EP0552443B1 - Machine à engrenages - Google Patents

Machine à engrenages Download PDF

Info

Publication number
EP0552443B1
EP0552443B1 EP92120410A EP92120410A EP0552443B1 EP 0552443 B1 EP0552443 B1 EP 0552443B1 EP 92120410 A EP92120410 A EP 92120410A EP 92120410 A EP92120410 A EP 92120410A EP 0552443 B1 EP0552443 B1 EP 0552443B1
Authority
EP
European Patent Office
Prior art keywords
teeth
gear
pinion
tooth
ring gear
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.)
Expired - Lifetime
Application number
EP92120410A
Other languages
German (de)
English (en)
Other versions
EP0552443A1 (fr
Inventor
Siegfried A. Dipl.-Ing. Eisenmann
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.)
Individual
Original Assignee
Individual
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 Individual filed Critical Individual
Publication of EP0552443A1 publication Critical patent/EP0552443A1/fr
Application granted granted Critical
Publication of EP0552443B1 publication Critical patent/EP0552443B1/fr
Anticipated expiration legal-status Critical
Expired - Lifetime legal-status Critical Current

Links

Images

Classifications

    • 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
    • F04C2/082—Details specially related to intermeshing engagement type machines or pumps
    • F04C2/084—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
    • 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
    • F04C18/10—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 internal-axis type with the outer member having more teeth or tooth equivalents, e.g. rollers, than the inner member
    • 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
    • F04C2/10—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of internal-axis type with the outer member having more teeth or tooth-equivalents, e.g. rollers, than the inner member

