EP0476499A1 - Radialrad für eine Turbomaschine - Google Patents

Radialrad für eine Turbomaschine Download PDF

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Publication number
EP0476499A1
EP0476499A1 EP91115357A EP91115357A EP0476499A1 EP 0476499 A1 EP0476499 A1 EP 0476499A1 EP 91115357 A EP91115357 A EP 91115357A EP 91115357 A EP91115357 A EP 91115357A EP 0476499 A1 EP0476499 A1 EP 0476499A1
Authority
EP
European Patent Office
Prior art keywords
blades
radial wheel
hub
radial
contour
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Withdrawn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
EP91115357A
Other languages
German (de)
English (en)
French (fr)
Inventor
Andreas Dr. Fiala
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.)
Fiala Andreas Dr
Original Assignee
MTU Motoren und Turbinen Union Muenchen GmbH
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 MTU Motoren und Turbinen Union Muenchen GmbH filed Critical MTU Motoren und Turbinen Union Muenchen GmbH
Publication of EP0476499A1 publication Critical patent/EP0476499A1/de
Withdrawn legal-status Critical Current

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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
    • F04D29/26—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
    • 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
    • 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
    • F04D29/30—Vanes
    • 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
    • F05D2200/00—Mathematical features
    • F05D2200/20—Special functions
    • F05D2200/25—Hyperbolic trigonometric, e.g. sinh, cosh, tanh
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y10—TECHNICAL SUBJECTS COVERED BY FORMER USPC
    • Y10S—TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y10S416/00—Fluid reaction surfaces, i.e. impellers
    • Y10S416/02—Formulas of curves

