US20120230858A1 - Screw pump - Google Patents

Screw pump Download PDF

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Publication number
US20120230858A1
US20120230858A1 US13/416,795 US201213416795A US2012230858A1 US 20120230858 A1 US20120230858 A1 US 20120230858A1 US 201213416795 A US201213416795 A US 201213416795A US 2012230858 A1 US2012230858 A1 US 2012230858A1
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United States
Prior art keywords
rotor
addendum
teeth
angle
screw pump
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.)
Abandoned
Application number
US13/416,795
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English (en)
Inventor
Yuya Izawa
Satoshi Umemura
Takahisa Ban
Taku Inoue
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Toyota Industries Corp
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Toyota Industries Corp
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Filing date
Publication date
Application filed by Toyota Industries Corp filed Critical Toyota Industries Corp
Assigned to KABUSHIKI KAISHA TOYOTA JIDOSHOKKI reassignment KABUSHIKI KAISHA TOYOTA JIDOSHOKKI ASSIGNMENT OF ASSIGNORS INTEREST (SEE DOCUMENT FOR DETAILS). Assignors: BAN, TAKAHISA, INOUE, TAKU, IZAWA, YUYA, UMEMURA, SATOSHI
Publication of US20120230858A1 publication Critical patent/US20120230858A1/en
Abandoned legal-status Critical Current

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
    • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C18/12—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type
    • F04C18/14—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons
    • F04C18/16—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons with helical teeth, e.g. chevron-shaped, screw type
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
    • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C18/082—Details specially related to intermeshing engagement type pumps
    • F04C18/084—Toothed wheels
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2/00—Rotary-piston machines or pumps
    • F04C2/08—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C2/12—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type
    • F04C2/14—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2220/00—Application
    • F04C2220/10—Vacuum