Definitions

  • the invention relates to a gear machine for liquids or gases with a housing which contains a gear chamber which has inlet and outlet openings, with an internally toothed toothed ring arranged in the gear chamber, and a pinion rotatably arranged within the toothed ring in the housing, which has one tooth less than the toothed ring has, is in engagement therewith and, when rotated between its teeth and the teeth of the toothed ring, forms circumferential enlarging and reducing liquid cells which conduct liquid from the inflow to the outflow, the tooth heads of the pinion and the tooth gaps of the toothed ring being in the form of epicycloids have, which are formed by rolling a first rolling circle (generating circle) on the pitch circle of the pinion or toothed ring, the tooth gaps of the pinion and the tooth tips of the toothed ring also having the shape of hypocycloids, which are formed by rolling a second rolling circle on the pitch circle of the pinion or toothed ring, and finally the radius of the first rolling
  • the gear machine according to the invention can be used both as a pump for liquids or gases and as a motor driven by pressurized liquids or gases.
  • the preferred field of application of the invention is the use as a liquid pump.
  • liquid is spoken of for the sake of simplicity.
  • the term liquid should therefore also include gases.
  • the gear machine according to the invention can be one in which the toothed ring is fixedly arranged in the housing, the pinion then rotating around the crank arm of a shaft, the latter being arranged centrally to the internal toothing of the pinion.
  • the machine according to the invention is preferably one in which the toothed ring rotates in the gear chamber and the pinion mounted eccentrically to the axis of the toothed ring and the gear chamber rotates with a stationary shaft or about such an axis.
  • a main area of application of the invention is the use of the machine designed as an internal ring gear pump as a lubricant and hydraulic fluid pump for internal combustion engines and automatic transmissions, where delivery pressures of up to 30 bar can occur.
  • the pump pinion is preferably arranged in the extension of the crankshaft of the engine or the main shaft of the transmission or is carried by this shaft
  • Internal gerotor pumps have proven themselves as quiet and low-vibration pumps.
  • the requirements for smooth running of such pumps are constantly increasing due to the ever better smooth running of the motors and gears.
  • Gear pumps of the type improved by the invention have long been known, for example from GB-PS 233 423 from 1925, or the publication "Kinematics of Gerotors” by Myron S. Hill, also from the 1920s.
  • a modern application of the cycloid gearing for the above-mentioned use in internal combustion engines and automatic transmissions is described in DE-PS 39 38 346 by the applicant.
  • the pump according to this German patent utilizes the excellent kinematic properties of the teeth and tooth gaps which have a complete cycloidal contour in an internal gear ring pump with the number of teeth difference one in order to support the gear ring with its toothing on that of the pinion, which is from the crankshaft of the engine or main shaft of the automatic transmission is carried.
  • the relatively strong radial movement of the crankshaft can be compensated for by selecting the circumferential bearing of the toothed ring with air sufficient for this compensation. It is also possible to store the toothed ring with little play and then to provide a correspondingly large play between the shaft carrying the pinion and the pinion, the pinion with its toothing then being mounted in that of the toothed ring.
  • Such pumps represent a preferred field of application of the present invention.
  • Pressure pulsations i.e. flow pulsations
  • the flow pulsations are intensified by pinch oil pressure peaks that lead to vibrations in the gear wheel set.
  • Gravitational noises act in the same sense, which arise primarily from the collapse of liquid vapor bubbles in the area of the pressure chamber of the pump.
  • the invention has in particular the object of making the known ring gear machines quieter, that is to say to reduce noise, which has a significant advantage when these machines are used as lubricating oil pumps in motor vehicle drive and transmission units. Another advantage achieved by this noise reduction is the improvement in efficiency and an increase in the service life of the gear ring machine.
  • the invention solves the problem according to the characterizing features of claim 1 in that the circumferential extent measured on the respective pitch circle of the pinion tooth gaps and toothed ring teeth limited by hypocycloids is 1.5 times to 3 times the circumferential extent measured on the respective pitch circle of the extent limited by epicycloids Sprocket teeth and tooth ring tooth gaps and that the epicycloids and hypocycloids are flattened towards their pitch circles to such an extent that the sum of the two flattenings corresponds to the required relatively large radial play between the tooth heads at the point opposite the point corresponds to the deepest meshing, while the gears mesh with each other at the point of deepest meshing with very little play.
  • the first-mentioned feature can also be formulated in such a way that the radius of the rolling circle producing the hypocycloids is equal to 1.5 times to 3 times the radius of the rolling circle producing the epicycloids.
  • the invention is based on the fact that the course of the instantaneous displacement volume is primarily responsible for the flow pulsations - at least with precise manufacture and little play - in gerotor machines according to the preamble of claim 1. This in turn depends primarily on the position of the sealing points between the pressure chamber and the suction chamber of the machine over the angle of rotation of the pinion or the zan ring. Theoretically, that is, with a perfect toothing without play, the sealing points coincide with the intersection of the tooth flanks with the tooth engagement line. The sealing points in the area above the pressure and suction openings are irrelevant since the liquid cells separated by the sealing points are connected to one another by the suction and pressure openings.
