WO2018101021A1 - Diffuser, discharge flow path, and centrifugal turbo machine - Google Patents

Diffuser, discharge flow path, and centrifugal turbo machine 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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French (fr)
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/en

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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

This diffuser (8) for a centrifugal turbo machine (E) comprises a radial flow path (3) and a bent flow path (4) which connects the radial flow path (3) and a discharge flow path (6) which circumferentially conducts a flow. If the inner peripheral-side radius of the bent flow path (4) is designated as (R), and if the axial flow path width of the radial flow path (3) is designated as (h), then the relationship of 0.2 ≤ R/h ≤ 0.5 is satisfied.

Description

ディフューザ、吐出流路、および遠心ターボ機械Diffuser, discharge flow path, and centrifugal turbomachine
　本発明は、ディフューザ、吐出流路、および遠心ターボ機械に関する。 The present invention relates to a diffuser, a discharge channel, and a centrifugal turbomachine.
　従来の遠心圧縮機のディフューザおよび吐出流路は、例えば特許文献１に記載の一般的なものがある。特許文献１に記載の遠心圧縮機は、軸方向の吸込流路と、遠心型羽根車と、半径方向流路および半径方向流路と吐出流路を繋ぐ曲がり流路で構成されたディフューザと、流路外径が一定で円形の周方向流路断面形状を有する吐出流路で構成される。 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.
　ただし、特許文献１の図１に示すように、半径方向流路と吐出流路を繋ぐ曲がり流路までディフューザベーンを延伸させた従来技術をディフューザに適用することで、ディフューザでの減速効果を従来よりも大きくしている。これにより、圧縮された流体が吐出流路へ流出する際の流速を従来よりも減速できるため、圧力損失を増加させることがなく高い効率が得られる。 However, as shown in FIG. 1 of Patent Document 1, the conventional technique in which the diffuser vane is extended to the curved flow path connecting the radial flow path and the discharge flow path is applied to the diffuser, so that the deceleration effect in the diffuser is conventionally achieved. Is bigger than. Thereby, since the flow velocity when the compressed fluid flows out to the discharge flow path can be reduced as compared with the conventional case, high efficiency can be obtained without increasing the pressure loss.
特開２００９－２６４１３６号公報（段落００２４、００２５、図１等）JP 2009-264136 A (paragraphs 0024, 0025, FIG. 1, etc.)
　ところで、特許文献１に記載のディフューザおよび吐出流路では、半径方向流路と吐出流路を繋ぐ曲がり流路までディフューザベーンを延伸させることで、ディフューザでの減速効果をより大きくしている。しかし、ディフューザベーンによる減速効果を大きくすると設計流量での効率は向上する。一方、遠心圧縮機を設計流量より小流量側で運転する際に、減速が大きいディフューザ部分で逆流が生じ易くなる。そのため、流れが乱れたり、逆流が生じる旋回失速や、逆流が高じて振動が発生するサージングで制限される設計点流量から小流量にかけての運転可能な流量範囲が狭くなる。 By the way, in the diffuser and the discharge flow path described in Patent Document 1, 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. However, increasing the deceleration effect of the diffuser vanes improves the efficiency at the design flow rate. On the other hand, when 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. Therefore, 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.
　本発明は上記実状に鑑み創案されたものであり、簡単な構造により運転可能な流量範囲を狭めることなく、遠心ターボ機械の効率を向上できるディフューザ、吐出流路、および遠心ターボ機械の提供を目的とする。 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. And
　前記課題を解決するため、第１の本発明の遠心ターボ機械のディフューザは、半径方向流路と、前記半径方向流路と周方向に流れを導く吐出流路を繋ぐ曲がり流路とで構成され、前記曲がり流路の内周側半径をＲとし、前記半径方向流路の軸方向流路幅をｈとした場合、　０．２≦Ｒ／ｈ≦０．５　の関係がある。 In order to solve the above problems, a diffuser for a centrifugal turbomachine according to the first 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 a flow in a circumferential direction. When the radius on the inner peripheral side of the curved flow path is R and the axial flow path width of the radial flow path is h, there is a relationship of 0.2 ≦ R / h ≦ 0.5.
　第２の本発明の遠心ターボ機械の吐出流路は、半径方向流路と、前記半径方向流路と周方向に流れを導く吐出流路を繋ぐ曲がり流路とで構成され、前記半径方向流路の出口端部の旋回流れ流速をＶdifとし、前記吐出流路の周方向断面積が最大となる部分における周方向流速をＶdとした場合、１．１≦Ｖdif/Ｖd≦２．０　の関係となるように前記吐出流路の周方向流路が形成されている。 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. 1.1 ≦ Vdif / Vd ≦ 2.0, where Vdif is the swirling flow velocity at the outlet end of the passage, and 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
　第３の本発明の遠心ターボ機械は、第１の本発明のディフューザと、羽根車とを具備している。 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.
　第４の本発明の遠心ターボ機械は第２の本発明の吐出流路と、羽根車とを具備している。 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.
　本発明によれば、簡単な構造により運転可能な流量範囲を狭めることなく、遠心ターボ機械の効率を向上できるディフューザ、吐出流路、および遠心ターボ機械を提供できる。 According to 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.