Definitions

  • the present invention relates to a screw pump.
  • a screw pump is provided with a first rotor and a second rotor having helical teeth.
  • the first rotor and the second rotor mesh with each other, thereby forming a working chamber.
  • the screw pump is configured to rotate the first rotor and the second rotor, so as to confine a fluid in the working chamber and transfer the fluid from an inlet to an outlet. It is known that the screw pump reduces its transfer efficiency of the fluid due to a blow-hole area or the like. For this reason, an improvement in efficiency is achieved by devising the geometries of the rotors, for example, as described in Japanese Patent Application Laid-open No. 2001-73959.
  • the Laid-open No. 2001-73959 discloses the screw pump in which a driving rotor (the first rotor) and a driven rotor (the second rotor) mesh with each other in a rotatable state.
  • a main tooth profile of the driving rotor consists of a cycloid drawn by an addendum (tooth top) tip of the driven rotor.
  • a main tooth profile of the driven rotor consists of a trochoid drawn by an addendum tip of the driving rotor. This prevents occurrence of blow-by of fluid (the blow-hole).
  • the screw pump since the addendum tips of the driven rotor are acute angles in the screw pump disclosed in the Laid-open No. 2001-73959, there is a problem of requiring considerable man-hours for processing steps and for ensuring quality. Furthermore, the screw pump generally has another problem of poor volumetric efficiency when compared to roots vacuum pumps. The rotors occupy a large volume relative to a rotor housing volume, resulting in a small discharge volume. For this reason, the screw pump has the size larger than the roots vacuum pumps and others with the same capacity.
  • An object of the present invention is to provide a screw pump reduced in acute-angled portions of rotors and achieving a high discharge volume ratio.
  • An aspect of the present invention is a screw pump comprising: a housing in which an inlet and an outlet are formed; a first rotor having helical teeth each including a first dedendum portion (tooth root portion), a first addendum portion (tooth top portion), and a first addendum tip portion in contact with the housing, and housed in a rotatable state in the housing; and a second rotor having helical teeth each including a second dedendum portion, a second addendum portion, and a second addendum tip portion in contact with the housing and meshing with the teeth of the first rotor, and housed in a synchronously rotatable state with the first rotor in the housing, wherein profiles of the first and second addendum portions are formed by cycloidal curves and profiles of the first and second dedendum portions are formed by trochoidal curves, wherein a pitch circle of the second rotor is larger than a pitch circle of the first rotor, wherein the number of teeth
  • the profiles of the first addendum portions are formed by the cycloidal curves and the profiles of the first dedendum portions are formed by the trochoidal curves.
  • the face width of the first rotor is narrow enough to reduce a rotor occupancy ratio which is a ratio of the rotor volume to the cylinder volume. Therefore, it is feasible to improve the discharge volume ratio of the screw pump.
  • obtuse-angled portions are formed at the second addendum tip portions of the second rotor, the acute-angled portions can be reduced when compared to the conventional screw pumps.
  • the discharge volume ratio is a ratio obtained by multiplying a theoretical discharge volume ratio, which is a ratio of a theoretical discharge volume to the cylinder volume, by a volumetric efficiency resulting from clearances between the housing and the rotors.
  • the screw pump may be configured as follows: on the pitch circles where the teeth of the first rotor and the teeth of the second rotor mesh with each other, the cycloidal curve of the first addendum portion varies to the trochoidal curve of the first dedendum portion and the cycloidal curve of the second addendum portion varies to the trochoidal curve of the second dedendum portion.
  • the width angle of the first rotor may be within 4° from an angle at which the discharge volume ratio is a maximum. Furthermore, the width angle of the first rotor may be not less than the minimum angle and not more than an angle 9° larger than the minimum angle.
  • the screw pump may be configured as follows: a ratio of an interval between a rotation axis of the first rotor and a rotation axis of the second rotor to a diameter of an outer circle of the first rotor is in the range of not less than a minimum of an establishment limit and not more than a value 0.02 larger than the minimum.
  • the screw pump may be configured as follows: the first rotor has three teeth, and the second rotor has four teeth.
  • a screw pump comprising: a housing in which an inlet and an outlet are formed; a first rotor having helical teeth each including a first dedendum portion, a first addendum portion, and a first addendum tip portion in contact with the housing, and housed in a rotatable state in the housing; and a second rotor having helical teeth each including a second dedendum portion, a second addendum portion, and a second addendum tip portion in contact with the housing and meshing with the teeth of the first rotor, and housed in a synchronously rotatable state with the first rotor in the housing, wherein profiles of the first addendum portions are formed by cycloidal curves and profiles of the first dedendum portions are formed by involute curves, wherein profiles of the second addendum portions are formed by involute curves and profiles of the second dedendum portions are formed by trochoidal curves, wherein a pitch circle of the second