  • the only decisive factor is the position of the sealing points in the area of the deepest tooth engagement and in the area opposite this point.
  • the theoretical line of engagement in gear ring machines is composed of three circles touching each other at the interface of the pitch circles and the connecting straight line of the two gear wheel centers, which are symmetrical to the connecting straight line of the two gear wheel centers and are halved by this straight line.
  • the cycloid gearing according to the preamble of claim 1 offers optimal engagement conditions in the most important area of deepest tooth engagement (in FIG. 1 above). However, this only applies if the play here is very small. The reduction of the backlash, however, is limited, among other things, by the fact that it is not possible for the series production to go below a certain degree of out-of-roundness of the toothed ring without excessive technical effort. The result of this is that, according to the prior art, the minimum play must still be large enough to prevent metallic contact between the pinion tooth tips and ring tooth tips opposite the point of deepest tooth engagement (in FIG. 1 below).
  • the flattening can of course also be distributed over the two cycloid groups mentioned above, i.e. the epicycloids and the hypocycloids. However, it is easier if you limit them to one of the two groups.
  • the gearwheels can mesh with each other in the area of deepest meshing with actually minimal play and very closely approximate theoretical maximum values.
  • An unfavorable influence of the deviation of the sealing points between intermeshing teeth in the area of the deepest tooth engagement from the theoretical course is thereby minimized.
  • the negative influence of such a deviation on the flow pulsation is thus reduced.
  • the delivery flow pulsation is reduced to a particularly high degree by the selected tooth thickness ratio according to the first feature of claim 1.
  • the flow pulsation is therefore not the fluctuation of the throughput per unit of time regardless of the tooth profile selected, which can be particularly easily changed with a cycloid toothing by changing the ratio of the tooth thicknesses of the inner ring gear and pinion to each other, without thereby the advantages the cycloid teeth are lost.
  • the first characteristic feature of claim 1 makes use of this fact. If one records the fluctuation of the instantaneous displacement volume, i.e.
  • the circumferential extent of the pinion gaps and toothed ring teeth is selected to be 1.75 to 2.25 times as large as the circumferential extent of the pinion teeth and toothed tooth gaps.
  • the pinion teeth are selected half as thick as the toothed ring teeth, that is to say that the rolling circle producing the epicycloids is made half as large as the rolling circle producing the zypocycloids.
  • only one of the two groups of cycloids that is to say either the epicycloids or the hypocycloids, is preferably flattened to the full extent of the required play, while the flattening of the other cycloid group is zero.
  • the epicycloids are flattened.
  • both the flattening of the tooth gaps and the flattening of the tooth heads interacting with these tooth gaps obey the same mathematical law.
  • the flattening can be brought about, for example, by reducing the radial height of the teeth and the radial depth of the gaps of the counter gear interacting with these teeth by a small amount, which is continuous from the center of the tooth or the center of the gap to the intersection of the tooth contour with the pitch circle decreases to zero.
  • this represents a deviation from the cycloid profile which is optimal per se.
  • the easiest way to achieve the flattening according to claim 6 is by a slight radial displacement of the point describing the cycloid from the circumference of the rolling circle towards its center. In this way a cycloid contour is maintained.
  • This gap is advantageously bridged according to claim 7 in that the starting point and the end point of the flattened cycloids are connected to the start and end point of the non-flattened cycloids on the pitch circle by a straight line.
  • the sum of the two cycloid displacements (the one displacement, as mentioned above, is also zero) can and is preferably also) measured in the middle of the cycloid is the 2000th to 500th part of the pitch circle diameter of the toothed ring.
  • the tooth play at the point of deepest engagement can be extremely small, it naturally cannot be zero.
  • the required minimal tooth flank play in the circumferential direction can be brought about by an equidistant withdrawal of the tooth contour.
  • the extent of this withdrawal can be, for example, 10 times the diameter of the toothed ring pitch circle. From this number you can see how small the backlash required in the invention.
  • the number of teeth of the pinion is advantageously chosen between 7 and 11.
  • a narrow axial groove is advantageously provided in the tooth gap base at least and preferably of the pinion.
  • the grooves are advantageously about a quarter to a sixth of the rolling circle circumference, preferably a fifth of the same width.
  • the grooves are advantageously 2 to 3 times as wide as deep.
  • the axial grooves in the base of the pinion tooth gaps ensure a certain dead space without, however, the optimal filling of the tooth gaps by the tooth heads of the toothed ring and thus also the optimal guidance of the gears to one another and thus the perfect seal between the teeth being impaired to a disruptive extent.
  • gravitation bubbles and squeeze oil filled with vapor of the operating liquid can collect without the bubbles being forced to collapse by the function of the pump or the motor. Since the gravitational bubbles collect due to their low mass under the influence of centrifugal force near the tooth base of the pinion, the negative dead space effect of the grooves provided according to the invention is reduced to a negligible residual minimum.
  • the edges between the side walls and the base of the grooves are advantageously rounded in order to avoid notching effects.
  • the edges between the side walls of the grooves and the subsequent tooth space base are advantageously angular in order to to maintain the full load-bearing capacity of the tooth space base as far as possible. However, these edges should not be sharp.
