WO1999061801A1 - Turbomachine - Google Patents

Turbomachine Download PDF

Info

Publication number
WO1999061801A1
WO1999061801A1 PCT/GB1998/001555 GB9801555W WO9961801A1 WO 1999061801 A1 WO1999061801 A1 WO 1999061801A1 GB 9801555 W GB9801555 W GB 9801555W WO 9961801 A1 WO9961801 A1 WO 9961801A1
Authority
WO
WIPO (PCT)
Prior art keywords
diffuser
vane
section
vanes
exit
Prior art date
Application number
PCT/GB1998/001555
Other languages
English (en)
Inventor
Hideomi Harada
Hiroyoshi Watanabe
Mehrdad Zangeneh
Original Assignee
Ebara Corporation
University College London
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Ebara Corporation, University College London filed Critical Ebara Corporation
Priority to PCT/GB1998/001555 priority Critical patent/WO1999061801A1/fr
Publication of WO1999061801A1 publication Critical patent/WO1999061801A1/fr

Links

Classifications

    • 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
    • F04D29/445—Fluid-guiding means, e.g. diffusers especially adapted for liquid pumps
    • F04D29/448—Fluid-guiding means, e.g. diffusers especially adapted for liquid pumps bladed 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
    • F04D29/40—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
    • F04D29/441—Fluid-guiding means, e.g. diffusers especially adapted for elastic fluid pumps
    • F04D29/444—Bladed diffusers
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
    • F05D—INDEXING SCHEME FOR ASPECTS RELATING TO NON-POSITIVE-DISPLACEMENT MACHINES OR ENGINES, GAS-TURBINES OR JET-PROPULSION PLANTS
    • F05D2250/00—Geometry
    • F05D2250/50—Inlet or outlet
    • F05D2250/52—Outlet

