranged on the outside of the impeller (31A) in the radial direction (Dr) and forms a spiral flow path (43, 44) which directs an exhaust gas to the impeller (31A) on the inside of the radial direction (Dr) while the exhaust gas is rotated in the circumferential direction (Dc), wherein a flow path width in the circumferential direction (Dc) of at least one of a plurality of inter-blade flow path sections (R, R21, R22) formed between the plurality of turbine blades (33) differs from a flow path width of another of the plurality of inter-blade flow path sections (R, R21, R22); wherein a radial dimension (ra) of the hub (32) of a first of the plurality of inter-blade flow path sections (R21) is larger than a radial dimension (rb) of the hub (32) of a second of the plurality of inter-blade flow path sections (R22) with a larger flow path width than that of the first of the plurality of inter-blade flow path sections (R21), wherein the hub (32) comprises a disk surface (32f, 32fa, 32fb) on which the majority of turbine blades (33) are arranged, and wherein the disk surface (32f, 32fa, 32fb) of the hub (32) in the first inter-blade flow path section (R21) is located further outwards in the radial direction (Dr) than the disk surface (32f, 32fa, 32fb) of the hub (32) in the second inter-blade flow path section (R22). [2] Turbocharger (10) according to claim 1, wherein the majority of turbine blades (33) comprise a first group of turbine blades (G1) arranged at equal intervals in the circumferential direction (Dc), and a second group of turbine blades (G2) arranged at equal intervals in the circumferential direction (Dc), and wherein each turbine blade (33) of the second group of turbine blades (G2) is arranged with a phase shift in the circumferential direction (Dc) with respect to each turbine blade (33) of the first group of turbine blades (G1). [3] Turbocharger (10) according to one of claims 1 or 2, wherein the turbine housing (40) comprises a plurality of spiral flow paths (43, 44) which each supply the exhaust gas to the impeller (31A) from a plurality of different positions of the impeller (31A) in the circumferential direction (Dc).

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