  • the grooves are also provided in the tooth space base of the ring gear.
  • the grooves cannot absorb gravitational bubbles, but can squeeze oil, which is an advantage in some cases.
  • These grooves can usually be kept smaller than those in the tooth gap base of the pinion.
  • the grooves can have a circular arc profile, for example, seen in axial section. For manufacturing reasons, however, it is preferred that the grooves run with a constant profile over the entire tooth width.
  • the gerotor pump shown in FIG. 1 has a housing 1 in which a cylindrical gerotor chamber 2 is recessed. On the circumferential surface of the toothed ring chamber 2, the toothed ring 3 is rotatably supported with its cylindrical circumferential surface.
  • the toothed ring 3 has eight teeth 4. These teeth mesh with the teeth 5 of the pinion 6, which is non-rotatably seated on a shaft 7 driving the pinion.
  • the axis of rotation of the ring gear 3 is designated 8; that of pinion 6 with 9.
  • the pump rotates clockwise as indicated by the arrow in FIG. 1. It has a suction opening 10 and an outlet opening 11. The contours of these two openings lie behind the gear wheels in FIG. 1 and are therefore shown in broken lines.
  • the pump is generally known.
  • the pump shown corresponds to a pump according to German Patent No. 39 38 346 or US Patent Application S.N. 593 135 of October 5, 1990.
  • the point of engagement is initially at the location EO (FIG. 3a). From there, the point of engagement moves along the semicircle E1 to the pitch point C, that is to say the point at which the two pitch circles TH and TR touch on the connecting line of the gear wheel centers 8 and 9. The point of engagement on the circle E3 moves from C in the direction of the arrow. If the point of engagement has reached the apex of this circle on the straight line through EO and C, then the center line of the pinion tooth which can be seen on the left in FIG. 3a is on the straight line EO-C.
  • the pitch circle of the gear to be corrected is designated by T there. In the following it is assumed that this is the pitch circle of the pinion.
  • point X1 also describes a cycloid FR1, the end point of which, however, is at a short distance from the pitch circle. This distance corresponds to the distance Z1-Z0 in FIG. 2.
  • the epicycloid FH which delimits the tooth tip of the pinion, can be flattened by rolling off the rolling circle RE.
  • the point X2 describing the flattened cycloid FH1 is in the starting position at Z2. In this way, both the large pinion tooth base located on the left was shifted radially outward to the pitch circle T, while the pinion tooth contour was flattened radially away from the cycloid FH toward the pitch circle T.
  • the teeth and tooth gaps of the inner tooth ring are flattened in the same way.
  • the construction is carried out as just described, except that the pitch circle T is the pitch circle of the inner toothed ring and the rolling circle RH generates the tooth contour and the rolling circle RE generates the tooth gap contour.
  • the flattened cycloids begin and end at a slight distance from the pitch circle T. In FIG. 2, this is the distance Z1-Z2. This distance can easily be bridged by a straight line, since it is very small compared to the greatly exaggerated illustration in FIG. 2. If you have constructed the teeth as just described, you first get an ideal toothing in the area of deepest tooth engagement, which corresponds to FIG.
  • FIG. 3b shows the toothing created by the invention in the same representation as FIG. 3a. It can be seen here that the minimal tooth play caused by the withdrawal of a tooth contour by, for example, a thousandth of the pitch circle diameter is filled by the liquid volume VR.
  • the play or gap thus generated between the two gearwheels in the position shown in FIG. 3b has the effect that the driving force exerted by the driven pinion is not transmitted, as in the theoretical case, at point E0, but rather is distributed over a fairly large area , which arises from the fact that the minimum gap is filled with fluid and this fluid fluid cushion transmits the driving force over a large width.
  • a force-transmitting tooth contact no longer takes place in the area of the engagement line parts E4 and E5 in FIG. 3a.
  • Fig. 3b finally shows that the inventive design with a minimum gap VR between the teeth in the position shown in Fig. 3 also ensures an excellent seal, since the remaining gap VR is extremely narrow over its entire length.
  • the circumferential extent of the tooth heads 4 or tooth spaces delimited by hypocycloids FR1 or tooth spaces delimited on the pitch circle T of the respective gear wheel 3, 6 is twice as large as the corresponding extent of the tooth spaces delimited by epicycloids FH1 or heads 5.
  • the rolling circle RH which describes the hypocycloid FR1
  • a particular advantage of the invention is that there are practically no radial and tangential accelerations and decelerations between the two gears.
  • the shortening of the tooth profiles which is also effective in the area of deepest tooth engagement, to an equidistant to the cycloid or flattened cycloid, which is one or a few hundredths of a millimeter behind, is generally one sixth to one third of the running clearance in the area compared to the point of deepest meshing.
  • the amount of residual squeezing oil in the invention which when the toothing is rotated further from the position shown in FIG. 3b to a position in which the center line of the pinion tooth on the axis spacing line covers - at least in the case of an oil pump - is not essential more than the thin oil film, which cannot be removed from the surface at all without excessively high pressures.
  • pinch oil hardly needs to be displaced, since the amount of oil remaining in the gap hardly exceeds the thin oil film that just fills the game.
  • the axial grooves 16 are provided in the center of the tooth space base of the pinion 6. As can be seen in the drawing, these grooves have a semicircular profile and merge angularly but not with sharp edges into the tooth space surface of the pinion.
  • Analog grooves can also be provided in the bottom of the tooth space of the internal gear at 17 for receiving pinch oil. These grooves are indicated by dashed lines in FIG. 5.