本発明の実施形態１に係る遠心圧縮機のディフューザおよび吐出流路を備えた単段遠心圧縮機の流路形状を示す断面図。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. ディフューザの下流の周方向に流れを導く吐出流路を示す模式図。The schematic diagram which shows the discharge flow path which guides a flow in the circumferential direction of the downstream of a diffuser. 実施形態１に係るディフューザを従来例に係るディフューザと比較して示すディフューザ部の断面図。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. 実施形態１に係るディフューザを従来例のディフューザと比較して示すディフューザの流路長さと断面積比の相関図の全体。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. 図３Ｂのディフューザの流路長さと断面積比の相関図の半径方向流路の後半部から下流側のみを示す図。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. 吐出流路における圧力損失の内訳を示すグラフ。The graph which shows the breakdown of the pressure loss in a discharge flow path. 比較例に係るＲ／ｈ＝１．１８の構成のディフューザと比較例に係るＶdif／Ｖd＝２．１１の流路断面積が比較的大きい周方向流路で構成される吐出流路を備えた単段遠心圧縮機の流路形状を示す断面図。A diffuser having a configuration of R / h = 1.18 according to the comparative example and a discharge flow path configured by a circumferential flow path having a relatively large flow path cross-sectional area of Vdif / Vd = 2.11 according to the comparative example are provided. Sectional drawing which shows the flow-path shape of a single stage centrifugal compressor. 実施形態１、実施形態２に係るディフューザおよび吐出流路を採用した場合の損失係数と、比較例の吐出流路の損失係数を質量流量に対してシミュレーションした結果を示す図。The figure which shows the result of having simulated the loss coefficient at the time of employ | adopting the diffuser and discharge flow path which concern on Embodiment 1, Embodiment 2, and the loss coefficient of the discharge flow path of a comparative example with respect to mass flow. 実施形態１、実施形態２に係るディフューザおよび吐出流路を採用した場合の効率と、比較例の遠心圧縮機の効率を質量流量に対してシミュレーションした結果を示す図。The figure which shows the result of having simulated the efficiency at the time of employ | adopting the diffuser and discharge flow path which concern on Embodiment 1, Embodiment 2, and the efficiency of the centrifugal compressor of a comparative example with respect to mass flow. 本発明の実施形態２に係るディフューザおよび吐出流路を備えた単段遠心圧縮機の流路形状を示す断面図。Sectional drawing which shows the flow-path shape of the single stage centrifugal compressor provided with the diffuser and discharge flow path which concern on Embodiment 2 of this invention.
　本発明は、作動流体を圧縮する遠心圧縮機などの遠心ターボ機械に係り、特にそのディフューザおよび吐出流路に関する。
　以下、本発明を実施するための形態について２つの実施形態を例示して詳細に説明する。なお、例示する２つの実施形態１、２は単段遠心圧縮機を例に挙げて説明するが、多段の場合や、類似構造を有するその他の遠心型流体機械にも本発明は適用可能である。 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.
Hereinafter, 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. .
＜＜実施形態１＞＞
　本発明の実施形態１に係る遠心圧縮機のディフューザおよび吐出流路について、図１ないし図３Ａ～図３Ｃを参照しつつ説明する。
　図１は本発明の実施形態１に係る遠心圧縮機のディフューザ８および吐出流路６を備えた単段遠心圧縮機Ｅの流路形状を示す断面図である。
　図２は、実施形態１のディフューザ８の下流の周方向に流れを導く吐出流路６を示す模式図である。 << Embodiment 1 >>
A diffuser and a discharge flow path of the centrifugal compressor according to the first embodiment of the present invention will be described with reference to FIGS. 1 to 3A to 3C.
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.
　実施形態１の単段遠心圧縮機Ｅは、羽根１ａが複数周状に形成された羽根車１が回転中心Ｃ１周りに回転自在に設けられている。羽根車１の中央には、作動流体１０１を単段遠心圧縮機Ｅ（以下、遠心圧縮機Ｅと称す）の内部に吸込む吸込み口１ｓが設けられている。羽根車１の外周部には、圧縮された作動流体１０１の動圧を静圧に変換するためのディフューザ８が設けられている。ディフューザ８は、上流から半径方向流路３、曲がり流路４および円筒流路５の順に構成される。ディフューザ８は、上流の流路より流路が狭く形成されている。 In the single-stage centrifugal compressor E of 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. In the center of the impeller 1, 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.
　図２に示すように、ディフューザ８の下流の吐出流路６には、周方向に流れを導く周方向流路９が円周状に上流から下流にいくに従って断面積が拡大して形成されている。周方向流路９の出口端部９ｏには、直線状の直管流路１０が接続されている。直管流路１０は、直線状に上流から下流にいくに従って断面積が拡大して形成されている。直管流路１０の出口は、圧縮機出口端部１０ｏとなる。 As shown in FIG. 2, 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. Yes. 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.
　＜遠心圧縮機Ｅの動作＞
　遠心圧縮機Ｅは、図１に示すように、羽根車１が回転中心Ｃ１周りに回転することで、圧縮機外部から中央付近の吸込み口１ｓを介して作動流体１０１が吸込まれる（図１の矢印α０）。吸込まれた作動流体１０１は、回転する羽根車１の羽根１ａにより遠心力で回転中心Ｃ１から外方に送られ圧縮される。 <Operation of centrifugal compressor E>
As shown in FIG. 1, the 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 α0). 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.
　圧縮された作動流体１０１は、図１の矢印α１に示すように、羽根車１の外周外方向に旋回流れとして放出される。放出された作動流体１０１は、半径方向流路３、曲がり流路４、円筒流路５をもつディフューザ８の流路を介して、図２に示す周方向流路９を有する吐出流路６に排出される。吐出流路６に送られた作動流体１０１は、周方向流路９、直管流路１０を流れて（図２の矢印α２）、遠心圧縮機Ｅ（圧縮機出口端部１０ｏ）に接続される配管ｈ１に排出される（図２の矢印α３）。 Compressed working fluid 101 is discharged as a swirl flow in the outer circumferential direction of the impeller 1 as indicated by an arrow α1 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 α2 in FIG. 2), and is connected to the centrifugal compressor E (compressor outlet end portion 10o). To the pipe h1 (arrow α3 in FIG. 2).