  • the profiles of the first addendum portions are formed by the cycloidal curves
  • the profiles of the first dedendum portions are formed by the involute curves
  • the profiles of the second addendum portions are formed by the involute curves
  • the profiles of the second dedendum portions are formed by the trochoidal curves.
  • FIG. 1 is a top cross-sectional view of a screw pump according to an embodiment of the present invention.
  • FIG. 2 is a cross-sectional view of a first rotor and a second rotor according to the embodiment.
  • FIG. 3 is a cross-sectional view showing the first rotor and second rotor used in simulations.
  • FIG. 4 is a drawing showing the simulation result.
  • FIG. 5 is a drawing showing the simulation result.
  • FIG. 6 is a drawing showing the simulation result.
  • FIG. 7 is a drawing showing the simulation result.
  • FIG. 8 is a cross-sectional view of rotors of a screw pump according to a modification example of the embodiment.
  • screw pump 10 is a horizontally-installed screw pump.
  • the screw pump 10 is used, for example, as an oil-free vacuum pump.
  • a housing of the screw pump 10 is composed of a rotor housing 11 , a front housing 12 joined to the front end portion of the rotor housing 11 , and a rear housing 13 joined to the rear end portion of the rotor housing 11 .
  • a first rotor 20 and a second rotor 30 meshing with each other are housed in a space inside the housing.
  • An inlet 14 for intake of a fluid into the housing is formed at one end of the rotor housing 11 (on the left side in FIG. 1 ).
  • An outlet 15 for discharge of the fluid in the housing to the outside is formed at the other end of the rotor housing 11 (on the right side in FIG. 1 ).
  • the inlet 14 has a nearly rectangular opening and is arranged nearer to the second rotor 30 .
  • the inlet 14 faces an engaging position of the two rotors 20 , 30 .
  • the outlet 15 opens on the side of the second rotor 30 .
  • the opening area of the outlet 15 is set smaller than that of the inlet 14 .
  • a first shaft 21 penetrates the first rotor 20 and is fixed to the first rotor 20 .
  • a second shaft 31 penetrates the second rotor 30 and is fixed to the second rotor 30 .
  • the rotation axis A 1 of the first rotor 20 and the rotation axis A 2 of the second rotor 30 are arranged in parallel at an interval L.
  • One end portion of the first shaft 21 (on the left side in FIG. 1 ) is passed through an axial hole 12 A formed in the front housing 12 and is supported through a bearing 18 on the front housing 12 .
  • the other end portion of the first shaft 21 (on the right side in FIG. 1 ) is passed through an axial hole 13 A formed in the rear housing 13 and is supported through a bearing 18 on the rear housing 13 .
  • One end portion of the second shaft 31 (on the left side in FIG. 1 ) is passed through an axial hole 12 B formed in the front housing 12 and is supported through a bearing 18 on the front housing 12 .
  • the other end portion of the second shaft 31 (on the right side in FIG. 1 ) is passed through an axial hole 13 B formed in the rear housing 13 and is supported through a bearing 18 on the rear housing 13 .
  • the first rotor 20 is rotatably supported in a both-ends supported state on the housing by the first shaft 21 projecting from the two ends thereof.
  • the second rotor 30 is rotatably supported in a both-ends supported state on the housing by the second shaft 31 projecting from the two ends thereof.
  • a gear housing 40 is joined to the rear housing 13 .
  • the gear housing 40 forms a gear chamber 41 together with the rear housing 13 .
  • the other end portion of the first shaft 21 penetrates the rear housing 13 and is fixed to a driving gear 42 in the gear housing 40 .
  • the other end portion of the second shaft 31 penetrates the rear housing 13 and is fixed to a driven gear 43 in the gear housing 40 .
  • An electric motor 45 as a drive source is arranged in the gear housing 40 and an output shaft 46 of the electric motor 45 is coupled through a coupling 47 to the other end portion of the first shaft 21 .
  • the driving gear 42 meshes with the driven gear 43 , whereby rotation of the first shaft 21 is transmitted through the driving gear 42 and the driven gear 43 to the second shaft 31 , thereby achieving synchronous rotation of the first rotor 20 and the second rotor 30 .
  • the first rotor 20 is a three-teeth male rotor.
  • the first rotor 20 has an axial periphery represented by an inner circle 20 B 1 .
  • the first rotor 20 is provided with a plurality of teeth (three teeth in the present embodiment) 20 A radiating in radial directions from the axial periphery and having a helical shape.
  • the inner circle 20 B 1 is a circle centered on the rotation axis A 1 of the first shaft 21 .
  • the teeth 20 A as shown in FIG. 2 , are arranged at equal intervals in a cross section of the first rotor 20 .
  • Each tooth 20 A of the first rotor 20 has a dedendum portion (tooth root portion) 20 A 1 located on the rotation axis A 1 side, and an addendum portion (tooth top portion) 20 A 2 located on the outer periphery side of the dedendum portion 20 A 1 .