Claims (13)

  1. Machine à engrenages (pompe ou moteur pour liquides ou gaz), comprenant un carter avec une chambre à engrenages (2) munie d'ouvertures d'entrée et de sortie (10, 11), une couronne dentée à denture intérieure (3) disposée dans la chambre à engrenages (2), et un pignon (6) monté dans le carter (1) de manière à pouvoir tourner à l'intérieur de la couronne dentée (3), ledit pignon présentant une dent (5) de moins que la couronne dentée, étant en prise avec celle-ci et formant lors de la rotation, entre ses dents (5) et les dents (4) de la couronne dentée (3), des cellules à liquide tournantes qui augmentent et diminuent de volume et conduisent le liquide de l'entrée vers la sortie, les sommets des dents du pignon (6) et les entredents de la couronne dentée (3) présentant la forme d'épicycloïdes (FH) formées par le déroulement d'un premier cercle primitif (RE) sur le cercle primitif de fonctionnement (T) du pignon (6) et respectivement de la couronne dentée (3) (fig. 2), les entredents du pignon (6) et les sommets des dents de la couronne dentée (3) présentant la forme d'hypocycloïdes (FR) formées par le déroulement d'un second cercle primitif (RH) sur le cercle primitif de fonctionnement (T) du pignon (6) et respectivement de la couronne dentée (3) (fig. 2), et le rayon du premier cercle primitif (RE) différant de celui du second cercle primitif (RH), caractérisée en ce que l'extension circonférentielle (BH) des entredents du pignon et des dents (4) de la couronne délimités par des hypocycloïdes (FR1), mesurée sur le cercle primitif de fonctionnement respectif (TH, TR), est égale à 1,5 à 3 fois l'extension circonférentielle (BE) des dents (5) du pignon et des entredents de la couronne délimités par des épicycloïdes (FH1), mesurée sur le cercle primitif de fonctionnement respectif (TH, TR), et que les épicycloïdes (FH1) et/ou les hypocycloïdes (FR1) sont aplaties en direction de leurs cercles primitifs de fonctionnement (TH, TR) de telle façon que l'aplatissement et respectivement la somme des deux aplatissements (Z0-Z1; Z0-Z2) correspond au jeu radial (SR) relativement grand nécessaire entre les sommets des dents dans la région opposée à l'engrenure la plus profonde, tandis que la couronne dentée et le pignon (3, 6) dans la région à l'engrenure la plus profonde coopèrent avec un jeu beaucoup plus faible.
  2. Machine à engrenages selon la revendication 1, caractérisée en ce que l'extension circonférentielle (BH) des entredents du pignon et des dents (4) de la couronne est de 1,75 à 2,25 fois l'extension circonférentielle (BE) des dents (5) du pignon et des entredents de la couronne.
  3. Machine à engrenages selon la revendication 2, caractérisée en ce que l'extension circonférentielle (BH) des entredents du pignon et des dents de la couronne est égale à 2 fois l'extension circonférentielle (BE) des dents du pignon et des entredents de la couronne.
  4. Machine à engrenages selon l'une des revendications 1 à 3, caractérisée en ce que les cycloïdes de l'un des deux groupes de cycloïdes (hypocycloïdes (FR) et épicycloïdes (FH)) sont aplaties de la valeur totale du jeu nécessaire, tandis que l'aplatissement des cycloïdes de l'autre groupe est nul.
  5. Machine à engrenages selon la revendication 4, caractérisée en ce que les épicycloïdes (FH) sont aplaties.
  6. Machine à engrenages selon l'une des revendications 1 à 5, caractérisée en ce que l'aplatissement des cycloïdes (FR, FH) est obtenu par un léger décalage radial du point décrivant respectivement les cycloïdes à partir du cercle primitif (RH, RE) en direction du centre de celui-ci. (fig. 2).
  7. Machine à engrenages selon la revendication 6, caractérisée en ce que le point de départ (Z1, Z2) et le point d'extrémité de chaque cycloïde aplatie (FR1, FH1) sont reliés, sur le cercle primitif de fonctionnement (T), au point de départ et respectivement au point d'extrémité (Z0) de la cycloïde initiale non aplatie (FR, FH) par une droite.
  8. Machine à engrenages selon l'une des revendications 1 à 7, caractérisée en ce que l'aplatissement et respectivement la somme des deux aplatissements de cycloïdes, mesurés au milieu de la cycloïde, sont égaux à 1/2000 à 1/500 du diamètre du cercle primitif de fonctionnement (TH) de la couronne dentée (3).
  9. Machine à engrenages selon l'une des revendications 1 à 8, caractérisée en ce que le jeu minimum entre les flancs des dents nécessaire dans la région à l'engrenure la plus profonde est obtenu par un retrait équidistant du contour des dents.
  10. Machine à engrenages selon l'une des revendications 1 à 9, caractérisée en ce que le pignon (6) comporte sept à onze dents.
  11. Machine à engrenages selon l'une des revendications 1 à 10 pour liquides, caractérisée en ce qu'au fond des entredents au moins du pignon (6) sont prévues d'étroites gorges axiales (16).
  12. Machine à engrenages selon la revendication 11, caractérisée en ce que les gorges (16) présentent une largeur d'environ un quart à un sixième de la circonférence du cercle primitif (RH, RE) générant l'entredent, et de préférence d'un cinquième de celle-ci.
  13. Machine à engrenages selon l'une des revendications 11 ou 12, caractérisée en ce que les gorges présentent une largeur deux à trois fois plus grande que la profondeur.
EP92120410A 1992-01-15 1992-11-30 Machine à engrenages Expired - Lifetime EP0552443B1 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE4200883 1992-01-15
DE4200883A DE4200883C1 (fr) 1992-01-15 1992-01-15

