US3764225A - Technique and blade arrangement to reduce the serpentine motion of a mass particle flowing through a turbomachine - Google Patents

Technique and blade arrangement to reduce the serpentine motion of a mass particle flowing through a turbomachine Download PDF

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Publication number
US3764225A
US3764225A US00139037A US3764225DA US3764225A US 3764225 A US3764225 A US 3764225A US 00139037 A US00139037 A US 00139037A US 3764225D A US3764225D A US 3764225DA US 3764225 A US3764225 A US 3764225A
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Prior art keywords
stage
blades
radial
blade
fixed
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US00139037A
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English (en)
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L Dzung
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BBC Brown Boveri AG Switzerland
BBC Brown Boveri France SA
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BBC Brown Boveri France SA
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    • 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
    • F01D5/12—Blades
    • F01D5/14—Form or construction
    • F01D5/141—Shape, i.e. outer, aerodynamic form
    • F01D5/142—Shape, i.e. outer, aerodynamic form of the blades of successive rotor or stator blade-rows
    • 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
    • F01D5/12—Blades
    • F01D5/14—Form or construction
    • F01D5/141—Shape, i.e. outer, aerodynamic form

Definitions

  • the present invention relates to an improved technique for reducing the serpentine motion, caused by the alternating peripheral component of the flow velocity, of the path, projected cylindrically on a meridional plane, of a mass particle in an axial-flow turbomachine, by means of the blading of the machine, and a blade arrangement to effect this technique.
  • the flow medium is subjected to a serpentine motion in the meridional plane owing to periodic variation of the peripheral component of the flow velocity and to the associated variation of the centrifugal forces.
  • This serpentine motion gives rise to additional energy losses and, by affecting the flow approach angles, makes correct determination of the blade profiles difficult.
  • a well-known countermeasure consists in twisting the blades so that the flow is free from eddies, at least in the axial direction.
  • An eddy-free flow can satisfy the condition of radial equilibrium without serpentine motion.
  • the disadvantage of this feature is that with machines of high volume through-put, and hence large blade-diameter ratios, the blades have to be very sharply twisted. This increases manufacturing costs and for reasons of strength is not always practicable.
  • the purpose of the present invention is to reduce the serpentine motion of the working medium on its way through the turbo-machine, and also the energy losses thus caused.
  • this is achieved by making the blades of such a shape that radial forces are induced in the working medium which at least partially compensate the radial forces resulting from the peripheral component.
  • a blade arrangement achieving this purpose is characterized by the fact that, when viewed in the axial direction and considered from base to tip of the blade, the leading and trailing edges of each fixed blade converge to at least such an extent that they coincide with a radius, and the leading and trailing edges of each moving blade, also viewed in the axial direction and considered from base to tip of the blade, deviate from the radii passing through the base points of the leading and trailing edges to at least such an extent that they are approximately parallel or even somewhat convergent.
  • FIG. 1 different forms of fixed and moving blades, viewed in the direction of the axis of the turbomachine,
  • FIG. 2 plane representation of part of cylindrical section II-II of FIG. 1, through one row each of fixed and moving blades,
  • FIG. 3-5 associated flow diagrams.
  • the cause of the serpentine motion of a mass particle must be examined more closely.
  • a typical stage of a multiple-stage, axial-flow turbo-machine with 50 percent reaction along the radius i.e., the fixed and moving blades each convert one half of the stage drop.
  • the peripheral component of the flow velocity In front of the fixed row of the stage, the peripheral component of the flow velocity has a certain value which is first considered arbitrarily to be in a negative direction. Within the blades of the fixed row the peripheral component of this negative value increases to a certain new, positive value. On flowing through the blades of the moving row the absolute peripheral component of velocity is converted into work in accordance with Eulers turbine law.
  • this component On leaving the moving row, this component again has the same negative value as in front of the fixed row, whereupon this process is repeated in the next stage.
  • This variation of the peripheral component causes corresponding variation of the radial acceleration which, with constant reaction, is proportional to this velocity component.
  • the radial acceleration in turn gives rise to a radial component of velocity, and hence to radial compression of the streamlines in the meridional plane. This is the phenomenon of serpentine motion.
  • FIG. 1 and 2 show a fixed blade 1 of a turbomachine with leading edge 3 and trailing edge 4, and also a mov ing blade 2 with leading edge 5 and trailing edge 6.
  • the leading and trailing edges of both blades coincide with radii, i.e., lines passing, through centre 0.
