Centrifugal compressor casing, centrifugal compressor and turbocharger

Abstract

This casing for a centrifugal compressor is configured to be capable of accommodating an impeller, the casing comprising: a casing body including at least one scroll channel that extends along a circumferential direction of the impeller and through which flows compressed gas that has passed by the impeller; a first discharge tube that is connected to the at least one scroll channel via a first branch port, and that forms a first discharge flow path through which the compressed gas guided from the scroll channel flows; and a second discharge tube that is connected to the at least one scroll channel via a second branch port, and that forms a second discharge flow path through which the compressed gas guided from the scroll channel flows. The first branch port and the second branch port are provided at mutually different positions in the circumferential direction of the impeller.

Description

[Technical Field]

[0001] The present disclosure relates to a casing for a centrifugal compressor, a centrifugal compressor including the casing, and a turbocharger. [Background technology]

[0002] Conventionally, turbochargers (superchargers) have been widely used as a technology to improve the output of engines (internal combustion engines) such as automobile engines. These devices compress the intake air drawn into the engine, increasing its density and supplying the engine with more oxygen.

[0003] A turbocharger, for example, includes a centrifugal compressor provided at one end of a rotating shaft and a turbine provided at the other end of the rotating shaft. The turbine rotor (turbine impeller) is rotated by the energy of exhaust gas sent from the engine, which in turn rotates the centrifugal compressor impeller, which rotates in conjunction with the rotation of the turbine rotor, compressing the intake air and supplying it to the engine.

[0004] The operating range of a centrifugal compressor is expressed by the relationship between the intake air flow rate (FR) and the pressure ratio (PR) of the centrifugal compressor (see, for example, Figure 6). In recent years, due to the electrification of powertrains, centrifugal compressors for turbochargers are required to have a wider operating range than conventional centrifugal compressors and to have improved compression efficiency near the surge line of the operating range. In particular, there is a need to widen the operating range and improve compression efficiency in the low peripheral speed region.

[0005] It is known that surge occurring in the high peripheral speed region is caused by stall of the impeller of the centrifugal compressor, while surge occurring in the low peripheral speed region is caused by backflow occurring in the stationary passage system (diffuser passage and scroll passage) downstream of the impeller of the centrifugal compressor. In order to expand the surge margin in the low peripheral speed region, it is effective to improve the flow in the stationary passage system.

[0006] In the scroll passage of a centrifugal compressor, the discharge rate and flow angle change depending on the operating point, which can cause distortion in the static pressure distribution in the circumferential direction of the impeller at non-design operating points. This distortion in the static pressure distribution can increase losses in the scroll passage due to pressure recovery losses caused by unnecessary acceleration and deceleration of the gas in the scroll passage. Furthermore, this distortion in the static pressure distribution can cause a drift in the diffuser passage, which can lead to separation due to the drift and change the amount of work at the impeller outlet, thereby reducing the efficiency of the impeller and diffuser passage. In order to increase the surge margin in the low peripheral speed region, it is effective to mitigate the effects of the distortion in the static pressure distribution. [Prior art documents] [Patent documents]

[0007] [Patent Document 1] Patent No. 5124636 [Patent Document 2] Patent No. 6977889 Summary of the Invention [Problem to be solved by the invention]

[0008] In the past, the problem of distortion of static pressure distribution in the scroll passage of a centrifugal compressor has been addressed mainly by adjusting the circumferential cross-sectional area distribution or the central diameter distribution of the scroll passage. For example, in Patent Document 1, the diameter or passage width of the diffuser passage is made asymmetric in the circumferential direction to alleviate the distortion of static pressure distribution in the scroll passage. The invention described in Patent Document 1 can improve efficiency at operating points where the conventional static pressure distribution is significantly distorted (e.g., operating points on the low flow rate side), but there is a risk that the distortion of static pressure distribution may be exacerbated at non-design operating points (e.g., operating points on the high flow rate side).

[0009] A variable inlet compressor is known as a technology for expanding surge margins in low peripheral speed regions (see, for example, Patent Document 2). This variable inlet compressor uses a movable element at the compressor inlet to reduce the inlet flow area when operating in low flow rate regions, allowing the compressor to operate as if it had a low blade height and a low optimal flow rate. However, the efficiency improvement achieved by the variable inlet compressor is limited to a narrow region near the surge line (e.g., a region where the original efficiency is significantly low, such as an efficiency of 70% or less). Furthermore, reducing the inlet flow area increases windage loss within the impeller, resulting in a significant decrease in the pressure ratio as the peripheral speed increases. As a result, the expansion of surge margins is limited to the low peripheral speed region. Furthermore, reducing the inlet flow area can cause turbulence in the inflowing gas, potentially resulting in performance degradation. For this reason, the mechanism for reducing the inlet flow area and the shape of the throttle section must be carefully designed.

[0010] In view of the above circumstances, at least one embodiment of the present disclosure aims to provide a casing for a centrifugal compressor that can mitigate the influence of distortion of static pressure distribution in a scroll flow path and achieve high efficiency and a wide range of the centrifugal compressor, and a centrifugal compressor and a turbocharger that include the casing. [Means for solving the problem]

[0011] A casing of a centrifugal compressor according to at least one embodiment of the present disclosure includes: A casing for a centrifugal compressor configured to be able to accommodate an impeller, a casing body having at least one scroll passage extending along the circumferential direction of the impeller and through which compressed gas that has passed through the impeller flows; a first discharge pipe portion that communicates with the at least one scroll passage via a first branch port and forms a first discharge passage through which the compressed gas introduced from the scroll passage flows; a second discharge pipe portion that communicates with the at least one scroll passage via a second branch port and forms a second discharge passage through which the compressed gas introduced from the scroll passage flows, The first branch port and the second branch port are provided at different positions in the circumferential direction of the impeller.

[0012] A centrifugal compressor according to at least one embodiment of the present disclosure comprises: The centrifugal compressor includes a casing.

[0013] A turbocharger according to at least one embodiment of the present disclosure includes: the centrifugal compressor; a turbine configured to drive the centrifugal compressor. [Effects of the Invention]

[0014] According to at least one embodiment of the present disclosure, there are provided a casing for a centrifugal compressor that can mitigate the influence of distortion of static pressure distribution in a scroll flow path and achieve high efficiency and a wide range of the centrifugal compressor, as well as a centrifugal compressor and a turbocharger that include the casing. [Brief explanation of the drawings]

[0015] [Figure 1] 1 is a schematic diagram of an internal combustion engine system including a turbocharger according to an embodiment; [Figure 2] 1 is a schematic cross-sectional view taken along the axis of a turbocharger according to an embodiment of the present invention; [Figure 3] 1 is a schematic cross-sectional view taken along the axis of a centrifugal compressor according to one embodiment. [Figure 4] FIG. 10 is a schematic contour diagram of static pressure distribution in a scroll passage at a large flow rate side operating point of a centrifugal compressor according to a comparative example. [Figure 5] FIG. 10 is a schematic contour diagram of static pressure distribution in a scroll passage at a small-volume operating point of a centrifugal compressor according to a comparative example. [Figure 6] FIG. 10 is a diagram showing the relationship between the flow rate of intake air and the pressure ratio in a centrifugal compressor according to a comparative example. [Figure 7] FIG. 10 is an explanatory diagram for explaining a circumferential static pressure distribution in a scroll passage in a centrifugal compressor according to a comparative example. [Figure 8] FIG. 3 is a schematic contour diagram of static pressure distribution in a scroll passage at a high flow rate side operating point of a centrifugal compressor according to an embodiment. [Figure 9] FIG. 2 is a schematic contour diagram of static pressure distribution in a scroll passage at a small-volume operating point of a centrifugal compressor according to an embodiment. [Figure 10] FIG. 2 is a schematic contour diagram of static pressure distribution in a scroll passage at a high flow rate side operating point of a centrifugal compressor according to an embodiment. [Figure 11] FIG. 2 is a schematic contour diagram of static pressure distribution in a scroll passage at a small-volume operating point of a centrifugal compressor according to an embodiment. [Figure 12] 1 is a schematic view of a centrifugal compressor according to an embodiment, taken perpendicular to the axis of a casing. FIG. [Figure 13] 1 is a schematic view of a centrifugal compressor according to an embodiment, taken perpendicular to the axis of a casing. FIG. [Figure 14] 1 is a schematic cross-sectional view taken along the axis of a centrifugal compressor according to one embodiment. [Figure 15] 1 is a schematic view of a centrifugal compressor according to an embodiment, taken along the axis of a casing; [Figure 16] 1 is a schematic view of a centrifugal compressor according to an embodiment, taken along the axis of a casing; [Figure 17] FIG. 2 is a schematic cross-sectional view of a connection portion of a casing of a centrifugal compressor according to one embodiment and a pipe connected to the connection portion. [Figure 18] FIG. 2 is a schematic cross-sectional view of a connection portion of a casing of a centrifugal compressor according to one embodiment and a pipe connected to the connection portion. [Figure 19] 1 is a schematic cross-sectional view perpendicular to a casing of a centrifugal compressor according to one embodiment. FIG. [Figure 20] FIG. 10 is a diagram showing the relationship between the circumferential angular position and the A / R ratio in a centrifugal compressor according to a comparative example. [Figure 21] FIG. 2 is a diagram illustrating a relationship between a circumferential angle position and an A / R ratio in a centrifugal compressor according to one embodiment. [Figure 22] FIG. 2 is a diagram illustrating a relationship between a circumferential angle position and an A / R ratio in a centrifugal compressor according to one embodiment. [Figure 23] FIG. 2 is a diagram illustrating a relationship between a circumferential angle position and an A / R ratio in a centrifugal compressor according to one embodiment. [Figure 24] FIG. 2 is a diagram illustrating a relationship between a circumferential angle position and an A / R ratio in a centrifugal compressor according to one embodiment. [Figure 25] 1 is a schematic view perpendicular to an axis of a centrifugal compressor including an outlet-side flow control device according to an embodiment; [Figure 26] 1 is a schematic view taken along the axis of a centrifugal compressor including an outlet-side flow control device according to an embodiment. [Figure 27] FIG. 2 is a diagram illustrating the relationship between the flow rate of intake air and the pressure ratio in a centrifugal compressor including an outlet-side flow control device according to one embodiment. [Figure 28] 1 is a schematic cross-sectional view taken along the axis of a centrifugal compressor including an inlet-side flow control device according to one embodiment. [Figure 29] 1 is a schematic cross-sectional view taken along the axis of a centrifugal compressor including an inlet-side flow control device according to one embodiment. [Figure 30] 1 is a schematic cross-sectional view taken along the axis of a centrifugal compressor including an inlet-side flow control device according to one embodiment. [Figure 31] FIG. 2 is a diagram illustrating the relationship between the flow rate of intake air and the pressure ratio in a centrifugal compressor including an inlet-side flow control device according to one embodiment. DETAILED DESCRIPTION OF THE INVENTION

[0016] Hereinafter, several embodiments of the present disclosure will be described with reference to the accompanying drawings. However, the dimensions, materials, shapes, relative arrangements, etc. of components described as embodiments or shown in the drawings are merely illustrative examples and are not intended to limit the scope of the present disclosure.

