WO2018101021A1 - Diffuseur, trajet d'écoulement d'évacuation et turbomachine centrifuge - Google Patents

Diffuseur, trajet d'écoulement d'évacuation et turbomachine centrifuge Download PDF

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Publication number
WO2018101021A1
WO2018101021A1 PCT/JP2017/040847 JP2017040847W WO2018101021A1 WO 2018101021 A1 WO2018101021 A1 WO 2018101021A1 JP 2017040847 W JP2017040847 W JP 2017040847W WO 2018101021 A1 WO2018101021 A1 WO 2018101021A1
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Prior art keywords
flow path
diffuser
discharge
circumferential
radial
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PCT/JP2017/040847
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English (en)
Japanese (ja)
Inventor
学 八木
光裕 成田
臣吾 木村
竜一 橋本
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株式会社日立製作所
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Publication of WO2018101021A1 publication Critical patent/WO2018101021A1/fr

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00—Details, component parts, or accessories
    • F04D29/40—Casings; Connections of working fluid
    • F04D29/42—Casings; Connections of working fluid for radial or helico-centrifugal pumps
    • F04D29/44—Fluid-guiding means, e.g. diffusers
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00—Details, component parts, or accessories
    • F04D29/66—Combating cavitation, whirls, noise, vibration or the like; Balancing

