US20110091323A1

US20110091323A1 - Compressor housing for turbocharger

Info

Publication number: US20110091323A1
Application number: US12/999,931
Authority: US
Inventors: Atsushi Koike; Kiyokazu Iizuka; Fumie Matsuhashi; Takahiro Kobayashi; Nobuo Takei; Yukio Takahashi
Original assignee: IHI Corp
Current assignee: IHI Corp
Priority date: 2008-06-17
Filing date: 2008-06-17
Publication date: 2011-04-21
Also published as: EP2290205A1; KR20100119565A; CN102066717A; WO2009153854A1

Abstract

A compressor impeller 2 has a plurality of long blades 2 a and short blades 2 b alternately arranged in a circumferential direction, and an inner surface of a compressor housing 10 is provided with an annular groove 12 which surrounds vicinities of leading edge tip portions of the short blades 2 b in a circumferential direction and is concave outward so as not to communicate with a suction port of a compressor.

Description

BACKGROUND OF THE INVENTION

1. Field of the Invention
The present invention relates to a compressor housing for a turbocharger capable of reducing abnormal noise (high-frequency noise).
2. Description of the Related Art
A supercharging operation means that air or air-fuel mixture supplied to a cylinder of an internal combustion engine is compressed in advance, and a compressor used for the supercharging operation is called a supercharger. In addition, a supercharger for performing the supercharging operation using exhaust gas of the internal combustion engine is simply called an exhaust turbine supercharger or a turbocharger.
FIG. 1 is an overall configuration diagram showing an example of a conventional turbocharger. In the drawing, the turbocharger includes a turbine rotor shaft 51, a compressor impeller 52, a bearing housing 53, a turbine housing 54, a compressor housing 55 a, and a seal plate 55 b.
The bearing housing 53, the turbine housing 54, the compressor housing 55 a, and the seal plate 55 b are connected to each other in an order shown in the drawing. In addition, the turbine rotor shaft 51 is configured by integrally forming a turbine impeller 51 a with a rotor shaft 51 b by welding, and is rotationally supported by the bearing housing 53 so as to be coaxially connected to the compressor impeller 52.
With such a configuration, the turbine impeller 51 a is rotationally driven using the exhaust gas of the internal combustion engine, and the rotation force is transmitted to the compressor impeller 52 via the rotor shaft 51 b to be rotationally driven so that air (or air-fuel mixture) is compressed and supplied to the internal combustion engine. Accordingly, it is possible to remarkably improve performance of the internal combustion engine.
FIG. 2 is a performance characteristic diagram showing an example of a compressor constituting the turbocharger. As shown in the drawing, in a general compressor, as the rpm N becomes high, a pressure ratio becomes high and a flow rate becomes large. In the drawing, the one-point dashed line indicates a surge line. When the flow rate arrives at the surge line at the same rpm, a surge is generated in a blade surface of the compressor impeller. As a result, the compressor hardly performs a compression operation and violent surge noise (intermittence noise) is observed.
Accordingly, as shown in the drawing, an operation line of the engine having the turbocharger is set so as to be sufficiently away from the surge line.
In addition, in the past, as means for shifting the surge line to a side of a small flow rate, there is a conventional compressor housing provided with a circulation passage for increasing an intake air amount in appearance, the circulation passage being used to communicate a suction port of the compressor with a part in the course of a compression passage of the compressor.
The compressor housing provided with the circulation passage is disclosed in, for example, Patent Documents 1 and 2.
In ‘Turbocharger Provided With Sliding Member’ disclosed in Patent Document 1, an object is to obtain high compression efficiency without deteriorating durability in a turbocharger configured to prevent surging, choking, etc.
As shown in FIG. 3, the turbocharger provided with the sliding member includes a new air passage 63 which guides fresh air to a compressor impeller in a compressor housing; a first air intake and exhaust port 68 which is formed on a part of a compressor housing wall facing to the compressor impeller; a second air intake and exhaust port 69 which faces the more upstream fresh air passage than the compressor impeller; and a bypass passage 60 which communicates the first and second air intake and exhaust ports with each other, wherein a sliding member 65 is attached to at least a part of the compressor housing wall facing the blade edge of the compressor impeller.
In ‘Inside of Compressor and Turbine and Related Improvement’ disclosed in Patent Document 2, an object is to retain breakage fragments of a compressor wheel (impeller) inside a diffuser flange by expectedly destructing the diffuser flange when the wheel is broken, to prevent serious accidents caused by the breakage fragments.
As shown in FIG. 4, a centrifugal compressor 71 includes a compressor housing 73 and a compressor wheel 74 mounted in the housing and having a compressor blade 75. The compressor housing 73 includes a cover plate 76 and a diffuser flange 79 fixed to both the cover plate 76 and a bearing housing. The diffuser flange 79 includes an outer edge portion attached to a cover member and a radial inside portion attached to the bearing housing. In the diffuser, a brittle groove portion is defined at a position halfway between the outer edge portion and the radial inside portion and thereby the expected breaking of the diffuser flange is made possible when the compressor wheel is broken.
Patent Document 1
Japanese published unexamined application No.11-173153, ‘Turbocharger Provided With Sliding Member’
Patent Document 2
Japanese published unexamined application No.11-190297, ‘Inside of Compressor and Turbine and Related Improvement’
In the above-described performance characteristic diagram, the dashed line indicates a pre-surge line, and when a flow rate becomes smaller than the pre-surge line at the same rpm, the abnormal noise (high-frequency noise) may be generated. The high-frequency noise may arrive at a sound pressure level of 90 dB or more, but does not deteriorate performance of the compressor. However, since the abnormal noise contributes to one of noise sources in an automobile requiring a silent state, it is necessary to prepare a countermeasure for the abnormal noise.
When the flow rate becomes small at the same rpm, the abnormal noise (high-frequency noise) is continuously generated up to the surge point. The abnormal noise is not generated in a surge state, but instead surge noise (intermittence noise) is observed.
In the past, since a generation cause of the abnormal noise (high-frequency noise) is not clear, in general, a sound isolator has been used in order to isolate the noise generated from the turbocharger, but a problem arises in that it takes a cost to perform the sound isolation.
The abnormal noise (high-frequency noise) can be reduced or removed by shifting the surge line to the side of the small flow rate by using the compressor housing provided with the above-described circulation passage, but in the compressor housing provided with the circulation passage, a problem arises in that a structure is complex and a manufacture cost is more expensive than that of the compressor housing without the circulation passage.

