JP6730917B2 - Centrifugal compressor and turbocharger - Google Patents

Description

The present disclosure relates to centrifugal compressors and turbochargers.

For example, in a turbocharger, a centrifugal compressor is used to increase the density of air taken in by the engine.
Centrifugal compressors have an impeller composed of multiple blades installed upright on the hub in order to give energy to the air flow and raise the pressure, and the air is rotated by rotating the impeller at high speed with a rotating shaft. It sucks the flow in the axial direction, gives a high speed velocity around the rotation axis to the air, and at the same time raises the pressure, causes it to flow out toward the outer circumference of the radius, and the diffuser installed at the outer circumference of the outlet of the impeller has high speed and increased pressure. An impeller with the function of sending in air is installed.
Below, it is referred to as a "centrifugal compressor impeller", or simply "impeller".

FIG. 26 is a diagram showing streamlines on the meridional plane near the highest efficiency point in the conventional centrifugal compressor impeller. 27 is a view taken along the line AA in FIG. 26, and is a diagram showing streamlines between blades on the rotating flow surface.

As shown in FIG. 27, in the centrifugal compressor impeller, generally, in the vicinity of the maximum efficiency point, the angle formed by the streamline flowing into the blade leading edge 030 and the centerline Lc of the blade cross section at the rotating flow surface at the inlet of the blade 012. The collision angle defined as is approximately 0, and the loss is small and the efficiency is high.

In the centrifugal compressor, when the flow rate decreases with the rotation speed kept constant, as shown in FIG. 28, the collision angle θ increases at the blade leading edge 030, and the flow near the blade tip edge separates from the blade surface. The blade 012 stalls and the efficiency decreases.

When the flow rate further decreases and the collision angle θ increases, the flow separation region near the blade tip edge expands toward the hub side, and a circulating flow Fc including backflow occurs near the blade leading edge 030. As shown in FIG. 29, when the circulation flow Fc is viewed from the meridian plane, the circulation flow Fc near the blade leading edge 030 flows outward in the radial direction of the impeller, and the pressure of the flow increases accordingly. Therefore, the flow goes toward the upstream side in the vicinity of the tip edge of the blade 012. The flow to the upstream side then flows toward the radially inner side of the impeller, the flow along the suction surface is separated at the blade leading edge, and the flow again goes to the radially outer side of the impeller. Therefore, the loss tends to increase as the flow rate decreases. In addition, below, the said circulation flow Fc is called "backflow of a meridian surface."
It is said that such a backflow on the meridian surface starts when the flow rate decreases from the highest efficiency point, the flow near the blade tip edge separates from the blade surface and stalls, and when the flow rate further decreases, it spreads to the region near the blade leading edge. Has characteristics.

When the flow rate further decreases, the backflow on the meridian surface expands. When the backflow on the meridian surface expands in the blade height direction and the axial direction, strain in the flow at the blade inlet reaches the trailing edge of the blade, causing the diffuser to stall and generate a surge.

In this way, when the flow rate decreases, blade stall and meridian backflow occur, resulting in reduced efficiency due to pressure drop and increased loss.When the flow rate further decreases, the meridional backflow expands and the diffuser stalls. There is a problem that surge occurs.

FIG. 30 shows a typical example of a flow rate-pressure ratio characteristic curve C0 showing the relationship between the flow rate and the pressure ratio at a constant rotation speed in the conventional centrifugal compressor.

As shown in FIG. 30, when the flow rate in the centrifugal compressor is lower than a certain flow rate, the pressure ratio (and efficiency) is reduced, and when the flow rate is further reduced, a surge occurs.

As shown in FIG. 30, in the flow rate-pressure ratio characteristic at a constant rotation speed, the flow rate difference between the design point (a point near the maximum efficiency point) and the surge point at that rotation speed is called a surge margin. Further, as shown in FIG. 8, the surge margin of operation of the compressor in a continuous rotation speed range refers to a flow rate difference between each operation point and a surge line which is an envelope of surge points at each rotation speed. ..

In centrifugal compressors for turbochargers, blade stall as described above narrows the surge margin, and pressure and efficiency decrease as the flow rate decreases, so the surge margin becomes wider and the pressure accompanying the flow rate decreases. A centrifugal compressor that is less likely to cause a decrease in efficiency and efficiency is desired.

Patent Document 1 discloses a centrifugal compressor intended to reduce the surging limit flow rate at a low flow rate. In the centrifugal compressor described in Patent Document 1, a resistor that radially narrows a passage cross section of an intake passage that communicates between an impeller of the centrifugal compressor and an intake port is provided to increase an inflow speed of the impeller into a blade. By doing so, the surging limit flow rate is reduced at a low flow rate.

International Publication No. 2014-128931

In the centrifugal compressor described in Patent Document 1, in order to reduce the surging limit flow rate at a low flow rate, it is necessary to provide a resistor configured as a perforated plate, a grid plate or a mesh plate, The configuration of the centrifugal compressor tends to be complicated.

The present invention has been made in view of the above-described conventional problems, and an object thereof is to provide a centrifugal compressor and a centrifugal compressor capable of suppressing a decrease in pressure and efficiency due to a decrease in flow rate with a simple configuration. It is to provide a turbocharger equipped with.

(1) A centrifugal compressor according to at least one embodiment of the present invention is a centrifugal compressor that includes an impeller and a first shroud casing that houses the impeller, the impeller being provided on a hub and an outer peripheral surface of the hub. A plurality of erected blades, and at least one blade of the plurality of blades has a first tip edge extending along the first shroud casing and an upstream end of the first tip edge. A first front edge extending inward in the radial direction of the impeller, a second tip edge extending upstream from an inner end of the first front edge in the radial direction, and an upstream end of the second tip edge. A second front edge extending inward in the radial direction.

In the centrifugal compressor according to (1) above, when the flow rate decreases, the flow near the first leading edge side portion of the first leading edge separates from the blade surface, and the above-mentioned “backflow of meridian surface” occurs. , A local inward flow due to the backflow is generated toward the radially inner second tip edge of the impeller. On the other hand, the flow passing through the second leading edge is subjected to a force in the impeller rotation direction by the blades to increase the speed in the circumferential direction, and the centrifugal force resulting from the speed in the circumferential direction causes the flow to flow downstream while flowing. It flows outward. The radially inward flow after backflowing upstream in the vicinity of the first leading edge side portion of the first leading edge is hindered from proceeding by the radially outward flow after passing through the second leading edge. Therefore, it does not flow further inward in the radial direction. As a result, the "backflow of the meridional plane" is suppressed in development and is confined in the region near the first leading edge and the first leading edge, and becomes a stable circulating flow.

Due to this stable circulation flow, the flow path is blocked at the position of the first front edge, so that the flow from the first front edge hardly flows into the impeller, and the flow from the second front edge mainly flows into the impeller. .. Further, since a smooth boundary surface is formed at the boundary between the stable circulation flow and the flow flowing into the impeller from the second leading edge, the flow passing through the second leading edge can flow smoothly to the trailing edge. .. As a result, the pressure drop and loss increase due to the blade stall and backflow described above are suppressed, and the pressure and efficiency decrease due to the flow rate decrease can be suppressed.
Further, as compared with the centrifugal compressor according to Patent Document 1, it is not necessary to provide a resistor having a complicated structure such as a perforated plate, so that the above effect can be obtained with a simple structure.