[0017] In the following embodiments, the centrifugal compressor of the present disclosure will be described as being provided in a turbocharger, but the centrifugal compressor of the present disclosure may also be an electric centrifugal compressor, etc. Furthermore, the gas to be compressed by the centrifugal compressor of the present disclosure does not need to be limited to air. In other words, the centrifugal compressor of the present disclosure may be configured as a single centrifugal compressor or may be configured in combination with mechanisms or devices other than a turbine as long as it is capable of compressing and sending gas. Furthermore, there is no need to limit its use, etc.

[0018] (Centrifugal compressor, turbocharger) FIG. 1 is a schematic diagram of an internal combustion engine system 11 including a turbocharger 10 according to one embodiment. FIG. 2 is a schematic cross-sectional view taken along an axis LA of the turbocharger 10 according to one embodiment. FIG. 3 is a schematic cross-sectional view taken along an axis LA of a centrifugal compressor 1 according to one embodiment. As shown in FIGS. 1 to 3, a centrifugal compressor 1 according to some embodiments includes an impeller 2 and a casing 3 configured to rotatably house the impeller 2. The centrifugal compressor 1 according to the present disclosure can be mounted on a turbocharger (supercharger) 10 for, for example, an automobile, a marine vehicle, or an industrial application (for example, for land-based power generation).

[0019] 1 and 2, the turbocharger 10 includes a centrifugal compressor 1 and a turbine 12 configured to drive the centrifugal compressor 1. The turbine 12 includes a turbine rotor 14 that rotates using the energy of exhaust gas discharged from an internal combustion engine 13 (engine, see FIG. 1), and a turbine housing 15 configured to rotatably accommodate the turbine rotor 14.

[0020] The turbocharger 10 further includes a rotating shaft 16 having the impeller 2 connected to one end thereof and the turbine rotor 14 connected to the other end thereof, and a bearing 17 that rotatably supports the rotating shaft 16 between the impeller 2 and the turbine rotor 14. The turbocharger 10 may further include a bearing housing 18 that is disposed between the casing 3 and the turbine housing 15 and is configured to house the rotating shaft 16 and the bearing 17.

[0021] The turbine 12 (turbocharger 10) is configured to rotate a turbine rotor 14 using the energy of exhaust gas discharged from an internal combustion engine 13. The impeller 2 is coaxially connected to the turbine rotor 14 via a rotating shaft 16, and is therefore driven to rotate about an axis LA in conjunction with the rotation of the turbine rotor 14. The centrifugal compressor 1 (turbocharger 10) is configured to draw air (intake air, gas) into the casing 3, compress the air, and send the compressed air to the internal combustion engine 13 as the impeller 2 is driven to rotate about an axis LA.

[0022] The compressed air sent from the centrifugal compressor 1 to the internal combustion engine 13 is used for combustion in the internal combustion engine 13. Exhaust gas generated by the combustion in the internal combustion engine 13 is sent from the internal combustion engine 13 to the turbine 12, causing the turbine rotor 14 to rotate.

[0023] Hereinafter, the direction in which the axis LA of the impeller 2 extends will be referred to as the axial direction of the impeller 2, the direction perpendicular to the axis LA will be referred to as the radial direction of the impeller 2, and the circumferential direction around the axis LA will be referred to as the circumferential direction of the impeller 2.

[0024] (Turbine rotor) 2, the turbine rotor 14 includes a hub 141 having a substantially truncated cone shape and a plurality of turbine blades 142 provided on the outer peripheral surface of the hub 141. The hub 141 and the plurality of turbine blades 142 are provided so as to be rotatable integrally with the rotating shaft 16 about the axis LA. The turbine rotor 14 is configured to guide exhaust gas introduced from the outside in the radial direction of the turbine rotor 14 to the front side of the turbine rotor 14 along the axial direction of the turbine rotor 14.

[0025] (turbine housing) A turbine scroll passage 151 for guiding exhaust gas discharged from the internal combustion engine 13 to the turbine rotor 14 and an exhaust gas discharge passage 152 for discharging exhaust gas that has passed through the turbine rotor 14 to the outside of the turbine housing 15 are formed inside the turbine housing 15. The turbine scroll passage 151 is provided on the outer periphery of the turbine rotor 14 and consists of a spiral passage extending along the circumferential direction of the turbine rotor 14. The exhaust gas discharge passage 152 extends along the axial direction of the turbine rotor 14.

[0026] The exhaust gas discharged from the internal combustion engine 13 is guided to the turbine rotor 14 via the turbine scroll passage 151, and drives the turbine rotor 14 to rotate. The exhaust gas that has driven the turbine rotor 14 to rotate is discharged to the outside of the turbine housing 15 via the exhaust gas discharge passage 152.

[0027] (impeller) As shown in FIGS. 2 and 3 , the impeller 2 includes a hub 21 having a generally truncated cone shape and a plurality of impeller vanes 23 provided on an outer peripheral surface 22 of the hub 21. The plurality of impeller vanes 23 are spaced apart from one another in the circumferential direction around the axis LA. The hub 21 is fixed to one end of the rotary shaft 16, so that the hub 21 and the plurality of impeller vanes 23 are rotatable integrally with the rotary shaft 16 about the axis LA of the impeller 2. The impeller 2 is configured to guide air introduced along the axial direction of the impeller 2 to the outside in the radial direction of the impeller 2. As shown in FIG. 3 , a gap (clearance) is formed between the tip ends (tips) 24 of the plurality of impeller vanes 23 and a shroud surface 31 that is convexly curved so as to face the tip ends 24. In other words, the impeller 2 is an open-type impeller that does not include an annular member covering the tip ends 24.

[0028] (Casing) As shown in Fig. 3, the casing 3 includes a casing main body 4. The casing main body 4 has at least one scroll passage 41 provided on the outer periphery of the impeller 2. Each of the scroll passages 41 is a spiral passage extending along the circumferential direction of the impeller 2. The casing main body 4 further includes the shroud surface 31, the gas introduction passage 32, and the diffuser passage 33 described above. That is, each of the scroll passages 41, the shroud surface 31, the gas introduction passage 32, and the diffuser passage 33 are formed inside the casing main body 4.

[0029] The gas introduction passage 32 is a passage for taking in air from outside the casing main body 4 (casing 3) and guiding the taken-in air (gas) to the impeller 2. The gas introduction passage 32 is provided on one side (forward side) of the impeller 2 in the axial direction of the impeller 2, and extends along the axial direction of the impeller 2. By driving the impeller 2 to rotate, air is taken in from outside the casing main body 4 (casing 3) into the gas introduction passage 32, and the taken-in air flows through the gas introduction passage 32 toward the other side (rear side) of the impeller 2 in the axial direction, and is guided to the impeller 2.

[0030] The diffuser passage 33 is a passage for guiding air that has passed through the impeller 2 and been compressed by the impeller 2 to each of the scroll passages 41. The diffuser passage 33 is provided between each of the scroll passages 41 and the impeller 2 in the radial direction of the impeller 2, and extends along the radial direction of the impeller 2. The diffuser passage 33 communicates with each of the scroll passages 41 at an outlet 34 provided at the downstream end (outer peripheral end) of the diffuser passage 33. In other words, the outlet 34 of the diffuser passage 33 is a communication port with each of the scroll passages 41. The compressed air (compressed gas) compressed by the impeller 2 flows into the diffuser passage 33, flows through the diffuser passage 33 toward the outside in the radial direction of the impeller 2, and is guided to each of the scroll passages 41.

[0031] (Centrifugal compressor according to comparative example) Fig. 4 is a schematic contour diagram of the static pressure distribution in the scroll passage 41 at a large flow rate side operating point of the centrifugal compressor 1A according to the comparative example. Fig. 5 is a schematic contour diagram of the static pressure distribution in the scroll passage 41 at a small flow rate side operating point of the centrifugal compressor 1A according to the comparative example. As shown in Figs. 4 and 5, the centrifugal compressor 1A according to the comparative example includes the above-described impeller 2 and the above-described casing 3.

[0032] The casing 3 of the centrifugal compressor 1A includes a casing body 4 having one scroll passage 41 with a tongue portion 42 provided at the spiral start position (θ=0°), and a discharge pipe portion 5 that communicates with the spiral end position (θ=360°) of the scroll passage 41 via a branch port 52 and forms a discharge passage 51 through which compressed air guided from the scroll passage 41 flows. Here, θ is a circumferential angle position in the circumferential direction of the impeller 2, with the spiral start position of the scroll passage 41 being 0° and the downstream side of the scroll passage 41 being positive.

[0033] In the centrifugal compressor 1A, no discharge port for discharging compressed air to the outside of the scroll flow path 41 is formed between the start position and the end position of the scroll flow path 41. The compressed air flowing through the scroll flow path 41 flows to the end position, and is then guided to the discharge flow path 51 via the branch port 52, and is then guided to the outside of the centrifugal compressor 1A from the discharge port 53 which is the outlet of the discharge flow path 51.

[0034] Fig. 6 is a diagram showing the relationship between the intake air flow rate FR and the pressure ratio PR in a centrifugal compressor 1A according to a comparative example. As shown in Fig. 6, in a high peripheral speed region (for example, region AH), a surge occurs due to stall of the impeller 2 of the centrifugal compressor 1A, so a high peripheral speed surge line SL1 is provided in response to this high peripheral speed surge. In a low peripheral speed region (for example, region AL), a surge occurs due to a backflow occurring in the stationary passage system (diffuser passage 33 and scroll passage 41) downstream of the impeller 2 of the centrifugal compressor 1A, so a low peripheral speed surge line SL2 is provided in response to this low peripheral speed surge.

[0035] FIG. 7 is an explanatory diagram for explaining the circumferential static pressure distribution in the scroll passage 41 of the centrifugal compressor 1A according to the comparative example. FIG. 7 shows a graph with the circumferential angular position θ as the horizontal axis and the static pressure (pressure) SP at the outlet 34 of the diffuser passage 33 as the vertical axis. The graph in FIG. 7 shows a curve C1 indicating the pressure change at the large flow rate side operating point (operating point PH in region AH, see FIG. 6 ) and a curve C2 indicating the pressure change at the small flow rate side operating point (operating point PL in region AL, see FIG. 6 ). As shown in FIG. 7 , at the beginning of the scroll passage 41 where the circumferential cross-sectional area A is relatively small, the pressure change is small regardless of the change in the operating point flow rate. In contrast, at the end of the scroll passage 41 where the circumferential cross-sectional area A is relatively large, a pressure change occurs corresponding to the change in the operating point flow rate (change in flow velocity), resulting in a large pressure change.

[0036] 4 and 7, at the high flow rate operating point, the static pressure SP decreases and the compressed air speed increases as the scroll passage 41 moves from the start position to the end position. Then, at the end position of the scroll passage 41, the static pressure SP increases and the compressed air speed decreases. Therefore, at the high flow rate operating point, there is a risk of pressure recovery loss occurring due to the increase or decrease in the speed of the compressed air in the scroll passage 41.

[0037] 5 and 7, at the low flow rate operating point, the static pressure SP increases and the compressed air decelerates as the scroll passage 41 moves from the beginning to the end of the scroll passage 41. The pressure difference at the tongue 42 of the scroll passage 41 increases the recirculation flow RF returning from the end of the scroll passage 41 to the beginning of the scroll passage 41, which may cause recirculation loss. This may also result in a decrease in surge resistance at the end of the scroll passage 41.