　ディフューザ８は、羽根車１の外周外方向に放出された作動流体１０１の流速を、減速させる効果を有している。ここで、遠心圧縮機Ｅの運転可能な流量範囲よりも効率を重視する場合は、放出された作動流体１０１の旋回流れ方向に沿って複数のディフューザベーン２（図１参照）が設けられる。複数のディフューザベーン２間に作動流体１０１が通過する流路を形成することにより、減速効果を増大させる。
　逆に、遠心圧縮機Ｅの効率よりも運転可能な流量範囲の広さを重視する場合などは、ディフューザベーン２は設けない場合もある。 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. Here, when the efficiency is more important than the operable flow range of the centrifugal compressor E, a plurality of diffuser vanes 2 (see FIG. 1) are provided along the swirl flow direction of the discharged working fluid 101. By forming a flow path through which the working fluid 101 passes between the plurality of diffuser vanes 2, the speed reduction effect is increased.
On the contrary, when importance is attached to the range of the flow rate that can be operated rather than the efficiency of the centrifugal compressor E, the diffuser vane 2 may not be provided.
　なお、実施形態１は、ディフューザベーン２の有無に関わらず、運転可能な流量範囲を狭くすることなく、遠心圧縮機Ｅの効率を向上できる。例えば、特許文献１に記載の遠心圧縮機のように、従来例のディフューザベーンの後縁部を曲がり流路まで延伸させたディフューザをもつ遠心圧縮機に、実施形態１（本発明）を適用できる。この場合も、運転可能な流量範囲を狭くすることなく、遠心圧縮機Ｅの効率を向上できる効果を得られる。つまり、流量範囲についても、効率についても良好な効果を得られる。 In addition, 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. For example, 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. . Also in this case, 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.
　実施形態１の特徴と効果について、従来の遠心圧縮機と比較するため、図５に示す比較例（従来技術）の半径方向流路出口端１０３ｏより内径側に後縁部を有するディフューザベーン１０２を設けたディフューザ１０８で構成された遠心圧縮機９Ｅを用いる。 In order to compare the features and effects of the first embodiment with a conventional centrifugal compressor, 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.
　＜ディフューザ８の曲がり流路４の内周側半径Ｒと半径方向流路３の軸方向流路幅ｈとの関係＞
　図１に示すディフューザ８において、曲がり流路４の内周側（曲率）半径Ｒを半径方向流路３の軸方向流路幅ｈで除した値であるＲ／ｈが０．２以上０．５以下（０．２≦Ｒ／ｈ≦０．５）となるように構成することが望ましい。ここで、ディフューザ８の流路の最大外径Ｄは変わらないとする。最大外径Ｄが変わると、遠心圧縮機Ｅの大きさが変わり、性能評価が困難になるからである。 <Relationship between the inner peripheral radius R of the curved flow path 4 of the diffuser 8 and the axial flow path width h of the radial flow path 3>
In the diffuser 8 shown in FIG. 1, 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). Here, it is assumed that 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が０．５より大きい場合には、Ｖdifの減速とディフューザ８の流路全長の短縮が十分でなくなるために、減速と流路全長の短縮の両者の相乗効果で得られる圧力損失の低減は小さなものになる。 When R / h is larger than 0.5, the reduction of Vdif and the shortening of the total length of the flow path of the diffuser 8 are not sufficient, and therefore the pressure loss obtained by the synergistic effect of both the speed reduction and the shortening of the total length of the flow path. The reduction is small.
　一方、Ｒ／ｈが０．２未満と小さい場合には、曲がり流路４の内周側半径Ｒが半径方向流路３の軸方向流路幅ｈに対して相対的に小さい。そのため、曲がり流路４において流れ方向が径方向の半径方向流路３から軸方向の円筒流路５に転向する際に、Ｒが小さ過ぎるために内周側で流れが剥離し易くなり、圧力損失が増加することが懸念される。従って、Ｒ／ｈが０．２以上０．５以下（０．２≦Ｒ／ｈ≦０．５）となるようにディフューザ８を構成することが望ましい。 On the other hand, when 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).
　０．２≦Ｒ／ｈ≦０．５とした場合の効果について図３Ａ～図３Ｃを用いて説明する。
　図３Ａは実施形態１に係るディフューザ８を比較例のディフューザ１０８（図５）と比較して示すディフューザの断面図である。図３Ｂは実施形態１のディフューザ８を比較例のディフューザ１０８と比較して示すディフューザ８、１０８の流路長さと断面積比の相関図の全体である。図３Ｃは図３Ｂのディフューザ８、１０８の流路長さと断面積比の相関図の半径方向流路３、１０３の後半部から下流側のみを示す図である。 The effect when 0.2 ≦ R / h ≦ 0.5 will be described with reference to FIGS. 3A to 3C.
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.
　図３Ａに示す半径方向流路３と曲がり流路４とで構成されたディフューザ８の断面図は、Ｒ／ｈ＝１．１８（０．２≦Ｒ／ｈ≦０．５の範囲外）の比較例に係る構成のディフューザ１０８と、０．２≦Ｒ／ｈ≦０．５の関係を満足するＲ／ｈ＝０．３６の実施形態１に係るディフューザ８を比較して示している。なお、上記したように、実施形態１と比較例の流路最大外径Ｄは同一寸法である。 The cross-sectional view of the diffuser 8 constituted by the radial flow path 3 and the curved flow path 4 shown in FIG. 3A is R / h = 1.18 (outside the range of 0.2 ≦ R / h ≦ 0.5). The diffuser 108 having the configuration according to the comparative example and the diffuser 8 according to the first embodiment with R / h = 0.36 satisfying the relationship of 0.2 ≦ R / h ≦ 0.5 are shown in comparison. As described above, the maximum flow path outer diameter D of the first embodiment and the comparative example is the same.