Publications (2)

Publication Number Publication Date
EP0552443A1 EP0552443A1 (fr) 1993-07-28
EP0552443B1 true EP0552443B1 (fr) 1995-09-27

Family

ID=6449558

Family Applications (1)

Application Number Title Priority Date Filing Date
EP92120410A Expired - Lifetime EP0552443B1 (fr) 1992-01-15 1992-11-30 Machine à engrenages

Country Status (5)

Country Link
US (1) US5368455A (fr)
EP (1) EP0552443B1 (fr)
JP (1) JP2818723B2 (fr)
KR (1) KR0150804B1 (fr)
DE (2) DE4200883C1 (fr)

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0785360A1 (fr) 1996-01-17 1997-07-23 Mitsubishi Materials Corporation Rotor pour pompe à huile
EP1340914A2 (fr) * 2002-03-01 2003-09-03 Mitsubishi Materials Corporation Pompe à huile à engrenages internes
DE10208408A1 (de) * 2002-02-27 2003-09-11 Schwaebische Huettenwerke Gmbh Zahnradverzahnung

Families Citing this family (37)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE4022500A1 (de) * 1990-07-14 1992-01-16 Gisbert Prof Dr Ing Lechner Zahnradpumpe oder motor
DE4311165C2 (de) * 1993-04-05 1995-02-02 Danfoss As Hydraulische Maschine
DE4311168C2 (de) * 1993-04-05 1995-01-12 Danfoss As Hydraulische Maschine
US5957762A (en) * 1994-09-01 1999-09-28 The Gleason Works Internally toothed tool for the precision machining of gear wheels
JP3481335B2 (ja) * 1995-01-06 2003-12-22 ティーエスコーポレーション株式会社 内接噛合型遊星歯車装置
US6077059A (en) * 1997-04-11 2000-06-20 Mitsubishi Materials Corporation Oil pump rotor
ES2205538T3 (es) * 1997-09-04 2004-05-01 Sumitomo Electric Industries, Ltd. Bomba de engranajes internos.
KR100763642B1 (ko) * 1998-07-31 2007-10-05 더 텍사스 에이 앤드 엠 유니버시티 시스템 준등온 브레이턴 사이클 기관
DE50202167D1 (de) * 2002-03-01 2005-03-10 Hermann Haerle Zahnringmaschine mit Zahnlaufspiel
DE10224784A1 (de) * 2002-06-04 2003-12-18 Siemens Ag G-Rotorpumpe
JP2004092637A (ja) * 2002-07-11 2004-03-25 Yamada Seisakusho Co Ltd トロコイドポンプ
JP4107895B2 (ja) * 2002-07-11 2008-06-25 株式会社日本自動車部品総合研究所 内接噛合遊星歯車機構
US7118359B2 (en) * 2002-07-18 2006-10-10 Mitsubishi Materials Corporation Oil pump rotor
DE10245814B3 (de) * 2002-10-01 2004-02-12 SCHWäBISCHE HüTTENWERKE GMBH Innenzahnradpumpe mit verbesserter Füllung
GB2394512A (en) * 2002-10-22 2004-04-28 Concentric Pumps Ltd Pump rotor set with increased fill limit
US20060239848A1 (en) * 2002-10-29 2006-10-26 Mitsubishi Materials Corporation Internal gear type oil pump rotor
JP4557514B2 (ja) * 2003-07-15 2010-10-06 住友電工焼結合金株式会社 内接歯車式ポンプ及びそのポンプのインナーロータ
JP4169724B2 (ja) 2003-07-17 2008-10-22 株式会社山田製作所 トロコイド型オイルポンプ
MY138173A (en) * 2003-08-12 2009-05-29 Diamet Corp Oil pump rotor assembly
JP4485770B2 (ja) * 2003-09-01 2010-06-23 株式会社ダイヤメット オイルポンプロータ
JP2006125391A (ja) * 2004-09-28 2006-05-18 Aisin Seiki Co Ltd 内接ギヤ型ポンプのロータ構造
JP2006152928A (ja) * 2004-11-30 2006-06-15 Hitachi Ltd 内接式歯車ポンプ
JP4319617B2 (ja) 2004-12-27 2009-08-26 株式会社山田製作所 トロコイド型オイルポンプ
JP4608365B2 (ja) * 2005-01-13 2011-01-12 住友電工焼結合金株式会社 内接歯車ポンプの歯形創生方法及び内接歯車
KR101263037B1 (ko) * 2005-02-16 2013-05-09 에스티티 테크놀로지스 인크., 어 조인트 벤쳐 오브 마그나 파워트레인 인크. 앤드 에스하베 게엠베하 신규의 회전자 세트를 갖는 초승달형 기어 펌프
JP2006233771A (ja) 2005-02-22 2006-09-07 Mitsubishi Materials Pmg Corp ポンプロータ
KR100754995B1 (ko) * 2005-08-19 2007-09-04 주식회사 해성산전 사이클로이드 치형을 갖는 유성기어 감속기의 외치기어 및가공방법
US8096795B2 (en) 2005-09-22 2012-01-17 Aisin Seiki Kabushki Kaisha Oil pump rotor
US20070092392A1 (en) * 2005-10-20 2007-04-26 Aisin Seiki Kabushiki Kaisha Internal gear pump
WO2008111270A1 (fr) * 2007-03-09 2008-09-18 Aisin Seiki Kabushiki Kaisha Rotor de pompe à huile
WO2009130433A1 (fr) * 2008-04-22 2009-10-29 Concentric Vfp Limited Pompe avec fentes de remplissage
ES2656432T3 (es) * 2008-08-08 2018-02-27 Sumitomo Electric Sintered Alloy, Ltd. Rotor de bomba de engranajes internos, y bomba de engranajes internos que utiliza el rotor
EP2469092B1 (fr) 2009-11-16 2018-08-15 Sumitomo Electric Sintered Alloy, Ltd. Rotor pour pompe et pompe à engrenages internes qui utilise celui-ci
DE102010002585A1 (de) * 2010-03-04 2011-09-08 Robert Bosch Gmbh Innenzahnradpumpe
KR101270892B1 (ko) * 2011-11-01 2013-06-05 명화공업주식회사 사이클로이드 기어 펌프
DE102012022787A1 (de) 2012-11-22 2014-05-22 Volkswagen Aktiengesellschaft Zahnradpumpe sowie Regelsystem mit Zahnradpumpe und Regelkolben
KR101704368B1 (ko) * 2013-10-01 2017-02-07 마아크펌프시스템즈아게 개선된 펌프 유입구를 가지는 기어 펌프