(2) In some embodiments, the centrifugal compressor according to (1) above further includes a second shroud casing provided inside an inner peripheral surface of the first shroud casing, and the second shroud casing. Is located upstream of the first front edge, forms a first flow path with the inner peripheral surface of the first shroud casing, and has a second flow inside the inner peripheral surface of the second shroud casing. Configured to form a passage.

In the centrifugal compressor according to (2) above, when the flow rate decreases, the flow near the first leading edge side portion of the first leading edge separates from the blade surface, and the above-mentioned “backflow of meridian surface” occurs. , A local inward flow due to the backflow is generated toward the radially inner second tip edge of the impeller. On the other hand, the flow passing through the second leading edge is subjected to a force in the impeller rotation direction by the blades to increase the speed in the circumferential direction, and the centrifugal force resulting from the speed in the circumferential direction causes the flow to flow downstream while flowing. It flows outward. The flow inward in the radial direction after backflowing upstream in the vicinity of the first leading edge side portion of the first front edge is prevented from proceeding by the outer peripheral surface of the second shroud casing, and therefore further inward in the radial direction. It doesn't flow to. As a result, the "backflow of the meridional plane" is suppressed in development and is confined in the region near the first leading edge and the first leading edge, and becomes a stable circulating flow.

Due to this stable circulation flow, the flow path is blocked at the position of the first leading edge, so that the flow from the first leading edge hardly flows into the impeller, and the flow from the second leading edge mainly flows into the impeller. .. Further, since a smooth boundary surface is formed at the boundary between the stable circulation flow and the flow flowing into the impeller from the second leading edge, the flow passing through the second leading edge can flow smoothly to the trailing edge. .. As a result, it is possible to suppress the pressure decrease and the loss increase due to the blade stall and the backflow described above, and it is possible to suppress the pressure and efficiency decrease due to the flow rate decrease.
Further, as compared with the centrifugal compressor according to Patent Document 1, it is not necessary to provide a resistor having a complicated structure such as a perforated plate, so that the above effect can be obtained with a simple structure.

(3) In some embodiments, the centrifugal compressor according to (1) or (2) further includes an annular second tip edge shroud fixed to the second tip edge.

According to the centrifugal compressor described in (3) above, it is possible to suppress a decrease in pressure and efficiency due to a decrease in flow rate.

(4) In some embodiments, in the centrifugal compressor according to (1), an annular second tip edge shroud fixed to the second tip edge and an inner peripheral surface of the first shroud casing are provided. A second shroud casing provided inside, the second shroud casing being located upstream of the first front edge, and having a first flow path between the second shroud casing and an inner peripheral surface of the first shroud casing. And forming a second flow path inside the inner peripheral surface of the second shroud casing, wherein the upstream end of the second tip edge shroud is upstream of the downstream end of the second shroud casing. Located on the side.

According to the centrifugal compressor described in (4) above, it is possible to suppress a decrease in pressure and efficiency due to a decrease in flow rate.

(5) In some embodiments, the centrifugal compressor according to any one of (1) to (4) above, further comprising an annular first tip edge shroud fixed to the first tip edge. ..

According to the centrifugal compressor described in (5) above, it is possible to suppress a decrease in pressure and efficiency due to a decrease in flow rate.

(6) In some embodiments, in the centrifugal compressor according to any one of (1) to (5), a downstream end of the second shroud casing has the second tip in the radial direction. Located outside the rim.

According to the centrifugal compressor described in (6) above, it is possible to prevent the flow of the "backflow of the meridian surface" inward in the radial direction from outside the second tip edge in the radial direction. The development of "reflux" can be effectively suppressed. As a result, it is possible to effectively suppress the decrease in pressure and efficiency due to the decrease in flow rate with a simple configuration.

(7) In some embodiments, in the centrifugal compressor according to any one of (1) to (6) above, a downstream end of the second shroud casing has the first end in the axial direction of the impeller. 2 Located downstream from the upstream end of the leading edge.

According to the centrifugal compressor described in (7) above, the progress of the flow inward in the "backflow of the meridional plane" in the radial direction is prevented at a position radially outward of the second leading edge and close to the first front edge. Therefore, it is possible to effectively suppress the development of “backflow of the meridian plane”. As a result, it is possible to effectively suppress the decrease in pressure and efficiency due to the decrease in flow rate with a simple configuration.

(8) In some embodiments, in the centrifugal compressor according to any one of (1) to (7), the second front edge is an outer end of the second front edge in the radial direction. Is inclined with respect to the radial direction so as to be positioned upstream of the inner end of the second front edge in the radial direction.

According to the centrifugal compressor described in (8) above, since the second leading edge is inclined in this manner, the pressure surface of the blade near the second leading edge moves in the outlet direction (radially outward) of the impeller. It will turn. A force in the direction perpendicular to the pressure surface of the blade acts on the flow flowing into the impeller, so that the action of turning the flow radially outward from the axial direction becomes large.

Therefore, it is possible to enhance the action of turning the flow flowing into the impeller from the second flow path to the radial outside in the vicinity of the second front edge. As a result, the expansion of the boundary layer in the vicinity of the inner peripheral surface of the first shroud casing can be suppressed, and the distribution of the flow velocity on the tip side at the trailing edge of the blade can be approximated to a uniform distribution.

As a result, the impeller efficiency is improved and the pressure recovery rate of the diffuser is improved, so that the centrifugal compressor can be made highly efficient.

(9) In some embodiments, in the centrifugal compressor according to any one of (1) to (8), the first front edge is an outer end of the first front edge in the radial direction. Is inclined with respect to the radial direction so as to be located upstream of the inner end of the first front edge in the radial direction.

According to the centrifugal compressor described in (9) above, since the first leading edge is thus inclined, the impeller is removed from the impeller from the first flow path for the same reason as when the second leading edge is inclined. It is possible to enhance the action of turning the flow that flows in to the outside in the radial direction near the first front edge. As a result, the expansion of the boundary layer in the vicinity of the inner peripheral surface of the first shroud casing can be suppressed, and the distribution of the flow velocity on the tip side at the trailing edge of the blade can be approximated to a uniform distribution.

As a result, the impeller efficiency is improved and the pressure recovery rate of the diffuser is improved, so that the centrifugal compressor 2 can be made highly efficient.

(10) In some embodiments, in the centrifugal compressor according to any one of (1) to (9), the first leading edge is an outer end of the first leading edge in the radial direction. Is inclined with respect to the radial direction so as to be located downstream of the inner end of the first front edge in the radial direction.

According to the centrifugal compressor described in (6) above, it is possible to increase the vibration strength of the blade.

(11) In some embodiments, in the centrifugal compressor according to any one of (1) to (10), the second shroud casing has a blade shape on a meridional surface of the impeller.

According to the centrifugal compressor described in (11) above, it is possible to suppress the pressure loss due to the resistance of the second shroud casing itself.

(12) In some embodiments, in the centrifugal compressor according to any one of (1) to (11), the inner peripheral surface of the second shroud casing has a radius that increases toward the upstream side. It has a bell mouth shape.

