EP0270723A1 - Rotor pour une turbomachine radiale - Google Patents

Rotor pour une turbomachine radiale Download PDF

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Publication number
EP0270723A1
EP0270723A1 EP86850424A EP86850424A EP0270723A1 EP 0270723 A1 EP0270723 A1 EP 0270723A1 EP 86850424 A EP86850424 A EP 86850424A EP 86850424 A EP86850424 A EP 86850424A EP 0270723 A1 EP0270723 A1 EP 0270723A1
Authority
EP
European Patent Office
Prior art keywords
impeller
pressure end
high pressure
channel
blades
Prior art date
Legal status (The legal status 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 status listed.)
Withdrawn
Application number
EP86850424A
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German (de)
English (en)
Inventor
Paul Jacques Jean Frigne
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Institut Cerac SA
Original Assignee
Institut Cerac SA
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 Institut Cerac SA filed Critical Institut Cerac SA
Priority to EP86850424A priority Critical patent/EP0270723A1/fr
Publication of EP0270723A1 publication Critical patent/EP0270723A1/fr
Withdrawn legal-status Critical Current

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01D—NON-POSITIVE DISPLACEMENT MACHINES OR ENGINES, e.g. STEAM TURBINES
    • F01D5/00—Blades; Blade-carrying members; Heating, heat-insulating, cooling or antivibration means on the blades or the members
    • F01D5/02—Blade-carrying members, e.g. rotors
    • F01D5/04—Blade-carrying members, e.g. rotors for radial-flow machines or engines
    • F01D5/043—Blade-carrying members, e.g. rotors for radial-flow machines or engines of the axial inlet- radial outlet, or vice versa, type
    • F01D5/048—Form or construction
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D17/00—Radial-flow pumps, e.g. centrifugal pumps; Helico-centrifugal pumps
    • F04D17/08—Centrifugal pumps
    • F04D17/10—Centrifugal pumps for compressing or evacuating
    • F04D17/105—Centrifugal pumps for compressing or evacuating with double suction
    • 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
    • F04D29/26—Rotors specially for elastic fluids
    • F04D29/28—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps
    • F04D29/284—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps for compressors