Definitions

  • the present invention relates to a turbomachinery having diffuser vanes which efficiently convert kinetic energy of the fluid flowing out of the impeller to a static pressure.
  • diffuser vanes can be provided downstream of the impeller to efficiently convert the kinetic energy of a turbomachinery (referred to as pumps hereinbelow) to a static pressure; however, to expand the operational range of the pump, parallel wall diffusers having no vanes are often used. Even in such a parallel wall type diffuser, since the cross sectional area increases as the radius increases from the entrance to the exit regions, the fluid velocity is decreased so that the kinetic energy is converted to a static pressure rise.
  • FIGS 1A-1G various designs of diffuser vanes have been devised and put into practice.
  • the techniques represented in Figures 1A to ID relate to reducing the fluid velocity by expanding the fluid passage so as to increase the static pressure.
  • the one shown in Figure ID relates to a known technique disclosed in a patent (USP3,973,872 ) in which the shape of the diffuser vanes is defined.
  • This patent defines a configuration of the vanes so as to have a linear section with a constant angle from the entrance to a center region and the curved section having a second order curve.
  • a velocity diagram shown by the dotted lines in Figure 2 indicates that the circumferential velocity component at the exit is lowered and consequently the pressure can be increased. Since the pressure increase is based on such a principle, there is no need to form fluid passages between the vanes, so that a throat section is not formed and thus, it has been considered that a wider operating range can be achieved without choking phenomenon.
  • vane loads are defined in terms of the pressure or velocity distribution (or Mach numbers at high velocity) on the vanes .
  • Figure 3 shows a distribution of Mach numbers on the vanes of the conventional design illustrated in Figure IE.
  • a large pressure difference is present near the leading edge between the pressure and suction surfaces of the vanes.
  • the conventional design shown in Figure 3 produces a large reduction in velocity on the suction surface, and the boundary layer becomes susceptible to separation.
  • boundary layers are formed on the casing wall surface at the upstream of the vane, and horseshoe vortex flow is produced at the root of the vanes on the pressure side and suction side surfaces when the fluid flow having boundary layer collides with the vane. The horseshoe vortex flows rise off the wall and move downstream.
  • the present invention presents a design of diffuser vanes which is neither the conventional design forming a fluid passage between the vanes nor comprising a small number of airfoil shaped vanes.
  • the present invention comprises a centrifugal or mixed flow type turbomachine comprising: a pressurizing section having a centrifugal or mixed flow type impeller housed in a casing; and a diffuser section having a plurality of and not more than ten diffuser vanes disposed in an outer circumferential area of the pressurizing section; wherein each the diffuser vane is formed so that a vane angle defined as an angle between a camber line of the diffuser vane and a circumferential direction at any point along the diffuser vane has a smaller rate of change with reference to a non-dimensional meridional length in a fore part region of the diffuser section than in an after part section of the diffuser section.
  • the present diffuser reduces the energy loss at high flow rates to expand the operating range of the turbomachinery to the higher flow rate region.
  • the vanes are shaped such that the load applied is low on the fore half of the vane (which is susceptible to form a horseshoe vortex) , thereby preventing increase in pressure loss by causing the movement of the vortex from the pressure side to the suction side. It is preferable to set a boundary point between the fore part section and the after part section at less than 0.5 of the non-dimensional meridional length, the vane angle is substantially constant in the fore part section and varies along a curve of at least second order in the after part section. And further, the boundary point can be set at 0.3 ⁇ 0.1 of the non-dimensional meridional length.
  • a vane angle of the diffuser vane at the exit of the diffuser section is determined so that an entrance/exit difference of products of a radius and circumferential velocity component is at 0.25 to 0.4.
  • FIGS 1A-1G show improvements of diffuser vanes in the conventional turbomachinery
  • Figure 2 is an illustration for explaining principle of improvements of diffuser vanes in the conventional turbomachinery
  • Figure 3 is a graph showing a distribution of Mach numbers on the vanes of the conventional design illustrated in Figure IE;
  • Figure 4 shows a horseshoe vortex flow formed on the diffuser vane wall surfaces
  • Figure 5 is a cross sectional view showing shape of a diffuser vane according to an embodiment of the present invention
  • Figure 6 shows vane angles at the camber line with reference to the entrance angle
  • Figure 7 shows a distribution of Mach number on the pressure side and the suction side surfaces of the diffuser vane with reference to the non-dimensional meridional length
  • Figure 8 shows computed results of velocity distribution at the suction of the vane shown in Figure 7 according to viscous flow analyses
  • Figure 9 shows the computed results of velocity distribution at the suction of the vane for conventional load distributions ;
  • Figure 10 shows contour diagrams formed by constant-loss lines in a 7-vaned diffuser operated at 80, 100 and 120 % of the design flow rate;
  • Figure 11 shows contour graphs formed by constant-loss lines in a 11-vaned diffuser operated at 80, 100 and 120 % of the design flow rate;
  • Figure 12 shows the effects of the number of vanes on the static pressure recovery coefficient and on the total pressure loss at 80, 100, 110 and 120 % of the design flow rate;
  • Figure 13 is a cross sectional view of a single shaft, multi-stage compressor;
  • Figure 14 shows a graph of static pressure recovery coefficient vs. angular momentum per unit flow volume at the entrance/exit of the diffuser vane for various ratios of the radii of the entrance/exit of the diffuser;
  • Figure 15 shows a graph of static pressure recovery coefficient vs . ratios of impeller exit diameter to diffuser vane thickness
  • Figure 16 shows the static pressure recovery coefficient at the diffuser inlet and exit when changing vane attachment radius ratio
  • Figure 17 shows the total pressure loss coefficient for various setting vane attachment radius ratio
  • Figures 18 to 22 show the velocity vector diagrams on the suction side surface of the vanes when the vane attachment radius ratio is altered.
  • Figure 23 shows the vane lengths along the meridian over the entire separation region in the cases of Figures 18 to 22.
  • Figure 5 shows the vanes of the present invention.