According to the centrifugal compressor described in (12) above, since the flow passage cross-sectional area of the second flow passage can be reduced toward the downstream side, the flow of the second flow passage can be accelerated in the axial direction. it can.

(13) In some embodiments, in the centrifugal compressor according to any one of (1) to (12), the outer peripheral surface of the second shroud casing has a curved diameter expansion whose radius increases toward the upstream side. A surface and a curved connection surface that connects the diameter expansion surface and the inner peripheral surface of the second shroud casing on the upstream side of the diameter expansion surface.

According to the centrifugal compressor described in (13) above, the flow can be smoothly guided to the first front edge along the connection surface and the expanded diameter surface.

(14) In some embodiments, in the centrifugal compressor according to any one of (1) to (13), the first flow path has a bellmouth-shaped inlet and is provided on a downstream side. It has a throttle channel portion whose channel cross-sectional area becomes smaller as it goes.

According to the centrifugal compressor described in (14) above, since the flow passage cross-sectional area of the throttle flow passage portion becomes smaller toward the downstream side, the flow in the first flow passage can be accelerated in the axial direction.

(15) In some embodiments, in the centrifugal compressor according to any one of (1) to (14), the impeller casing supports the first flow so as to support the second shroud casing. The support includes a support provided in the passage, and the support has a blade shape in a cross section orthogonal to the radial direction of the impeller.

According to the centrifugal compressor described in (15) above, since the cross-sectional shape of the support column has a small loss, the pressure loss due to the resistance of the support column itself can be reduced.

(16) In some embodiments, in the centrifugal compressor according to any one of (1) to (15), the impeller casing supports the first flow so as to support the second shroud casing. A second shroud casing, the support pillar, and the annular support member, each of which includes a support pillar provided in the passage, and an annular support portion provided to support the support pillar from the outside of the first flow path in the radial direction. The support part is integrated, and the annular support part is configured to be attachable to and detachable from the first shroud casing.

According to the centrifugal compressor described in (16) above, if the second shroud casing is formed as another component of the first shroud casing and can be assembled coaxially, the height of the second tip edge of the impeller can be increased. After being set by casting or adjusted by machining, it is possible to manufacture a component constituted by the second shroud casing, the pillar and the annular supporting portion depending on the height. Therefore, the degree of freedom in performance design and the degree of freedom in structure design of the centrifugal compressor can be increased.

(17) A turbocharger according to at least one embodiment of the present invention includes the centrifugal compressor according to any one of (1) to (16) above.

According to the turbocharger described in (17) above, since the centrifugal compressor that can suppress the decrease in pressure and efficiency due to the decrease in flow rate with a simple configuration is provided, a simple and high-performance turbocharger can be provided. it can.

According to at least one embodiment of the present invention, there is provided a centrifugal compressor and a turbocharger including the centrifugal compressor, which can suppress the pressure and the efficiency from being reduced due to the reduction in the flow rate with a simple configuration.

It is a meridional view which shows schematically a part of centrifugal compressor 2 in the turbocharger 100 which concerns on one Embodiment. It is a front view which shows schematically some impellers 4 in FIG. It is a figure which shows the flow on the meridional surface at the time of a large flow rate near the maximum efficiency point in the centrifugal compressor 2 (henceforth "flow rate A time"). It is a figure which shows the flow on a meridian surface at the time of a small flow volume (henceforth "flow volume B time") whose flow volume is smaller than the flow volume A in the centrifugal compressor 2. Flow of the inter-blade flow surface in the vicinity of the first tip edge 24 when the flow rate B of the centrifugal compressor 2 is B (arrow B-B view in FIG. 4) and in the vicinity of the second tip edge 28 when the flow rate B of the centrifugal compressor 2 is B. FIG. 5 is a diagram showing the flow on the flow surface between blades of FIG. It is a figure which shows the flow volume-pressure ratio characteristic curve C1 of the centrifugal compressor 2. It is a figure which shows the flow volume-efficiency characteristic curve C2 of the centrifugal compressor 2. The flow rate-pressure ratio characteristic curves C1a, C1b, C1c of the centrifugal compressor 2 for each impeller rotation speed, which increase substantially in proportion to the increase in the engine rotation speed in the turbocharger 100, and the corresponding conventional impeller rotation speeds for each impeller rotation speed. It is a figure which shows the flow volume-pressure ratio characteristic curve C0a, C0b, C0c of a centrifugal compressor. It is a figure which shows the relationship between the flow rate-pressure ratio characteristic curve C0a, C0b, C0c in a conventional centrifugal compressor, and an engine operating characteristic. It is a figure which shows the flow volume-pressure ratio characteristic curve C1a, C1b, C1c in the centrifugal compressor 2, and the relationship of an engine operating characteristic. It is a figure which shows roughly a part of meridional surface of the centrifugal compressor 2 in the turbocharger 100 which concerns on one Embodiment. FIG. 12 is a front view schematically showing a part of the impeller 4 in FIG. 11. It is a figure which shows roughly a part of meridional surface of the centrifugal compressor 2 in the turbocharger 100 which concerns on one Embodiment. FIG. 14 is a front view schematically showing a part of the impeller 4 in FIG. 13. The annular support portion 54 is a diagram showing how the annular support portion 54 is attached to and detached from the first shroud casing 14. It is a figure which shows roughly a part of meridional surface of the centrifugal compressor 2 in the turbocharger 100 which concerns on one Embodiment. FIG. 17 is a front view schematically showing a part of the impeller 4 in FIG. 16. FIG. 7 is a meridional view showing a configuration in which a valve 62 is provided in the first flow path 16. It is a meridional view which shows schematically a part of centrifugal compressor 2 in the turbocharger 100 which concerns on one Embodiment. FIG. 20 is a diagram showing a flow on a meridional surface at a small flow rate (hereinafter, referred to as “flow rate B”) in which the flow rate is smaller than the flow rate A in the centrifugal compressor 2 shown in FIG. 19. It is a meridional view which shows schematically a part of centrifugal compressor 2 in the turbocharger 100 which concerns on one Embodiment. It is a meridional view which shows schematically a part of centrifugal compressor 2 in the turbocharger 100 which concerns on one Embodiment. It is a meridional view which shows schematically a part of centrifugal compressor 2 in the turbocharger 100 which concerns on one Embodiment. It is a meridional view which shows schematically a part of centrifugal compressor 2 in the turbocharger 100 which concerns on one Embodiment. It is a figure for demonstrating the definition of the inner side edge 36 of the 1st front edge, and the definition of the downstream end 70 of the 2nd front edge 28. It is a figure which shows the streamline on the meridional surface in the vicinity of the maximum efficiency in the conventional centrifugal compressor. FIG. 20 is a view taken along the arrow AA in FIG. 19 and is a diagram showing streamlines between blades on a rotating flow surface. FIG. 21 is a diagram showing streamlines between blades on a rotating flow surface in the case where the flow rate decreases from the state shown in FIG. 20 and the flow enters with a collision angle. It is a figure which shows the streamline on a meridian surface when the flow volume decreases from the state shown in FIG. 19 and a flow inflows with a collision angle. It is a figure which shows the flow volume-pressure ratio characteristic curve C0 in the conventional centrifugal compressor.