Definitions

  • the invention relates to a radial wheel for a turbomachine with a hub and blades distributed on the outer circumference of the hub.
  • the main disadvantage of a radial compressor is that it only achieves an isentropic step efficiency of around 80-84%. In addition to the growth and detachment of the boundary layer in the housing area, this is due to the fact that the friction losses between the fluid and the radial wheel and the adjoining diffuser are significantly greater than with the axial compressor.
  • a radial wheel is also understood to mean an impeller in which the direction of flow at the outlet is not purely radial, but also has an axial component.
  • a disadvantage of conventional radial wheels is that the surfaces of the wheel against which the fluid flows are relatively large, which increases the friction-related flow losses.
  • the meridian cut contour of the hub outer surface is a chain line.
  • the differential equation for minimal areas is from Dr. Bernhard Baule, the mathematics of the natural scientist and engineer, part 2, published by Harri Deutsch, Frankfurt / Main, ⁇ 11, p. 46, 1979.
  • the constants c and d are determined with the help of two parameters from the 4 possible parameters, flow angle, inclination angle, axial distance or radius for the boundary conditions at the entry and exit of the radial wheel.
  • Both surfaces that is to say the hub outer surface and the envelope surface of the outer contour of the blades on the housing, are preferably provided with a chain line as a meridian cut contour in order to obtain the smallest possible surface (minimum surface) both towards the housing and towards the hub. If necessary, however, only one of the two surfaces of the hub outer surface or the envelope surface can be shaped according to the invention, while the other surface is of conventional design.
  • the entry or exit area of the outer hub surface or the outer surface of the outer contour of the blades of a radial wheel can also deviate in its contour from the ideal chain line in order to meet certain entry or exit requirements.
  • Such screw surfaces are solutions of the differential equation for minimal surfaces.
  • the vector function according to the invention in a Cartesian coordinate system describes the surfaces formed by the blades in order to minimize the surfaces flowed against by the flow medium while maintaining predetermined, known contours for the hub-side outer surface or the envelope surface of the outer contour of the blades on the housing side and the number of blades.
  • This design of the blade surfaces as screw surfaces enables a reduction in the flow losses and thus an increase in the efficiency for a radial wheel according to the invention.
  • the blade surfaces can in turn deviate from this contour with minimal surface in partial areas, in particular in the blade inlet or outlet area, without leaving the scope of the invention.
  • the blades preferably have an approximately screw surface as the surface from the inlet to at least half the flow thread length of the radial wheel.
  • Such a design of the blade surfaces of a radial wheel advantageously uses the inventive concept to improve the efficiency.
  • a further preferred embodiment of the invention provides that the surfaces of the blades approximately have screw surfaces on the outlet side, these extending over at least half the length of the filament of the radial wheel.
  • a further advantageous embodiment of the invention provides that the spatial curves of the cutting line of the blade surface and the hub outer surface and / or the cutting line of the blade surface and the outer contour of the blades on the housing side is a chain screw line.
  • the chain screw line is a vector function with the circumferential angle (4 »as scalar variable and with where c, d, 1, and k are constants that are determined from the boundary conditions at the entry and exit of the radial wheel.
  • This design combines the advantages of the two embodiments described above in such a way that both a minimization of the surfaces on the hub and housing side can be achieved, and that at the same time the blade surfaces have minimized surfaces.
  • a greater reduction in the frictional losses can be achieved overall than with the minimal surface formation of the hub-side outer surface or the envelope surface of the housing-side outer contours of the blades or the blade surface.
  • a chain-screw line is understood to mean a spatial curve which, with the angle as an independent parameter, only depends on the angle (4 »itself. It results from a screw line in the front view and a chain line in the meridian section.
  • the number of blades in a generic radial wheel for a turbomachine, can be changed in the axial direction, the blades being arranged one behind the other in the flow direction and in each meridian normal section along the flow channel at the angle of inclination ( E ) to the radial direction with a hub radius ( R N ) and a housing radius (R G ) the number (n) of blades in the flow direction is approximately determined by the following equation for n:
  • the design according to the invention has the advantage that a minimal circumference (U) is achieved for a given flow channel cross section (A) which is limited by two blades, a hub-side and a casing-side envelope surface of the blades.
  • the number of blades preferably doubles step-by-step in certain axial points.
  • two axial points are provided at which the number of blades doubles. That is, at a first axial position, the leading edges of an equal number of shorter blades are spaced between the blades starting at the inlet. At a second axial position, this is the case again, so that in the area of the radial wheel outlet of a compressor wheel or the radial wheel inlet of a turbine wheel there are four times the number of blades as at the inlet of a compressor wheel or at the outlet of a turbine wheel. If the number of blades is doubled at three axial points, the number of blades at the outlet is eight times as high as at the inlet.
  • the person skilled in the art will determine the axial points at which the front edges of the vanes displaced to the rear are located in coordination with the other required flow properties.
  • the axial point can be provided at the position at which the optimal number of blades according to the above formula has reached twice the value of the blades that were actually present up to that point. Expediently, however, the axial point will be moved further forward in order to achieve the lowest possible loss of efficiency.
  • At least two successive axial sections are provided with blades distributed over the circumference and extending only over the axial length of one section, the leading edges of the next blade group adjoining the rear edges of the previous blade group offset in the circumferential direction.
  • the blade groups can also overlap slightly axially.
  • three or four sections lying one behind the other are provided.
  • This design has the essential advantage that, instead of doubling the number of blades, any increase in the number of blades is possible.
  • the number of blades can be gradually increased in four sections from 9 to 13 to 23 and finally to 56.
  • the blades are usually only as long as the axial section in question, i.e. there are no, or very few, blades extending over the entire radial wheel length.
  • the sections are preferably of the same length. If necessary, however, a different extent of the sections can also be provided. This development of the inventive concept enables the best possible adaptation of the number of blades, which necessarily changes in discrete steps, to the number of blades n which is optimal with regard to surface minimization according to the above equation for n.