  • the vane is designed so that the fore part of the center line of the vane (camber line) has a constant angle relative to the circumferential direction, and the after part varying angle relative to the circumferential direction by following a second (or higher) order curve.
  • Figure 6 shows a graph with the x-axis relating non- dimensional length of the diffuser vanes , and the y-axis relating the vane angle at the camber line with reference to the entrance angle.
  • This graph shows that in the fore part, up to about non-dimensional meridional distance of 0.3, the vane angle is about constant, and the loading is low on that area of the vane.
  • the curve is of at least a second degree so that the load is increased and the pressure is increased. Therefore, the vane of the present invention has an angle changing point X shown at Figure 5 and 6.
  • Figure 7 shows a distribution of Mach number on the pressure side and the suction side surfaces of the diffuser vane.
  • the distribution corresponds to a pressure distribution on the vane surface, indicating the loading condition on the vane. From this figure, it can be seen that the fore part has a low loading and the loading increases towards the tail end of the vane towards the exit side. Because the load on the vane is low in the inlet region, pressure loss increase due to the movement of horseshoe vortex flow on the pressure side towards the suction side is prevented. In the boundary layer produced at the suction side, if the main flow decelerate too much, a flow separation can occur.
  • Figure 8 shows the computed results of velocity distribution at the suction surface of the vane shown in Figure 7 according to viscous flow analyses, which show that the separation region exists only in a small section of the after half of the vane.
  • Figure 3 the results from a diffuser designed to generate the conventional load distribution (refer to Figure 3) are shown in Figure 9, and it can be seen that separation starts from a mid-section of the vane, which is a reason for lowering the pump performance.
  • the number of vanes required is roughly twenty vanes in a passage type diffuser or more than ten vanes even in a diffuser with low-solidity vanes .
  • Figure 10 shows contour diagrams obtained by drawing constant energy loss lines in a 7-vaned diffuser at the flow rates of 80, 100 and 120 % of the design flow rate.
  • Figure 11 shows the results for a 11-vaned diffuser.
  • the flow rate is equal to 80 % of the design flow rate, a large region of energy loss, thought to be the result of separation of the boundary layer, can be observed on the suction surface side, conversely, at a 120 % flow rate, the loss region is observed on the pressure surface side.
  • Figure 12 shows the results of comparing the performance of a present pump operated at 80, 100, 110 and 120 % flow rates, in terms of the number of vanes plotted on the horizontal axis, and the static pressure recovery coefficient and the total pressure plotted on the vertical axis. It can be seen that, at 80 and 100%, the effects of the number of vanes are not so obvious , but at 120 %, the performance is superior for the pump with lesser number of vanes. Considering the performances at 80 and 100 % in addition, it can be understood that less than ten vanes are preferable for this type of pump.
  • the circumferential angle at the design flow rate is determined.
  • a product of the circumferential component of the absolute velocity and the radius, r.Cu which represents an angular momentum per unit flow rate, should be made zero at the diffuser exit by employing the vanes and deflecting the flow.
  • a return path is generally provided to guide the flow to the next stage in a downstream location of the diffuser.
  • Figure 14 summarizes the various aspects of diffuser performance.
  • the vertical axis represents the static pressure recovery coefficient and the horizontal axis represents angular momentum per unit flow rate at entrance/exit of the diffuser and the results are grouped under different ratios of the radii at the entrance/exit of the diffuser. For all the radius ratios, it can be seen that optimum design would be to generate a difference in the angular momentum in a range of 0.25 to 0.4.
  • the present invention has put an emphasis on the importance of the velocity distribution on the vane surface in order to determine the diffuser performance as has been demonstrated above.
  • airfoil-shaped vanes have generally been used.
  • the velocity at the vane entrance becomes high because of large thickness of the vanes, and it was found that the performance suffered. Therefore, it was necessary to select optimum vane thickness to prevent the enhanced velocity distribution at the vane entrance.
  • Figure 15 shows the results of investigation of the variation of the static pressure recovery coefficient (vertical axis) with the ratio of the diffuser vane thickness t to the exit diameter of the impeller d 2 (horizontal axis) for different values of the vane thickness . It has been found that the smaller the ratio the better the performance, and the ratio less than 0.02 is recommendable. However, there is a certain lower limitation due to structural or manufacturing reasons.
  • Figure 16 shows the static pressure recovery coefficient at the diffuser inlet and exit when changing vane attachment radius ratio r 5 /r 2 , which is a ratio of radius r 5 of the rear edge of the vane to the radius r 2 of the exit of the impeller (refer to Figure 13), for a constant solidity. Also shown are the results from vane-less diffuser. This shows that the present invention gave at the vane attachment radius ratio of about 1.5 the same pressure recovery as that given by the vane-less diffuser at the radius ratio of 1.8.
  • Figure 17 shows the total pressure loss coefficient for various setting vane attachment radius ratio, and when the ratio is between 1.45 to 1.7, the separation at the diffuser exit is the least, thus providing low loss.
  • Figures 18 to 22 show the velocity vector diagrams on the suction side surface of the vanes when the vane attachment radius ratio is altered. When the ratio is between 1.5 and 1.6, the separation at the diffuser exit is the least and good performances can be obtained.
  • Figure 23 shows the vane lengths along the meridian over the entire separation region in the cases of Figures 18 to 22. From the results, it can be seen that the best performance is obtained when the setting radius ratio at the diffuser exit is between 1 .45 to 1 . 7 .
  • the vane angle from the inlet to the center region of the vane has a constant gradient curve along the camber line from the entrance to the radius ratio 0.3, and the after part of the vane follows a more than second degree curve.
  • the result is that the pressure difference at the inlet region between the pressure side and suction side of the vane is small thus enabling to prevent the formation of vortex flow flowing from the pressure side to the suction side of the vane.
  • the velocity distribution at the suction side of the vane is about constant so that the boundary layer experiences favorable condition at the vane wall surfaces, so minimizing flow separation and improving the performance of the turbomachinery.