[0038] Fig. 8 is a schematic contour diagram of the static pressure distribution in the scroll passage 41 at a large flow rate side operating point of the centrifugal compressor 1 (1B) with a double scroll structure according to one embodiment. Fig. 9 is a schematic contour diagram of the static pressure distribution in the scroll passage 41 at a small flow rate side operating point of the centrifugal compressor 1 (1B) with a double scroll structure according to one embodiment. Fig. 10 is a schematic contour diagram of the static pressure distribution in the scroll passage 41 at a large flow rate side operating point of the twin-scroll centrifugal compressor 1 (1C) according to one embodiment. Fig. 11 is a schematic contour diagram of the static pressure distribution in the scroll passage 41 at a small flow rate side operating point of the twin-scroll centrifugal compressor 1 (1C) according to one embodiment.

[0039] (1st discharge pipe section, 2nd discharge pipe section) As shown in Figures 8 to 13, the casing 3 of the centrifugal compressor 1 according to some embodiments includes at least a casing main body 4 having at least one scroll passage 41 described above, a first discharge pipe section 5 forming a first discharge passage 51, and a second discharge pipe section 6 forming a second discharge passage 61.

[0040] The first discharge flow path 51 communicates with at least one scroll flow path 41 via a first branch port 52 provided at the upstream end (one end) of the first discharge pipe section 5. The first branch port 52 is a communication port with the scroll flow path 41. A first discharge port 53, which is an outlet of the first discharge flow path 51, is provided at the downstream end (the other end) of the first discharge pipe section 5. The compressed air guided from the scroll flow path 41 to the first discharge flow path 51 via the first branch port 52 is guided to the outside of the centrifugal compressor 1 via the first discharge port 53.

[0041] The second discharge flow path 61 communicates with at least one scroll flow path 41 via a second branch port 62 provided at the upstream end (one end) of the second discharge pipe section 6. The second branch port 62 is a communication port with the scroll flow path 41. A second discharge port 63, which is an outlet of the second discharge flow path 61, is provided at the downstream end (the other end) of the second discharge pipe section 6. The compressed air guided from the scroll flow path 41 to the second discharge flow path 61 via the second branch port 62 is guided to the outside of the centrifugal compressor 1 via the second discharge port 63.

[0042] 8 to 13, the first branch port 52 and the second branch port 62 are provided at different positions in the circumferential direction of the impeller 2. In this case, by discharging the compressed gas flowing through the scroll passage 41 from each of the multiple branch ports (such as the first branch port 52, the second branch port 62, and a third branch port 72 described later) provided at different positions in the circumferential direction of the impeller 2, it is possible to mitigate the influence of the circumferential static pressure distribution caused by the circumferential integration of the flow rate flowing from the diffuser passage 33 into the scroll passage 41. This makes it possible to mitigate the influence of distortion in the static pressure distribution in the scroll passage 41.

[0043] As shown in Figures 8 and 10, the centrifugal compressor 1 can suppress the pressure recovery loss caused by the acceleration / deceleration of the compressed air in the scroll passage 41 at a large flow rate operating point, compared to the centrifugal compressor 1A according to the comparative example. Also, as shown in Figures 9 and 11, the centrifugal compressor 1 can reduce the static pressure difference between the start and end of the scroll passage 41 at a small flow rate operating point, compared to the centrifugal compressor 1A according to the comparative example, and can suppress the recirculation flow RF.

[0044] By mitigating the effects of distortion in the static pressure distribution within the scroll flow path 41, the centrifugal compressor 1 can suppress the occurrence of losses in the centrifugal compressor 1 caused by distortion in the static pressure distribution, thereby achieving high efficiency and a wide range (expansion of the operating range) of the centrifugal compressor 1.

[0045] Furthermore, according to the above configuration, the flow path cross-sectional areas of the scroll flow path 41 and the discharge flow path 51 can be made smaller than those of the centrifugal compressor 1A according to the comparative example, so that the centrifugal compressor 1 can be made smaller and lighter, and the mountability on the internal combustion engine 13 can be improved.

[0046] 4 and 5, at least one of the first branch port 52 and the second branch port 62 (62A) may be provided midway through the scroll passage 41 (specifically, a portion excluding the start and end of the winding). In the embodiment shown in FIGS. 4 and 5, the second branch port 62 (62A) is provided midway through the scroll passage 41, and the first branch port 52 is provided at the end of the winding of the scroll passage 41.

[0047] 8 to 13, the at least one scroll passage 41 includes a first scroll passage 41A having a first tongue portion 42A at a start position of the spiral, and a second scroll passage 41B having a second tongue portion 42B at a start position of the spiral. The first discharge pipe section 5 has a first discharge passage 51 that communicates with the first scroll passage 41A via a first branch port 52, and the second discharge pipe section 6 has a second discharge passage 61 that communicates with the second scroll passage 41B via a second branch port 62. The first tongue portion 42A and the second tongue portion 42B are provided at different positions in the circumferential direction of the impeller 2.

[0048] According to the above configuration, the compressed gas flowing through the first scroll passage 41A can be discharged from the first branch port 52, and the compressed gas flowing through the second scroll passage 41B can be discharged from the second branch port 62. In this case, it is possible to suppress interference between the compressed gas flowing through the first scroll passage 41A and the first discharge passage 51 and the compressed gas flowing through the second scroll passage 41B and the second discharge passage 61. This makes it possible to suppress the occurrence of loss in the centrifugal compressor 1 caused by the above interference, thereby achieving high efficiency of the centrifugal compressor 1.

[0049] In some embodiments, as shown in Figures 8 to 13, the above-mentioned casing body 4 is formed so that the first scroll passage 41A and the second scroll passage 41B are divided in the axial or circumferential direction of the impeller 2.

[0050] 8, 9, and 12, the casing body 4 is formed so that the first scroll passage 41A and the second scroll passage 41B are divided in the circumferential direction of the impeller 2. The first scroll passage 41A and the second scroll passage 41B are formed in the same region in the axial direction of the impeller 2, and are formed at positions offset from each other in the circumferential direction of the impeller 2. Therefore, the first scroll passage 41A and the second scroll passage 41B communicate with the outlet 34 of the diffuser passage 33 at positions offset from each other in the circumferential direction of the impeller 2. Note that "formed in the same region in the axial direction of the impeller 2" means that the scroll shapes of the first scroll passage 41A and the second scroll passage 41B are displayed together in a cross section perpendicular to the axis LA of the impeller 2, as shown in FIGS.

[0051] 10, 11, and 13, the casing body 4 is formed so that the first scroll passage 41A and the second scroll passage 41B are separated in the axial direction of the impeller 2. The first scroll passage 41A and the second scroll passage 41B are formed at positions offset from each other in the axial direction of the impeller 2. Therefore, the first scroll passage 41A and the second scroll passage 41B communicate with the outlet 34 of the diffuser passage 33 at positions offset from each other in the axial direction of the impeller 2.

[0052] Fig. 14 is a schematic cross-sectional view along the axis LA of a centrifugal compressor 1 (1C) with a twin-scroll structure according to one embodiment. As shown in Fig. 14, the above-mentioned casing body 4 may include a partition wall 35 that protrudes inward in the radial direction of the impeller 2 from a wall surface that forms the scroll passage 41 and divides the scroll passage 41 into two in the axial direction of the impeller 2. The partition wall 35 divides the scroll passage 41 into a first scroll passage 41A and a second scroll passage 41B.

[0053] According to the above configuration, by dividing the first scroll passage 41A and the second scroll passage 41B in the axial or circumferential direction of the impeller 2, it is possible to suppress abrupt turning of the compressed gas flow in each of the scroll passages 41A, 41B and separation of the compressed gas from the wall surfaces of the scroll passages 41A, 41B due to the turning. This makes it possible to suppress the occurrence of losses in the centrifugal compressor 1 caused by the turning and separation, thereby improving the efficiency of the centrifugal compressor 1.

[0054] When the first scroll passage 41A and the second scroll passage 41B are divided in the circumferential direction of the impeller 2, the influence of the circumferential static pressure distribution can be alleviated. When the first scroll passage 41A and the second scroll passage 41B are divided in the axial direction of the impeller 2, the static pressure distribution generated in the second scroll passage 41B can be shifted in the circumferential direction relative to the static pressure distribution generated in the first scroll passage 41A, thereby making the static pressure at the outlet 34 of the diffuser passage 33 uniform in the circumferential direction. In these cases, the occurrence of losses in the centrifugal compressor 1 caused by the circumferential static pressure distribution can be suppressed, and the centrifugal compressor 1 can be made more efficient and have a wider operating range.

[0055] (Triple scroll structure, trio scroll structure) Fig. 12 is a schematic view perpendicular to the axis LA of the casing 3 of a centrifugal compressor 1 (1D) having a triple scroll structure according to one embodiment. Fig. 13 is a schematic view perpendicular to the axis LA of the casing 3 of a centrifugal compressor 1 (1E) having a trio scroll structure according to one embodiment. In some embodiments, as shown in Figs. 12 and 13, the scroll passage 41 further includes a third scroll passage 41C in which a third tongue portion 42C is provided at a winding start position, and the casing 3 further includes a third discharge pipe portion 7 that forms a third discharge passage 71.

[0056] The third discharge passage 71 communicates with the third scroll passage 41C via a third branch port 72 provided at the upstream end (one end) of the third discharge pipe section 7. The third branch port 72 is a communication port with the third scroll passage 41C. A third discharge port 73, which is an outlet of the third discharge passage 71, is provided at the downstream end (the other end) of the third discharge pipe section 7. The compressed air guided from the third scroll passage 41C to the third discharge passage 71 via the third branch port 72 is guided to the outside of the centrifugal compressor 1 via the third discharge port 73. The third discharge port 73 is provided at a position different from the first branch port 52 and the second branch port 62 in the circumferential direction of the impeller 2.

[0057] 12, the casing body 4 is formed so that the first scroll passage 41A, the second scroll passage 41B, and the third scroll passage 41C are divided in the circumferential direction of the impeller 2. The third scroll passage 41C is formed in the same region in the axial direction of the impeller 2 for the first scroll passage 41A and the second scroll passage 41B, and is formed at positions shifted from each other in the circumferential direction of the impeller 2. Therefore, the third scroll passage 41C communicates with the outlet 34 of the diffuser passage 33 at positions shifted from each other in the circumferential direction of the impeller 2 for the first scroll passage 41A and the second scroll passage 41B.

[0058] 13, the casing body 4 is formed so that the first scroll passage 41A, the second scroll passage 41B, and the third scroll passage 41C are divided in the axial direction of the impeller 2. The third scroll passage 41C is formed at positions where the first scroll passage 41A and the second scroll passage 41B are shifted from each other in the axial direction of the impeller 2. Therefore, the third scroll passage 41C communicates with the outlet 34 of the diffuser passage 33 at positions where the first scroll passage 41A and the second scroll passage 41B are shifted from each other in the axial direction of the impeller 2.

[0059] According to the above configuration, by dividing and arranging the three scroll passages 41 (41A, 41B, 41C) in the axial or circumferential direction of the impeller 2, it is possible to suppress abrupt turning of the compressed gas flow in each of the scroll passages 41A, 41B, 41C and separation of the compressed gas from the wall surfaces of the scroll passages 41A, 41B, 41C due to the turning. This makes it possible to suppress the occurrence of losses in the centrifugal compressor 1 caused by the turning and separation, thereby improving the efficiency of the centrifugal compressor 1.