Definitions

  • the present invention relates to fluid machines i.e. machines for conversion of fluid energy to mechanical energy or vice versa. More precisely the invention relates to an impeller in such a machine.
  • radial turbomachine stands for a turbomachine in which the flow at the high pressure end of the impeller is in a plane orthogonal to the axis of the impeller. In the following description this designation is also used for machines in which the flow at the outlet of the impeller is slightly deviating from this plane.
  • This type of machines includes centrifugal compressors and pumps as well as radial inflow turbines.
  • inventive idea which will be described below is generally applicable to all the different radial turbomachines it will be discussed mainly in connection with dynamic radial compressors. Therefore, the discussion of the background art will also mainly be restricted to such compressors.
  • This type of compressor operates according to the principle of dynamic compression i.e. dynamic energy, or head, is created by means of an impeller and subsequently converted into pressure energy by means of a diffusion process which takes place partly within the flow channels of the impeller and then in a down-stream stationary diffuser.
  • the dual inlet impeller is another known concept within the technical field.
  • This type of impeller could be considered comprising two ordinary impellers, designed for opposite rotational direction, put together in a back-to-back arrangement on the same axis.
  • This impeller design eliminates the axial thrust problem due to the symmetry.
  • the hub portion of such an impeller i.e. the body portion supporting the impeller blades has at the periphery of the impeller the form of a disc.
  • This disc portion creates a wake in the middle of the combined output flow from the two impeller halves. Due to the fact that this wake occurs at the high speed section of the flow before it enters the diffuser, the shear and mixing losses thus created are considerable.
  • the problem created by this wake has partly been solved in an impeller design according to EP-A-0153066. Aerodynamically shaped saw tooth - ­profiled reliefs are provided in the hub radial portion between adjacent blade tips. The two flows could therefore mix when they are still inside the impeller and the impact of the wake is decreased.
  • One object of the invention is to provide an impeller which solves the problem with the intrinsic incompatibility between the optimum blade spacing at the inlet and the outlet in order to improve the overall diffusion in the impeller channels without introducing splitter vanes.
  • Another object of the invention is to provide an impeller which allows an extension of the stable flow range according to the principles disclosed in the FLYNN-WEBER paper mentioned above. At the same time the drawbacks inherent in their design created by the thickness of the blade trailing edges should be avoided.
  • a further object of the invention is to provide a solution to the wake-problem at the high speed section of previously known impellers of the dual inlet type.
  • One advantage resides in the significant improvement of the hydraulic section of the impeller channels as compared to conventional back-to-back configurations, which results in a considerable reduction of the friction inside the channels.
  • Unshrouded impellers according to the invention show the advantage of decreased leakage flow from pressure to suction side of each blade compared to known dual inlet impellers.
  • a further advantage resides in the fact that the hub and shroud extremities of each blade of the impeller are linked to the main core of the impeller which will significantly reduce blade bending during operation or machining compared to unshrouded conventional or back-to-back wheels.
  • FIG. 1 shows the main features of an impeller according to the invention.
  • P.F. Flynn and H.G. Weber in the paper mentioned above. They introduced the concept of gradually thickening the blades from the low pressure end towards the high pressure end in an ordinary single sided impeller leading to a blunt trailing edge design. As mentioned above they observed some positive effects mainly in connection with the extension of the stable flow range. However, the design was linked to a significant decrease in the overall efficiency compared to ordinary single sided thin-bladed impellers.
  • the basic idea of the present invention resides in the design of a dual inlet radial impeller in a back-to-back configuration of two blunt trailing edge impeller halves with common high pressure end.
  • the channel outlet tangential width i.e. the outlet blade-to-blade width, which is the same for the two impeller halves, must be smaller than the blunt width.
  • the high pressure ends of the two impeller halves are intermeshed in such a way that the through-flow areas of the one impeller half, i.e. the channel openings of that half, are embedded in or entirely substitute the blunt portions of the other half. Seen from the low pressure ends the impeller-halves are designed for clockwise and anti-clockwise rotation respectively.
  • the flow follows the channels between adjacent blades 4, 4 ⁇ from the low pressure ends 2, 2 ⁇ to the common high pressure end 5.
  • the designations 1, 1 ⁇ refer to the shaded shroud portions of the left hand and right hand impeller halves respectively.
  • the hub side 3 of the channels in the left hand impeller half coincides at the outlet with the shroud portions 1 ⁇ of the right hand impeller half and vice versa.
  • the divergency of the flow channel in the blade-to-blade plane can be obtained when the outlet-to-inlet diameter ratio of the impeller exceeds a critical value, which depends on the number of blades and the blade lean angle.
  • the curves in Figure 2 are the limit lines for a nested channel impeller with strictly parallel blades.
  • the hub inlet diameter is normally chosen on basis of stress considerations, and for the following discussion we assume that it equals about 20% of the impeller outlet tip diameter. We further assume that the blades are still radial and straight. For a 3-dimensional impeller of course this assumption cannot be valid, as the blade shape turns out to be rather complex when turning from axial to radial. However, these simplifications are only meant to explain the concept's basic ideas with simple relations.
  • the specific speed gives a very universal relation between the capacity, head and rotational speed of a turbocompressor. According to empirical rules, a well designed conventional impeller should have a specific speed within the range from 80 to 110 when expressed according to Baljé (cf. Baljé, O., "A study of design criteria and matching of turbomachines; Part B: Compressor and pump performance and matching of turbocomponents"; The American Society of Mechanical Engineers, paper 60 WA 231, 1961).
  • Figure 3 shows the maximum specific speed, at which the blade-to-blade width of the channels are divergent over the whole blade height, in function of the blade outlet lean angle and the number of blades.
  • the specific speed in this case is related to the volume flow at the inlet eye of one impeller half.
  • the points on the curves indicate when the shroud section becomes parallel. It is evident that the divergency condition can easier be obtained at the hub than at the shroud section of the channels, as the outlet-to-inlet diameter ratio is bigger.
  • shaded shroud portions 1 for instance, between adjacent flow channels in the left impeller half represent a favourable dam to prevent blade tip leakage from the pressure to the suction sides of the blades. The same goes for portions 1 ⁇ on the right hand side.
  • the blades are typically overhang which means that they have one extremity which is free to move and another extremity which is bound to the hub of the impeller.
  • This arrangement creates potential modes of vibration which of course are related to fatigue problems.
  • the freedom of the blade extremities is reduced as the blades are linked to the main core of the impeller which means that the resonant frequency of the blades will be advantageously raised.
  • the blade bending during machining compared to unshrouded conventional single or back-to-back wheels will be significantly reduced.
  • the flow channels are at the low pressure ends, as in ordinary dual inlet impellers, substantially axial and at the common high pressure end along the periphery of the impeller substantially radial. As for ordinary dual inlet impellers the thrust is balanced and there is no need for axial high pressure seals.
  • trailing edge structure resulting from the alternance of hub and shroud flow will improve the symmetry of flow at the inlet of the diffuser. And this same trailing edge structure is likely to avoid wake interaction instabilities occasionally observed in back-to-back impellers.
  • Figure 4a, b shows a further possibility, introduced by the inventive concept, to control the inlet flow to the diffuser 6. Due to the fact that the hub high pressure edges 3 have a certain thickness these edges could be designed to produce either a purely radial or slightly converging flow to improve the diffuser performance or reduce the exit angle difference between the hub flow and adjacent channel shroud flows.
  • the hub angle at the channel exit i.e. the angle between hub and shroud sections 3, 1 ⁇ and 3 ⁇ , 1, is not necessarily constant over the width of the channel. This angle could e.g. be smaller at the center portion of the channel opening making the hub section 3, 3 ⁇ concave over part of the high pressure portion of the channel. Such a configuration could release some of the geometrical constraints in the intermeshed part of the channels, and introduces an additional parameter for optimising the variation of the cross sectional area.
  • Figure 4c shows a cross section through a channel at the high pressure end in which the flow by means of adding material at the hub side is made slightly diagonal.
  • a substantional improvement, in relation to the basic nested channel impeller described above, can be obtained by the introduction of a rake angle. If one considers the intersection line between the blade and a coaxial cylinder through the impeller radial part, the rake angle is defined as the angle between this intersection line and the axial direction. A positive rake angle gives a channel opening at the high pressure side which is divergent in the direction from the hub to the shroud side. An impeller having such a rake angle is shown in Figure 5. The opposite sign of the rake angle gives an impeller according to Figure 7.
  • the divergency condition for the blade-to-blade width of the channel is, according to the above, more difficult to achieve for the shroud side than for the hub side of the channel.
  • a positive rake angle it will be possible to equalise the divergency at the two channel sides. It would for example be possible to maintain the trapezoidal shape of the channel all the way from the inlet to the outlet. In other words, by controlling the rake angle the distribution of the diffusion process over the channel cross section, in the hub to shroud direction, could be controlled.
  • a positive rake angle for instance, increases the blade loading at the hub section and decreases it at the shroud section of each flow channel. In this way, it is possible to select a rake angle having and equalising effect on the blade loading at the shroud section of one impeller side and the blade loading at the hub section of the opposite impeller side. This results in a smoother operation of the compressor.
  • Another type of recesses in the form of cavities in the same surfaces 1, 1 ⁇ , communicating with the flow, could be designed to delay stall and surge in the same way as casing treatment is employed in conventional compressor technology.
  • FIG. 7 An impeller having a negative rake angle is shown in Figure 7.
  • This type impeller could be advantageous in high pressure ratio and/or high pressure level applications due to the fact that the dam to prevent the blade tip leakage across the blades from the pressure to the suction sides of the blades has been further increased, in that the shroud surfaces 1, 1 ⁇ are made broader.
  • the dam function of these surfaces 1, 1 ⁇ could, for any rake angle, be increased by means of the arrangement of labyrinth-type sealing elements on the surface. The sealing effect could in this way be increased both across the blade and along the blade. This could also contribute to an improvement of the loading at the blade tip.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Structures Of Non-Positive Displacement Pumps (AREA)
EP86850424A 1986-12-05 1986-12-05 Rotor pour une turbomachine radiale Withdrawn EP0270723A1 (fr)