Family Cites Families (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB223257A (en) * 1923-04-16 1924-10-16 Hill Engineering Company Inc Improvements in rotors for rotary compressors and the like
GB233423A (en) * 1924-02-07 1925-05-07 Hill Compressor & Pump Co Inc Improvements in or relating to rotary pumps or the like
US2344628A (en) * 1940-12-26 1944-03-21 Gar Wood Ind Inc Gear pump
CN1007545B (zh) * 1985-08-24 1990-04-11 沈培基 摆线等距线齿轮传动副及其装置
JPH01249971A (ja) * 1988-03-31 1989-10-05 Suzuki Motor Co Ltd トロコイドポンプ
DE3938346C1 (fr) * 1989-11-17 1991-04-25 Siegfried A. Dipl.-Ing. 7960 Aulendorf De Eisenmann
US5226798A (en) * 1989-11-17 1993-07-13 Eisenmann Siegfried A Gear ring pump for internal-combustion engines and automatic transmissions
US5163826A (en) * 1990-10-23 1992-11-17 Cozens Eric E Crescent gear pump with hypo cycloidal and epi cycloidal tooth shapes

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0785360A1 (fr) 1996-01-17 1997-07-23 Mitsubishi Materials Corporation Rotor pour pompe à huile
DE10208408A1 (de) * 2002-02-27 2003-09-11 Schwaebische Huettenwerke Gmbh Zahnradverzahnung
EP1340914A2 (fr) * 2002-03-01 2003-09-03 Mitsubishi Materials Corporation Pompe à huile à engrenages internes

Also Published As

Publication number Publication date
DE4200883C1 (fr) 1993-04-15
JP2818723B2 (ja) 1998-10-30
US5368455A (en) 1994-11-29
JPH05256268A (ja) 1993-10-05
EP0552443A1 (fr) 1993-07-28
KR930016665A (ko) 1993-08-26
KR0150804B1 (ko) 1998-11-02
DE59203844D1 (de) 1995-11-02