  • the blades are made to have minimal surfaces, i.e. that the blade surfaces, or at least essential parts thereof, are designed as screw surfaces. Furthermore, it is particularly advantageous if at the same time the hub-side outer surface and / or the housing-side rotation surface is shaped in such a way that they have a chain-like contour in a meridian section. Such a radial wheel is optimized from the standpoint of frictional resistance, i.e. it has the smallest possible surface.
  • the exposed blade edges experience along a part or along the entire extent a circumferential curvature that is the same or more pronounced than the meridian curvature. This design reduces the risk of boundary layer detachments in the area of the blade tips, which also result in reduced efficiency.
  • the blades are preferably curved backwards.
  • Backward curvature means on the one hand that the direction of rotation of the impeller is opposite to the direction of rotation of a particle flowing through the impeller, and on the other hand that at the impeller outlet the peripheral component of the mean relative speed vector has the opposite direction to the peripheral speed.
  • the backward curvature has the advantage that the aerodynamic load is also reduced.
  • a backward curved radial wheel 1a is shown in the meridian section (circular projection).
  • Two blades 2a and 2b rotated in the sectional plane can be seen, wherein in the blade 2b shown below in the drawing, radial generatrix of the blade 3 which is evenly spaced are shown.
  • a flow channel outer housing 4 is provided radially outside of the blades 2a and 2b, it also being possible to provide a shroud attached to the free blade edges.
  • the blade 2a has a hub section contour 5 of the hub surface and a housing section contour 6 of the housing surface.
  • These two meridian cut contours have a curve that can be called a chain line. This means that the contour corresponds to the line that a chain suspended between points 7a and 7b or 8a and 8b would take.
  • the radial wheel 1 has three blades of different lengths, the front edges 9a, 9b and 9c of which can be seen in the meridian section plane.
  • a first group of blades extends over the entire streamline length of the radial wheel 1 b, that is to say that the front edges 9 a begin at the inlet 10 of the radial wheel 1.
  • the blades beginning with the front edge 9b are set back by a distance Z 1 from the entry of the radial wheel 1 b, twice as many such blades being provided.
  • This second group of blades ends exactly like the first group of blades at the outlet 11 of the radial wheel 1 b, so that they are shorter overall than the blades of the first group.
  • leading edges 9c of a third group of even shorter blades beginning at the axial position Z 2 are provided, of which in turn there are twice as many blades as those of the second group and thus four times as many blades as those of the first group.
  • the radial wheel 1 according to FIG. 2 is shown in a front view and in a 3-d view, respectively, with the (11) blades 2a of the first group, i.e. connect the longest blades 2a, which extend over the entire arc length of the radial wheel 1, the twice as many (22) blades 2b of the second group and the four times as many (44) blades 2c of the third group.
  • the blades 2a, 2b and 2c are in particular designed to be curved like a helix.
  • FIG. 3 shows a Cartesian coordinate system x-y-z with the two independent parameters r and ⁇ , which describe point 21 of the blade surface.
  • the backward curvature of the blades can be recognized from the fact that the positive direction 17 at point 8b of the housing section contour 6 has the opposite sign to the positive direction of rotation 18 of the impeller.
  • the direction of rotation of the impeller is opposite to the direction of rotation that a particle has to traverse along the housing section contour 6.
  • the radial wheel 1 also has the following dimensions, based on the meridian intersection points shown in FIG. 1:
  • FIG. 5 shows a meridian section of a further backward-curved radial wheel 1c, which differs from the previous versions in that it consists of four sections 12a, 12b, 12c and 12d lying axially one behind the other, the blades 13a, 13b, 13c and 13d of the respective sections extend over the relevant axial length of the sections 12a-d and are arranged axially slightly overlapping.
  • the front and rear edges of the blades 13a-d preferably run radially so that no bending moments caused by centrifugal force act in the blade root.
  • the generatrices of the blades 2a and 2b and 13a-d are radial straight lines, likewise in order to avoid bending moments.
  • the axial sections 12a-d are of equal length in the embodiment shown, can be in adaptation to other required flow conditions, however, they can also be of different lengths.
  • the radial wheel 1 according to FIG. 5 is shown in a 3-d view in FIG. 6. In the first section 12a there are nine blades 13a, in the second section 12b there are thirteen blades 13b, in the third section 12c there are twenty-three blades 13c and in the rearmost section 12d there are fifty-six blades 13d.
  • FIG. 7 shows a diagram in which the course of the surface efficiency of various radial wheels is plotted over the axial length Z, where Zo denotes the blade entry and Z the axial end of the blade.
  • the surface efficiency compares the hydraulic diameter d h y dr of a flow channel cross section (A) (meridian normal section) with the circle diameter d theo for the same cross-sectional area, since the circle is the function with the smallest circumference for (U) a given cross-sectional area.
  • the two diameters can be determined from:
  • the theoretically maximum achievable surface efficiency for the rectangular cross-section and the circular ring section is 86.6% of the circular cross-section and corresponds to the increase in circumference due to the square cross-sectional contour compared to the circumference. This value can only be achieved theoretically, since it would apply to continuously increasing blade numbers, but in fact the number of blades can only be increased in discrete steps.
  • the dashed line 14 is assigned to a radial wheel 1 (FIG. 1) with twenty-two blades distributed around the circumference. It can be seen that the surface efficiency only approaches the theoretical value at a single point, namely where the flow channel delimited by the hub and the housing contour and the blades has a square cross section. The surface efficiency drops significantly towards the impeller inlet to the left and towards the impeller outlet to the right.
  • Line 15 is assigned to radial wheel 1 according to the invention according to FIGS. 2 and 3. Two positions can be seen, which correspond to the axial points Z 1 and Z 2 according to FIG. 2, at which the number of blades doubles, which is accompanied by a change in the cross-sectional contour. Compared to the radial wheel 1a with only one number of blades, the theoretical value of the surface efficiency is now achieved three times. Overall, this radial wheel 1 has a significantly improved surface efficiency compared to the radial wheel 1 a.
  • a further improvement can be achieved by the radial wheel 1 shown in line 16 according to FIGS. 5 and 6. There are three positions at which the number of blades increases, which means that the theoretical value of the surface efficiency is now reached four times.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Structures Of Non-Positive Displacement Pumps (AREA)
  • Turbine Rotor Nozzle Sealing (AREA)
EP91115357A 1990-09-15 1991-09-11 Radialrad für eine Turbomaschine Withdrawn EP0476499A1 (de)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE4029331A DE4029331C1 (enrdf_load_stackoverflow) 1990-09-15 1990-09-15
DE4029331 1990-09-15