SUMMARY OF THE INVENTION

The present invention is contrived to solve the above-described problems. That is, an object of the invention is to provide a compressor housing for a turbocharger capable of remarkably reducing or removing abnormal noise (high-frequency noise) generated from a pre-surge line to a surge line without a sound isolator or a circulation passage.
According to an aspect of the invention, there is provided a compressor housing for a turbocharger for rotationally driving a turbine impeller using exhaust gas of an internal combustion engine and transmitting the rotation force to a compressor impeller to be rotationally driven so that air or air-fuel mixture is compressed and supplied to the internal combustion engine, wherein the compressor impeller has a plurality of long blades and short blades alternately arranged in a circumferential direction, and wherein an inner surface of the compressor housing is provided with an annular groove which surrounds vicinities of leading edge tip portions of the short blades in a circumferential direction and is concave outward so as not to communicate with a suction port of a compressor.
According to a preferred embodiment of the invention, an axial center of the annular groove is located within 5 mm from the leading edge tip portions of the short blades in an axial direction on an upstream or downstream side thereof, an axial groove width is no less than 2.5 mm and no more than 10 mm, and a maximum diameter of the annular groove is less than 1.2 times a diameter of each leading edge tip portion of the short blades.
The annular groove extends outward from the inner surface of the compressor housing so as to be perpendicular or inclined with respect to a rotary shaft of a compressor.
The inventors of the invention have independently examined and found out that the abnormal noise is generated by a rotating stall of the impeller and peeling becomes large to thereby contact with the short blade. The invention is based on the above-described new viewpoints.
That is, according to the above-described configuration of the invention, since the inner surface of the compressor housing is provided with the annular groove which surrounds vicinities of the leading edge tip portions of the short blades in a circumferential direction and is concave outward so as not to communicate with the suction port of the compressor, a sectional area of a flow passage in the annular groove is suddenly enlarged. Accordingly, it is possible to silence the noise.
In addition, the silencing advantage is shown in the embodiments described below.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is an overall configuration diagram showing an example of a conventional turbocharger.

FIG. 2 is a performance characteristic diagram of a compressor constituting the turbocharger.