　図３Ｂの横軸は、（実施形態１、比較例の流路中央部における半径方向流路入口端３ｉ、１０３ｉからの距離）／（比較例の流路中央部における半径方向流路入口端１０３ｉから曲がり流路出口端１０４ｏまでの距離）、つまり（実施形態１の半径方向流路３と曲がり流路４の長さ）と（比較例の半径方向流路１０３と曲がり流路１０４の長さ）とを比較する。図３Ｂの縦軸は、（実施形態１、比較例の流路距離の流路断面積）／（実施形態１、比較例の曲がり流路の出口端４ｏ、１０４ｏの流路断面積）をとっている。 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. ). The vertical axis in FIG. 3B is taken as (flow path cross-sectional area of the flow path distance of the first embodiment, comparative example) / (flow path cross-sectional area of the outlet ends 4o, 104o of the curved flow path of the first embodiment, comparative example). ing.
　図３Ｂは、図３Ａに示す実施形態１の半径方向流路入口端３ｉから曲がり流路出口端４ｏまでの流路距離と比較例の半径方向流路入口端１０３ｉから曲がり流路出口端１０４ｏまでの流路距離との比較を行い、各流路距離での流路断面積の比較を行っている。 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.
　図３Ｃの横軸は、（実施形態１と比較例の流路中央部の半径方向流路入口端３ｉ、１０３ｉからの距離）／（比較例の流路中央部の半径方向流路入口端１０３ｉから曲がり流路出口端１０４ｏまでの距離）をとって、実施形態１と比較例の半径方向流路３、１０３および曲がり流路４、１０４の流路長さを比較している。図３Ｃの縦軸は、（実施形態１、比較例の半径方向流路入口端３ｉ、１０３ｉから流路距離の流路断面積）／（実施形態１、比較例の曲がり流路出口端４ｏ、１０４ｏの断面積）をとっている 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. The vertical axis of FIG. 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)
　図３Ｃは、比較例と実施形態１の半径方向流路入口端３ｉ、１０３ｉ（図３Ａ参照）から曲がり流路出口端４ｏ、１０４ｏまでの流路距離の後半部分の図３Ｂでの比較を行っている。すなわち、図３Ｃは図３Ｂの流路の後半部分の拡大図である。
　図３Ｂおよび図３Ｃの実施形態１、比較例のディフューザ８、１０８の流路長さと断面積比の相関図は、実施形態１と比較例とを比較して示している。 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.
　図３Ｃに示すように、実施形態１の半径方向流路出口端３oにおける流路断面積と、比較例の半径方向流路出口端１０３oにおける流路断面積とでは、約０．９２と約０．８４であり、実施形態１の方が比較例よりも相対的に約９％大きくなっている。従って、実施形態１は比較例よりもディフューザ８の半径方向流路３での断面積拡大率を増加させている。 As shown in FIG. 3C, the flow path cross-sectional area at the radial flow path outlet end 3o of Embodiment 1 and the flow path cross-sectional area at the radial flow path outlet end 103o of the comparative example are about 0.92 and about 0. .84, and the first embodiment is about 9% larger than the comparative example. Therefore, 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.
　ここで、流量が一定で密度が同一の場合、Ａ１×Ｖ１＝Ａ２×Ｖ２　の関係がある。Ａ１、Ａ２は流路の断面積であり、Ｖ１、Ｖ２はそれぞれの断面積Ａ１、Ａ２での流速である。
　そのため、曲がり流路４に流入する際の作動流体１０１の旋回流れ流速Ｖdifは、実施形態１の方が従来技術よりも小さくなる。 Here, when the flow rate is constant and the density is the same, there is a relationship of A1 × V1 = A2 × V2. A1 and A2 are the cross-sectional areas of the flow paths, and V1 and V2 are the flow velocities at the respective cross-sectional areas A1 and A2.
Therefore, the swirl flow velocity Vdif of the working fluid 101 when flowing into the bent flow path 4 is smaller in the first embodiment than in the prior art.
　Ｖdifが小さいということは、流速の二乗に比例する下記（１）式で表される摩擦損失も小さくなり、作動流体１０１が曲がり流路４を通過する際に生じる摩擦損失に起因する圧力損失は、実施形態１の方が従来技術よりも小さくなる。
摩擦損失に起因する圧力損失＝１／２・（τ0・ρ・ｖ2・ｄ／Δｌ）　（１）
　　　　　　ここで、τ0は摩擦係数、ρは入口密度、ｖは流速、ｄは流路内径、Δｌは流路長さである。 When Vdif is small, the friction loss represented by the following equation (1) proportional to the square of the flow velocity is also small, and the pressure loss due to the friction loss generated when the working fluid 101 passes through the curved flow path 4 is The first embodiment is smaller than the prior art.
Pressure loss due to friction loss = 1/2 · (τ 0 · ρ · v 2 · d / Δl) (1)
Here, τ 0 is the friction coefficient, ρ is the inlet density, v is the flow velocity, d is the channel inner diameter, and Δl is the channel length.
　図３Ｂおよび図３Ｃのディフューザ８、１０８の流路長さと断面積比の相関図は、従来技術と実施形態１を比較して示している。
　図３Ｃに示すように、流路中央部における半径方向流路入口端３ｉ、１０３ｉから曲がり流路出口端４ｏ、１０４ｏまでの距離において、実施形態１の方が従来技術よりも相対的に約７％短くなっていることが分る。 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.
As shown in FIG. 3C, in the distance from the radial flow path inlet ends 3i, 103i to the curved flow path outlet ends 4o, 104o in the central portion of the flow path, the first embodiment is relatively about 7% more than the prior art. % Is shorter.
　従って、実施形態１は従来技術よりも半径方向流路３と前記曲がり流路４で構成されたディフューザ８の流路全長を短縮しているので、作動流体１０１がディフューザ８を通過する際に生じる摩擦損失に起因する圧力損失は、（１）式に示すように、流路全長に比例して実施形態１の方が比較例よりも小さくなる。 Accordingly, in the first embodiment, 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. As shown in the equation (1), 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.