  • a boundary between the dedendum portion 20 A 1 and the addendum portion 20 A 2 is located on a pitch circle of the first rotor 20 represented by a medium circle 20 B 2 .
  • Each tooth 20 A has an addendum (tooth top) tip portion 20 A 3 represented by an outer circle 20 B 3 , at a radial distal end of the addendum portion 20 A 2 . As shown in FIG.
  • the dedendum portion 20 A 1 is a portion located between the inner circle 20 B 1 and the medium circle 20 B 2 in the radial direction.
  • the addendum portion 20 A 2 is a portion located between the medium circle 20 B 2 and the outer circle 20 B 3 in the radial direction.
  • the diameter of the medium circle 20 B 2 is set to be larger than that of the inner circle 20 B 1 .
  • the diameter of the outer circle 20 B 3 is set to be larger than that of the medium circle 20 B 2 .
  • a portion of an inner periphery of the rotor housing 11 is positioned along the outer circle 20 B 3 .
  • the addendum tip portion 20 A 3 is in contact with the inner periphery of the rotor housing 11 on the outer circle 20 B 3 . Since the first rotor 20 has the teeth 20 A formed in the helical shape, the addendum tip portions 20 A 3 are in line contact with the inner periphery of the rotor housing 11 .
  • the second rotor 30 is a four-teeth female rotor.
  • the second rotor 30 has an axial periphery represented by an inner circle 30 B 1 .
  • the inner circle 30 B 1 is a circle centered on the rotation axis A 2 of the second shaft 31 .
  • the second rotor 30 is provided with a plurality of teeth (four teeth in the present embodiment) 30 A radiating in radial directions from the axial periphery and having a helical shape.
  • the teeth 30 A mesh with the teeth 20 A of the first rotor 20 .
  • the teeth 30 A as shown in FIG. 2 , are arranged at equal intervals in a cross section of the second rotor 30 .
  • Each tooth 30 A of the second rotor 30 has a dedendum portion 30 A 1 located on the rotation axis A 2 side, and an addendum portion 30 A 2 located on the outer periphery side of the dedendum portion 30 A 1 .
  • a boundary between the dedendum portion 30 A 1 and the addendum portion 30 A 2 is located on a pitch circle of the second rotor 30 represented by a medium circle 30 B 2 .
  • Each tooth 30 A has an addendum tip portion 30 A 3 represented by an outer circle 30 B 3 , at a radial distal end of the addendum portion 30 A 2 . As shown in FIG.
  • the dedendum portion 30 A 1 is a portion located between the inner circle 30 B 1 and the medium circle 30 B 2 in the radial direction.
  • the addendum portion 30 A 2 is a portion located between the medium circle 30 B 2 and the outer circle 30 B 3 in the radial direction.
  • the addendum tip portion 30 A 3 is formed in an arcuate shape so as to extend along the outer circle 30 B 3 .
  • the diameter of the medium circle 30 B 2 is set to be larger than that of the inner circle 30 B 1 .
  • the diameter of the outer circle 30 B 3 is set to be larger than that of the medium circle 30 B 2 .
  • the outer circle 20 B 3 of the first rotor 20 and the outer circle 30 B 3 of the second rotor 30 have the same diameter.
  • a portion of the inner periphery of the rotor housing 11 is positioned along the outer circle 30 B 3 .
  • Each addendum tip portion 30 A 3 has a certain surface in contact with the inner periphery of the rotor housing 11 on the outer circle 30 B 3 .
  • Profiles of the dedendum portions 20 A 1 of the first rotor 20 are formed by trochoidal curves based on the outer circle 30 B 3 of the second rotor 30 .
  • the profile of each dedendum portion 20 A 1 is formed as a trochoidal curve drawn from a given point on the medium circle 20 B 2 being the pitch circle of the first rotor 20 , by an arbitrary point on the outer circle 30 B 3 of the second rotor 30 .
  • the profiles of the dedendum portions 20 A 1 all are trochoidal curves.
  • Profiles of the addendum portions 20 A 2 of the first rotor 20 are formed by cycloidal curves based on the medium circle 30 B 2 of the second rotor 30 .
  • each addendum portion 20 A 2 is formed as a cycloidal curve drawn from a given point on the medium circle 20 B 2 being the pitch circle of the first rotor 20 , by an arbitrary point on the medium circle 30 B 2 of the second rotor 30 .
  • Two cycloidal curves intersect with each other at the addendum tip portion 20 A 3 (on the outer circle 20 B 3 ).
  • the profiles of the addendum portions 20 A 2 all are cycloidal curves.
  • the face width of each tooth 20 A of the first rotor 20 is defined by a width angle ⁇ shown in FIG. 2 .
  • the width angle ⁇ is an angle made between lines connecting the width on the medium circle 20 B 2 of the tooth 20 A from the rotation axis A 1 of the first rotor 20 .
  • the three teeth 20 A are formed each with the same width angle ⁇ . If the width angle ⁇ is larger than the angle shown in FIG. 2 , the addendum tip portions 20 A 3 will expand so as to be in face contact with the rotor housing 11 on the outer circle 20 B 3 . If the width angle ⁇ is smaller than the angle shown in FIG.
  • the width angle ⁇ in the state in which the addendum tip portions 20 A 3 are in line contact with the inner periphery of the rotor housing 11 on the outer circle 20 B 3 as shown in FIG. 2 is a minimum angle of the width angle ⁇ .
  • the width angle ⁇ in the state in which the addendum tip portions 20 A 3 are in line contact with the outer circle 20 B 3 is the minimum angle of the width angle ⁇ .
  • the width angle ⁇ of the first rotor 20 is set at this minimum angle.