Landscapes

  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Structures Of Non-Positive Displacement Pumps (AREA)

Abstract

L'invention concerne une pompe centrifuge ou à pignons, dont les aubes de diffuseur ne sont pas conçues comme les aubes classiques formant un passage de liquide entre elles et ne sont pas constituées d'un petit nombre d'aubes en forme d'ailettes. La turbomachine du type à pignons ou centrifuge comporte une partie de pressurisation possédant une roue centrifuge ou hélicocentrifuge logée dans un carter, et une partie diffuseur possédant au plus dix aubes de diffuseur, placées dans une zone circonférentielle extérieure de la partie de pressurisation. Chaque aube de diffuseur est formée de sorte que l'angle de l'aube défini comme l'angle entre la ligne moyenne de l'aube et l'axe circonférentiel en n'importe quel point le long de l'aube, présente une vitesse de variation inférieure, par rapport à une longueur méridionale non dimensionnelle se trouvant dans la partie avant de la partie diffuseur, à la vitesse de variation dans une partie arrière de la partie diffuseur.
PCT/GB1998/001555 1998-05-28 1998-05-28 Turbomachine WO1999061801A1 (fr)

Priority Applications (1)

Application Number Priority Date Filing Date Title
PCT/GB1998/001555 WO1999061801A1 (fr) 1998-05-28 1998-05-28 Turbomachine

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
PCT/GB1998/001555 WO1999061801A1 (fr) 1998-05-28 1998-05-28 Turbomachine

Publications (1)

Publication Number Publication Date
WO1999061801A1 true WO1999061801A1 (fr) 1999-12-02

Family

ID=10825850

Family Applications (1)

Application Number Title Priority Date Filing Date
PCT/GB1998/001555 WO1999061801A1 (fr) 1998-05-28 1998-05-28 Turbomachine

Country Status (1)

Country Link
WO (1) WO1999061801A1 (fr)

Cited By (6)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP1431586A1 (fr) * 2002-12-17 2004-06-23 Nuovo Pignone Holding S.P.A. Diffuseur pour compresseur centrifuge
EP1568891A1 (fr) * 2002-12-04 2005-08-31 Mitsubishi Heavy Industries, Ltd. Diffuseur pour compresseur centrifuge et son procede de production
CN102042266A (zh) * 2009-10-22 2011-05-04 株式会社日立工业设备技术 涡轮式流体机械
WO2012019650A1 (fr) * 2010-08-12 2012-02-16 Nuovo Pignone S.P.A. Aube de diffuseur radiale destinée à des compresseurs centrifuges
US10527059B2 (en) 2013-10-21 2020-01-07 Williams International Co., L.L.C. Turbomachine diffuser
CN113931879A (zh) * 2021-10-19 2022-01-14 中国科学院工程热物理研究所 一种高压比离心压气机径向叶片扩压器参数化设计方法

Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE275924C (fr) *
FR440487A (fr) * 1912-02-22 1912-07-11 Franz Lawaczeck Turbomachine
US3973872A (en) 1975-08-01 1976-08-10 Konstantin Pavlovich Seleznev Centrifugal compressor
DE3926152A1 (de) * 1989-07-19 1991-01-24 Escher Wyss Gmbh Radialkompressor
DE4309479A1 (de) * 1993-03-24 1994-09-29 Wilo Gmbh Radialkreiselpumpe