　＜半径方向流路出口端部３ｏにおける旋回流れ流速Ｖdifと吐出流路６の周方向断面積が最大となる部分９ｏの周方向断面平均流速Ｖd＞
　遠心圧縮機Ｅでは、ディフューザ８と同時に、ディフューザ８の半径方向流路出口端部３ｏにおける旋回流れ流速Ｖdifを、吐出流路６の周方向断面積が最大となる部分９ｏにおける体積流量を周方向断面積（部分９ｏの断面積）で除して求めた周方向断面平均流速（周方向流速）Ｖdで除した値のＶdif／Ｖdが１．１以上２．０以下となるように、吐出流路６の周方向流路断面形状を形成している。 <Surrounding Flow Velocity Vdif at Radial Channel Outlet End 3o and Circumferential Cross Section Average Flow Vd of Portion 9o Where Circumferential Cross-Sectional Area of Discharge Channel 6 is Maximum>
In the centrifugal compressor E, simultaneously with the diffuser 8, the swirl flow velocity Vdif at the radial flow path outlet end 3 o of the diffuser 8 and the volume flow rate at the portion 9 o where the circumferential cross-sectional area of the discharge flow path 6 is maximum are measured in the circumferential direction. 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.
　ここで、吐出流路６の周方向断面積が最大となる部分９ｏとは、周方向に流れを導く吐出流路６において周方向流路９と直管流路１０との接続部分で流路断面積が他の周方向流路９の流路断面積より大きい部分をいう。 Here, 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.
　圧力損失を小さくするためには、周方向に流れを導く吐出流路６の周方向流路９の形状は円形断面が好ましい。
　また、図２に示す吐出流路６において周方向流路９と直管流路５（図１参照）との接続部分９ｉでの流路断面積が小さい部分から、流路断面積が大きい部分９ｏまでの間の周方向流路部分（９ｉ～９ｏ）は、流路断面積が周方向流路長さに比例や２次関数などの相関をもって漸増するように形成すればよい。 In order to reduce the pressure loss, 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.
Further, in the discharge flow channel 6 shown in FIG. 2, 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.
　また、周方向流路９と直管流路５との接続部分９ｉでの流路断面積が小さい部分からの周方向流路長さ（９ｉ～）が、周方向流路９と直管流路５との接続部分９ｉで流路断面積が小さい部分から大きい部分９ｏまでの間の周方向流路長さ（９ｉ～９ｏ）に対して０～１５％までの間は、例えば周方向流路９と直管流路１０との接続部分での流路断面積が大きい部分９ｏの流路断面積の１／４の面積で同一の流路断面積を有するように吐出流路６の周方向流路形状を形成してもよい。 In addition, 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. Between 0 and 15% of the circumferential flow path length (9i to 9o) between the portion where the flow passage cross-sectional area is small to the large portion 9o in the connecting portion 9i with the passage 5, for example, the circumferential 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.
　図４は、吐出流路６における圧力損失の内訳を示すグラフである。図４の横軸に吐出流路６における（質量流量）／（設計点の質量流量）をとり、図４の縦軸に吐出流路６の圧力損失をとっている。
　吐出流路６における損失には、摩擦損失と減速損失とがある。図４の一点鎖線で示す摩擦損失は、吐出流路６における質量流量が大きくなるに従って増加する傾向をもつ。一方、図４の破線で示す減速損失は、吐出流路６における質量流量が大きくなるに従って減少する傾向をもつ。 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. On the other hand, 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.
　吐出流路６における損失は、摩擦損失と減速損失との和で表される。そのため、当該和を最も小さくすると吐出流路６の圧力損失が最小になる。換言すれば、効率が最大となる。
　上述の関係を用いて、Ｖdif／Ｖdの値に関して評価する。図４の摩擦損失と減速損失の関係から下記のように評価される。
　Ｖdif／Ｖdが１．１未満と小さい場合には、Ｖdif からＶdの減速が十分でないために、減速損失は減少するものの、Ｖdの２乗に比例する吐出流路６の周方向流路面で発生する摩擦損失が増加して、吐出流路６で発生する圧力損失が大きくなる。 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.
Using the relationship described above, the value of 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.
　これに対して、Ｖdif／Ｖdが２．０より大きい場合には、半径方向流路出口端３ｏから吐出流路６の周方向流路９の間におけるＶdif からＶdの減速が大き過ぎるために、摩擦損失は減少するものの、流れの運動エネルギの急減に伴う減速損失が増大して、結果として圧力損失が大きくなる。従って、Ｖdif／Ｖdが１．１以上２．０以下（１．１≦Ｖdif／Ｖd≦２．０）となるように、吐出流路６の周方向流路９の断面形状を形成することが、圧力損失を抑えるために適している。 On the other hand, when Vdif / Vd is larger than 2.0, the deceleration of Vdif to Vd between the radial flow path outlet end 3o and the circumferential flow path 9 of the discharge flow path 6 is too large. Although the friction loss is reduced, the deceleration loss accompanying the sudden decrease in the kinetic energy of the flow is increased, and as a result, the pressure loss is increased. Accordingly, 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.
　次に、実施形態１に係る遠心圧縮機Ｅの従来例に係る遠心圧縮機９Ｅに対する効果について、図５、図６、および図７を参照しつつ説明する。
　図５は比較例（従来例）に係るＲ／ｈ＝１．１８（０．２≦Ｒ／ｈ≦０．５の範囲外）の構成のディフューザと比較例に係るＶdif／Ｖd＝２．１１（１．１≦Ｖdif／Ｖd≦２．０の範囲外）の流路断面積が比較的大きい周方向流路で構成される吐出流路１０６を備えた単段遠心圧縮機９Ｅの流路形状を示す断面図である。図５に示す単段遠心圧縮機９Ｅは図６および図７に示す比較例に相当する。 Next, the effect of the centrifugal compressor E according to the first embodiment on the centrifugal compressor 9E according to the conventional example will be described with reference to FIG. 5, FIG. 6, and FIG.