Priority Applications (1)

Application Number Priority Date Filing Date Title
EP86850424A EP0270723A1 (fr) 1986-12-05 1986-12-05 Rotor pour une turbomachine radiale

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
EP86850424A EP0270723A1 (fr) 1986-12-05 1986-12-05 Rotor pour une turbomachine radiale

Publications (1)

Publication Number Publication Date
EP0270723A1 true EP0270723A1 (fr) 1988-06-15

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ID=8196533

Family Applications (1)

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EP86850424A Withdrawn EP0270723A1 (fr) 1986-12-05 1986-12-05 Rotor pour une turbomachine radiale

Country Status (1)

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EP (1) EP0270723A1 (fr)

Cited By (11)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
FR2728314A1 (fr) * 1994-12-14 1996-06-21 Bosch Gmbh Robert Ventilateur radial a simple flux dont la construction permet de prevoir un ventilateur a double flux
US6398853B1 (en) 1998-12-16 2002-06-04 Quest Air Gases Inc. Gas separation with split stream centrifugal turbomachinery
GB2398842A (en) * 2002-11-15 2004-09-01 Visteon Global Tech Inc Cooling fan for alternator
EP2781760A1 (fr) 2011-11-17 2014-09-24 Hitachi, Ltd. Machine à fluide centrifuge
EP2843236A1 (fr) * 2013-08-27 2015-03-04 Honeywell International Inc. Roue de turbocompresseur à deux faces fonctionnellement asymétriques et diffuseur
CN104421201A (zh) * 2013-08-27 2015-03-18 霍尼韦尔国际公司 结构非对称的双侧涡轮增压器叶轮
CN105041712A (zh) * 2015-08-12 2015-11-11 苏州圆能动力科技有限公司 一种叶轮
EP3001038A1 (fr) * 2014-09-26 2016-03-30 CLAAS Selbstfahrende Erntemaschinen GmbH Machine radiale
US10718346B2 (en) 2015-12-21 2020-07-21 General Electric Company Apparatus for pressurizing a fluid within a turbomachine and method of operating the same
CN112211829A (zh) * 2020-10-12 2021-01-12 林海涛 一种水泵内置调控压力装置
US11421702B2 (en) 2019-08-21 2022-08-23 Pratt & Whitney Canada Corp. Impeller with chordwise vane thickness variation

Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1865918A (en) * 1928-06-30 1932-07-05 Junkers Hugo Impeller and method of making same
GB537916A (en) * 1940-01-02 1941-07-11 Fritz Albert Max Heppner Improvements in rotors of centrifugal compressors, turbines and the like
GB613892A (en) * 1945-07-09 1948-12-03 Bbc Brown Boveri & Cie Method of producing centrifugal blowers for supercharging internal combustion engines
DE861142C (de) * 1950-06-25 1952-12-29 Licentia Gmbh Doppelseitig beaufschlagtes Laufrad fuer Kreiselmaschinen, insbesondere fuer Radialverdichter
FR1188110A (fr) * 1957-12-04 1959-09-18 Snecma Compresseur centrifuge supersonique
US3010642A (en) * 1955-02-16 1961-11-28 Rheinische Maschinen Und App G Radial flow supersonic compressor
GB1036486A (en) * 1964-06-05 1966-07-20 Bristol Siddeley Engines Ltd Double-sided radial flow rotor in or for a centripetal turbine or centrifugal compressor
GB2027816A (en) * 1978-08-15 1980-02-27 Sugiura E Centrifugal pump

Patent Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1865918A (en) * 1928-06-30 1932-07-05 Junkers Hugo Impeller and method of making same
GB537916A (en) * 1940-01-02 1941-07-11 Fritz Albert Max Heppner Improvements in rotors of centrifugal compressors, turbines and the like
GB613892A (en) * 1945-07-09 1948-12-03 Bbc Brown Boveri & Cie Method of producing centrifugal blowers for supercharging internal combustion engines
DE861142C (de) * 1950-06-25 1952-12-29 Licentia Gmbh Doppelseitig beaufschlagtes Laufrad fuer Kreiselmaschinen, insbesondere fuer Radialverdichter
US3010642A (en) * 1955-02-16 1961-11-28 Rheinische Maschinen Und App G Radial flow supersonic compressor
FR1188110A (fr) * 1957-12-04 1959-09-18 Snecma Compresseur centrifuge supersonique
GB1036486A (en) * 1964-06-05 1966-07-20 Bristol Siddeley Engines Ltd Double-sided radial flow rotor in or for a centripetal turbine or centrifugal compressor
GB2027816A (en) * 1978-08-15 1980-02-27 Sugiura E Centrifugal pump

Cited By (14)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
FR2728314A1 (fr) * 1994-12-14 1996-06-21 Bosch Gmbh Robert Ventilateur radial a simple flux dont la construction permet de prevoir un ventilateur a double flux
US6398853B1 (en) 1998-12-16 2002-06-04 Quest Air Gases Inc. Gas separation with split stream centrifugal turbomachinery
GB2398842A (en) * 2002-11-15 2004-09-01 Visteon Global Tech Inc Cooling fan for alternator
GB2398842B (en) * 2002-11-15 2005-04-06 Visteon Global Tech Inc Alternator fan
EP2781760A1 (fr) 2011-11-17 2014-09-24 Hitachi, Ltd. Machine à fluide centrifuge
CN104421201A (zh) * 2013-08-27 2015-03-18 霍尼韦尔国际公司 结构非对称的双侧涡轮增压器叶轮
EP2843236A1 (fr) * 2013-08-27 2015-03-04 Honeywell International Inc. Roue de turbocompresseur à deux faces fonctionnellement asymétriques et diffuseur
US10006290B2 (en) 2013-08-27 2018-06-26 Honeywell International Inc. Functionally asymmetric two-sided turbocharger wheel and diffuser
US10233756B2 (en) 2013-08-27 2019-03-19 Garrett Transportation I Inc. Two-sided turbocharger wheel with differing blade parameters
EP3001038A1 (fr) * 2014-09-26 2016-03-30 CLAAS Selbstfahrende Erntemaschinen GmbH Machine radiale
CN105041712A (zh) * 2015-08-12 2015-11-11 苏州圆能动力科技有限公司 一种叶轮
US10718346B2 (en) 2015-12-21 2020-07-21 General Electric Company Apparatus for pressurizing a fluid within a turbomachine and method of operating the same
US11421702B2 (en) 2019-08-21 2022-08-23 Pratt & Whitney Canada Corp. Impeller with chordwise vane thickness variation
CN112211829A (zh) * 2020-10-12 2021-01-12 林海涛 一种水泵内置调控压力装置

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