[0060] In some other embodiments, the casing 3 described above may have four or more branch ports (such as the first branch port 52). However, the more branch ports there are, the greater the impact of disadvantages such as increased friction loss due to an increased leakage area, increased recirculation loss due to an increased number of tongues, and increased weight and manufacturing costs of the casing 3. Therefore, it is preferable to provide two or three branch ports in the casing 3.

[0061] (Circumferential arrangement of multiple tongues) 8 to 13, the casing body 4 has a plurality of tongue portions 42 provided at different positions in the circumferential direction of the impeller 2. When the number of the plurality of tongue portions 42 is n and the angle obtained by dividing 360° by n is θ1, each of the plurality of tongue portions 42 is configured such that the angle formed with an adjacent tongue portion 42 satisfies the condition θ1±10°.

[0062] In the embodiment shown in Figures 8 to 11, the casing body 4 has two tongue portions 42A, 42B provided at different positions in the circumferential direction of the impeller 2. The angle formed by the two tongue portions 42A, 42B satisfies the condition of 180° (θ1) ± 10°. In the embodiment shown in Figures 12 and 13, the casing body 4 has three tongue portions 42A, 42B, 42C provided at different positions in the circumferential direction of the impeller 2. The angle formed by each of the three tongue portions 42A, 42B, 42C with the adjacent tongue portion 42 satisfies the condition of 120° (θ1) ± 10°.

[0063] According to the above configuration, by arranging each of the plurality of tongue portions 42 at approximately equal intervals in the circumferential direction so as to satisfy the above condition, the circumferential static pressure distribution can be made more uniform in the circumferential direction. This makes it possible to effectively suppress the occurrence of losses in the centrifugal compressor 1 caused by the circumferential static pressure distribution. Note that in some other embodiments of the present disclosure, each of the plurality of tongue portions 42 does not have to satisfy the above condition.

[0064] (Bend in the discharge pipe) 15 and 16 are schematic views along the axis LA of the casing 3 of the centrifugal compressor 1 according to one embodiment. In some embodiments, as shown in Fig. 15 and 16, at least one of the first discharge pipe section 5 or the second discharge pipe section 6 described above has a bending section 54, 64 that bends the discharge flow passage 51, 61 through which the compressed gas guided from the scroll flow passage 41 flows. In the discharge pipe sections 5, 6 having the bending sections 54, 64, axes L53, L63 at discharge ports 53, 63 that are outlets of the discharge flow passages 51, 61 extend along a direction intersecting with a plane OP perpendicular to the axis LA of the impeller 2.

[0065] If neither the first discharge pipe section 5 nor the second discharge pipe section 6 had the bent sections 54, 64, when the first branch port 52 and the second branch port 62 were provided at different positions in the circumferential direction of the impeller 2, the discharge port 53 of the first discharge pipe section 5 and the discharge port 63 of the second discharge pipe section 6 would face in different directions, which could result in a complicated piping layout for the first discharge pipe section 5 or the second discharge pipe section 6. According to the above configuration, by configuring the first discharge pipe section 5 or the second discharge pipe section 6 to have the bent sections 54, 64 that bend the discharge flow paths 52, 62, it is possible to prevent the piping layout from becoming complicated.

[0066] 15 and 16 , the first discharge pipe section 5 has a first bent portion 54 that bends the first discharge flow path 51, and an axis L53 at a first discharge port 53 that is an outlet of the first discharge flow path 51 extends along a direction intersecting with a plane OP that is perpendicular to the axis LA of the impeller 2. The second discharge pipe section 6 has a second bent portion 64 that bends the second discharge flow path 61, and an axis L63 at a second discharge port 63 that is an outlet of the second discharge flow path 61 extends along a direction intersecting with a plane OP that is perpendicular to the axis LA of the impeller 2. As shown in FIG. 16 , the axis L63 at the second discharge port 63 extends in a direction that is not parallel to the extension direction of the axis L53 at the first discharge port 53.

[0067] According to the above configuration, the axis L53 of the first discharge port 53 of the first discharge pipe section 5 having the first bent portion 54 and the axis L63 of the second discharge port 63 of the second discharge pipe section 6 having the second bent portion 64 extend in directions that are not parallel to each other, thereby increasing the degree of freedom in the piping layout of the first discharge pipe section 5 and the second discharge pipe section 6. In some other embodiments of the present disclosure, the axis L63 may extend along a direction parallel to the extension direction of the axis L53.

[0068] The above-described casing 3 has the first bent portion 54 and the second bent portion 64, so that both the first discharge port 53 and the second discharge port 54 may be located forward of the scroll passage 41 in the axial direction of the impeller 2 (see FIG. 15 ), or may be located rearward. In these cases, by arranging the first discharge port 53 and the second discharge port 54 on the same side of the scroll passage 41, it is possible to prevent the piping layout from becoming complicated.

[0069] Furthermore, by having the first bent portion 54 and the second bent portion 64, the casing 3 described above may have one of the first discharge port 53 and the second discharge port 54 located forward of the scroll passage 41 in the axial direction of the impeller 2, and the other located rearward of the scroll passage 41 (see FIG. 16). By arranging the second discharge port 54 on the opposite side of the scroll passage 41 from the side on which the first discharge port 53 is located, the degree of freedom in the piping layout can be increased.

[0070] (Outlet structure of discharge pipe) 17 and 18 are schematic cross-sectional views of a connection portion 36 of a casing 3 of a centrifugal compressor 1 according to one embodiment and a pipe 19 connected to the connection portion 36. In some embodiments, as shown in Fig. 17 and 18, the first discharge pipe portion 5 and the second discharge pipe portion 6 described above have a connection portion 36 connected to the same pipe 19. The connection portion 36 is provided at the downstream end of the first discharge pipe portion 5 and the downstream end of the second discharge pipe portion 6.

[0071] The piping 19 has a conduit 191 that communicates with the first discharge flow path 51 and the second discharge flow path 61, and through which the compressed gas that has flowed through the first discharge flow path 51 and the compressed gas that has flowed through the second discharge flow path 61 can flow. The piping 19 constitutes a part of a compressed gas supply system that guides the compressed gas to the internal combustion engine 13. The piping 19 may be an intake manifold of the internal combustion engine 13.

[0072] The piping 19 may have a piping-side flange portion 193 that protrudes radially outward from the piping 19 at a connection portion 192 that is connected to the connection portion 36 of the casing 3, and the connection portion 36 may have a flange portion 361 that protrudes radially outward from the connection portion 36 and abuts against the piping-side flange portion 193. The connection portion 36 of the casing 3 and the connection portion 192 of the piping 19 are connected by fastening members 37 (for example, bolts, nuts) that fasten the piping-side flange portion 193 and the flange portion 361 together.

[0073] 17 , the connecting portion 36 has a junction flow path 362 formed therein. The junction flow path 362 communicates with the first discharge flow path 51 via the first discharge port 52 and with the second discharge flow path 61 via the second discharge port 62. Therefore, the junction flow path 362 is configured to merge the compressed gas discharged from the first discharge flow path 51 via the first discharge port 52 and the compressed gas discharged from the second discharge flow path 61 via the second discharge port 62. By connecting the connecting portion 36 of the casing 3 to the connecting portion 192 of the piping 19, the junction flow path 362 and the conduit 191 are communicated with each other, and the compressed gas flows from the junction flow path 362 to the conduit 191.

[0074] As shown in FIGS. 17 and 18, the first discharge pipe section 5 is disposed adjacent to the second discharge pipe section 6 over a predetermined section S1 on the upstream side from the connecting section .

[0075] According to the above configuration, the discharge port 53 of the first discharge pipe section 5 and the discharge port 63 of the second discharge pipe section 6 can be connected to the same pipe 19, making it possible to have a piping layout similar to that of a conventional casing having a single discharge port 53. Furthermore, by arranging the first discharge pipe section 5 and the second discharge pipe section 6 adjacent to each other over a predetermined section S1 upstream from the connecting section, it is possible to prevent the piping layout from becoming too complicated.

[0076] In some embodiments, as shown in FIG. 18, the first discharge pipe section 5 and the second discharge pipe section 6 described above are configured so that the first discharge port 53, which is the outlet of the first discharge flow path 51, and the second discharge port 63, which is the outlet of the second discharge flow path 61, are located on the same plane PS.

[0077] 18, the downstream end of the first discharge flow path 51 and the downstream end of the second discharge flow path 61 are formed inside the connecting portion 36, and the first discharge port 53 and the second discharge port 63 are formed on a flat abutment surface 363 of the connecting portion 36 that abuts on the connecting portion 192 of the pipe 19. In other words, the abutment surface 363 exists on the plane PS. As shown in FIG. 18, the conduit 191 may be configured so that the compressed gas discharged from the first discharge flow path 51 via the first discharge port 52 and the compressed gas discharged from the second discharge flow path 61 via the second discharge port 62 join together.

[0078] According to the above configuration, the first discharge outlet 53 and the second discharge outlet 63 are located on the same plane PS, which makes it easy to connect the first discharge outlet 53 and the second discharge outlet 63 to the same piping 19, thereby preventing the piping layout from becoming too complicated.

[0079] (Shape of the scroll passage) 19 is a schematic cross-sectional view perpendicular to the axis LA of the casing 3 of a centrifugal compressor 1 according to one embodiment. As shown in Fig. 19, the cross-sectional area of ​​the flow passage cross section CS of at least one scroll flow passage 41 at a certain circumferential angular position is defined as A(θ), and the radius (shortest distance) from the axis LA of the impeller 2 to the center (e.g., centroid) CP of the flow passage cross section CS is defined as R(θ).

[0080] In the circumferential direction of the impeller 2, the starting position of the winding of the first scroll passage 41A is defined as 0° and the first circumferential angular position, with the downstream side of the first scroll passage 41A being positive, is defined as α, the flow path cross-sectional area of ​​the first scroll passage 41A at the first circumferential angular position α is defined as A1(α), and the radius (shortest distance) from the axis LA of the impeller 2 to the center CP of the flow path cross section CS of the first scroll passage 41A is defined as R1(α).

[0081] In the circumferential direction of the impeller 2, the starting position of the winding of the second scroll passage 41B is defined as 0° and the second circumferential angular position, with the downstream side of the second scroll passage 41B being positive, is defined as β.The flow path cross-sectional area of ​​the second scroll passage 41B at the second circumferential angular position β is defined as A2(β), and the radius (shortest distance) from the axis LA of the impeller 2 to the center CP of the flow path cross section CS of the second scroll passage 41B is defined as R2(β).

[0082] In some embodiments, the casing body 4 (casing 3) described above is configured to satisfy either the following formula (1) or (2) at least when α=β. A1(α)≠A2(β) (1) A1(α) / R1(α)≠A2(β) / R2(β) (2)

[0083] 19, the casing main body 4 (casing 3) is configured to satisfy either of the above formulas (1) or (2) at at least one point in an intermediate range (0°<α<180°) between the start of winding (α=θ=0°) and the end of winding (α=θ=180°) of the first scroll passage 41A. For example, the casing main body 4 (casing 3) is configured to satisfy either of the above formulas (1) or (2) at an intermediate position P1 (α=θ=90°, an intermediate position between the first and second tongues in the circumferential direction) between the start of winding (α=θ=0°) and the end of winding (α=θ=180°) of the first scroll passage 41A, and at an intermediate position P2 (β=90° (θ=270°), an intermediate position between the second and first tongues in the circumferential direction) of the second scroll passage 41B corresponding to the intermediate position P1.