Hereinafter, some embodiments of the present invention will be described with reference to the accompanying drawings. However, the dimensions, materials, shapes, relative arrangements, and the like of the components described as the embodiments or shown in the drawings are not intended to limit the scope of the present invention thereto, but are merely illustrative examples. Absent.
For example, the expression "relative or absolute" such as "in a certain direction", "along a certain direction", "parallel", "orthogonal", "center", "concentric", or "coaxial" is strict. In addition to representing such an arrangement, it also represents a state in which the components are relatively displaced by a tolerance or an angle or a distance at which the same function can be obtained.
For example, expressions such as "identical", "equal", and "homogeneous" that indicate that they are in the same state are not limited to a state in which they are exactly equal. It also represents the existing state.
For example, the representation of a shape such as a quadrangle or a cylinder does not only represent the shape of a quadrangle or a cylinder in a geometrically strict sense, but also an uneven portion or a chamfer within a range in which the same effect can be obtained. The shape including parts and the like is also shown.
On the other hand, the expressions “comprising”, “comprising”, “comprising”, “including”, or “having” one element are not exclusive expressions excluding the existence of other elements.

FIG. 1 is a meridional view schematically showing a part of a centrifugal compressor 2 in a turbocharger 100 according to an embodiment of the present invention. FIG. 2 is a front view schematically showing a part of the impeller 4 in the centrifugal compressor 2 according to the embodiment.

As shown in FIG. 1, the centrifugal compressor 2 includes an impeller 4 and an impeller casing 6 that houses the impeller 4. The impeller 4 is provided coaxially with a rotor of a turbine (not shown) that rotates by utilizing a flow of exhaust gas of the engine (not shown), and is rotationally driven by the rotor of the turbine to compress air supplied to the engine. Is configured. The air that has passed through the impeller 4 is supplied to the engine after passing through the diffuser 5 and a scroll chamber (not shown).

The impeller 4 includes a hub 8 and a plurality of blades 12 provided upright on an outer peripheral surface 10 of the hub 8. In the following, the radial direction of the impeller 4 is simply referred to as “radial direction”, the axial direction of the impeller 4 is simply referred to as “axial direction”, and the circumferential direction of the impeller 4 is simply referred to as “circumferential direction”. Further, hereinafter, the upstream side of the air flow in the axial direction is simply referred to as “upstream side”, and the downstream side of the air flow in the axial direction is simply referred to as “downstream side”.

The impeller casing 6 is provided inside the first shroud casing 14 and the first shroud casing 14, forms an annular first flow passage 16 between the impeller casing 6 and the second shroud casing 14, and 18, a second shroud casing 20 forming the first shroud 18, a column 52 provided at an upstream portion in the flow path length of the first flow path 16 so as to support the second shroud casing 20, and the first flow path 16 in the radial direction. An annular support portion 54 provided so as to support the column 52 from the outside. The second shroud casing 20 is provided on the upstream side of the impeller 4.

As shown in FIGS. 1 and 2, each of the wings 12 includes a trailing edge 22, a first leading edge 24, a first leading edge 26, a second leading edge 28, and a second leading edge 30.

As shown in FIG. 1, the first tip edge 24 extends in a curved shape from the tip 32 of the trailing edge 22 along the inner peripheral surface 66 of the first shroud casing 14. A slight gap is provided between the first tip edge 24 and the inner peripheral surface 66 of the first shroud casing 14.

The first front edge 26 is provided on the downstream side of the second shroud casing 20 so as to face the downstream end 48 of the second shroud casing 20 with a gap in the axial direction. The first front edge 26 extends inward in the radial direction from the upstream end 34 of the first tip edge 24, and connects the downstream end 70 of the second tip edge 28 and the upstream end 34 of the first tip edge 24. The first front edge 26 extends in the radial direction so as to face the first flow path 16, and is provided on the downstream side of the downstream end 48 of the second shroud casing 20 and adjacent to the downstream end 48. In the exemplary form shown in FIG. 1, the first leading edge 26 extends linearly along a direction orthogonal to the axial direction.

The second tip edge 28 extends along the inner peripheral surface 40 so as to face a portion of the inner peripheral surface 40 of the second shroud casing 20 in the vicinity of the downstream end 48 with a gap, and extends in the radial direction. It is provided inside the two-shroud casing 20. The second tip edge 28 extends linearly or in a curved shape from the inner end 36 of the first front edge 26 in the radial direction toward the upstream side, and is located upstream of the first tip edge 24. Further, since the first leading edge 26 is interposed between the first tip edge 24 and the second tip edge 28, the leading edge radius of the first tip edge 24 is smaller than the radius of the downstream end 70 of the second tip edge 28. Is large (that is, the distance between the downstream end 70 of the second tip edge 28 and the rotation axis O is larger than the distance between the upstream end 34 of the first tip edge 24 and the rotation axis O). In the illustrated exemplary embodiment, the second tip edge 28 extends linearly along the axial direction substantially parallel to an imaginary extension line to the upstream side of the first tip edge 24.

The length by which the second tip edge 28 and the second shroud casing 20 overlap in the axial direction is preferably at least half the radial length of the first front edge 26. Further, the downstream end 48 of the second shroud casing 20 is preferably located inside the center of the first front edge 26 in the radial direction. Further, the downstream end 48 of the second shroud casing 20 and the inner peripheral surface 40 of the second shroud casing 20 are formed by curved lines that face the second tip edge 28 and the first front edge 26 of the impeller 4 with a substantially constant gap therebetween. It may be done.

In addition, in the exemplary embodiment shown in FIG. 1 and the like, the first front edge 26 and the second tip edge 28 are configured to intersect each other at substantially right angles, and the inner end 36 and the second tip edge of the first front edge 26 are configured. The downstream ends 70 of 28 coincide. However, in an actual structure, the first front edge 26 and the second tip edge 28 may be connected by a smooth curve, as shown by the broken line in FIG. In this case, the intersection point between the virtual extension line of the linear first front edge 26 toward the inner side in the radial direction and the virtual extension line of the linear second front edge 28 toward the downstream side is defined as the first front edge. It is defined as an inner end 36 of 26 and a downstream end 70 of the second tip edge 28.

The second front edge 30 is located on the upstream side of the first front edge 26 in the axial direction, extends linearly or curvedly inward in the radial direction from the upstream end 38 of the second tip edge 28, and Connect to the outer peripheral surface 10. The second front edge 30 is provided so as to face the second flow path 18. In the illustrated exemplary form, the second leading edge 30 extends linearly along a direction orthogonal to the axial direction.

The second shroud casing 20 is configured in an annular shape, is provided on the inner side in the radial direction of the inner peripheral surface 66 of the first shroud casing 14, and is closer to the inner peripheral surface 66 of the first shroud casing 14 than the first front edge 26. The above-described first flow path 16 is formed between the inner wall surface 67 on the upstream side and the outer peripheral surface 42 of the second shroud casing 20. Further, the above-mentioned second flow path 18 is formed in a circular tube shape or an annular shape inside the inner peripheral surface 40 of the second shroud casing 20. The second shroud casing 20 has a blade shape (wing cross-sectional shape) on the meridian surface. The inner peripheral surface 40 of the second shroud casing 20 has a bell mouth shape whose radius increases toward the upstream side. The outer peripheral surface 42 of the second shroud casing 20 has a curved surface-shaped expanding surface 44 whose radius increases toward the upstream side, and the expanding surface 44 and the inside of the second shroud casing 20 on the upstream side of the expanding surface 44. A curved connecting surface 46 that connects the peripheral surface 40 is included. Further, the downstream end 48 of the second shroud casing 20 is located outside the second tip edge 28 in the radial direction, and is located downstream of the upstream end 38 of the second tip edge 28 in the axial direction.