FIG. 5 shows a diffuser having a configuration of R / h = 1.18 (outside the range of 0.2 ≦ R / h ≦ 0.5) according to the comparative example (conventional example) and Vdif / Vd = 2.11 according to the comparative example. Flow path shape of single-stage centrifugal compressor 9E provided with discharge flow path 106 composed of a circumferential flow path having a relatively large flow path cross-sectional area of 1.1 ≦ Vdif / Vd ≦ 2.0. FIG. A single-stage centrifugal compressor 9E shown in FIG. 5 corresponds to the comparative example shown in FIGS.
　図６は実施形態１、後記の実施形態２に係るディフューザ８（図１参照）、２８（図８参照）および吐出流路６、２６を採用した場合の損失係数と、比較例（図５参照）の吐出流路１０６の損失係数を質量流量に対してシミュレーションした結果を示す図である。
　図７は実施形態１、後記の実施形態２に係るディフューザ８、２８および吐出流路６、２６を採用した場合の効率と、比較例の遠心圧縮機９Ｅの効率を質量流量に対してシミュレーションした結果を示す図である。 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. ) Is a diagram showing a simulation result of the loss coefficient of the discharge flow path 106 with respect to the mass flow rate.
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.
　ここでは、実施形態１の望ましい０．２≦Ｒ／ｈ≦０．５を満たすＲ／ｈ＝０．３６の構成のディフューザ８と、望ましい１．１≦Ｖdif／Ｖd≦２．０を満たすＶｄｉｆ／Ｖｄ＝１．１９の流路断面積の好適な周方向流路９で構成される吐出流路６を備えた単段遠心圧縮機Ｅ（図６の実線）と比較例の遠心圧縮機９Ｅ（図６の破線）との比較に着目して説明する。 Here, the diffuser 8 having the configuration of R / h = 0.36 that satisfies the desirable 0.2 ≦ R / h ≦ 0.5 of the first embodiment and the Vdif that satisfies the desirable 1.1 ≦ Vdif / Vd ≦ 2.0. Single-stage centrifugal compressor E (solid line in FIG. 6) having a discharge flow path 6 constituted by a preferred circumferential flow path 9 having a flow path cross-sectional area of /Vd=1.19 and a comparative centrifugal compressor 9E Description will be made by paying attention to the comparison with (broken line in FIG. 6).
　図６の縦軸である吐出流路６、１０６の損失係数は、曲がり流路出口端４ｏ、１０４ｏから吐出流路６の出口端６ｏ（図２参照）までの全圧差を曲がり流路出口端４ｏ（図１参照）、１０４ｏ（図５参照）における動圧で除した値である。 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).
　図６において、設計点の質量流量における吐出流路６、１０６の損失係数は、実施形態１が比較例よりも相対的に３８％小さく、設計点の質量流量に対して質量流量割合８０～１２０％の全範囲で実施形態１が比較例よりも小さい。これは、実施形態１のＶdif／Ｖdを比較例（従来技術）の２．１１から、好適な１．１９に変更したことにより減速を抑え、吐出流路６の周方向流路９における運動エネルギの急減に伴う減速損失が小さくなった（図４参照）ためと考えられる。 In FIG. 6, 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. In the entire range of%, 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).
　図７において、設計点の質量流量における効率は、実施形態１（図７の実線）が比較例（図７の破線）よりも１．５％向上している。また、設計点との質量流量割合８０～１２０％の全範囲で効率が実施形態１が比較例よりも向上している。これは、Ｒ／ｈを比較例の１．１８から、好適な０．３６（０．２≦Ｒ／ｈ≦０．５の範囲内）に変更したことによりディフューザ８の流路長さが、Ｒを小さくすることで約７％短縮した（図３Ｃ参照）のに加えて、Ｖdifも約７％減速してディフューザ８の摩擦損失に伴う圧力損失が減少し、図６に示した実施形態１の吐出流路６の損失係数が従来技術よりも相対的に３８％小さくなったためと考えられる。 7, 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.
　実施形態１によれば、１．１≦Ｖdif／Ｖd≦２．０を満たすＶｄｉｆ／Ｖｄ＝１．１９の流路断面積の周方向流路９で構成される吐出流路６を備える。そのため、吐出流路６の損失係数が設計点で比較例よりも相対的に３８％小さく、設計点の質量流量に対して質量流量割合８０～１２０％の全範囲で比較例よりも小さい。 According to the first embodiment, the discharge flow path 6 including the circumferential flow path 9 having a flow path cross-sectional area of Vdif / Vd = 1.19 satisfying 1.1 ≦ Vdif / Vd ≦ 2.0 is provided. Therefore, the loss coefficient of the discharge flow path 6 is 38% smaller than the comparative example at the design point, and smaller than the comparative example in the entire range of the mass flow rate ratio of 80 to 120% with respect to the mass flow rate at the design point.
　加えて、０．２≦Ｒ／ｈ≦０．５を満たすＲ／ｈ＝０．３６をもつ。そのため、ディフューザ８の流路の最大外径Ｄを比較例と同一とした場合、Ｒが小さくディフューザ８の流路長さが、約７％短縮（図３Ｃ参照）する。また、Ｖdifも約７％減速して摩擦損失に伴う圧力損失が減少し、効率が比較例よりも向上する。 In addition, it has R / h = 0.36 satisfying 0.2 ≦ R / h ≦ 0.5. Therefore, when the maximum outer diameter D of the flow path of the diffuser 8 is the same as that in the comparative example, R is small and the flow path length of the diffuser 8 is shortened by about 7% (see FIG. 3C). Also, Vdif is reduced by about 7%, the pressure loss due to friction loss is reduced, and the efficiency is improved as compared with the comparative example.
　ディフューザベーン２の下流側に設けられる曲がり流路４が、Ｒ／ｈが０．２以上０．５以下となるように形成されている。そのため、ディフューザ８の最大外径Ｄを増加させることなく、半径方向流路３における断面積拡大率を増加させるとともに、半径方向流路３と曲がり流路４で構成されたディフューザ８の流路全長を短縮することができる。
　従って、断面積拡大率の増加によりディフューザ８の減速効果を適切に大きくしつつも、ディフューザ８の圧力損失を小さくできる。 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.