[0084] In one embodiment, the casing main body 4 (casing 3) may be configured to satisfy either of the above formulas (1) or (2) when α = β throughout the entire intermediate range between the beginning of the first scroll flow path 41A (α = θ = 0°) and the end of the first scroll flow path 41A (α = θ = 180°).

[0085] According to the above configuration, the circumferential cross-sectional area distribution and central diameter distribution of the scroll passage 41 significantly affect the design point of the scroll passage 41, i.e., the operating point at which maximum efficiency is achieved. Therefore, by making the circumferential cross-sectional area distribution and central diameter distribution of the two scroll passages 41A and 41B different, robust efficiency characteristics can be obtained over a wide operating range.

[0086] The circumferential cross-sectional area distribution and the central diameter distribution of the two scroll flow paths 41A, 41B will be specifically described with reference to Figs. 20 to 24. Fig. 20 is a diagram showing the relationship between the circumferential angular position and the A / R ratio in a centrifugal compressor 1A (see Fig. 4) according to a comparative example. Figs. 21 and 22 are diagrams showing the relationship between the circumferential angular position and the A / R ratio in a centrifugal compressor 1B (see Fig. 8) having a double-scroll structure according to one embodiment. Figs. 23 and 24 are diagrams showing the relationship between the circumferential angular position and the A / R ratio in a centrifugal compressor 1C (see Fig. 10) having a twin-scroll structure according to one embodiment. Figs. 20 to 24 show graphs with the above-mentioned circumferential angular positions θ, α, and β on the horizontal axis and the A / R ratio, which is the ratio of the circumferential cross-sectional area A to the radius R, on the vertical axis. As shown in Figures 20 to 24, the A / R ratio gradually increases from the beginning of the scroll flow path 41 (41A, 41B) to the end of the winding, with the A / R ratio at the beginning of the winding being the minimum value and the A / R ratio at the end of the winding being the maximum value.

[0087] As shown in FIG. 20, in the centrifugal compressor 1A according to the comparative example, the A / R ratio is at a minimum value MIN1 at the beginning of the scroll flow path 41 (θ=0°) and at a maximum value MAX1 at the end of the scroll flow path 41 (θ=360°).

[0088] In FIGS. 21 and 23, the first scroll passage 41A and the second scroll passage 41B have the same shape, while in FIGS. 22 and 24, the first scroll passage 41A and the second scroll passage 41B have different shapes.

[0089] The centrifugal compressor 1B shown in Fig. 21 has a minimum A / R ratio MIN2 at the start of winding (α = 0°, β = 0°) of the first scroll flow path 41A and the second scroll flow path 41B, and a maximum A / R ratio MAX2 at the end of winding (α = 180°, β = 180°) of the first scroll flow path 41A and the second scroll flow path 41B. Here, the maximum value MAX2 is approximately half that of the above-mentioned maximum value MAX1.

[0090] The centrifugal compressor 1C shown in Fig. 23 has a minimum A / R ratio MIN4 at the start of winding (α = 0°, β = 0°) of the first scroll flow path 41A and the second scroll flow path 41B, and a maximum A / R ratio MAX4 at the end of winding (α = 360°, β = 360°) of the first scroll flow path 41A and the second scroll flow path 41B. Here, the maximum value MAX4 is approximately half that of the above-mentioned maximum value MAX1.

[0091] The centrifugal compressor 1B shown in Fig. 22 and the centrifugal compressor 1C shown in Fig. 24 have a shape (A / R ratio) of the first scroll flow path 41A suitable for low-flow operation conditions and a shape (A / R ratio) of the second scroll flow path 41B suitable for high-flow operation conditions.

[0092] For the centrifugal compressor 1B shown in Fig. 22, the A / R ratio (minimum value MIN3) at the start of winding (β = 0°) of the second scroll flow path 41B is larger than the A / R ratio (minimum value MIN2) at the start of winding (α = 0°) of the first scroll flow path 41A. Also, the A / R ratio (maximum value MAX3) at the end of winding (β = 180°) of the second scroll flow path 41B is larger than the A / R ratio (maximum value MAX2) at the end of winding (α = 180°) of the first scroll flow path 41A. That is, in the entire range of the intermediate range between the start of winding (α = 0°) and the end of winding (α = 180°) of the first scroll flow path 41A, when α = β, A1(α) / R1(α) < A2(β) / R2(β) is satisfied.

[0093] The centrifugal compressor 1C shown in Fig. 24 has an A / R ratio (minimum value MIN5) at the start of winding (β = 0°) of the second scroll flow path 41B that is larger than the A / R ratio (minimum value MIN4) at the start of winding (α = 0°) of the first scroll flow path 41A. Also, the A / R ratio (maximum value MAX5) at the end of winding (β = 360°) of the second scroll flow path 41B is larger than the A / R ratio (maximum value MAX4) at the end of winding (α = 360°) of the first scroll flow path 41A. That is, in the entire range of the intermediate range between the start of winding (α = 0°) and the end of winding (α = 360°) of the first scroll flow path 41A, when α = β, A1(α) / R1(α) < A2(β) / R2(β) is satisfied.

[0094] In some embodiments, as shown in Figs. 22 and 24, the above-described casing main body 4 (casing 3) is configured to satisfy at least one of the following formulas (3) or (4) when α = β. A1(α) < A2(β) ··· (3) A1(α) / R1(α) < A2(β) / R2(β) ··· (4)

[0095] The casing main body 4 (casing 3) is configured to satisfy at least one of the above formulas (3) or (4) at at least one point in the intermediate range (0° < α < 180°) between the start of winding (α = θ = 0°) and the end of winding (α = θ = 180°) of the first scroll flow path 41A, as shown in Fig. 19 for example. For example, at the intermediate position P1 (α = θ = 90°, the intermediate position between the first tongue portion and the second tongue portion in the above circumferential direction) of the start of winding (α = θ = 0°) and the end of winding (α = θ = 180°) of the first scroll flow path 41A, and at the intermediate position P2 (β = 90° (θ = 270°), the intermediate position between the second tongue portion and the first tongue portion in the above circumferential direction) of the second scroll flow path 41B corresponding to this intermediate position P1, it is configured to satisfy at least one of the above formulas (3) or (4).

[0096] [[ID=**15**]] In one embodiment, the casing main body 4 (casing 3) may be configured to satisfy either of the above formulas (3) or (4) when α = β throughout the entire intermediate range between the beginning of the first scroll flow path 41A (α = θ = 0°) and the end of the first scroll flow path 41A (α = θ = 180°).

[0097] According to the above configuration, the second scroll passage 41B is given a circumferential cross-sectional area distribution and a central diameter distribution that maximize efficiency on the small flow rate side, and the first scroll passage 41A is given a circumferential cross-sectional area distribution and a central diameter distribution that maximize efficiency on the large flow rate side, thereby improving efficiency on both the small flow rate side and the large flow rate side.

[0098] (Outlet side flow rate adjustment device) Fig. 25 is a schematic view perpendicular to the axis LA of a centrifugal compressor 1 including an outlet-side flow control device 9 according to one embodiment. Fig. 26 is a schematic view along the axis LA of a centrifugal compressor 1 including an outlet-side flow control device 9 according to one embodiment.

[0099] In some embodiments, the casing 3 described above comprises the casing main body 4 described above, the first discharge pipe section 5 described above, the second discharge pipe section 6 described above, and an outlet-side flow control device 9 configured to be able to adjust the flow rate of at least one of the compressed gas flowing through the first discharge flow path 51 or the compressed gas flowing through the second discharge flow path 61, as shown in Figures 25 and 26.

[0100] In the illustrated embodiment, the outlet-side flow control device 9 may be a flow control valve including a valve element 91 capable of opening and closing the first discharge flow path 51 or the second discharge flow path 61, such as a butterfly valve, a swing valve, a ball valve, or a nozzle-shaped throttle mechanism. This flow control valve is preferably a swing valve or a ball valve, which reduces pressure loss when the valve element 91 is closed when a large amount of compressed gas flows through the flow path that the valve element 91 opens and closes. The valve element 91 may open and close the first discharge port 53 or the second discharge port 63. The outlet-side flow control device 9 may further include a valve actuator 92 configured to drive the valve element 91 to open and close. The outlet-side flow control device 9 may be an on-off valve whose opening degree can be adjusted between fully closed and fully open, or an opening degree adjustment valve whose opening degree can be adjusted between fully closed and fully open and at least one intermediate opening degree therebetween. In addition, the outlet-side flow control device 9 may be one that partially restricts the flow path area of ​​the first discharge flow path 51 or the second discharge flow path 61 (one that only partially blocks the flow path area when half-opened or fully closed).

[0101] FIG. 27 is a diagram showing the relationship between the flow rate (FR) of intake air and the pressure ratio (PR) in a centrifugal compressor 1 including an outlet-side flow control device 9 according to an embodiment. FIG. 28 is a diagram showing the relationship between the flow rate (FR) of intake air and the pressure ratio (PR) in a centrifugal compressor 1 including an inlet-side flow control device 8 (described later) according to an embodiment. FIG. 27 shows a region EA1 where the efficiency of the centrifugal compressor 1 including the outlet-side flow control device 9 is improved and a region RE1 where the operating range is expanded compared to the centrifugal compressor 1A according to the comparative example. FIG. 28 shows a region EA2 where the efficiency of the centrifugal compressor 1 including the inlet-side flow control device 8 is improved and a region RE2 where the operating range is expanded compared to the centrifugal compressor 1A according to the comparative example.

[0102] 27 and 28, the centrifugal compressor 1 equipped with the outlet-side flow control device 9 has a wider operating range from the low peripheral speed side to the high peripheral speed side, and the region EA1 in which efficiency is improved is wider, compared to the centrifugal compressor 1 equipped with the inlet-side flow control device 8. In this way, the centrifugal compressor 1 equipped with the outlet-side flow control device 9 can achieve a wider operating range and improved compression efficiency, compared to the centrifugal compressor 1 equipped with the inlet-side flow control device 8. Furthermore, the outlet-side flow control device 9 has a simpler structure than the inlet-side flow control device 8, and is easier to manufacture and install. Furthermore, the outlet-side flow control device 9 is less likely to adversely affect the performance of the impeller 2, compared to the inlet-side flow control device 8.

[0103] In some embodiments, the casing 3 described above comprises the casing main body 4 described above, the first discharge pipe section 5 described above, and the second discharge pipe section 6 described above, as shown in Figures 25 and 26, and is configured to satisfy either of the following equations (3) or (4) at least when α = β. A1(α) <A2(β)···(3) A1(α) / R1(α) <A2(β) / R2(β)···(4) The casing 3 further includes an outlet-side flow rate adjustment device (flow rate adjustment valve) 9 configured to be able to adjust the flow rate of the compressed gas flowing through the first discharge flow path 51.

[0104] In this embodiment, the first scroll passage 41A has a shape (A or A / R ratio) suitable for low-flow rate operating conditions, and the second scroll passage 41B has a shape (A or A / R ratio) suitable for high-flow rate operating conditions. The outlet-side flow control device (flow control valve) 9 includes a valve body 91 that can open and close the first discharge passage 51. This valve body 91 may open and close the first discharge port 53.