The inner wall surface 67 of the first shroud casing 14 on the upstream side of the first front edge 26 has a bellmouth-shaped inlet, and the first flow passage 16 has a flow passage cross-sectional area that increases toward the downstream side. It has a narrowed throttle channel section 50. An inner wall surface 68 of the first shroud casing 14 on the downstream side of the first front edge 26 is curved outward in the radial direction toward the downstream side along the first tip edge 24. The support column 52 is installed near the upstream end of the bellmouth-shaped inlet in the first flow path 16, and has a wing shape in a cross section orthogonal to the radial direction.

In such a configuration, the flow velocity around the support column 52 that supports the second shroud casing 20 is slower than the flow velocity near the first front edge 26, and the cross-sectional shape of the support column 52 uses a small loss cross-sectional shape. The pressure loss due to the resistance of the column 52 itself can be reduced. In another embodiment, the support column 52 may be formed of a flat plate, a cylinder or the like having arcs at its front and rear edges in a cross section orthogonal to the radial direction of the impeller 4.

Further, as the flow rate decreases, the region in which the stable circulation flow Fc is generated expands, so that the second shroud casing 20 extends upstream with respect to the second front edge 30 to expand the circulation flow Fc. It is possible to prevent Fc from affecting the flow near the second front edge 30 and separating the flow near the second front edge 30.

Further, the length in which the upstream end of the second shroud casing 20 extends upstream of the second front edge 30 is the upstream end 34 of the first tip edge 24, the second front edge 30, and the hub 8. It is desirable that the distance be equal to or greater than the radial distance from the connection position. Accordingly, the radius of the upstream end of the second shroud casing 20 can be increased, the cross-sectional area of the first flow path 16 can be increased, and the flow velocity around the support column 52 can be sufficiently reduced. Therefore, the pressure loss due to the resistance of the support column 52 itself can be reduced, and the separation of the flow in the vicinity of the second front edge 30 due to the stable expansion of the circulation flow Fc can be effectively suppressed.

In the illustrated exemplary embodiment, the radius at the inlet position of the impeller 4 (the distance between the upstream end 34 of the first tip edge 24 and the rotation axis O of the impeller 4) at the outlet position of the impeller 4 (the trailing edge of the blade 12). A value obtained by dividing the distance 22 between the impeller 4 and the rotation axis O of the impeller 4 is about 1.1 to 1.3. Further, the angle θ formed by the outer peripheral surface 10 of the hub 8 with respect to the radial direction at the outlet position of the impeller 4 is set to 5°≦θ≦15°.

FIG. 3 is a diagram showing a flow on the meridian plane at a large flow rate (hereinafter, referred to as “flow rate A time”) near the maximum efficiency point in the centrifugal compressor 2. FIG. 4 is a diagram showing a flow on the meridian plane when the flow rate is smaller than the flow rate A in the centrifugal compressor 2 (hereinafter referred to as “flow rate B”). FIG. 5 shows the flow of the inter-blade flow surface in the vicinity of the first tip edge 24 when the flow rate B of the centrifugal compressor 2 is B (the arrow B-B in FIG. 4), and the second flow rate of the centrifugal compressor 2 when the flow rate B is B. It is a figure which overlaps and shows the flow (CC arrow line view in FIG. 4) of the flow surface between blades near the front edge 28.

When the flow rate A is in the vicinity of the maximum efficiency point shown in FIG. 3, the flow upstream of the impeller 4 has a first flow path 16 such that the axial flow velocity is substantially constant at each of the first leading edge 26 and the second leading edge 30. And flows into the second flow path 18. At a flow rate A in the vicinity of the maximum efficiency point, a collision-free flow (flow in the direction along the center line of the blade cross section, that is, the camber line) is formed with respect to the first leading edge 26 and the second leading edge 30, so that high efficiency is achieved. In addition, the centrifugal compressor 2 can be operated at high pressure.

On the other hand, the flow rate A near the maximum efficiency is the ratio of the flow passage cross-sectional area S2 of the second flow passage 18 to the sum of the flow passage cross-sectional area S1 of the first flow passage 16 and the flow passage cross-sectional area S2 of the second flow passage 18. When the flow rate B is reduced to about S2/(S1+S2), almost no flow flows into the first flow path 16 and only the second flow path 18 flows, as shown in FIG.

This phenomenon will be described below.
As shown in FIG. 4, in the centrifugal compressor 2 also, the above-mentioned “backflow of the meridian plane” occurs, but in the vicinity of the first tip edge 24 on the blade inlet side, the backflow toward the upstream side is followed by the inward radial direction. The flow is blocked by the outer peripheral surface 42 of the second shroud casing 20, and therefore does not flow further inward in the radial direction. As a result, the "meridian backflow" is restrained from developing and is confined in the region between the first leading edge 26 and the first tip edge 24, and becomes a stable circulating flow Fc.

Due to this stable circulation flow Fc, the flow passage is closed at the position of the first leading edge 26, so that the flow from the first flow passage 16 hardly flows into the impeller 4, and the flow from the second flow passage 18 only. Inflow to 4. Further, since a smooth boundary surface is formed at the boundary between the stable circulation flow Fc and the flow flowing into the impeller 4 from the second front edge 30, the flow passing through the second front edge 30 is smooth to the rear edge 22. Can flow to. As a result, the pressure drop and the loss increase due to the blade stall and backflow described with reference to FIGS. 21 and 22 and the like can be suppressed, and the pressure and the efficiency decrease due to the flow rate decrease can be suppressed.

Further, since the flow passage at the position of the first front edge 26 is closed as described above, the inlet area of the impeller 4 is substantially only the flow passage area corresponding to the radial range of the second front edge 30. Therefore, even if the flow rate A is decreased to the flow rate B, the decrease in the flow rate of the second flow path 18 is suppressed, and the flow without collision with the second front edge 30 can be maintained. That is, in FIG. 5, the collision angle of the flow with respect to the second leading edge 30 (the angle between the streamline flowing into the second leading edge 30 and the center line of the blade cross section, that is, the camber line) is maintained at approximately 0, and the flow thereof is maintained. Can reach the exit of the impeller 4 with high efficiency.

Further, as shown in FIG. 4, even if the inlet area of the impeller 4 is substantially only the flow passage cross-sectional area S2 corresponding to the position of the second front edge 30, the outlet area S3 of the impeller 4 does not change. The amount of deceleration and the amount of work from the inlet to the outlet of the impeller 4 increase, and the discharge pressure increases.

FIG. 6 is a diagram showing a flow rate-pressure ratio characteristic curve C1 of the centrifugal compressor 2. FIG. 7 is a diagram showing a flow rate-efficiency characteristic curve C2 of the centrifugal compressor 2.