Patent Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE275924C (fr) *
FR440487A (fr) * 1912-02-22 1912-07-11 Franz Lawaczeck Turbomachine
US3973872A (en) 1975-08-01 1976-08-10 Konstantin Pavlovich Seleznev Centrifugal compressor
DE3926152A1 (de) * 1989-07-19 1991-01-24 Escher Wyss Gmbh Radialkompressor
DE4309479A1 (de) * 1993-03-24 1994-09-29 Wilo Gmbh Radialkreiselpumpe

Cited By (10)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP1568891A1 (fr) * 2002-12-04 2005-08-31 Mitsubishi Heavy Industries, Ltd. Diffuseur pour compresseur centrifuge et son procede de production
EP1568891A4 (fr) * 2002-12-04 2006-01-04 Mitsubishi Heavy Ind Ltd Diffuseur pour compresseur centrifuge et son procede de production
EP1431586A1 (fr) * 2002-12-17 2004-06-23 Nuovo Pignone Holding S.P.A. Diffuseur pour compresseur centrifuge
CN102042266A (zh) * 2009-10-22 2011-05-04 株式会社日立工业设备技术 涡轮式流体机械
WO2012019650A1 (fr) * 2010-08-12 2012-02-16 Nuovo Pignone S.P.A. Aube de diffuseur radiale destinée à des compresseurs centrifuges
CN103154526A (zh) * 2010-08-12 2013-06-12 诺沃皮尼奥内有限公司 用于离心压缩机的径向扩散器导叶
RU2581686C2 (ru) * 2010-08-12 2016-04-20 Нуово Пиньоне С.п.А. Радиальная диффузорная лопатка для центробежных компрессоров
US10527059B2 (en) 2013-10-21 2020-01-07 Williams International Co., L.L.C. Turbomachine diffuser
CN113931879A (zh) * 2021-10-19 2022-01-14 中国科学院工程热物理研究所 一种高压比离心压气机径向叶片扩压器参数化设计方法
CN113931879B (zh) * 2021-10-19 2023-12-22 中国科学院工程热物理研究所 一种高压比离心压气机径向叶片扩压器参数化设计方法

Similar Documents

Publication Publication Date Title
EP1082545B1 (fr) Turbines pour turbomachines
US6203275B1 (en) Centrifugal compressor and diffuser for centrifugal compressor
US7229248B2 (en) Blade structure in a gas turbine
JP5233436B2 (ja) 羽根無しディフューザを備えた遠心圧縮機および羽根無しディフューザ
US5228832A (en) Mixed flow compressor
KR101286344B1 (ko) 경사식 원심 압축기 에어포일 확산기
JP5608062B2 (ja) 遠心型ターボ機械
JP5678066B2 (ja) 軸流コンプレッサー用のコンプレッサーブレード
EP1046783A2 (fr) Aubes de turbine
JPH0646035B2 (ja) 多段遠心圧縮機
US20170009781A1 (en) Turbomachine component or collection of components and associated turbomachine
EP1260674B1 (fr) Turbine et aube de turbine
JPH086711B2 (ja) 遠心圧縮機
JP6034162B2 (ja) 遠心式流体機械
JP2009057959A (ja) 遠心圧縮機とその羽根車およびその運転方法
JPH11257272A (ja) 多段遠心ターボ機械
US11982204B2 (en) Turbomachine part or assembly of parts
WO1999061801A1 (fr) Turbomachine
JP3899829B2 (ja) ポンプ
JPH01318790A (ja) 多段ポンプの水返し羽根
JP3005839B2 (ja) 軸流タービン
JP2000204903A (ja) 軸流型タ―ビン
JPH1182389A (ja) ターボ形流体機械
JPH0972222A (ja) 軸流圧縮機
RU2794951C2 (ru) Лопатка газотурбинного двигателя с правилом максимальной толщины с большим запасом прочности при флаттере

Legal Events

Date Code Title Description
AK Designated states

Kind code of ref document: A1

Designated state(s): CN JP KR US

AL Designated countries for regional patents

Kind code of ref document: A1

Designated state(s): AT BE CH CY DE DK ES FI FR GB GR IE IT LU MC NL PT SE

DFPE Request for preliminary examination filed prior to expiration of 19th month from priority date (pct application filed before 20040101)
121 Ep: the epo has been informed by wipo that ep was designated in this application
NENP Non-entry into the national phase

Ref country code: KR

122 Ep: pct application non-entry in european phase