  • Curve 21 in FIG. 3 represents schematically the radial acceleration of a mass particle in a stage as caused by the radial forces resulting from the peripheral component of the flow velocity.
  • Curve 22 illustrates the velocity, and curve 23 the radial displacement, where in each case, as also in FIG. 4 and 5, the left-hand half of a diagram represents conditions in the fixed row, and the right-hand half conditions in the moving row of a stage.
  • the abscissae are in each case the time taken by a mass particle of the working medium to flow through the stage. Except for the influence of the constriction due to the blade thickness, this time is proportional to the axial distance, so that curve 23 also represents the meridional streamline projected cylindrically in a meridional plane.
  • Curves 22 and 23 are obtained by single and double integration, respectively, of curve 21, with the aid of suitable integration constants. The above statements require modification if account is taken of the finite thickness of the blade.
  • the traditional form of fixed and moving blades is cylindrical, i.e., with practically parallel leading edges, as fixed blade 7 and moving blade 8 in FIG. 1.
  • the leading and trailing edges then make the angle iv with the radii passing through their base points, such that the leading edge 9 of fixed blade 7 and the trailing edge 12 of moving blade 8 form negative angles, while the trailing edge 10 of fixed blade 7 and the leading edge 11 of moving blade 8 form positive angles, as is shown for moving blade 8, with base points 13 and 14 for leading edge 11 and trailing edge 12, respectively.
  • angles 7 are at the same time the angles between the radial and peripheral components of the forces exerted by the blades on the working medium. Since the peripheral components of the flow velocity in the fixed and moving blades are always in mutually opposed directions, the additional radial blade-force components in the fixed and moving blades, and hence also the additional radial accelerations, always follow a similar course, e.g. as curve 31 in FIG. 4. Integration of curve 31 again yields velocity curve 32 and displacement curve 33.
  • FIG. 3 and 4 shows that the radial forces in the fixed blades are added to each other, although they compensate each other in' the moving blades. This applies to both a turbine and a compressor. As in the case of an oscillation of different frequencies the two causes cannot cancel out each other. Indeed, the dissipation loss will always be cumulative.
  • the fixed blades are so formed that their leading edges and trailing edges are essentially radial as indicated by blade I in FIG. 1 so that there are no radial forces acting on the fluid.
  • the moving blades e.g., blade 8 in FIG. 1 are so formed that the leading edges, e.g., edge 11, are inclined at a positive angle of y with respect to the radial line so that the concave side of each blade, i.e., the pressure side, as represented by the left-hand side of the moving blade in FIG. 2 is slightly facing toward the axis, while the trailing edge, e.g., edge 12 of blade 8 is inclined at an opposite, i.e., a negative angle of 7 against the radial line.
  • the effect of compensation can be increased if the leading and trailing edges of the moving blades viewed in the axial direction and considered from base to tip of the blade converge as shown by in FIG. 1. In this way the relationship between the two causes of radial displacement of a mass particle can be altered. It must be expressly emphasized here that this blade form is not to be confused with the well-known technique of tapering the blades. This is, as a rule, such that in axial section the blades are indeed broader at the base than at the tip, but that the relationships in the peripheral direction are reversed.
  • a compensating force can be applied if the leading and trailing edges, considered from base to tip of the blade, are inclined towards each other even more than the corresponding radii, e.g., as 16 in FIG. 1. If the radial force in the fixed blades acts in the appropriate direction, almost com plete compensation can be achieved.
  • a similar increase in effectiveness can be achieved with a blade arrangement in accordance with the invention such that the moving blades are broader than the fixed blades in either the peripheral direction or the axial direction.
  • a compensating effect occurs only in the case of the moving blades.
  • the moving row is wider in the axial direction, the time taken in flowing through the moving blades is greater; the same radial acceleration then gives rise to a higher value of the radial velocity and of the radial displacement. If the moving blades are broader in the peripheral direction, the value of angle 1 'y in FIG. 1 is higher, thus reinforcing the compensating effect.
  • a multi-stage axial-flow turbo-machine each stage thereof comprising a row of fixed blades having a concave curved configuration and an adjacent row of movable blades having an oppositely concave curved configuration, the leading and trailing edges of said fixed blades of each stage extending along radial lines as viewed in the direction of the axis, the leading edges of said movable blades of each stage being inclined at a positive angle with respect to a radial line whereby the concave side of each movable blade faces slightly towards the axis, and the trailing edges of said movable blades of each stage being inclined at the same but negative angle with respect to a radial line, thereby to effect a reduction in the serpentine motion of a mass particle as it passes through successive stages of said multistage turbo-machine and which is caused by the alternating peripheral component of the flow velocity of the particle path projected cylindrically on a meridonal