　以上より、０．２≦Ｒ／ｈ≦０．５または／および１．１≦Ｖdif／Ｖd≦２．０を満たす簡単な構造により運転可能な流量範囲を狭くすることなく、効率を向上できる遠心型ターボ機械を得られる。
＜＜実施形態２＞＞ From the above, a centrifugal structure capable of improving efficiency without narrowing the operable flow range with a simple structure satisfying 0.2 ≦ R / h ≦ 0.5 or / and 1.1 ≦ Vdif / Vd ≦ 2.0. A type turbomachine can be obtained.
<< Embodiment 2 >>
　図８は本発明の実施形態２に係るディフューザ２８および吐出流路２６を備えた単段遠心圧縮機２Ｅの流路形状を示す断面図である。
　本発明の実施形態２に係る遠心圧縮機２Ｅのディフューザ２８および吐出流路２６について、図８を参照しつつ、以下説明する。 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.
　実施形態２に係る遠心圧縮機２Ｅは、基本的な構成は実施形態１と同様であるが、ディフューザ２８における曲がり流路２４の内周側半径Ｒが半径方向流路２３の軸方向流路幅ｈで除した値であるＲ／ｈが、０．２≦Ｒ／ｈ≦０．５を満足しない０．５より大きいディフューザ２８で構成されている点のみが異なる。
　その他の構成は、同様であるから同様な構成要素には、同一の符号を付して示し、詳細な説明は省略する。 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. The only difference is that 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.
　実施形態２に係る遠心圧縮機２Ｅ（図８参照）の比較例に係る遠心圧縮機９Ｅ（図５参照）に対する効果について、比較例の図５と、損失係数を質量流量に対してシミュレーションした結果の図６、および遠心圧縮機２Ｅ、９Ｅの効率を質量流量に対してシミュレーションした結果を示す図７を参照しつつ以下に説明する。 About the effect with respect to the centrifugal compressor 9E (refer 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.
　実施形態２に係るＲ／ｈ＝１．１８（０．２≦Ｒ／ｈ≦０．５の範囲外）の構成のディフューザ２８とＶdif/Ｖd＝１．２７の流路断面積が最適な周方向流路９で構成される吐出流路６を備えた単段遠心圧縮機２Ｅと、比較例との比較に着目して説明する。 The diffuser 28 having the configuration of R / h = 1.18 (outside the range of 0.2 ≦ R / h ≦ 0.5) and the flow path cross-sectional area of Vdif / Vd = 1.27 according to the second embodiment are optimal. Description will be made by paying attention to the comparison between the single-stage centrifugal compressor 2E having the discharge flow path 6 constituted by the directional flow path 9 and the comparative example.
　図６において、設計点の質量流量における吐出流路の損失係数は、実施形態２（図６の一点鎖線）が比較例（図６の破線）よりも相対的に２２％小さく、かつ設計点の質量に対して質量流量割合８０～１２０％の全範囲で実施形態２が比較例よりも小さくなっている。これは、実施形態２のＶdif／Ｖdを比較例の２．１１から、好適な１．１≦Ｖdif／Ｖd≦２．０を満足する１．２７に変更したことにより吐出流路６の周方向流路９（図２参照）における運動エネルギの急減に伴う減速損失が小さくなったためと考えられる。 In FIG. 6, 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. This is because 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.
　図７において、設計点の質量流量における効率は、実施形態２（図７の一点鎖線）が比較例（図７の破線）よりも０．９％向上しており、設計点の質量に対して質量流量割合８０～１２０％の全範囲で効率が比較例よりも実施形態２で向上している。 In FIG. 7, 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%.
　実施形態２と比較例との違いは、実施形態２は、１．１≦Ｖdif／Ｖd≦２．０の範囲内のＶdif／Ｖd＝１．２７の流路断面積が好適な周方向流路９で構成される吐出流路６を採用しているのに対して、比較例は、１．１≦Ｖdif／Ｖd≦２．０の範囲外のＶdif／Ｖd＝２．１１の流路断面積が比較的大きい周方向流路で構成される吐出流路１０６（図５参照）を採用している点のみである。従って、０．９％の効率向上は図６に示した実施形態２の吐出流路２６の損失係数が比較例よりも相対的に２２％小さくなったためと考えられる。 The difference between the second embodiment and the comparative example is that the second embodiment has a circumferential flow channel in which a flow channel cross-sectional area of Vdif / Vd = 1.27 within a range of 1.1 ≦ Vdif / Vd ≦ 2.0 is preferable. 9 is employed, whereas the comparative example has a channel cross-sectional area of Vdif / Vd = 2.11 outside the range of 1.1 ≦ Vdif / Vd ≦ 2.0. This is only the point that the discharge flow path 106 (see FIG. 5) configured by a relatively large circumferential flow path is employed. Therefore, the efficiency improvement of 0.9% is considered to be due to the fact that the loss coefficient of the discharge flow path 26 of the second embodiment shown in FIG. 6 is 22% smaller than that of the comparative example.
　実施形態２によれば、１．１≦Ｖdif／Ｖd≦２．０の範囲内のＶdif/Ｖd＝１．２７としたので、吐出流路６の損失係数が比較例よりも２２％向上し、設計点の質量に対して質量流量割合８０～１２０％の全範囲で効率が比較例よりも向上する。 According to the second embodiment, since Vdif / Vd = 1.27 within the range of 1.1 ≦ Vdif / Vd ≦ 2.0, 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.
　従って、１．１≦Ｖdif／Ｖd≦２．０を満足する簡単な構造により運転可能な流量範囲を狭くすることなく、効率を向上できる単段遠心圧縮機（遠心型ターボ機械）２Ｅを得られる。 Therefore, a single-stage centrifugal compressor (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. .