Definitions

  • the present invention relates to a diffuser, a discharge channel, and a centrifugal turbomachine.
  • centrifugal compressors Conventional diffusers and discharge passages of centrifugal compressors include, for example, a general one described in Patent Document 1.
  • the centrifugal compressor described in Patent Document 1 includes an axial suction flow path, a centrifugal impeller, a diffuser composed of a radial flow path and a curved flow path that connects the radial flow path and the discharge flow path,
  • the discharge channel has a circular circumferential channel cross-section with a constant channel outer diameter.
  • JP 2009-264136 A paragraphs 0024, 0025, FIG. 1, etc.
  • the diffuser vane is extended to the curved flow path connecting the radial flow path and the discharge flow path, thereby increasing the deceleration effect in the diffuser.
  • increasing the deceleration effect of the diffuser vanes improves the efficiency at the design flow rate.
  • the centrifugal compressor is operated on the smaller flow rate side than the design flow rate, backflow tends to occur in the diffuser portion where the deceleration is large.
  • the operable flow rate range from a design point flow rate to a small flow rate, which is limited by swirling stall in which the flow is disturbed or a reverse flow is generated, or surging in which the reverse flow is high and vibration is generated, is narrowed.
  • the present invention was devised in view of the above circumstances, and it is an object of the present invention to provide a diffuser, a discharge flow path, and a centrifugal turbomachine that can improve the efficiency of the centrifugal turbomachine without narrowing the operable flow range with a simple structure.
  • a diffuser for a centrifugal turbomachine includes a radial flow path and a curved flow path that connects the radial flow path and a discharge flow path that guides a flow in a circumferential direction.
  • the discharge flow path of the centrifugal turbomachine of the second aspect of the present invention includes a radial flow path and a curved flow path that connects the radial flow path and a discharge flow path that guides the flow in the circumferential direction.
  • Vdif is the swirling flow velocity at the outlet end of the passage
  • Vd is the circumferential flow velocity at the portion where the circumferential cross-sectional area of the discharge flow path is maximum.
  • the circumferential flow path of the discharge flow path is formed so that
  • the centrifugal turbomachine of the third aspect of the present invention includes the diffuser of the first aspect of the present invention and an impeller.
  • a centrifugal turbomachine of the fourth aspect of the present invention includes the discharge flow path of the second aspect of the present invention and an impeller.
  • the present invention it is possible to provide a diffuser, a discharge channel, and a centrifugal turbomachine that can improve the efficiency of the centrifugal turbomachine without narrowing the operable flow range with a simple structure.
  • Sectional drawing which shows the flow-path shape of the single stage centrifugal compressor provided with the diffuser and discharge flow path of the centrifugal compressor which concerns on Embodiment 1 of this invention.
  • the schematic diagram which shows the discharge flow path which guides a flow in the circumferential direction of the downstream of a diffuser.
  • Sectional drawing of the diffuser part which compares the diffuser which concerns on Embodiment 1 with the diffuser which concerns on a prior art example.
  • the whole correlation diagram of the flow path length and cross-sectional area ratio of a diffuser which shows the diffuser which concerns on Embodiment 1 in comparison with the diffuser of a prior art example.
  • FIG. 3B The figure which shows only the downstream from the latter half part of the radial direction flow path of the correlation diagram of the flow path length and cross-sectional area ratio of the diffuser of FIG. 3B.
  • the graph which shows the breakdown of the pressure loss in a discharge flow path.
  • Sectional drawing which shows the flow-path shape of a single stage centrifugal compressor.
  • the present invention relates to a centrifugal turbomachine such as a centrifugal compressor that compresses a working fluid, and more particularly to a diffuser and a discharge flow path thereof.
  • a centrifugal turbomachine such as a centrifugal compressor that compresses a working fluid
  • a diffuser and a discharge flow path thereof a centrifugal turbomachine
  • the embodiment for carrying out the present invention will be described in detail by exemplifying two embodiments. Although the two illustrated embodiments 1 and 2 are described by taking a single-stage centrifugal compressor as an example, the present invention can be applied to a multi-stage or other centrifugal fluid machine having a similar structure. .
  • FIG. 1 is a cross-sectional view showing a flow path shape of a single-stage centrifugal compressor E provided with a diffuser 8 and a discharge flow path 6 of a centrifugal compressor according to Embodiment 1 of the present invention.
  • FIG. 2 is a schematic diagram illustrating the discharge flow path 6 that guides the flow in the circumferential direction downstream of the diffuser 8 according to the first embodiment.
  • an impeller 1 in which blades 1a are formed in a plurality of circumferential shapes is provided to be rotatable around a rotation center C1.
  • a suction port 1 s is provided for sucking the working fluid 101 into a single-stage centrifugal compressor E (hereinafter referred to as a centrifugal compressor E).
  • a diffuser 8 for converting the dynamic pressure of the compressed working fluid 101 into a static pressure is provided on the outer peripheral portion of the impeller 1.
  • the diffuser 8 is configured in the order of the radial flow path 3, the curved flow path 4, and the cylindrical flow path 5 from the upstream.