  • Profiles of the dedendum portions 30 A 1 of the second rotor 30 are formed by trochoidal curves based on the outer circle 20 B 3 of the first rotor 20 .
  • the profile of each dedendum portion 30 A 1 is formed as a trochoidal curve drawn from a given point on the medium circle 30 B 2 being the pitch circle of the second rotor 30 , by an arbitrary point on the outer circle 20 B 3 of the first rotor 20 .
  • the profiles of the dedendum portions 30 A 1 all are trochoidal curves.
  • Profiles of the addendum portions 30 A 2 of the second rotor 30 are formed by cycloidal curves based on the medium circle 20 B 2 of the first rotor 20 .
  • each addendum portion 30 A 2 is formed as a cycloidal curve drawn from a given point on the medium circle 30 B 2 being the pitch circle of the second rotor 30 , by an arbitrary point on the medium circle 20 B 2 of the first rotor 20 .
  • the trochoidal curves of the dedendum portions 30 A 1 are in contact with the inner circle 30 B 1 of the second rotor 30 .
  • two adjacent teeth 30 A have their respective dedendum portions 30 A 1 formed by one trochoidal curve.
  • the addendum portions 30 A 2 are formed by the cycloidal curves up to portions reaching the inner periphery of the rotor housing 11 (outer circle 30 B 3 ).
  • FIG. 3 is a drawing showing a cross section of the first rotor 20 and the second rotor 30 of the screw pump 10 used in the simulations below.
  • the diameter of the outer circle 20 B 3 of the first rotor 20 is 100 millimeters.
  • the interval L between the rotation axis A 1 of the first rotor 20 and the rotation axis A 2 of the second rotor 30 is 70 millimeters.
  • the first rotor 20 is provided with three teeth and thus the number of teeth is 3.
  • the second rotor 30 is provided with four teeth and thus the number of teeth is 4.
  • the profiles of the dedendum portions 20 A 1 , 30 A 1 are formed by trochoidal curves.
  • the profiles of the addendum portions 20 A 2 , 30 A 2 are formed by cycloidal curves.
  • FIG. 4 shows a change of theoretical discharge volume ratio in the screw pump 10 , with change in the width angle ⁇ of the first rotor 20 shown in FIG. 3 .
  • the theoretical discharge volume ratio herein is defined by Formula (1) below.
  • Theoretical discharge volume ratio theoretical discharge volume/cylinder volume (1)
  • the theoretical discharge volume is a discharge volume per complete rotation of the first rotor 20 .
  • the cylinder volume is the volume of the rotor housing 11 in which the first rotor 20 and the second rotor 30 are housed.
  • the simulation was conducted with the width angle ⁇ of the first rotor 20 being set at each of 51°, 60°, 70°, and 80°.
  • the theoretical discharge volume ratio decreased with increase in the width angle ⁇ of the first rotor 20 .
  • the theoretical discharge volume ratio is the largest, about 0.50, when the width angle ⁇ is 51°. Therefore, the width angle ⁇ of the first rotor 20 is preferably set as small as possible. In the conditions of this simulation, the minimum angle of the width angle ⁇ of the first rotor 20 is 51°.
  • the width angle ⁇ is set below 51°, the first rotor 20 will have a gap relative to the inner periphery of the rotor housing 11 so as to increase the fluid leakage, thereby resulting in significant reduction of efficiency of the screw pump 10 .
  • FIG. 4 shows the simulation result of the change of theoretical discharge volume ratio in the screw pump 10 , with change in the width angle ⁇ of the first rotor 20 , in a state in which the addendum tip portions 20 A 3 and the addendum tip portions 30 A 3 are in contact with the inner periphery of the rotor housing 11 on the outer circle 20 B 3 , 30 B 3 .
  • clearances are needed between the addendum tip portions 20 A 3 of the first rotor 20 and the rotor housing 11 and between the addendum tip portions 30 A 3 of the second rotor 30 and the rotor housing 11 because of friction resistance and manufacturing tolerance. Since these clearances change the volumetric efficiency, a real discharge volume ratio is defined by Formula (2) below.
  • FIG. 5 shows a change of discharge volume ratio in the screw pump 10 , with change in the width angle ⁇ of the first rotor 20 shown in FIG. 3 , in a configuration in which ordinary clearances in the size of the screw pump 10 are set.
  • the simulation was conducted with the width angle ⁇ of the first rotor 20 being set at each of 51°, 60°, 70°, and 80°.
  • the discharge volume ratio once increased and then decreased with increase in the width angle ⁇ of the first rotor 20 .
  • the discharge volume ratio is about 0.461 at the width angle ⁇ of 51° and about 0.460 at the width angle ⁇ of 60°.
  • the discharge volume ratio reached a maximum at the width angle ⁇ of 55° and then decreased at the width angles ⁇ of more than 55°.
  • the discharge volume ratios in the range of width angle ⁇ of up to 59° are approximately equal to that at the width angle ⁇ of 51°.
  • the minimum of the width angle ⁇ of the first rotor 20 is also 51°.
  • the conditions of the present simulation are the same as the simulation conditions in FIG. 4 . Namely, the diameter of the outer circle 20 B 3 is 100 millimeters and the interval L 70 millimeters. Then the theoretical discharge volume ratio was measured with change in each of the number of teeth of the first rotor 20 and the number of teeth of the second rotor 30 . The width angle ⁇ of the first rotor 20 was always set at the minimum angle even with change in the number of teeth.