[0105] The outlet-side flow control device (flow control valve) 9 may be configured so that when operating in a small flow rate region, the valve body 91 closes the first discharge port 53 (first discharge flow path 51), and when operating in a large flow rate region, the valve body 91 opens to open the first discharge port 53 (first discharge flow path 51). When operating in the small flow rate region, the compressed gas is discharged to the outside of the casing 3 through the second discharge port 63, which is always open, out of the first discharge port 53 and the second discharge port 63. When operating in the large flow rate region, the compressed gas is discharged to the outside of the casing 3 through the first discharge port 53 and the second discharge port 63.

[0106] According to the above configuration, under high flow rate operating conditions, the outlet-side flow control device 9 adjusts the flow rate of the compressed gas flowing through the first discharge passage 51, thereby making it possible to maintain appropriate flow rates of the compressed gas flowing through the first scroll passage 41A and the second scroll passage 41B. The outlet-side flow control device 9 is preferably provided in a scroll passage 41 having a small cross-sectional area A or a small ratio A / R among the plurality of scroll passages 41. When the plurality of scroll passages 41 each have a tongue portion 42 at the beginning of their spiral, the law of conservation of angular momentum dictates that the larger the radius R, the smaller the circumferential flow velocity. Therefore, the larger the radius R for the same cross-sectional area A, the smaller the flow rate of the compressed gas passing through the scroll passage 41. Therefore, it is more preferable to provide the outlet-side flow control device 9 in a scroll passage 41 having a small ratio A / R among the plurality of scroll passages 41. In other embodiments, instead of the cross-sectional area A or the ratio A / R of the scroll flow path 41, the outlet-side flow control device 9 may be provided at the one of the first discharge port 53 and the second discharge port 63 that has the smaller cross-sectional area (cross-sectional area in a cross section perpendicular to the direction in which the axis of the discharge port extends (the direction in which the discharge pipeline at the discharge port extends)) (in the illustrated example, the first discharge port 53).

[0107] In some embodiments, the casing 3 described above, as shown in Figures 25 and 26, comprises the casing main body 4 described above, the first discharge pipe section 5 described above, the second discharge pipe section 6 described above, and an outlet-side flow control device (flow control valve) 9 configured to adjust the flow rate of compressed gas flowing through the first discharge flow path 51.

[0108] Here, in the case of a centrifugal compressor 1 having a double scroll structure as shown in Fig. 25, when the first discharge port 53 (first discharge flow path 51) is closed by the valve body 91, there is a risk of loss occurring due to a sudden change in area occurring near the second tongue portion 42B close to the first discharge port 53. For this reason, it is preferable that the flow path area of ​​the first discharge flow path 51 is smaller than the flow path area of ​​the second discharge flow path 61. Furthermore, it is preferable that the area SA1 of the first discharge port 53 is smaller than the area SA2 of the second discharge port 63. This makes it possible to reduce loss due to the above-mentioned sudden change in area.

[0109] By providing the outlet-side flow control device 9 in the first discharge passage 51 communicating with the first scroll passage 41A, in which the flow passage cross-sectional area A of the scroll passage 41 or the ratio A / R of the flow passage cross-sectional area to the radius is relatively small, it is possible to reduce losses caused by a sudden change in area near the tongue 42B near the discharge passage in which the outlet-side flow control device 9 is provided when the flow rate of the compressed gas is reduced by the outlet-side flow control device 9, compared to when the outlet-side flow control device 9 is provided in the second discharge passage 61 communicating with the second scroll passage 41B, in which the flow passage cross-sectional area A of the scroll passage 41 or the ratio A / R of the flow passage cross-sectional area to the radius is relatively large.

[0110] 25 and 26, the casing 3 includes the casing main body 4, the first discharge pipe 5, the second discharge pipe 6, and an outlet-side flow rate control device (flow rate control valve) 9 configured to be able to adjust the flow rate of compressed gas flowing through the first discharge flow path 51. When the area of ​​the first discharge port 53 is defined as SA1 and the area of ​​the second discharge port 63 is defined as SA2, the casing 3 is configured to satisfy the following formula (5): 55%≦SA2 / (SA1+SA2)≦65% (5)

[0111] If the area SA1 of the first discharge port 53 and the area SA2 of the second discharge port 63 were approximately the same, the outlet for the compressed gas (the total area of ​​the open discharge ports) would suddenly become smaller when the outlet-side flow control device 9 reduces the flow rate of the compressed gas (when the first discharge flow path 51 is closed), which could cause a sudden change in performance of the centrifugal compressor 1. According to the above configuration, the outlet for the compressed gas (the total area of ​​the open discharge ports) can be prevented from suddenly becoming smaller when the outlet-side flow control device 9 reduces the flow rate of the compressed gas, which makes it possible to smoothly change the characteristics of the centrifugal compressor 1 at this time.

[0112] (Inlet side flow rate adjustment device) 29 to 31 are schematic cross-sectional views along the axis LA of a centrifugal compressor 1 including an inlet-side flow control device 8 according to one embodiment. In some embodiments, as shown in FIGS. 25 and 26, the casing 3 includes the casing main body 4, the first discharge pipe 5, the second discharge pipe 6, and an outlet-side flow control device (flow control valve) 9 configured to adjust the flow rate of at least one of the compressed gas flowing through the first discharge flow path 51 and the compressed gas flowing through the second discharge flow path 61. The casing 3 further includes an inlet-side flow control device 8 configured to adjust the flow rate of the gas introduced into the impeller 2 via the gas introduction flow path 32, as shown in FIGS. 29 to 31.

[0113] 29, the inlet-side flow control device 8 (8A) is provided in the above-described gas introduction passage 32 and includes a throttle mechanism 81 that is capable of reducing the flow path area of ​​the gas introduction passage 32 by covering the outer periphery of the impeller blades 23. In the illustrated embodiment, the casing main body 4 has a recess 82 formed in the wall surface that forms the gas introduction passage 32, the recess 82 being capable of accommodating the throttle mechanism 81. By reducing the flow path area of ​​the gas introduction passage 32 with the throttle mechanism 81, it is possible to adjust the flow rate of the gas introduced into the impeller 2 via the gas introduction passage 32.

[0114] 30, the casing main body 4 has a recirculation flow path 83 for returning the gas introduced to the impeller 2 from downstream to upstream of the leading edge of the impeller blades 23, the recirculation flow path 83 being formed on the outer periphery of the leading edge of the impeller blades 23. The inlet-side flow rate control device 8 (8B) includes a cylindrical body 84 that separates the gas introduction flow path 32 from the recirculation flow path 83 provided on the outer periphery of the gas introduction flow path 32, the cylindrical body 84 being configured to be movable along the axial direction of the impeller 2, and an actuator 85 that moves the cylindrical body 84 along the axial direction of the impeller 2. The actuator 85 moves the cylindrical body 84 along the axial direction of the impeller 2, increasing or decreasing the opening areas of the inlet and outlet of the recirculation flow path 83, thereby adjusting the recirculation flow rate introduced to the impeller 2 through the recirculation flow path 83.

[0115] In the embodiment shown in Figure 31, the inlet side flow control device 8 (8C) is provided in the above-mentioned gas introduction flow path 32 and is a plurality of inlet guide vanes 86 rotatably supported on the casing 3. By rotating the inlet guide vanes 86 around the axis LB and adjusting their opening, the flow rate of the gas introduced into the impeller 2 via the gas introduction flow path 32 and the flow angle to the leading edge of the impeller 2 can be adjusted.

[0116] According to the above configuration, it is possible to selectively operate either the inlet-side flow control device 8 or the outlet-side flow control device 9, or to operate both simultaneously, depending on the change in the operating point. This allows for a wider operating range and improved efficiency compared to when only one of the inlet-side flow control device 8 or the outlet-side flow control device 9 is provided.

[0117] A centrifugal compressor 1 according to some embodiments includes the casing 3 described above. In this case, the centrifugal compressor 1 can be made highly efficient and have a wide operating range. Furthermore, a turbocharger 10 according to some embodiments includes the centrifugal compressor 1 described above and the turbine 12 described above, as shown in Figs. 1 and 2. In this case, the centrifugal compressor 1 can be made highly efficient and have a wide operating range.

[0118] In this specification, expressions expressing relative or absolute arrangement such as "in a certain direction," "along a certain direction," "parallel," "orthogonal," "center," "concentric," or "coaxial" not only express such an arrangement strictly, but also express a state in which there is a relative displacement with a tolerance or an angle or distance to the extent that the same function is obtained. For example, expressions such as "identical," "equal," and "homogeneous" that indicate that something is in an equal state not only indicate a state of strict equality, but also indicate a state in which there is a tolerance or a difference to the extent that the same function is obtained. Furthermore, in this specification, expressions representing shapes such as a rectangular shape or a cylindrical shape not only represent rectangular shapes or cylindrical shapes in the strict geometric sense, but also represent shapes including uneven portions, chamfered portions, etc., to the extent that the same effect can be obtained. Furthermore, in this specification, the expressions "comprise," "include," or "have" a component are not exclusive expressions that exclude the presence of other components.

[0119] The present disclosure is not limited to the above-described embodiments, but also includes modifications to the above-described embodiments and appropriate combinations of these modifications.

[0120] The contents of the above-described embodiments can be understood, for example, as follows.

[0121] 1) The casing (3) of the centrifugal compressor (1) according to at least one embodiment of the present disclosure comprises: A casing (3) of a centrifugal compressor configured to accommodate an impeller (2), a casing body (4) having at least one scroll passage (41) extending along the circumferential direction of the impeller (2) and through which compressed gas that has passed through the impeller (2) flows; a first discharge pipe section (5) communicating with the at least one scroll flow path (41) via a first branch port (52) and forming a first discharge flow path (51) through which the compressed gas introduced from the scroll flow path (41) flows; a second discharge pipe section (6) communicating with the at least one scroll flow path (41) via a second branch port (62) to form a second discharge flow path (61) through which the compressed gas introduced from the scroll flow path (41) flows, The first branch port (52) and the second branch port (62) are provided at different positions in the circumferential direction of the impeller (2).

[0122] According to the configuration 1), the compressed gas flowing through the scroll flow path (41) is discharged from each of the branch ports (the first branch port 51, the second branch port 61) provided at different positions in the circumferential direction of the impeller (2), thereby mitigating the influence of the circumferential static pressure distribution caused by the circumferential accumulation of the flow rate flowing into the scroll flow path (41). By mitigating the influence of the circumferential static pressure distribution, it is possible to suppress the occurrence of losses in the centrifugal compressor (1) caused by the circumferential static pressure distribution, and it is possible to improve the efficiency and widen the operating range (expansion of the operating range) of the centrifugal compressor (1).

[0123] 2) In some embodiments, the casing (3) of the centrifugal compressor (1) described in 1) above, the at least one scroll flow path (41) includes a first scroll flow path (41A) in which a first tongue portion (42A) is provided at a winding start position, and a second scroll flow path (41B) in which a second tongue portion (42B) is provided at a winding start position, The first discharge pipe portion (5) is configured such that the first discharge flow path (51) communicates with the first scroll flow path (41A) through the first branch port (52), The second discharge pipe portion (6) is configured such that the second discharge flow path (61) communicates with the second scroll flow path (41B) via the second branch port (62), The first tongue portion (42A) and the second tongue portion (42B) are provided at different positions in the circumferential direction of the impeller (2).