Publications (1)

Publication Number Publication Date
EP0476499A1 true EP0476499A1 (de) 1992-03-25

Family

ID=6414337

Family Applications (1)

Application Number Title Priority Date Filing Date
EP91115357A Withdrawn EP0476499A1 (de) 1990-09-15 1991-09-11 Radialrad für eine Turbomaschine

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US (1) US5213473A (enrdf_load_stackoverflow)
EP (1) EP0476499A1 (enrdf_load_stackoverflow)
DE (1) DE4029331C1 (enrdf_load_stackoverflow)

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DE102005019937B3 (de) * 2005-04-29 2006-05-18 Daimlerchrysler Ag Turbine mit einem Turbinenrad für einen Abgasturbolader einer Brennkraftmaschine und Abgasturbolader für eine Brennkraftmaschine
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WO2009065894A1 (de) * 2007-11-20 2009-05-28 Mann+Hummel Gmbh Verdichterrad eines radialverdichters und verfahren zur herstellung eines solchen verdichterrades
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FR2944060B1 (fr) * 2009-04-06 2013-07-19 Turbomeca Systeme d'air secondaire pour compresseur centrifuge ou mixte
US8523530B2 (en) * 2010-12-21 2013-09-03 Hamilton Sundstrand Corporation Turbine rotor for air cycle machine
US8529210B2 (en) * 2010-12-21 2013-09-10 Hamilton Sundstrand Corporation Air cycle machine compressor rotor
JP5665535B2 (ja) * 2010-12-28 2015-02-04 三菱重工業株式会社 遠心圧縮機
US8997486B2 (en) * 2012-03-23 2015-04-07 Bullseye Power LLC Compressor wheel
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US9920768B2 (en) * 2015-03-26 2018-03-20 Deere & Company Centrifugal fan assembly
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US10087947B2 (en) * 2016-01-04 2018-10-02 Caterpillar Inc. Turbocharger compressor and method
US10167875B2 (en) * 2016-01-04 2019-01-01 Caterpillar Inc. Turbocharger compressor and method
US10167876B2 (en) * 2016-01-04 2019-01-01 Caterpillar Inc. Turbocharger compressor and method
US10100841B2 (en) * 2016-03-21 2018-10-16 General Electric Company Centrifugal compressor and system
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DE102017114679A1 (de) * 2017-06-30 2019-01-03 Ebm-Papst Mulfingen Gmbh & Co. Kg Gebläserad
RU2019106653A (ru) * 2018-03-14 2020-09-11 Кэрриер Корпорейшн Открытое рабочее колесо центробежного компрессора
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CN114412828A (zh) * 2021-12-24 2022-04-29 中国北方发动机研究所(天津) 一种拓宽压气机堵塞流量的叶轮结构
JP2024177929A (ja) * 2023-06-12 2024-12-24 株式会社東芝 遠心送風機

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DE4029331C1 (enrdf_load_stackoverflow) 1992-01-30
US5213473A (en) 1993-05-25

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