　なお、前記の実施形態１、２は、本発明の一例を示したものであり、特許請求の範囲内で、様々な具体的形態、変形形態が可能である。 The first and second embodiments are examples of the present invention, and various specific forms and modifications are possible within the scope of the claims.
　本発明の活用例として、遠心ブロワ（送風機）や遠心圧縮機などに代表される遠心ターボ機械に適用できる。なお、多段の遠心ターボ機械は、最後段に本発明を適用すればよい。 As an application example of the present invention, it can be applied to a centrifugal turbomachine represented by a centrifugal blower (blower) or a centrifugal compressor. In the multi-stage centrifugal turbomachine, the present invention may be applied to the last stage.
　１　　　羽根車
　２　　　ディフューザベーン
　３、２３　半径方向流路
　３ｉ　　半径方向流路入口端
　３ｏ　　半径方向流路出口端
　４　　　曲がり流路
　４ｏ　　曲がり流路出口端
　５　　　円筒流路
　６、２６　吐出流路
　８、２８　ディフューザ
　９　　　周方向流路
　９ｏ　　吐出流路の周方向断面積が最大となる部分
　１０１　作動流体
　Ｅ、２Ｅ　遠心圧縮機（遠心ターボ機械）
　ｈ　　　半径方向流路の軸方向流路幅
　Ｒ　　　曲がり流路の内周側曲率半径（内周側半径）
　Ｖd    吐出流路の周方向断面積が最大となる部分における体積流量を周方向断面積で除する一次元計算で求めた周方向断面平均流速（吐出流路の周方向断面積が最大となる部分における周方向断面平均流速）
　Ｖdif  半径方向流路の出口端部の旋回流れ流速 DESCRIPTION OF SYMBOLS 1 Impeller 2 Diffuser vane 3, 23 Radial flow path 3i Radial flow path inlet end 3o Radial flow path outlet end 4 Curved flow path 4o Curved flow path outlet end 5 Cylindrical flow path 6, 26 Discharge flow path 8, 28 Diffuser 9 Circumferential flow path 9o Portion where the circumferential cross-sectional area of the discharge flow path is maximized 101 Working fluid E, 2E Centrifugal compressor (centrifugal turbomachine)
h Axial channel width of the radial channel R Inner radius of curvature of the curved channel (inner radius)
Vd Average cross-sectional flow velocity obtained by one-dimensional calculation by dividing the volumetric flow rate at the portion where the circumferential cross-sectional area of the discharge flow path is maximum by the circumferential cross-sectional area (the portion where the circumferential cross-sectional area of the discharge flow path is maximum) (Circumferential average velocity in circumferential direction)
Vdif Swirl flow velocity at the outlet end of the radial flow path

Claims (5)

  1. 　半径方向流路と、前記半径方向流路と周方向に流れを導く吐出流路を繋ぐ曲がり流路とで、構成され、
    　前記曲がり流路の内周側半径をＲとし、前記半径方向流路の軸方向流路幅をｈとした場合、
    　　　０．２≦Ｒ／ｈ≦０．５
    　の関係がある
    　ことを特徴とする遠心ターボ機械のディフューザ。     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,
    When the inner peripheral radius of the curved flow path is R and the axial flow path width of the radial flow path is h,
    0.2 ≦ R / h ≦ 0.5
    A centrifugal turbomachine diffuser characterized by the following relationship:
  2. 　半径方向流路と、前記半径方向流路と周方向に流れを導く吐出流路を繋ぐ曲がり流路とで、構成され、
    　前記半径方向流路の出口端部の旋回流れ流速をＶdifとし、前記吐出流路の周方向断面積が最大となる部分における周方向流速をＶdとした場合、
    　　　　１．１≦Ｖdif/Ｖd≦２．０
    　の関係となるように前記吐出流路の周方向流路が形成されている
    　ことを特徴とする遠心ターボ機械の吐出流路。     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,
    When the swirl flow velocity at the outlet end of the radial flow channel is Vdif, and the circumferential flow velocity at the portion where the circumferential cross-sectional area of the discharge flow channel is maximum is Vd,
    1.1 ≦ Vdif / Vd ≦ 2.0
    The discharge flow path of the centrifugal turbomachine is characterized in that a circumferential flow path of the discharge flow path is formed so as to satisfy the following relationship.
  3. 　前記半径方向流路の出口端部の旋回流れ流速をＶdifとし、前記吐出流路の周方向断面積が最大となる部分における周方向流速をＶdとした場合、
    　　　１．１≦Ｖdif/Ｖd≦２．０
    　の関係となるように前記吐出流路の周方向流路が形成されている
    　ことを特徴とする請求項１に記載の遠心ターボ機械のディフューザ。     When the swirl flow velocity at the outlet end of the radial flow channel is Vdif, and the circumferential flow velocity at the portion where the circumferential cross-sectional area of the discharge flow channel is maximum is Vd,
    1.1 ≦ Vdif / Vd ≦ 2.0
    The diffuser of the centrifugal turbomachine according to claim 1, wherein a circumferential flow path of the discharge flow path is formed so as to satisfy the following relationship.
  4. 　請求項１または請求項３に記載の前記ディフューザを具備することを特徴とする遠心ターボ機械。 A centrifugal turbomachine comprising the diffuser according to claim 1 or 3.
  5. 　請求項２に記載の前記吐出流路を具備することを特徴とする遠心ターボ機械。 A centrifugal turbomachine comprising the discharge passage according to claim 2.
PCT/JP2017/040847 2016-11-29 2017-11-14 Diffuser, discharge flow path, and centrifugal turbo machine WO2018101021A1 (en)

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* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2020240608A1 (en) * 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 Centrifugal compressor and turbocharger

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JPWO2020240608A1 (en) * 2019-05-24 2020-12-03
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