  • the diffuser 8 has a channel narrower than the upstream channel.
  • a circumferential flow path 9 that guides the flow in the circumferential direction is formed in the discharge flow path 6 downstream of the diffuser 8 so that its cross-sectional area increases in the circumferential direction from upstream to downstream.
  • a straight straight pipe channel 10 is connected to the outlet end portion 9 o of the circumferential channel 9.
  • the straight pipe flow path 10 is formed so that the cross-sectional area increases linearly from upstream to downstream.
  • the outlet of the straight pipe channel 10 is a compressor outlet end portion 10o.
  • centrifugal compressor E rotates the impeller 1 around the rotation center C1, and the working fluid 101 is sucked from the outside of the compressor through the suction port 1s near the center (FIG. 1). Arrow ⁇ 0). The sucked working fluid 101 is sent outward from the rotation center C1 by the centrifugal force by the blade 1a of the rotating impeller 1 and compressed.
  • Compressed working fluid 101 is discharged as a swirl flow in the outer circumferential direction of the impeller 1 as indicated by an arrow ⁇ 1 in FIG.
  • the discharged working fluid 101 is discharged into the discharge flow path 6 having the circumferential flow path 9 shown in FIG. 2 through the flow path of the diffuser 8 having the radial flow path 3, the curved flow path 4, and the cylindrical flow path 5. Discharged.
  • the working fluid 101 sent to the discharge flow path 6 flows through the circumferential flow path 9 and the straight pipe flow path 10 (arrow ⁇ 2 in FIG. 2), and is connected to the centrifugal compressor E (compressor outlet end portion 10o).
  • the pipe h1 (arrow ⁇ 3 in FIG. 2).
  • the diffuser 8 has the effect of decelerating the flow velocity of the working fluid 101 released in the outer peripheral direction of the impeller 1.
  • a plurality of diffuser vanes 2 are provided along the swirl flow direction of the discharged working fluid 101.
  • the speed reduction effect is increased.
  • the diffuser vane 2 may not be provided.
  • Embodiment 1 can improve the efficiency of the centrifugal compressor E without narrowing the operable flow range regardless of the presence or absence of the diffuser vane 2.
  • Embodiment 1 (the present invention) can be applied to a centrifugal compressor having a diffuser in which the rear edge portion of a conventional diffuser vane is bent to a curved flow path, such as the centrifugal compressor described in Patent Document 1. .
  • the effect of improving the efficiency of the centrifugal compressor E can be obtained without narrowing the operable flow range. In other words, good effects can be obtained in both the flow rate range and efficiency.
  • a diffuser vane 102 having a rear edge portion on the inner diameter side from the radial flow path outlet end 103o of the comparative example (prior art) shown in FIG.
  • a centrifugal compressor 9E composed of the provided diffuser 108 is used.
  • R / h which is a value obtained by dividing the inner peripheral side (curvature) radius R of the curved flow path 4 by the axial flow path width h of the radial flow path 3, is 0.2 or more and 0.00. It is desirable that the configuration be 5 or less (0.2 ⁇ R / h ⁇ 0.5).
  • the maximum outer diameter D of the flow path of the diffuser 8 does not change. This is because when the maximum outer diameter D changes, the size of the centrifugal compressor E changes, and performance evaluation becomes difficult.
  • R / h is as small as less than 0.2
  • the inner peripheral radius R of the curved channel 4 is relatively small with respect to the axial channel width h of the radial channel 3. Therefore, when the flow direction of the curved flow path 4 is turned from the radial flow path 3 in the radial direction to the cylindrical flow path 5 in the axial direction, the flow is easily separated on the inner peripheral side because R is too small. There is concern that the loss will increase. Therefore, it is desirable to configure the diffuser 8 so that R / h is 0.2 or more and 0.5 or less (0.2 ⁇ R / h ⁇ 0.5).
  • FIG. 3A is a cross-sectional view of a diffuser showing the diffuser 8 according to Embodiment 1 in comparison with a diffuser 108 (FIG. 5) of a comparative example.
  • FIG. 3B is an entire correlation diagram of flow path lengths and cross-sectional area ratios of the diffusers 8 and 108 showing the diffuser 8 of Embodiment 1 in comparison with the diffuser 108 of the comparative example.
  • 3C is a diagram showing only the downstream side from the latter half of the radial flow paths 3 and 103 in the correlation diagram of the flow path length and the cross-sectional area ratio of the diffusers 8 and 108 in FIG. 3B.
  • the maximum flow path outer diameter D of the first embodiment and the comparative example is the same.
  • the horizontal axis of FIG. 3B is (distance from radial flow path inlet ends 3i and 103i in the flow path central part of the first embodiment, comparative example) / (radial flow path inlet end 103i in the flow path central part of the comparative example).
  • To the bent flow path outlet end 104o that is, (the length of the radial flow path 3 and the curved flow path 4 of the first embodiment) and the length of the radial flow path 103 and the curved flow path 104 of the comparative example. ).
  • FIG. 3B shows the flow path distance from the radial flow path inlet end 3i to the curved flow path outlet end 4o of the first embodiment shown in FIG. 3A and the radial flow path inlet end 103i of the comparative example to the curved flow path outlet end 104o. And the cross-sectional area of each channel at each channel distance are compared.
  • the horizontal axis of FIG. 3C represents (distance from the radial flow path inlet ends 3i and 103i of the central portion of the flow path in the first embodiment and the comparative example) / (radial flow path inlet end 103i of the central portion of the flow path of the comparative example.