As shown in FIG. 6, the flow rate-pressure ratio characteristic curve C1 of the centrifugal compressor 2 has a relatively large flow rate range WL, and therefore flows in both the first flow path 16 and the second flow path 18, so that FIG. It has substantially the same characteristics as the flow rate-pressure ratio characteristic curve C0 of the conventional centrifugal compressor described with reference to FIG.

Further, the flow rate-pressure ratio characteristic curve C1 of the centrifugal compressor 2 is a virtual flow chart including a choke flow rate and a surge flow rate in the second flow path 18 because the flow flows only in the second flow path 18 in the flow rate range WS that is relatively small. Has substantially the same characteristics as the flow rate-pressure ratio characteristic curve C3 (flow rate-pressure ratio characteristic curve in a virtual centrifugal compressor in which the flow flows from only the second flow path 18 into the impeller 4). Therefore, in the small flow rate range WS, the pressure ratio becomes higher than that of the conventional centrifugal compressor as shown by the arrow P1. Further, the surge flow rate of the second flow path 18 can be made smaller than the surge flow rate of the conventional centrifugal compressor, and the surge margin can be expanded as shown by the arrow P2.

As described above, the flow rate-pressure ratio characteristic curve C1 of the centrifugal compressor 2 is a virtual flow rate including the flow rate-pressure ratio characteristic curve C0 of the conventional centrifugal compressor and the choke flow rate and the surge flow rate of the second flow path 18. -It has a characteristic that it is combined with the pressure ratio characteristic curve C3.

Further, as shown in FIG. 7, the flow rate-efficiency characteristic curve C2 of the centrifugal compressor 2 similarly shows the flow rate-efficiency characteristic curve C4 of the conventional centrifugal compressor, the choke flow rate and the surge flow rate of the second flow path 18. It has a characteristic as if it were combined with a virtual efficiency-pressure ratio characteristic curve C5.

For this reason, in the flow rate-efficiency characteristic curve C2 of the centrifugal compressor 2, the flow rate-efficiency characteristic curve C2 of the conventional centrifugal compressor is made to correspond to the flow rate-pressure ratio characteristic curve C5 of the virtual centrifugal compressor in which the flow flows only in the second flow path 18. An efficiency peak point occurs on the small flow rate side, and as shown by an arrow P3, high efficiency can be obtained up to a small flow rate as compared with the flow rate-efficiency characteristic curve C4 of the conventional centrifugal compressor.

As described above, according to the centrifugal compressor 2, it is possible to increase the surge margin as compared with the conventional centrifugal compressor, increase the pressure ratio in the small flow rate region and improve the efficiency in the small flow rate region. It becomes possible to do. Further, as compared with the centrifugal compressor according to Patent Document 1, there is no need to provide a resistor having a complicated structure including a perforated plate or the like. Therefore, it is possible to suppress a decrease in pressure and efficiency due to a decrease in flow rate with a simple configuration.

FIG. 8 is a flow rate-pressure ratio characteristic curve C1a, C1b, C1c of the centrifugal compressor 2 for each engine speed in the turbocharger 100, and a corresponding flow rate-pressure ratio of the conventional centrifugal compressor for each engine speed corresponding thereto. It is a figure which shows characteristic curve C0a, C0b, C0c. In FIG. 8, the corresponding engine speed increases in the order of the characteristic curves C1a, C1b, C1c.

According to the centrifugal compressor 2, as shown by an arrow P4, it is possible to reach a high pressure ratio at a low rotation speed. Further, since the surge flow rate can be made smaller than that of the conventional centrifugal compressor, the surge margin can be expanded when operating on the same engine operating line L0 as the conventional one. Further, when the surge margin is set as in the conventional case, as shown by an arrow P5, it is possible to widen the flow rate region where the pressure is constant, and it is possible to cope with a wide flow rate range required by the engine.

FIG. 9 is a diagram showing a relationship between flow rate-pressure ratio characteristic curves C0a, C0b, C0c and engine operating characteristics in a conventional centrifugal compressor. FIG. 10 is a diagram showing the relationship between the flow rate-pressure ratio characteristic curves C1a, C1b, C1c and the engine operating characteristics in the centrifugal compressor 2.

As shown in FIG. 9, conventionally, when the centrifugal compressor is operated on the engine operating line L0, the operation is performed in the region near the surge margin at a small flow rate corresponding to the time of start-up or acceleration from a low speed, so that the efficiency peak is reached. However, it had to be used in areas where efficiency was low.

On the other hand, according to the centrifugal compressor 2, as shown in FIG. 10, a high-efficiency region is generated in a small flow rate region, so that a high-efficiency operating point can be used even at a small flow rate, and a conventional pressure can be achieved at a low rotation speed. it can. Therefore, the flow rate and the pressure ratio can be quickly increased during acceleration, and the response can be improved with respect to the increase in the flow rate and the pressure ratio.

Further, since the leading edge portion of the blade 12 has a notched shape composed of the first leading edge 26 and the second leading edge 28, the weight of the leading end portion of the blade 12 on the inlet side is reduced, so that the vibration of the blade 12 is reduced. It is possible to increase the eigenvalue. Therefore, the strength against vibration can be increased, the centrifugal force acting on the blade 12 can be reduced, and the stress generated in the blade 12 can be reduced. Further, when the strength and the vibration eigenvalue are set to the conventional values, the blade 12 can be made thin, so that the peak efficiency can be improved.

In one embodiment, as shown in FIGS. 11 and 12, the second front edge 30 of the centrifugal compressor 2 has the outer end 39 (38) of the second front edge 30 in the radial direction being the second front edge in the radial direction. It may be inclined with respect to the radial direction so as to be located upstream of the inner end 56 of 30.

Since the second leading edge 30 is inclined in this way, the pressure surface of the blade 12 near the second leading edge 30 faces the outlet direction (radially outward) of the impeller 4. Since a force in a direction perpendicular to the pressure surface of the blade 12 acts on the flow flowing into the impeller 4, the action of diverting the flow from the axial direction to the radially outer side becomes large.

Therefore, it is possible to enhance the action of turning the flow flowing into the impeller 4 from the second flow path 18 outward in the radial direction in the vicinity of the second front edge 30 as shown by the arrow P6. As a result, the expansion of the boundary layer in the vicinity of the inner peripheral surface 58 of the first shroud casing 14 can be suppressed, and the distribution of the flow velocity on the tip side at the trailing edge 22 of the blade 12 can be made close to a uniform distribution.

As a result, the impeller efficiency is improved and the pressure recovery rate of the diffuser 5 is improved, so that the centrifugal compressor 2 can be made highly efficient.

In one embodiment, as shown in FIGS. 13 and 14, the first leading edge 26 has a radially outer edge 35 (34) of the first leading edge 26 that is radially inner edge 36 of the first leading edge 26. It may be inclined with respect to the radial direction so as to be located on the more upstream side.

The inclination of the first front edge 26 in this way indicates the flow flowing from the first flow path 16 into the impeller 4 for the same reason as when the second front edge 30 is inclined, as indicated by an arrow P7. As described above, it is possible to enhance the action of turning outward in the radial direction in the vicinity of the first front edge 26. As a result, the expansion of the boundary layer in the vicinity of the inner peripheral surface 58 of the first shroud casing 14 can be suppressed, and the distribution of the flow velocity on the tip side at the trailing edge 22 of the blade 12 can be made close to a uniform distribution.