FIG. 3 is a schematic diagram showing ‘Turbocharger Provided With Sliding Member’ disclosed in Japanese published unexamined application No.11-173153.

FIG. 4 is a schematic diagram showing an apparatus disclosed in Japanese published unexamined application No.11-190297.

FIG. 5 is an overall configuration diagram showing a turbocharger having a compressor housing according to the invention.

FIG. 6 is a sectional diagram showing the compressor housing in FIG. 5.

FIG. 7 is a perspective diagram showing a compressor impeller.

FIG. 8A is a diagram showing the compressor housing according to a first embodiment of the invention.

FIG. 8B is a diagram showing the conventional compressor housing provided with a circulation passage.

FIG. 9A is a test result of the conventional compressor housing, where a turbo rpm is 160,000 rpm and a flow rate is 6 m³/min.

FIG. 9B is a test result of the conventional compressor housing, where a turbo rpm is 160,000 rpm and a flow rate is 5 m³/min.

FIG. 9C is a test result of the conventional compressor housing, where a turbo rpm is 160,000 rpm and a flow rate is 4.3 m³/min.

FIG. 10A is a test result of the compressor housing according to the invention, where a turbo rpm is 160,000 rpm and a flow rate is 6 m³/min.

FIG. 10B is a test result of the compressor housing according to the invention, where a turbo rpm is 160,000 rpm and a flow rate is 5 m³/min.

FIG. 10C is a test result of the compressor housing according to the invention, where a turbo rpm is 160,000 rpm and a flow rate is 4.3 m3/min.

FIG. 11A is a diagram showing an original compressor housing.

FIG. 11B is a diagram showing a second compressor housing according to a second embodiment of the invention.

FIG. 11C is a diagram showing a third compressor housing according to the second embodiment of the invention.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Hereinafter, preferred embodiments of the invention will be described with reference to the accompanying drawings. In the respective drawings, the same reference numerals are given to the same components, and the repetitive description will be omitted.
FIG. 5 is an overall configuration diagram showing a turbocharger having a compressor housing according to the invention.
The turbocharger rotationally drives a turbine impeller 1 by use of exhaust gas of an internal combustion engine, and a rotation force thereof is transmitted to a compressor impeller 2 to rotationally drive the compressor impeller 2 so that air or air-fuel mixture is compressed and supplied to the internal combustion engine.
FIG. 6 is an enlarged sectional diagram showing the compressor housing in FIG. 5, and FIG. 7 is a perspective diagram showing the compressor impeller 2.
As shown in FIG. 7, in this invention, the compressor impeller 2 includes a plurality of long blades 2 a and short blades 2 b alternately arranged in a circumferential direction.
As shown in FIG. 6, a compressor housing 10 according to the invention has an inner surface in which an annular groove 12 is formed. The annular groove 12 surrounds vicinities of the leading edge tip portions of the short blades 2 b of the compressor impeller 2 and is concave outward so as not to communicate with a suction port of the compressor.
An axial center a of the annular groove 12 is located within 5 mm from the leading edge tip portions of the short blades 2 b in an axial direction on an upstream or downstream side thereof. An axial groove width b of the annular groove 12 is no less than 2.5 mm and no more than 10 mm. A maximum diameter d of the annular groove 12 is desirably less than 1.2 times a diameter of each leading edge tip portion of the short blades 2 b.
In this example, the annular groove 12 extends outward from the inner surface of the compressor housing 10 so as to be perpendicular with respect to a rotary shaft of the compressor, but may extend so as to be inclined thereto.