  • the theoretical discharge volume ratio is lower with the number of teeth of the second rotor 30 being 5 than that with the number being 4.
  • the theoretical discharge volume ratio is lower with the number of teeth of the second rotor 30 being 6 than that with the number being 5. It is seen about the numbers of teeth of the first rotor 20 and the second rotor 30 from this result that the theoretical discharge volume ratio can be made large when the difference between the number of teeth of the first rotor 20 and the number of teeth of the second rotor 30 is the smallest. Under the conditions of the present simulation, where the number of teeth of the first rotor 20 was 5, the second rotor 30 could not be formed with the number of teeth being 7.
  • the theoretical discharge volume ratio increases with increase in the number of teeth of the first rotor 20 from 3 to 4 and 5. According to this simulation result, the theoretical discharge volume ratio is the highest, about 0.535, when the number of teeth of the first rotor 20 is 5 and the number of teeth of the second rotor 30 is 6.
  • the width angle ⁇ of the first rotor 20 was always set at the minimum angle.
  • the diameter of the outer circle 20 B 3 is 100 millimeters.
  • a change of theoretical discharge volume ratio was measured with change in the numbers of teeth of the first rotor 20 and the second rotor 30 and with change in the interval L between the rotation axis A 1 of the first rotor 20 and the rotation axis A 2 of the second rotor 30 .
  • the number of teeth of the second rotor 30 is set to be larger than the number of teeth of the first rotor 20 and the difference between them is set to the smallest. Namely, the number of teeth of the second rotor 30 is one larger than that of the first rotor 20 .
  • the theoretical discharge volume ratio increases with decrease in the interval L from 70 millimeters.
  • the theoretical discharge volume ratio is the largest, about 0.63, when a ratio of the interval L to the outer circle 20 B 3 (interval L/diameter of outer circle 20 B 3 ) is 0.62, in the configuration wherein the number of teeth of the first rotor 20 is 3 and that of the second rotor 30 is 4.
  • an establishment limit of the screw pump 10 about the ratio of the interval L (62 mm) to the outer circle 20 B 3 (100 mm) is 0.62 in the configuration wherein the number of teeth of the first rotor 20 is 3 and that of the second rotor 30 is 4.
  • the ratio of the interval L to the outer circle 20 B 3 can be at least 0.62 being the establishment limit.
  • a minimum value as an establishment limit about the ratio of the interval L (66 mm) to the outer circle 20 B 3 (100 mm) is 0.66.
  • the ratio of the interval L to the outer circle 20 B 3 is preferably set in the range from the establishment limit value of the first rotor 20 and the second rotor 30 to a value 0.02 larger than the establishment limit value, with a high theoretical discharge volume ratio.
  • the number of teeth of the first rotor 20 is 3.
  • the number of teeth of the second rotor 30 is larger than that of the first rotor 20 and is 4 with the smallest difference in the number of teeth.
  • the diameter of the outer circle 20 B 3 of the first rotor 20 is set at 100 mm.
  • the interval L between the rotation axis A 1 of the first rotor 20 and the rotation axis A 2 of the second rotor 30 is preferably set at 62 mm.
  • the width angle ⁇ of the first rotor 20 is preferably set between 51° and 59°.
  • the width angle ⁇ is preferably set between 51° of the minimum angle and 59°, approximately 1° can be a permissible range in view of manufacturing tolerance. Namely, the width angle ⁇ can be not less than 51° and not more than 60°, with a high discharge volume ratio.
  • the width angle ⁇ of the first rotor 20 is preferably set within 4° from the angle at which the discharge volume ratio is the maximum. Since the rotor housing 11 has no involvement in the simulations shown in FIGS. 6 and 7 , an increase in theoretical discharge volume ratio leads directly to an increase in discharge volume ratio.
  • the profiles of the respective dedendum portions 20 A 1 , 30 A 1 are formed by the trochoidal curves and the profiles of the respective addendum portions 20 A 2 , 30 A 2 are formed by the cycloidal curves. This reduces the blow-hole area and thus enables provision of high-efficiency screw pump 10 .
  • the second rotor 30 has the number of teeth larger than that of the first rotor 20 and the pitch circle of the second rotor 30 is set larger than that of the first rotor 20 . Since the width angle ⁇ of the first rotor 20 is set at an angle to achieve the discharge volume ratio approximately equal to that at the minimum angle at which the addendum tip portions 20 A 3 are in line contact with the outer circle 20 B 3 , the screw pump 10 can have a high discharge volume ratio. Since the discharge volume ratio is high, the screw pump 10 can be constructed in a reduced size.
  • the second rotor 30 is formed by the cycloidal curves and trochoidal curves, acute-angled portions of the second rotor 30 are reduced. The reduction in acute-angled portions facilitates processing. Furthermore, it improves the quality of the second rotor 30 .