As a result, the impeller efficiency is improved and the pressure recovery rate of the diffuser 5 is improved, so that the centrifugal compressor 2 can be made highly efficient.

The first front edge 26 is arranged in the radial direction such that the outer end 35 (34) of the first front edge 26 in the radial direction is located downstream of the inner end 36 of the first front edge 26 in the radial direction. May be inclined. As a result, the vibration strength of the blade 12 can be increased.

In one embodiment, as shown in FIG. 15, the second shroud casing 20, the column 52, and the annular support portion 54 are integrated (configured as one component), and the annular support portion 54 is The first shroud casing 14 may be detachably attached to the first shroud casing 14. In this case, the annular support portion 54 may be attached to the first shroud casing 14 by, for example, a bolt (not shown).

In this way, if the second shroud casing 20 is used as another component of the first shroud casing 14 so that the second shroud casing 20 and the first shroud casing 14 can be assembled coaxially, the second shroud casing 20 After the height of the tip edge 28 is set by casting or adjusted by machining, a component constituted by the second shroud casing 20, the support column 52 and the annular support portion 54 can be manufactured according to the height. .. Therefore, the degree of freedom in the performance design of the centrifugal compressor 2 and the degree of freedom in the structural design can be increased.

The present invention is not limited to the above-described embodiment, and includes a modified form of the above-described embodiment and a combination of these forms as appropriate. In the embodiment described below, configurations having the same functions as those already described are given the same reference numerals and the description thereof will be omitted.

For example, in the form shown in FIG. 1, the second shroud casing 20 and the second tip edge 28 do not have to overlap in the axial direction. The downstream end 48 of the second shroud casing 20 may be radially outside or inside the second tip edge 28. When the downstream end 48 of the second shroud casing 20 is radially inward of the second tip edge 28, the downstream end 48 of the second shroud casing 20 is located upstream of the second front edge 30.

For example, as shown in FIGS. 16 and 17, the impeller 4 may include a splitter blade 60 provided between adjacent blades 12 in addition to the blade 12 (main blade) described above. The splitter blade 60 is shorter than the blade 12, and the leading edge position of the splitter blade 60 is set downstream of the leading edge position of the blade 12. In this case, the present invention can be applied to either or both of the blade 12 as the main blade and the splitter blade 60, and in particular, by being applied to the blade 12 as the main blade, the pressure and The decrease in efficiency can be effectively suppressed.

Further, in one embodiment, as shown in FIG. 18, the centrifugal compressor 2 may further include a valve 62 configured to be able to open and close the first flow path 16. In this case, the valve 62 may be configured to be opened/closed by the actuator 64 or may be configured to be opened/closed manually. With such a configuration, the valve 62 is opened when the flow rate is large, and the valve 62 is closed when the flow rate is small, so that it is possible to suppress a decrease in pressure and efficiency due to a decrease in flow rate.

Further, in one embodiment, as shown in FIG. 19, for example, the centrifugal compressor 2 may not include the second shroud casing 20.

In this case, when the flow rate decreases, the flow in the vicinity of the first leading edge 24 side of the first leading edge 26 separates from the blade surface, and the above-mentioned “backflow of the meridional surface” occurs and the radial direction of the impeller 4 is generated. A local inward flow due to backflow occurs toward the inner second tip edge 28 (see FIG. 20 ). On the other hand, the flow passing through the second front edge 30 receives a force in the rotation direction of the impeller 4 by the blades 12 to increase the speed in the circumferential direction, and the centrifugal force due to the speed in the circumferential direction causes the flow to flow downstream. Flows radially outward while flowing. The radially inward flow after backflowing upstream in the vicinity of the first tip edge 24 side portion of the first front edge 26 is the radially outward flow Fo after passing through the second front edge 30. Therefore, it does not flow further inward in the radial direction because the movement is hindered by. As a result, the "meridian backflow" is restrained from developing and is confined in the region near the first leading edge 26 and the first tip edge 24, and becomes a stable circulating flow Fc.

Due to the stable circulation flow Fc, the flow path is blocked at the position of the first leading edge 26, so that the flow from the first leading edge 26 hardly flows into the impeller 4, and the flow mainly from the second leading edge 30. Flows into the impeller 4. Further, since a smooth boundary surface is formed at the boundary between the stable circulation flow Fc and the flow flowing into the impeller 4 from the second front edge 30, the flow passing through the second front edge 30 is smooth to the rear edge 22. Can flow to. As a result, the decrease in pressure and the increase in loss due to the stall of the blade 12 and the backflow described above are suppressed, and the decrease in pressure and efficiency due to the decrease in flow rate can be suppressed. Further, as compared with the centrifugal compressor according to Patent Document 1, it is not necessary to provide a resistor having a complicated structure, such as a perforated plate, so that the above effect can be obtained with a simple structure.

In one embodiment, for example, in FIG. 20, the second leading edge 28 forms an angle within 10 degrees radially outward or within 20 degrees radially inward with respect to an imaginary extension line that extends f24 upstream. It is desirable that it is formed or extends parallel to the virtual extension line.

In one embodiment, as shown in, for example, FIG. 21, the second shroud casing 20 has an annular structure, is provided inside the inner peripheral surface 66 of the first shroud casing 14, and is closer to the first front edge 26. It is arranged on the upstream side and radially outside of the second tip edge 28. The second shroud casing 20 forms the first flow passage 16 between the inner peripheral surface 66 of the first shroud casing 14 and the wall surface 67 on the upstream side of the first front edge 26, and at the same time, the second shroud casing 20. The second flow path 18 is formed inside the inner peripheral surface of the.

In the configuration shown in FIG. 21, the second shroud casing 20 is a plate that smoothly guides the flow of the first flow path 16 and the flow of the second flow path 18 to the first front edge 26 and the second front edge 30, respectively. It has a meridional cross section. Further, the column 52 that supports the second shroud casing 20 has a plate-shaped or blade-shaped cross-sectional shape, and the center line of the cross-sectional shape is substantially rotated so as not to obstruct the flow of the first flow path 16. It is formed along the axial direction. The column 52 extends inward in the radial direction from the wall surface 67 of the first shroud casing 14 and is connected to the outer peripheral surface of the second shroud casing 20. In addition, regarding such a form, the second shroud casing 20 may have a blade-shaped cross-sectional shape.

In addition, in the form shown in FIG. 21, the center line of the cross-sectional shape of the support column 52 is arranged in the rotational axis direction so that the incident becomes almost 0 when the flow of the first flow path 16 flows into the first front edge 26. It may have an inclination with respect to it. Further, it is desirable that the downstream end 48 of the second shroud casing 20 be disposed downstream of the second front edge 30 and upstream of the first front edge 26 with a gap from the second tip edge 28.

In one embodiment, as shown in FIG. 22, for example, an annular second leading edge shroud 72 fixed to the second leading edge 28 is provided. In the illustrated form, the second leading edge shroud 72 has a longer axial length than the second leading edge 28 so as to project upstream of the second front edge 30. The second tip edge shroud 72 has a meridional cross-sectional shape that is, for example, a rectangle or a wing shape.