  • the distance from the curved flow path outlet end 104o to the curved flow path outlet end 104o) is compared with the flow lengths of the radial flow paths 3, 103 and the curved flow paths 4, 104 of the first embodiment and the comparative example.
  • 3C represents (the cross-sectional area of the flow path distance from the radial direction flow path inlet ends 3i and 103i of the first embodiment, comparative example) / (the first embodiment, the curved flow path outlet end 4o of the comparative example, 104o cross-sectional area)
  • FIG. 3C compares the second half of the flow path distance from the radial flow path inlet ends 3i and 103i (see FIG. 3A) to the bent flow path outlet ends 4o and 104o of the comparative example and Embodiment 1 in FIG. 3B. ing. That is, FIG. 3C is an enlarged view of the latter half of the flow path of FIG. 3B.
  • the correlation diagram between the channel length and the cross-sectional area ratio of the diffusers 8 and 108 of the first embodiment and the comparative example in FIGS. 3B and 3C shows a comparison between the first embodiment and the comparative example.
  • Embodiment 1 is increasing the cross-sectional area expansion rate in the radial flow path 3 of the diffuser 8 rather than the comparative example.
  • the correlation diagram between the channel length and the cross-sectional area ratio of the diffusers 8 and 108 in FIGS. 3B and 3C shows a comparison between the prior art and the first embodiment.
  • the first embodiment is relatively about 7% more than the prior art. % Is shorter.
  • the total length of the diffuser 8 constituted by the radial flow path 3 and the curved flow path 4 is shortened as compared with the prior art, so that it occurs when the working fluid 101 passes through the diffuser 8.
  • the pressure loss due to the friction loss is smaller in the first embodiment than in the comparative example in proportion to the total length of the flow path.
  • the discharge flow is such that Vdif / Vd of the value obtained by dividing by the cross-sectional area (cross-sectional area of the portion 9o) divided by the circumferential cross-sectional average flow velocity (circumferential flow velocity) Vd is 1.1 or more and 2.0 or less.
  • the cross-sectional shape of the channel 6 in the circumferential direction is formed.
  • the portion 9o having the largest circumferential cross-sectional area of the discharge flow path 6 is a flow path at a connection portion between the circumferential flow path 9 and the straight pipe flow path 10 in the discharge flow path 6 that guides the flow in the circumferential direction.
  • the cross-sectional area is a portion larger than the cross-sectional area of the other circumferential flow path 9.
  • the shape of the circumferential flow path 9 of the discharge flow path 6 that guides the flow in the circumferential direction is preferably a circular cross section.
  • a portion having a larger flow path cross-sectional area than a portion having a smaller flow path cross-sectional area at the connection portion 9i between the circumferential flow path 9 and the straight pipe flow path 5 (see FIG. 1).
  • the circumferential channel portions (9i to 9o) up to 9o may be formed such that the channel cross-sectional area gradually increases with a correlation such as a proportional or quadratic function to the circumferential channel length.
  • the circumferential flow path length (9i ⁇ ) from the portion where the flow passage cross-sectional area at the connection portion 9i between the circumferential flow passage 9 and the straight pipe flow passage 5 is small is equal to the circumferential flow passage 9 and the straight pipe flow.
  • the circumference of the discharge flow path 6 has the same flow path cross-sectional area with a quarter of the flow path cross-sectional area of the portion 9o where the flow path cross-sectional area at the connection portion between the path 9 and the straight pipe flow path 10 is large.
  • a directional channel shape may be formed.
  • FIG. 4 is a graph showing a breakdown of pressure loss in the discharge flow path 6.
  • the horizontal axis in FIG. 4 represents (mass flow rate) / (design point mass flow rate) in the discharge flow path 6, and the vertical axis in FIG. 4 represents pressure loss in the discharge flow path 6.
  • the loss in the discharge flow path 6 includes a friction loss and a deceleration loss.
  • the friction loss indicated by the one-dot chain line in FIG. 4 tends to increase as the mass flow rate in the discharge flow path 6 increases.
  • the deceleration loss indicated by the broken line in FIG. 4 tends to decrease as the mass flow rate in the discharge flow path 6 increases.
  • the loss in the discharge channel 6 is represented by the sum of friction loss and deceleration loss. Therefore, when the sum is minimized, the pressure loss in the discharge flow path 6 is minimized. In other words, efficiency is maximized.
  • Vdif / Vd is evaluated. Evaluation is made as follows from the relationship between the friction loss and the deceleration loss in FIG. When Vdif / Vd is less than 1.1, since deceleration from Vdif to Vd is not sufficient, deceleration loss is reduced, but it occurs on the circumferential flow surface of the discharge flow channel 6 proportional to the square of Vd. The friction loss is increased, and the pressure loss generated in the discharge passage 6 is increased.
  • the cross-sectional shape of the circumferential flow path 9 of the discharge flow path 6 may be formed so that Vdif / Vd is 1.1 or more and 2.0 or less (1.1 ⁇ Vdif / Vd ⁇ 2.0). Suitable for suppressing pressure loss.
  • FIG. A single-stage centrifugal compressor 9E shown in FIG. 5 corresponds to the comparative example shown in FIGS.
  • FIG. 6 shows a loss factor and a comparative example (see FIG. 5) when the diffusers 8 (see FIG. 1) and 28 (see FIG. 8) and the discharge passages 6 and 26 according to Embodiment 1 and Embodiment 2 described later are employed.