[0124] According to the configuration 2), the compressed gas flowing through the first scroll flow path (41A) can be discharged from the first branch port (52), and the compressed gas flowing through the second scroll flow path (41B) can be discharged from the second branch port (62). In this case, interference between the compressed gas flowing through the first scroll flow path (41A) or the first discharge flow path (51) and the compressed gas flowing through the second scroll flow path (41B) or the second discharge flow path (61) can be suppressed. This can suppress the occurrence of loss in the centrifugal compressor (1) caused by the interference, thereby improving the efficiency of the centrifugal compressor (1).

[0125] 3) In some embodiments, the casing (3) of the centrifugal compressor (1) described in 2) above, The casing body (4) is formed so that the first scroll passage (41A) and the second scroll passage (41B) are divided in the axial direction or the circumferential direction of the impeller (2).

[0126] According to the configuration 3), by dividing the first scroll passage (41A) and the second scroll passage (41B) in the axial direction or the circumferential direction of the impeller (2), it is possible to suppress a sudden turning of the compressed gas flow in each scroll passage (41A, 41B) and separation of the compressed gas from the wall surfaces of the scroll passages (41A, 41B) due to the turning. This makes it possible to suppress losses in the centrifugal compressor (1) caused by the turning and separation, thereby improving the efficiency of the centrifugal compressor (1).

[0127] When the first scroll flow path (41A) and the second scroll flow path (41B) are divided in the circumferential direction of the impeller (2), the influence of the circumferential static pressure distribution can be alleviated. When the first scroll flow path (41A) and the second scroll flow path (41B) are divided in the axial direction of the impeller (2), the static pressure distribution generated in the second scroll flow path (41B) can be shifted in the circumferential direction relative to the static pressure distribution generated in the first scroll flow path (41A). As a result, the static pressure at the outlet (34) of the diffuser flow path (33) can be made uniform in the circumferential direction. In these cases, the loss of the centrifugal compressor (1) caused by the circumferential static pressure distribution can be suppressed, and the centrifugal compressor (1) can be made more efficient and have a wider operating range.

[0128] 4) In some embodiments, the casing (3) of the centrifugal compressor (1) described in any one of 1) to 3) above, The casing body (4) has a plurality of tongues (42) provided at different positions in the circumferential direction of the impeller (2), When the number of the plurality of tongue portions (42) is n and the angle obtained by dividing 360° by n is θ1, each of the plurality of tongue portions (42) is configured so that the angle formed with an adjacent tongue portion (42) satisfies the condition θ1±10°.

[0129] According to the configuration 4), the plurality of tongue portions 42 are arranged at approximately equal intervals in the circumferential direction so as to satisfy the above condition, thereby making the circumferential static pressure distribution more uniform in the circumferential direction, thereby effectively suppressing the occurrence of losses in the centrifugal compressor 1 caused by the circumferential static pressure distribution.

[0130] 5) In some embodiments, the casing (3) of the centrifugal compressor (1) described in any one of 1) to 4) above, At least one of the first discharge pipe section (5) and the second discharge pipe section (6) has a bending section (54, 64) that bends a discharge flow path (51, 61) through which the compressed gas introduced from the scroll flow path (41) flows, and an axis (L53, L63) at a discharge port (53, 63) that is the outlet of the discharge flow path (51, 61) extends along a direction intersecting with a plane (OP) perpendicular to the axis (LA) of the impeller (2).

[0131] If neither the first discharge pipe section (5) nor the second discharge pipe section (6) had the bent section (54, 64), when the first branch port (52) and the second branch port (62) were provided at different positions in the circumferential direction of the impeller (2), the discharge port (53) of the first discharge pipe section (5) and the discharge port (63) of the second discharge pipe section (6) would face in different directions, which could result in a complicated piping layout of the first discharge pipe section (5) or the second discharge pipe section (6). According to the configuration of 5), the first discharge pipe section (5) or the second discharge pipe section (6) is configured to have the bent section (54, 64) that bends the discharge flow path (52, 62), thereby making it possible to prevent the piping layout from becoming complicated.

[0132] 6) In some embodiments, the casing (3) of the centrifugal compressor (1) described in any one of 1) to 4) above, the first discharge pipe portion (5) has a first bent portion (54) that bends the first discharge flow path (51), and an axis (L53) at a first discharge port (53) that is an outlet of the first discharge flow path (51) extends along a direction that intersects with a plane (OP) that is perpendicular to the axis (LA) of the impeller (2); the second discharge pipe portion (6) has a second bent portion (64) that bends the second discharge flow path (61), and an axis (L63) at a second discharge port (63) that is an outlet of the second discharge flow path (61) extends along a direction that intersects with a plane (OP) that is perpendicular to the axis (LA) of the impeller (2); The axis (L63) of the second discharge port (63) extends in a direction not parallel to the direction in which the axis (L53) of the first discharge port (53) extends.

[0133] According to the configuration of 6), the axis (L53) of the first discharge port (53) of the first discharge pipe section (5) having the first bent portion (54) and the axis (L63) of the second discharge port (63) of the second discharge pipe section (6) having the second bent portion (64) are extended in directions that are not parallel to each other, thereby increasing the degree of freedom in the piping layout of the first discharge pipe section (5) and the second discharge pipe section (6).

[0134] 7) In some embodiments, the casing (3) of the centrifugal compressor (1) described in any one of 1) to 6) above, the first discharge pipe portion (5) and the second discharge pipe portion (6) have a connection portion (36) connected to the same pipe (19); The first discharge pipe section (5) is disposed adjacent to the second discharge pipe section (6) over a predetermined section (S1) upstream from the connection section (36).

[0135] According to the configuration of 7), the discharge port (53) of the first discharge pipe section (5) and the discharge port (63) of the second discharge pipe section (6) can be connected to the same pipe (19), thereby enabling a piping layout similar to that of a conventional casing having one discharge port (53). In addition, by arranging the first discharge pipe section (5) and the second discharge pipe section (6) adjacent to each other over a predetermined section (S1) upstream from the connecting section, it is possible to prevent the piping layout from becoming complicated.

[0136] 8) In some embodiments, the casing (3) of the centrifugal compressor (1) described in 7) above, The first discharge pipe section (5) and the second discharge pipe section (6) are configured so that a first discharge port (53), which is the outlet of the first discharge flow path (51), and a second discharge port (63), which is the outlet of the second discharge flow path (61), are on the same plane (PS).

[0137] According to the configuration of 8), the first discharge port (53) and the second discharge port (63) are located on the same plane (PS), which makes it easy to connect the first discharge port (53) and the second discharge port (63) to the same pipe (19), thereby preventing the piping layout from becoming complicated.

[0138] 9) In some embodiments, the casing (3) of the centrifugal compressor (1) described in 2) or 3) above, In the circumferential direction of the impeller (2), a first circumferential angular position, in which a winding start position of the first scroll passage (41A) is defined as 0° and a downstream side of the first scroll passage (41A) is defined as a positive angle, is defined as α, a flow path cross-sectional area of ​​the first scroll passage (41A) at the first circumferential angular position α is defined as A1(α), and a radius from an axis (LA) of the impeller (2) to a center of a flow path cross-section of the first scroll passage (41A) is defined as R1(α), In the circumferential direction of the impeller (2), a winding start position of the second scroll passage (41B) is defined as 0° and a second circumferential angular position, in which the downstream side of the second scroll passage (41B) is positive, is defined as β, a flow path cross-sectional area of ​​the second scroll passage (41B) at the second circumferential angular position β is defined as A2(β), and a radius from an axis (LA) of the impeller (2) to a center of the flow path cross-section of the second scroll passage (41B) is defined as R2(β), It is configured to satisfy either the following formula (1) or (2) at least when α=β. A1(α)≠A2(β) (1) A1(α) / R1(α)≠A2(β) / R2(β) (2)

[0139] In the configuration of 9), the circumferential cross-sectional area distribution and the central diameter distribution of the scroll passage (41) significantly affect the design point of the scroll passage (41), i.e., the operating point at which the efficiency is maximized. Therefore, by differentiating the circumferential cross-sectional area distribution and the central diameter distribution of the two scroll passages (41A, 41B), robust efficiency characteristics can be obtained over a wide operating range.

[0140] 10) In some embodiments, the casing (3) of the centrifugal compressor (1) described in 9) above, It is configured to satisfy either the following formula (3) or (4) at least when α=β. A1(α) <A2(β)···(3) A1(α) / R1(α) <A2(β) / R2(β)···(4)

[0141] According to the above configuration 10), the second scroll flow path (41B) is given a circumferential cross-sectional area distribution and a central diameter distribution that maximize the efficiency on the small flow rate side, and the first scroll flow path (41A) is given a circumferential cross-sectional area distribution and a central diameter distribution that maximize the efficiency on the large flow rate side, thereby improving the efficiency on both the small flow rate side and the large flow rate side.

[0142] 11) In some embodiments, the casing (3) of the centrifugal compressor (1) described in 10) above, The gas supply system further includes an outlet-side flow rate control device (9) configured to be able to adjust the flow rate of the compressed gas flowing through the first discharge flow path (51).

[0143] According to the above configuration 11), under a high flow rate operating condition, the flow rate of the compressed gas flowing through the first discharge flow path (51) is adjusted by the outlet-side flow control device (9), whereby the flow rates of the compressed gas flowing through the first scroll flow path (41A) and the second scroll flow path (41B) can be made appropriate. Furthermore, by providing the outlet-side flow control device (9) in the first discharge flow path (51) communicating with the first scroll flow path (41A) in which the flow path cross-sectional area (A) or the ratio of the flow path cross-sectional area to the radius (A / R) of the scroll flow path (41) is relatively small, it is possible to reduce loss caused by a sudden change in area near the tongue portion (42B) close to the discharge flow path in which the outlet-side flow control device (9) is provided when the flow rate of the compressed gas is reduced by the outlet-side flow control device (9), compared to a case in which the outlet-side flow control device (9) is provided in the second discharge flow path (61) communicating with the second scroll flow path (41B) in which the flow path cross-sectional area (A) or the ratio of the flow path cross-sectional area to the radius (A / R) of the scroll flow path (41) is relatively large.

[0144] 12) In some embodiments, the casing (3) of the centrifugal compressor (1) described in 11) above, The area (SA1) of the first discharge port (53) which is the outlet of the first discharge flow path (51) is smaller than the area (SA2) of the second discharge port (63) which is the outlet of the second discharge flow path (61).

[0145] According to the configuration 12), it is possible to reduce loss caused by a sudden change in area near the tongue (second tongue 42B) close to the first discharge port (53) when the first discharge port (531) is closed by the valve body (91) of the outlet-side flow control device (9).

[0146] 13) In some embodiments, the casing (3) of the centrifugal compressor (1) described in 12) above, The area of ​​the first discharge port (53) is defined as SA1, When the area of ​​the second discharge port (63) is defined as SA2, the second discharge port (63) is configured to satisfy the following formula (5). 55%≦SA2 / (SA1+SA2)≦65% (5)

[0147] If the area (SA1) of the first discharge port (53) and the area (SA2) of the second discharge port (63) were approximately the same, when the flow rate of the compressed gas was reduced by the outlet-side flow control device (9) (when the first discharge flow path (51) was closed), the outlet for the compressed gas (the total area of ​​the open discharge ports) would suddenly become small, which could cause a sudden change in performance of the centrifugal compressor (1). According to the configuration of 13), when the flow rate of the compressed gas was reduced by the outlet-side flow control device (9), the outlet for the compressed gas (the total area of ​​the open discharge ports) can be prevented from suddenly becoming small. Therefore, the characteristics of the centrifugal compressor (1) can be smoothly changed.