In one embodiment, for example, as shown in FIG. 23, an annular second tip edge shroud 72 fixed to the second tip edge 28 and a second shroud installed axially upstream of the second tip edge shroud 72. And a casing 20. In the embodiment shown in FIG. 23, the second shroud casing 20 and the second tip edge shroud 72 have smooth wall surfaces on the first flow path 16 side and smooth wall surfaces on the second flow path 18 side. It is configured to be continuous. That is, the outer peripheral surface 42 of the second shroud casing 20 and the outer peripheral surface 82 of the second tip edge shroud 72 are smoothly connected through a gap, and the inner peripheral surface 40 of the second shroud casing 20 and the second tip edge shroud 72 are connected. It is configured to be smoothly connected to the inner peripheral surface 84 via a gap. Further, the downstream end portion 74 of the second shroud casing 20 and the upstream end portion 76 of the second tip edge shroud 72 are prevented from interfering with the rotation of the impeller 4 through a slight gap in the axial direction and the radial direction. It is arranged opposite. In the illustrated embodiment, the downstream end portion 74 of the second shroud casing 20 has a cutout shape on the inner peripheral side thereof, and the upstream end portion 76 of the second tip edge shroud 72 has a space within the cutout shape. Have invaded. Further, the upstream end 78 of the second leading edge shroud 72 is located upstream of the downstream end 48 of the second shroud casing 20.

In one embodiment, for example, as shown in FIG. 24, an annular first leading edge shroud 80 secured to the first leading edge 24 is provided. The inner surface of the cross section of the first tip edge shroud 80 has the meridional surface shape of the first tip edge 24.

2 Centrifugal Compressor 4 Impeller 6 Impeller Casing 8 Hub 10 Outer Surface 12 Blade 14 First Shroud Casing 16 First Flow Path 18 Second Flow Path 20 Second Shroud Casing 22 Trailing Edge 24 First Tip Edge 26 First Leading Edge 28 Second tip edge 30 Second front edge 32 Tip 34 Upstream end 35 Outer end 36 Outer end 36 Inner end 38 Upstream end 39 Outer end 40 Inner peripheral surface 42 Outer peripheral surface 44 Expanded surface 46 Connection surface 48 Downstream end 50 Throttling flow path portion 52 Strut 54 annular support portion 56 inner end 58 inner peripheral surface 60 splitter blade 62 valve 64 actuator 66 inner peripheral surface 67 wall surface 68 wall surface 70 downstream end 72 second tip edge shroud 74 downstream end 76 upstream end 78 upstream end 80 first Tip edge shroud 82 Outer peripheral surface 84 Inner peripheral surface 100 Turbocharger

Claims (13)

A centrifugal compressor comprising an impeller and a first shroud casing containing the impeller,
The impeller includes a hub and a plurality of blades erected on the outer peripheral surface of the hub,
At least one wing of the plurality of wings,
A first tip edge extending along the first shroud casing;
A first front edge extending inward in the radial direction of the impeller from an upstream end of the first tip edge;
A second tip edge extending from the inner end of the first front edge in the radial direction to the upstream side;
A second front edge extending inward in the radial direction from an upstream end of the second tip edge;
Including
Further comprising a second shroud casing provided inside the inner peripheral surface of the first shroud casing,
The second shroud casing is located on the upstream side of the first front edge, forms a first flow path between the second shroud casing and the inner peripheral surface of the first shroud casing, and forms an inner peripheral surface of the second shroud casing. Configured to form a second flow path inside,
The inner peripheral surface of the second shroud casing has a parallel surface extending from the downstream end of the second shroud casing toward the upstream side in parallel to the rotation axis of the impeller, and an upstream side of the parallel surface. Centrifugal compression, including a bell mouth surface having a bell mouth shape whose radius increases toward the side, and a curved connecting surface connecting the upstream end of the bell mouth surface and the outer peripheral surface of the second shroud casing. Machine. A centrifugal compressor comprising an impeller and a first shroud casing containing the impeller,
The impeller includes a hub and a plurality of blades erected on the outer peripheral surface of the hub,
At least one wing of the plurality of wings,
A first tip edge extending along the first shroud casing;
A first front edge extending inward in the radial direction of the impeller from an upstream end of the first tip edge;
A second tip edge extending from the inner end of the first front edge in the radial direction to the upstream side;
A second front edge extending inward in the radial direction from an upstream end of the second tip edge;
Including
Further comprising a second shroud casing provided inside the inner peripheral surface of the first shroud casing,
The second shroud casing is located on the upstream side of the first front edge, forms a first flow path between the second shroud casing and the inner peripheral surface of the first shroud casing, and forms an inner peripheral surface of the second shroud casing. Configured to form a second flow path inside,
The outer peripheral surface of the second shroud casing is a curved surface-shaped expanded surface whose radius increases from the downstream end of the second shroud casing toward the upstream side toward the upstream side, and is concave toward the inner side in the radial direction. A centrifugal compressor, comprising: a diameter-expanded surface formed so as to have a curved surface; The centrifugal compressor according to claim 1 or 2, wherein a downstream end of the second shroud casing is located outside the second tip edge in the radial direction. The centrifugal compressor according to any one of claims 1 to 3, wherein a downstream end of the second shroud casing is located downstream of an upstream end of the second tip edge in the axial direction of the impeller. The lengths of the upstream end of the second shroud casing and the second front edge along the extending direction of the rotation axis of the impeller are the upstream end of the first tip edge, the second front edge, and the hub. Greater than the length along the radial direction with the connection position with
The centrifugal compressor according to any one of claims 1 to 4. Further comprising a valve configured to open and close the first flow path by advancing and retracting along an extending direction of a rotation axis of the impeller,
The centrifugal compressor according to any one of claims 1 to 5. The centrifugal compressor according to any one of claims 1 to 6, wherein the second shroud casing has a blade shape on a meridional surface of the impeller. An inner peripheral surface of the second shroud casing has a parallel surface extending from a downstream end of the second shroud casing toward an upstream side in parallel to a rotation axis of the impeller, and an upstream end of the parallel surface. A bell mouth surface having a bell mouth shape whose radius increases toward the side, and a curved connecting surface connecting the upstream end of the bell mouth surface and the outer peripheral surface of the second shroud casing,
The centrifugal compressor according to claim 2. The outer peripheral surface of the second shroud casing is a curved surface-shaped expanded surface whose radius increases from the downstream end of the second shroud casing toward the upstream side toward the upstream side, and is concave toward the inner side in the radial direction. And a curved connecting surface that connects the upstream end of the expanded surface and the inner peripheral surface of the second shroud casing.
The centrifugal compressor according to claim 1. The centrifugal compressor according to any one of claims 1 to 9, wherein the first flow passage includes a throttle flow passage portion having a flow passage cross-sectional area that decreases toward a downstream side. Further comprising a column provided in the first flow path so as to support the second shroud casing,
The centrifugal compressor according to any one of claims 1 to 10, wherein the support column has a blade shape in a cross section orthogonal to a radial direction of the impeller. A pillar provided in the first flow path so as to support the second shroud casing,
An annular support portion provided to support the pillar from the outside of the first flow path in the radial direction,
Including
The second shroud casing, the pillar, and the annular support portion are integrated,
The centrifugal compressor according to any one of claims 1 to 11, wherein the annular support portion is configured to be attachable to and detachable from the first shroud casing. A turbocharger comprising the centrifugal compressor according to claim 1.

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