  • the first rotor 20 Since in the first rotor 20 the three teeth 20 A are provided at equal intervals in the circumferential direction and radiate in radial directions from the axial periphery, the first rotor 20 is well-balanced during rotation. Since in the second rotor 30 the four teeth 30 A are provided at equal intervals in the circumferential direction and radially in radial directions from the axial periphery, the second rotor 30 is well-balanced during rotation. Then the screw pump 10 is well-balanced during rotation of the first rotor 20 and the second rotor 30 in the intermeshing state.
  • the ratio of the interval L between the rotation axis A 1 of the first rotor 20 and the rotation axis A 2 of the second rotor 30 to the outer circle 20 B 3 of the first rotor 20 is set in the range of not less than the minimum of the establishment limit and not more than the value 0.02 larger than the minimum. This can enhance the theoretical discharge efficiency of the screw pump 10 . Furthermore, it permits reduction in the size of the screw pump 10 .
  • the present invention is by no means limited to the above embodiment.
  • the profiles of the first rotor 20 and the second rotor 30 are formed by only the cycloidal curves and trochoidal curves, but the present invention is not limited to this example.
  • profiles of dedendum portions 60 A 1 of first rotor 60 may be formed in part by involute curves.
  • FIG. 8 is a cross-sectional view of a screw pump in which the first rotor 60 with four teeth 60 A having the width angle ⁇ meshes with a second rotor 70 with six teeth 70 A.
  • Each tooth 60 A has a dedendum portion 60 A 1 located between an inner circle 60 B 1 and a medium circle 60 B 2 , an addendum portion 60 A 2 located between the medium circle 60 B 2 and an outer circle 60 B 3 , and an addendum tip portion 60 A 3 being a tip of the addendum portion 60 A 2 .
  • Each tooth 70 A has a dedendum portion 70 A 1 located between an inner circle 70 B 1 and a medium circle 70 B 2 , an addendum portion 70 A 2 located between the medium circle 70 B 2 and an outer circle 70 B 3 , and an addendum tip portion 70 A 3 being a tip of the addendum portion 70 A 2 .
  • the profiles of the dedendum portions 60 A 1 of the first rotor 60 are formed by involute curves and trochoidal curves.
  • the profiles of the dedendum portions 60 A 1 are the involute curves from the medium circle 60 B 2 toward the inner circle 60 B 1 and transfer to the trochoidal curves near the inner circle 60 B 1 .
  • the profiles of the addendum portions 70 A 2 of the second rotor 70 are formed by cycloidal curves and involute curves.
  • the involute curves in the dedendum portions 60 A 1 of the first rotor 60 correspond to those in the addendum portions 70 A 2 of the second rotor 70 .
  • the screw pump can also be realized with a large discharge volume ratio and in a reduced size in the configuration wherein the involute curves are adopted for the profiles of the first rotor 20 and the second rotor 30 .
  • the width angle ⁇ of the first rotor 60 is set at an angle at which the discharge volume ratio of the screw pump is approximately equal to that at the minimum angle.
  • the screw pump shown in FIG. 8 which is not shown, has the housing composed of the rotor housing 11 , the front housing 12 , and the rear housing 13 as the screw pump 10 shown in FIG. 1 does.
  • the first rotor 60 and the second rotor 70 are housed in the space inside the housing.
  • Profiles of the addendum portions 60 A 2 of the first rotor 60 are formed by cycloidal curves as the profiles of the addendum portions 20 A 2 of the first rotor 20 are.
  • Profiles of the dedendum portions 70 A 1 of the second rotor 70 are formed by trochoidal curves as the profiles of the dedendum portions 30 A 1 of the second rotor 30 are.
  • the pitch circle of the second rotor 70 is set larger than that of the first rotor 60 .
  • the cycloidal curve of the addendum portion 60 A 2 varies to the involute curve of the dedendum portion 60 A 1 and the involute curve of the addendum portion 70 A 2 varies to the trochoidal curve of the dedendum portion 70 A 1 .
  • the ratio of the interval L between the rotation axis A 1 of the first rotor 60 and the rotation axis A 2 of the second rotor 70 to the outer circle 60 B 3 of the first rotor 60 is set in the range of not less than the minimum of the establishment limit and not more than the value 0.02 larger than the minimum.
  • the width angle ⁇ of the first rotor 60 is within 4° from the angle at which the discharge volume ratio is a maximum.
  • the width angle ⁇ of the first rotor 60 is not less than the minimum angle and not more than an angle 9° larger than the minimum angle.
  • the profiles of the dedendum portions 20 A 1 , 30 A 1 and the addendum portions 20 A 2 , 30 A 2 do not have to be formed completely by the cycloidal curves or by the trochoidal curves.
  • the teeth 20 A, 30 A may be formed by partially modifying arcuate curves, near the addendum tip portions 20 A 3 , 30 A 3 .
  • the medium circles 20 B 2 , 30 B 2 do not always have to be on the pitch circles.
  • the medium circles 20 B 2 , 30 B 2 can be larger or smaller than the respective pitch circles, while achieving a large theoretical discharge volume ratio and a reduction in size of the screw pump.
  • outer circle 20 B 3 of the first rotor 20 and the outer circle 30 B 3 of the second rotor 30 have the same size, but the present invention is not limited to this example.
  • the outer circles 20 B 3 , 30 B 3 have the same size in the configuration of the present embodiment shown in FIG. 2 , but the sizes of the respective outer circles 20 B 3 , 30 B 3 may be varied depending upon applications and places for use of the screw pump 10 .