[0148] 14) In some embodiments, the casing (3) of the centrifugal compressor (1) described in any one of 1) to 9) above, The compressor further includes an outlet-side flow rate control device (9) configured to be able to adjust the flow rate of at least one of the compressed gas flowing through the first discharge flow path (51) and the compressed gas flowing through the second discharge flow path (61).

[0149] According to the configuration of 14), the centrifugal compressor (1) including the outlet-side flow control device (9) can achieve a wider operating range and improved compression efficiency compared to the centrifugal compressor (1) including the inlet-side flow control device (8). Moreover, the outlet-side flow control device (9) has a simpler structure than the inlet-side flow control device (8), and is easier to manufacture and install. Moreover, the outlet-side flow control device (9) is less likely to adversely affect the performance of the impeller (2) compared to the inlet-side flow control device (8).

[0150] 15) In some embodiments, the casing (3) of the centrifugal compressor (1) described in 12) above, The casing body (4) has a gas introduction passage (32) for introducing gas to the impeller (2), The casing (3) of the centrifugal compressor (1) further includes an inlet flow rate control device (8) configured to be able to adjust the flow rate of the gas introduced into the impeller (2) through the gas introduction passage (32).

[0151] According to the configuration of 15), either the inlet-side flow control device (8) or the outlet-side flow control device (9) can be selectively operated, or both can be operated simultaneously, depending on the change in the operating point. This allows for a wider operating range and improved efficiency compared to when only one of the inlet-side flow control device (8) or the outlet-side flow control device (9) is provided.

[0152] 16) The centrifugal compressor (1) according to at least one embodiment of the present disclosure comprises: The centrifugal compressor (1) includes a casing (3) according to any one of 1) to 15) above.

[0153] According to the above configuration 16), the centrifugal compressor (1) can be made highly efficient and have a wide operating range.

[0154] 17) A turbocharger (10) according to at least one embodiment of the present disclosure includes: The centrifugal compressor (1) according to 16) above; and a turbine (12) configured to drive the centrifugal compressor (1).

[0155] According to the configuration of 17) above, the centrifugal compressor (1) can be made highly efficient and have a wide operating range. [Explanation of symbols]

[0156] 1,1A centrifugal compressor 2 impellers 3 Casing 4 Casing body 5 1st discharge pipe section 6 Second discharge pipe section 7 Third discharge pipe section 8 Inlet side flow rate adjustment device 9 Outlet side flow rate adjustment device 10. Turbocharger 11 Internal combustion engine system 12 Turbine 13 Internal combustion engine 14 Turbine rotor 15 Turbine housing 16 Rotating shaft 17 Bearings 18 Bearing housing 31 Shroud surface 32 Gas introduction channel 33 Diffuser passage 34 Exit 35 Bulkhead 41 Scroll flow passage 41A First scroll passage 41B Second scroll passage 41C Third scroll passage 42 Tongue 42A First Tongue 42B 2nd tongue 42C Third Tongue 51 first discharge flow path 52 First Branch 53 1st discharge port 54 First bend 61 Second discharge flow path 62 Second Branch 63 2nd outlet 64 Second bend 71 Second discharge flow path 72 Second Branch 73 2nd outlet

Claims

1. A casing for a centrifugal compressor configured to be able to accommodate an impeller, The impeller extends along the circumferential direction of the impeller. a casing body having at least one scroll passage through which the compressed gas flows; a first discharge pipe portion that communicates with the at least one scroll passage via a first branch port and forms a first discharge passage through which the compressed gas introduced from the scroll passage flows; a second discharge pipe portion that communicates with the at least one scroll passage via a second branch port and forms a second discharge passage through which the compressed gas introduced from the scroll passage flows, the first branch port and the second branch port are provided at different positions in a circumferential direction of the impeller, the at least one scroll passage includes a first scroll passage having a first tongue portion provided at a winding start position and a second scroll passage having a second tongue portion provided at a winding start position, the first discharge pipe portion is configured such that the first discharge flow path communicates with the first scroll flow path via the first branch port; the second discharge pipe portion is configured such that the second discharge flow path communicates with the second scroll flow path via the second branch port, the first tongue portion and the second tongue portion are provided at different positions in the circumferential direction of the impeller, In the circumferential direction of the impeller, a first circumferential angular position, in which a winding start position of the first scroll passage is defined as 0° and the downstream side of the first scroll passage is defined as positive, is defined as α, a flow path cross-sectional area of ​​the first scroll passage at the first circumferential angular position α is defined as A1(α), and a radius from the axis of the impeller to the center of the flow path cross section of the first scroll passage is defined as R1(α), In the circumferential direction of the impeller, a winding start position of the second scroll passage is defined as 0°, and a second circumferential angular position in which the downstream side of the second scroll passage is positive is defined as β, a flow path cross-sectional area of ​​the second scroll passage at the second circumferential angular position β is defined as A2(β), and a radius from the axis of the impeller to the center of the flow path cross section of the second scroll passage is defined as R2(β), It is configured to satisfy either of the following formulas (1) or (2) at least when α = β: Centrifugal compressor casing. A1(α)≠A2(β)...(1) A1(α) / R1(α)≠A2(β) / R2(β)...(2)

2. A casing for a centrifugal compressor configured to accommodate an impeller, The impeller extends along the circumferential direction of the impeller. a casing body having at least one scroll passage through which the compressed gas flows; a first discharge pipe portion that communicates with the at least one scroll passage via a first branch port and forms a first discharge passage through which the compressed gas introduced from the scroll passage flows; a second discharge pipe portion that communicates with the at least one scroll passage via a second branch port and forms a second discharge passage through which the compressed gas introduced from the scroll passage flows, the first branch port and the second branch port are provided at different positions in a circumferential direction of the impeller, the at least one scroll passage includes a first scroll passage having a first tongue portion provided at a winding start position and a second scroll passage having a second tongue portion provided at a winding start position, the first discharge pipe portion is configured such that the first discharge flow path communicates with the first scroll flow path via the first branch port; the second discharge pipe portion is configured such that the second discharge flow path communicates with the second scroll flow path via the second branch port, The first tongue portion and the second tongue portion are provided at different positions in the circumferential direction of the impeller. The casing body is formed so that the first scroll passage and the second scroll passage are separated in the axial direction of the impeller. Centrifugal compressor casing.

3. The casing main body is formed so that the first scroll passage and the second scroll passage are divided in the axial direction or the circumferential direction of the impeller. A casing for a centrifugal compressor according to claim 1.

4. the casing body has a plurality of tongues provided at different positions in the circumferential direction of the impeller, When the number of the plurality of tongue portions is n and the angle obtained by dividing 360° by n is θ1, each of the plurality of tongue portions is configured so that the angle formed with an adjacent tongue portion satisfies the condition of θ1±10°. A casing for a centrifugal compressor according to any one of claims 1 to 3.

5. At least one of the first discharge pipe section and the second discharge pipe section has a bending section that bends a discharge flow path through which the compressed gas introduced from the scroll flow path flows, and an axis of a discharge port that is an outlet of the discharge flow path extends along a direction intersecting a plane perpendicular to the axis of the impeller. A casing for a centrifugal compressor according to any one of claims 1 to 3.

6. the first discharge pipe portion has a first bending portion that bends the first discharge flow path, and an axis of a first discharge port that is an outlet of the first discharge flow path extends along a direction that intersects with a plane that is perpendicular to the axis of the impeller, the second discharge pipe portion has a second bent portion that bends the second discharge flow path, and an axis of a second discharge port that is an outlet of the second discharge flow path extends along a direction that intersects with a plane that is perpendicular to the axis of the impeller, The axis of the second outlet extends in a direction that is not parallel to the direction in which the axis of the first outlet extends. A casing for a centrifugal compressor according to any one of claims 1 to 3.

7. the first discharge pipe section and the second discharge pipe section have connection sections connected to the same pipe, The first discharge pipe section is disposed adjacent to the second discharge pipe section over a predetermined section upstream from the connecting section. A casing for a centrifugal compressor according to any one of claims 1 to 3.

8. the first discharge pipe portion and the second discharge pipe portion are configured such that a first discharge port, which is an outlet of the first discharge flow path, and a second discharge port, which is an outlet of the second discharge flow path, are on the same plane; A casing for a centrifugal compressor according to claim 7.

9. In the circumferential direction of the impeller, a first circumferential angular position, in which a winding start position of the first scroll passage is defined as 0° and the downstream side of the first scroll passage is defined as positive, is defined as α, a flow path cross-sectional area of ​​the first scroll passage at the first circumferential angular position α is defined as A1(α), and a radius from the axis of the impeller to the center of the flow path cross section of the first scroll passage is defined as R1(α), In the circumferential direction of the impeller, a winding start position of the second scroll passage is defined as 0°, and a second circumferential angular position in which the downstream side of the second scroll passage is positive is defined as β, a flow path cross-sectional area of ​​the second scroll passage at the second circumferential angular position β is defined as A2(β), and a radius from the axis of the impeller to the center of the flow path cross section of the second scroll passage is defined as R2(β), It is configured to satisfy either of the following formulas (1) or (2) at least when α = β: A casing for a centrifugal compressor according to claim 2. A1(α)≠A2(β)...(1) A1(α) / R1(α)≠A2(β) / R2(β)...(2)

10. It is configured to satisfy either of the following formulas (3) or (4) at least when α = β: A casing for a centrifugal compressor according to claim 1 or 9. A1(α)<A2(β)...(3) A1(α) / R1(α)<A2(β) / R2(β)...(4)

11. an outlet-side flow rate control device configured to adjust the flow rate of the compressed gas flowing through the first discharge flow path; A casing for a centrifugal compressor according to claim 10.

12. an area of ​​a first discharge port which is an outlet of the first discharge flow path is smaller than an area of ​​a second discharge port which is an outlet of the second discharge flow path; A casing for a centrifugal compressor according to claim 11.

13. The area of ​​the first outlet is defined as SA1, When the area of ​​the second discharge port is defined as SA2, the ink jet head is configured to satisfy the following formula (5): A casing for a centrifugal compressor according to claim 12. 55%≦SA2 / (SA1+SA2)≦65%...(5)

14. further comprising an outlet-side flow rate control device configured to be able to adjust the flow rate of at least one of the compressed gas flowing through the first discharge flow path and the compressed gas flowing through the second discharge flow path; A casing for a centrifugal compressor according to any one of claims 1 to 3.

15. the casing body has a gas introduction passage for introducing gas to the impeller, the casing of the centrifugal compressor further includes an inlet-side flow rate control device configured to adjust the flow rate of the gas introduced into the impeller through the gas introduction passage, A casing for a centrifugal compressor according to claim 12.

16. A centrifugal compressor comprising the casing of any one of claims 1 to 3.

17. A centrifugal compressor according to claim 16; a turbine configured to drive the centrifugal compressor; and A turbocharger comprising:

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