First Embodiment

In the turbocharger mounted to an engine for an automobile, the above-described abnormal noise (high-frequency noise) was observed during an engine test. For this reason, a component performance test of a turbo unit was carried out. Even in this component performance test, the abnormal noise was observed at a turbo rpm of 160,000 rpm and 180,000 rpm, and the generation frequency was about 2.3 kHz equivalent to an engine.
The inventors of the invention have independently examined and found out that the abnormal noise is generated by a rotating stall of the impeller. That is, as shown in FIG. 7 in which a flow state of a compressed fluid is schematically illustrated, peeling which is generated in the vicinity of an inlet or a tip of the long blade becomes large to thereby contact with the short blade at a position indicated by a dashed line in the drawing.
The reason is because an abnormal noise frequency does not depend on a first rotation, an abnormal noise start point is identical with a pre-surge point (a pressure variation start point of the inlet), the abnormal noise is continuously generated in an unstable region of the compressor (a pressure characteristic in which a pressure increases rightward), and the compressor housing is vibrated.
In addition, the inventors of the invention have considered that a high abnormal noise generation frequency of 2.3 kHZ is generated by a characteristic of the impeller, that is, generated by the fact that the number of stall cells (the number of stall blades) is large.
On the basis of the above-described new viewpoints, the inventors of the invention have prepared two types of compressor housings for delaying a stall and have carried out the component performance test of the turbo unit.
FIG. 8A is a diagram showing the compressor housing according to a first embodiment of the invention. FIG. 8A shows the compressor housing according to the invention and FIG. 8B shows the conventional compressor housing provided with a circulation passage.
The compressor housing 10 according to the invention is configured by additionally forming the annular groove 12 in an inner surface of the compressor housing of the turbocharger in which the abnormal noise (high-frequency noise) was observed.
In this example, an axial center of the annular groove 12 is located at a position shifted by 4 mm from the leading edge tip portions of the short blades 2 b in an axial direction on the upstream side thereof, an axial groove width b of the annular groove 12 is 2.5 mm, and a depth c of the annular groove 12 is 4 mm. In addition, in this example, the annular groove 12 extends outward from the inner surface of the compressor housing 10 so as to be perpendicular with respect to the rotary shaft of the compressor.
Meanwhile, a shape of the inner surface of the conventional compressor housing provided with the circulation passage is identical with that of the compressor housing of the turbocharger in which the abnormal noise (high-frequency noise) was observed, but its molding is newly manufactured. The circulation passage of the compressor housing communicates a suction port of the compressor with the same position as the position of the compressor housing 10 of the present invention. A groove width b′ of the circulation passage is 2.5 mm and a groove width e of an outlet port is 6 mm.
FIGS. 9A, 9B, and 9C are test results of the conventional compressor housing. FIGS. 9A, 9B, and 9C show cases where a turbo rpm is 160,000 rpm, and flow rates are about 6, 5, and 4.3 m³/min in order of the diagrams. In each of FIGS. 9A, 9B, and 9C, the left side indicates a noise measurement value and the right side indicates a pressure variation measurement value.
In these diagrams, at positions when a turbo rpm is 160,000 rpm, a flow rate is about 6 m³/min, and a frequency is about 2.3 kHz, large peaks are generated in the noise and the pressure variation, which corresponds to the above-described abnormal noise (high-frequency noise).
FIGS. 10A, 10B, and 10C are test results of the compressor housing according to the invention. FIGS. 10A, 10B, and 10C show cases where a turbo rpm is 160,000 rpm, and flow rates are about 6, 5, and 4.3 m³/min in order of the diagrams. In each of FIGS. 10A, 10B, and 10C the left side indicates a noise measurement value and the right side indicates a pressure variation measurement value.
In these diagrams, at positions when a turbo rpm is 160,000 rpm, a flow rate is about 6 m³/min, and a frequency is about 2.3 kHz, peaks are hardly generated in the noise and the pressure variation, which shows that the above-described abnormal noise (high-frequency noise) is hardly generated.
Table 1 shows a measurement result of an abnormal noise output at an abnormal noise start point in this embodiment.
In this Table, the conventional compressor housing without a countermeasure for abnormal noise corresponds to ‘Original’, the compressor housing according to the invention corresponds to ‘With Annular Groove’, and a reference comparative example corresponds to ‘With Circulation Passage’ as another countermeasure for abnormal noise.

TABLE 1

ABNORMAL NOISE START POINT
IN CASE OF NOISE COUNTERMEASURE

NOISE

REVOLUTION Nt

FLOW RATE Q

PRESSURE

ABNORMAL NOISE

MEASURE TYPE	(rpm)	(m3/min)	RATIO π	OUTPUT (V)	FREQUENCY (kHz)

WITH ANNULAR	160,000 (2.67 kHz)	(6.015)	(2.627)	(0.0153)	(2.21)
GROOVE	180,000 (3.00 kHz)	7.526	3.190	0.1856	2.39
WITH	160,000 (2.67 kHz)	—	—	—	—
CIRCULATION	180,000 (3.00 kHz)	—	—	—	—
PASSAGE
ORIGINAL	160,000 (2.67 kHz)	6.177	2.597	0.3706	2.15
(NO MEASURE)	180,000 (3.00 kHz)	7.503	3.099	0.5468	2.26