  • FIG. 7 simulates the efficiency when the diffusers 8 and 28 and the discharge passages 6 and 26 according to the first embodiment and the second embodiment described later are employed and the efficiency of the centrifugal compressor 9E of the comparative example with respect to the mass flow rate. It is a figure which shows a result.
  • the loss factor of the discharge flow paths 6 and 106 which is the vertical axis in FIG. 6, is the total pressure difference from the bent flow path outlet ends 4o and 104o to the outlet end 6o of the discharge flow path 6 (see FIG. 2). The value is divided by the dynamic pressure at 4o (see FIG. 1) and 104o (see FIG. 5).
  • the loss coefficient of the discharge passages 6 and 106 at the design point mass flow rate is 38% smaller than that of the comparative example in the first embodiment, and the mass flow rate ratio of 80 to 120 with respect to the design point mass flow rate.
  • Embodiment 1 is smaller than the comparative example. This is because the Vdif / Vd of the first embodiment is changed from 2.11 of the comparative example (prior art) to a preferable 1.19 to suppress deceleration, and the kinetic energy in the circumferential flow path 9 of the discharge flow path 6 is reduced. This is thought to be because the deceleration loss accompanying the sudden decrease in the frequency became smaller (see FIG. 4).
  • the efficiency at the mass flow rate at the design point is 1.5% higher in Embodiment 1 (solid line in FIG. 7) than in the comparative example (broken line in FIG. 7). Further, the efficiency of the first embodiment is improved over that of the comparative example over the entire range of the mass flow rate ratio of 80 to 120% with respect to the design point. This is because the flow path length of the diffuser 8 is changed by changing R / h from 1.18 of the comparative example to a preferable 0.36 (within a range of 0.2 ⁇ R / h ⁇ 0.5). In addition to shortening R by about 7% (see FIG. 3C), Vdif is also decelerated by about 7% to reduce the pressure loss associated with the friction loss of the diffuser 8, and the first embodiment shown in FIG. This is probably because the loss coefficient of the discharge flow path 6 is 38% smaller than that of the prior art.
  • the curved flow path 4 provided on the downstream side of the diffuser vane 2 is formed so that R / h is 0.2 or more and 0.5 or less. Therefore, the cross-sectional area enlargement ratio in the radial flow path 3 is increased without increasing the maximum outer diameter D of the diffuser 8, and the overall flow path length of the diffuser 8 constituted by the radial flow path 3 and the curved flow path 4. Can be shortened. Therefore, the pressure loss of the diffuser 8 can be reduced while appropriately increasing the deceleration effect of the diffuser 8 by increasing the cross-sectional area enlargement ratio.
  • FIG. 8 is a cross-sectional view showing the channel shape of a single-stage centrifugal compressor 2E provided with a diffuser 28 and a discharge channel 26 according to Embodiment 2 of the present invention.
  • the diffuser 28 and the discharge flow path 26 of the centrifugal compressor 2E according to the second embodiment of the present invention will be described below with reference to FIG.
  • the basic configuration of the centrifugal compressor 2E according to the second embodiment is the same as that of the first embodiment, but the inner radius R of the curved flow path 24 in the diffuser 28 is the axial flow width of the radial flow path 23.
  • R / h which is a value divided by h, is constituted by a diffuser 28 that is larger than 0.5 and does not satisfy 0.2 ⁇ R / h ⁇ 0.5. Since other configurations are the same, the same components are denoted by the same reference numerals, and detailed description thereof is omitted.
  • FIG. 5 which concerns on the comparative example of the centrifugal compressor 2E (refer FIG. 8) which concerns on Embodiment 2
  • FIG. 5 of a comparative example and the result of having simulated the loss coefficient with respect to mass flow rate 6 and FIG. 7 showing the result of simulating the efficiency of the centrifugal compressors 2E and 9E with respect to the mass flow rate will be described below.
  • the loss coefficient of the discharge flow path at the design point mass flow rate is 22% smaller than that of the comparative example (dashed line in FIG. 6) in the second embodiment (the dashed line in FIG. 6), and
  • the embodiment 2 is smaller than the comparative example in the entire range of the mass flow rate ratio of 80 to 120% with respect to the mass.
  • the Vdif / Vd of the second embodiment is changed from 2.11 of the comparative example to 1.27 that satisfies the preferable 1.1 ⁇ Vdif / Vd ⁇ 2.0, and thus the circumferential direction of the discharge flow path 6 This is presumably because the deceleration loss accompanying the sudden decrease in kinetic energy in the flow path 9 (see FIG. 2) is reduced.
  • the efficiency of the design point in terms of mass flow rate is 0.9% higher in Embodiment 2 (the dashed line in FIG. 7) than in the comparative example (broken line in FIG. 7).
  • the efficiency is improved in the second embodiment over the comparative example over the entire range of the mass flow rate ratio of 80 to 120%.
  • the loss coefficient of the discharge flow path 6 is improved by 22% over the comparative example, The efficiency is improved as compared with the comparative example in the entire range of the mass flow rate ratio of 80 to 120% with respect to the design point mass.
  • centrifugal turbomachine 2E that can improve the efficiency without narrowing the operable flow range with a simple structure that satisfies 1.1 ⁇ Vdif / Vd ⁇ 2.0 can be obtained. .
  • the present invention can be applied to a centrifugal turbomachine represented by a centrifugal blower (blower) or a centrifugal compressor.
  • a centrifugal turbomachine represented by a centrifugal blower (blower) or a centrifugal compressor.
  • the present invention may be applied to the last stage.