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Fluid Mechanics (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Structures Of Non-Positive Displacement Pumps (AREA)
US00139037A 1970-05-27 1971-04-30 Technique and blade arrangement to reduce the serpentine motion of a mass particle flowing through a turbomachine Expired - Lifetime US3764225A (en)

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
CH784170A CH516747A (de) 1970-05-27 1970-05-27 Verfahren und Beschaufelung zur Verminderung der schlängelnden Bewegung eines Massenteilchens beim Durchströmen einer Turbomaschine

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US3764225A true US3764225A (en) 1973-10-09

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US00139037A Expired - Lifetime US3764225A (en) 1970-05-27 1971-04-30 Technique and blade arrangement to reduce the serpentine motion of a mass particle flowing through a turbomachine

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US (1) US3764225A (de)
CH (1) CH516747A (de)
DE (1) DE2043083C3 (de)
FR (1) FR2093626A5 (de)
GB (1) GB1306891A (de)

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4253800A (en) * 1978-08-12 1981-03-03 Hitachi, Ltd. Wheel or rotor with a plurality of blades
US4474534A (en) * 1982-05-17 1984-10-02 General Dynamics Corp. Axial flow fan
US5000660A (en) * 1989-08-11 1991-03-19 Airflow Research And Manufacturing Corporation Variable skew fan

Families Citing this family (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP6134628B2 (ja) * 2013-10-17 2017-05-24 三菱重工業株式会社 軸流式の圧縮機、及びガスタービン
CN107965352A (zh) * 2017-12-26 2018-04-27 北京全四维动力科技有限公司 能降低水蚀危险性的弯叶片、采用其的叶栅及工业汽轮机

Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE311606C (de) *
US2224519A (en) * 1938-03-05 1940-12-10 Macard Screws Ltd Screw type fluid propelling apparatus
US2298576A (en) * 1941-07-17 1942-10-13 Internat Engineering Inc Air handling apparatus
US2320733A (en) * 1938-01-07 1943-06-01 Macard Screws Ltd Screw type fluid propelling apparatus
GB631231A (en) * 1947-12-10 1949-10-28 Aerex Ltd Improvements relating to cased screw fans
GB722001A (en) * 1951-03-09 1955-01-19 Power Jets Res & Dev Ltd Turbines
US3045969A (en) * 1958-09-26 1962-07-24 Escher Wyss Ag Vibration damping device for turbo-machine
US3270953A (en) * 1963-05-21 1966-09-06 Jerie Jan Axial flow compressor, blower or ventilator with reduced noise production

Patent Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE311606C (de) *
US2320733A (en) * 1938-01-07 1943-06-01 Macard Screws Ltd Screw type fluid propelling apparatus
US2224519A (en) * 1938-03-05 1940-12-10 Macard Screws Ltd Screw type fluid propelling apparatus
US2298576A (en) * 1941-07-17 1942-10-13 Internat Engineering Inc Air handling apparatus
GB631231A (en) * 1947-12-10 1949-10-28 Aerex Ltd Improvements relating to cased screw fans
GB722001A (en) * 1951-03-09 1955-01-19 Power Jets Res & Dev Ltd Turbines
US3045969A (en) * 1958-09-26 1962-07-24 Escher Wyss Ag Vibration damping device for turbo-machine
US3270953A (en) * 1963-05-21 1966-09-06 Jerie Jan Axial flow compressor, blower or ventilator with reduced noise production

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4253800A (en) * 1978-08-12 1981-03-03 Hitachi, Ltd. Wheel or rotor with a plurality of blades
US4474534A (en) * 1982-05-17 1984-10-02 General Dynamics Corp. Axial flow fan
US5000660A (en) * 1989-08-11 1991-03-19 Airflow Research And Manufacturing Corporation Variable skew fan

Also Published As

Publication number Publication date
FR2093626A5 (de) 1972-01-28
CH516747A (de) 1971-12-15
DE2043083C3 (de) 1979-07-26
DE2043083B2 (de) 1975-10-09
GB1306891A (en) 1973-02-14
DE2043083A1 (de) 1971-12-09

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