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Rotary Pumps (AREA)
  • Details And Applications Of Rotary Liquid Pumps (AREA)
  • Applications Or Details Of Rotary Compressors (AREA)
US13/416,795 2011-03-11 2012-03-09 Screw pump Abandoned US20120230858A1 (en)

Applications Claiming Priority (4)

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JP2011054168 2011-03-11
JP2011-054168 2011-03-11
JP2012-034454 2012-02-20
JP2012034454A JP2012207660A (ja) 2011-03-11 2012-02-20 スクリュポンプ

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US20120230858A1 true US20120230858A1 (en) 2012-09-13

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US13/416,795 Abandoned US20120230858A1 (en) 2011-03-11 2012-03-09 Screw pump

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CN (1) CN102678542A (ja)

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WO2021022352A1 (en) * 2019-08-02 2021-02-11 Rt Hamilton And Associates Limited Cooled dry vacuum screw pump
US20220145886A1 (en) * 2019-03-14 2022-05-12 Ateliers Busch Sa Dry pump for gas and set of a plurality of dry pumps for gas

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CN103850931A (zh) * 2012-12-05 2014-06-11 上海易昆机械工程有限公司 一种无脉动转子泵
DE102013021902B4 (de) 2013-12-26 2017-06-14 HENKE Property UG (haftungsbeschränkt) Schmelzepumpe zum Aufbau von Druck zwecks Durchdrücken von Kunststoffschmelze durch ein Werkzeug
WO2016031413A1 (ja) * 2014-08-28 2016-03-03 株式会社Ihi スクリューロータ
CN107110156B (zh) * 2015-01-05 2018-08-24 株式会社爱发科 螺杆真空泵
ES2883556T3 (es) * 2018-09-11 2021-12-09 Common Spolka Akcyjna Medidor de flujo rotativo para medir el flujo de gas
CN109441811A (zh) * 2018-12-26 2019-03-08 东莞赫升机电有限公司 堆叠转子式螺杆压缩机

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Publication number Priority date Publication date Assignee Title
US20220145886A1 (en) * 2019-03-14 2022-05-12 Ateliers Busch Sa Dry pump for gas and set of a plurality of dry pumps for gas
US11920592B2 (en) * 2019-03-14 2024-03-05 Ateliers Busch Sa Dry pump for gas and set of a plurality of dry pumps for gas
WO2021022352A1 (en) * 2019-08-02 2021-02-11 Rt Hamilton And Associates Limited Cooled dry vacuum screw pump
US11708832B2 (en) 2019-08-02 2023-07-25 Fruvac Ltd. Cooled dry vacuum screw pump

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JP2012207660A (ja) 2012-10-25
KR20120104113A (ko) 2012-09-20

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