WITH ANNULAR GROOVE: NO ABNORMAL NOISE IS OBSERVED AT 160,000 rpm.
VALUES IN PARENTHESES ARE REFERENCE VALES OF PRE-SURGE POINT

In this Table, at both turbo rpm of 160,000 rpm and 180,000 rpm, it is observed that an abnormal noise output is more reduced than that of ‘Original’.
In addition, in ‘With Circulation Passage’ as a reference comparative example, it is possible to obtain the same advantage, but the compressor housing provided with the circulation passage has a problem that a structure is complex and a manufacture cost is more expensive than that of the compressor housing without the circulation passage. Accordingly, it is not possible to obtain the object of the invention.

Second Embodiment

FIGS. 11B and 11C are diagrams showing the compressor housing according to a second embodiment of the invention. FIG. 11A shows an original compressor housing, FIG. 11B shows a second compressor housing according to the invention, and FIG. 11C shows a third compressor housing according to the invention.
An original turbocharger is different from that of the first embodiment, and in these diagrams, diameters d1 and d2 are 62 mm and 82 mm, respectively.
The second compressor housing shown in FIG. 11B is configured in such a manner that the annular groove 12 is additionally formed in the inner surface of the original compressor housing shown in FIG. 11A.
In this example, an axial center of the annular groove 12 is identical with that of each leading edge tip portion of the short blades 2 b, an axial groove width b1 of the annular groove 12 is 3.5 mm, and an outer diameter d3 of the annular groove 12 is 80 mm. In addition, in this example, the annular groove 12 extends outward upstream from the inner surface of the compressor housing 10 so as to be inclined at 60 degree with respect to the rotary shaft of the compressor.
The third compressor housing shown in FIG. 11C is configured in such a manner that an annular cavity having an inner diameter (d4) of 75 mm, an outer diameter (d5) of 90 mm, and a length (b2) of 22.5 mm is formed in an outer side of the annular groove 12 of the second compressor housing shown in FIG. 11B.
As a test result of the turbocharger having the compressor housings shown in FIGS. 11A, 11B, and 11C, at a pressure ratio of 2.4, a surge flow rate can be reduced from about 8.4 m³/min of the original compressor housing shown in FIG. 11A to about 7.7 m³/min in the case of FIG. 11B and about 7.6 m³/min in the case of FIG. 11C.
Accordingly, as obviously shown in the results, it is possible to reduce a surge line by using the compressor housing according to the invention including those shown in FIGS. 11B and 11C, and thus to reduce an abnormal noise output.
As described above, according to the configuration of the invention, since the inner surface of the compressor housing 10 is provided with the annular groove 12 which surrounds vicinities of the leading edge tip portions of the short blades in a circumferential direction and is concave outward so as not to communicate with the suction port of the compressor, a sectional area of the flow passage is suddenly enlarged by the annular groove 12. As a result, it is possible to silence noise and to remarkably reduce or remove the abnormal noise (high-frequency noise) generated from a pre-surge line to a surge line without a sound isolator or a circulation passage.
The invention is not limited to the preferred embodiments, but may be, of course, modified into various forms without departing from the spirit and the scope of the invention.

Claims

1. A compressor housing for a turbocharger for rotationally driving a turbine impeller using exhaust gas of an internal combustion engine and transmitting the rotation force to a compressor impeller to be rotationally driven so that air or air-fuel mixture is compressed and supplied to the internal combustion engine,

wherein the compressor impeller has a plurality of long blades and short blades alternately arranged in a circumferential direction, and

wherein an inner surface of the compressor housing is provided with an annular groove which surrounds vicinities of leading edge tip portions of the short blades in a circumferential direction and is concave outward so as not to communicate with a suction port of a compressor.

2. The compressor housing for the turbocharger according to claim 1,

wherein an axial center of the annular groove is located within 5 mm from the leading edge tip portions of the short blades in an axial direction on an upstream or downstream side thereof, an axial groove width is no less than 2.5 mm and no more than 10 mm, and a maximum diameter of the annular groove is less than 1.2 times a diameter of each leading edge tip portion of the short blades.

3. The compressor housing for the turbocharger according to claim 1,

wherein the annular groove extends outward from the inner surface of the compressor housing so as to be perpendicular or inclined with respect to a rotary shaft of a compressor.