Abstract

L'invention concerne un diffuseur (8) destiné à une turbomachine centrifuge (E) comprenant un trajet d'écoulement radial (3) et un trajet d'écoulement courbé (4) qui relie le trajet d'écoulement radial (3) et un trajet d'écoulement d'évacuation (6) qui conduit un écoulement de manière circonférentielle. Si le rayon côté périphérique interne du trajet d'écoulement courbé (4) est désigné par (R), et si la largeur de trajet d'écoulement axial du trajet d'écoulement radial (3) est désignée par (h), alors la relation 0,2 ≤ R/h ≤ 0,5 est satisfaite.
PCT/JP2017/040847 2016-11-29 2017-11-14 Diffuseur, trajet d'écoulement d'évacuation et turbomachine centrifuge WO2018101021A1 (fr)

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JP2016-230924 2016-11-29
JP2016230924A JP6860331B2 (ja) 2016-11-29 2016-11-29 ディフューザ、吐出流路、および遠心ターボ機械

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Cited By (1)

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Publication number Priority date Publication date Assignee Title
JPWO2020240608A1 (fr) * 2019-05-24 2020-12-03

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JPS56138499A (en) * 1980-04-01 1981-10-29 Toyota Motor Corp Compressor structure for turbocharger
JPS5753300Y2 (fr) * 1978-08-17 1982-11-18
JPH03217699A (ja) * 1990-01-23 1991-09-25 Nissan Motor Co Ltd 圧縮機のスクロール構造
JP2000081000A (ja) * 1998-09-04 2000-03-21 Ishikawajima Harima Heavy Ind Co Ltd ターボ圧縮機
JP2001317485A (ja) * 2000-05-08 2001-11-16 Mitsuya Soufuuki Seisakusho:Kk 遠心式送風機
JP2004232637A (ja) * 2003-01-28 2004-08-19 Dresser-Rand Co ノイズ減衰化ガス圧縮装置および方法
US20100158722A1 (en) * 2007-06-21 2010-06-24 Siegfried Sumser Air supplier, particularly for an air supply system for fuel cells

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JPS5753300Y2 (fr) * 1978-08-17 1982-11-18
JPS56138499A (en) * 1980-04-01 1981-10-29 Toyota Motor Corp Compressor structure for turbocharger
JPH03217699A (ja) * 1990-01-23 1991-09-25 Nissan Motor Co Ltd 圧縮機のスクロール構造
JP2000081000A (ja) * 1998-09-04 2000-03-21 Ishikawajima Harima Heavy Ind Co Ltd ターボ圧縮機
JP2001317485A (ja) * 2000-05-08 2001-11-16 Mitsuya Soufuuki Seisakusho:Kk 遠心式送風機
JP2004232637A (ja) * 2003-01-28 2004-08-19 Dresser-Rand Co ノイズ減衰化ガス圧縮装置および方法
US20100158722A1 (en) * 2007-06-21 2010-06-24 Siegfried Sumser Air supplier, particularly for an air supply system for fuel cells

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* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPWO2020240608A1 (fr) * 2019-05-24 2020-12-03
WO2020240608A1 (fr) * 2019-05-24 2020-12-03 三菱重工エンジン＆ターボチャージャ株式会社 Compresseur centrifuge et turbocompresseur
JP7198923B2 (ja) 2019-05-24 2023-01-04 三菱重工エンジン＆ターボチャージャ株式会社 遠心圧縮機及びターボチャージャ

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