Similar Documents

Publication Publication Date Title
EP0552443B1 (fr) Machine à engrenages
EP0043899B1 (fr) Pompe à engrenages annulaires
EP1340912B1 (fr) Machine à engrenage interne avec jeu de dentures
DE2644531C2 (de) Hydrostatische Zahnradmaschine mit einem Trochoidenzahnradpaar
EP0619430B1 (fr) Pompe à engrenage internes pour gamme de vitesses rotatives élévées
EP3507496B1 (fr) Pompe à vide à compression sèche
DE2758376C2 (fr)
EP3384159B1 (fr) Pompe à engrenages extérieurs
DE10208408A1 (de) Zahnradverzahnung
EP0433576B1 (fr) Pompe à engrenages annulaires pour moteurs à combustion interne et transmissions automatiques
EP2235374A2 (fr) Pompe à engrenage intérieur, à volume variable
EP1252444B1 (fr) Systeme d'entrainement de pompe a vis
EP0737812B1 (fr) Pompe à engrenages
DE3117651A1 (de) Drehkolbenmotor mit im durchmesser vergroesserter antriebswelle
EP1406015B1 (fr) Pompe à engrenage interne avec remplissage amélioré
DE4311165C2 (de) Hydraulische Maschine
DE3209945A1 (de) Antriebswellenanordnung fuer kreiskolbenmaschinen
DE102016207093B4 (de) Zahnradfluidmaschine
DE602107C (de) Als Pumpe, Motor oder einfaches UEbertragungsgetriebe verwendbare Vorrichtung aus zwei ineinander angeordneten Teilen mit dauernd in Beruehrung stehenden Zaehnen
DE3225790A1 (de) Pumpe oder motor
DE2203868A1 (de) Zahnradpumpe mit Abdichtungsplatten
DE4440782C5 (de) Innenzahnradpumpe mit Verdrängervorsprüngen
DE202014006761U1 (de) Hydrostatische Kreiskolbenmaschine nach dem Orbitprinzip
EP1182349A2 (fr) Pompe à engrenages internes avec des roues à denture hélicoidale
AT311178B (de) Gas- und Hydromotor oder Pumpe

Legal Events

Date Code Title Description
PUAI Public reference made under article 153(3) epc to a published international application that has entered the european phase

Free format text: ORIGINAL CODE: 0009012

AK Designated contracting states

Kind code of ref document: A1

Designated state(s): CH DE FR GB IT LI SE

17P Request for examination filed

Effective date: 19931104

17Q First examination report despatched

Effective date: 19941222

GRAA (expected) grant

Free format text: ORIGINAL CODE: 0009210

ITF It: translation for a ep patent filed

Owner name: BARZANO' E ZANARDO MILANO S.P.A.

AK Designated contracting states

Kind code of ref document: B1

Designated state(s): CH DE FR GB IT LI SE

REF Corresponds to:

Ref document number: 59203844

Country of ref document: DE

Date of ref document: 19951102

GBT Gb: translation of ep patent filed (gb section 77(6)(a)/1977)

Effective date: 19951017

ET Fr: translation filed
PLBE No opposition filed within time limit

Free format text: ORIGINAL CODE: 0009261

STAA Information on the status of an ep patent application or granted ep patent

Free format text: STATUS: NO OPPOSITION FILED WITHIN TIME LIMIT

26N No opposition filed
REG Reference to a national code

Ref country code: GB

Ref legal event code: IF02

REG Reference to a national code

Ref country code: CH

Ref legal event code: NV

Representative=s name: RIEDERER HASLER & PARTNER PATENTANWAELTE AG

PGFP Annual fee paid to national office [announced via postgrant information from national office to epo]

Ref country code: DE

Payment date: 20101119

Year of fee payment: 19

PGFP Annual fee paid to national office [announced via postgrant information from national office to epo]

Ref country code: GB

Payment date: 20101118

Year of fee payment: 19

Ref country code: IT

Payment date: 20101126

Year of fee payment: 19

PGFP Annual fee paid to national office [announced via postgrant information from national office to epo]

Ref country code: FR

Payment date: 20111130

Year of fee payment: 20

Ref country code: SE

Payment date: 20111128

Year of fee payment: 20

Ref country code: CH

Payment date: 20111123

Year of fee payment: 20

REG Reference to a national code

Ref country code: DE

Ref legal event code: R071

Ref document number: 59203844

Country of ref document: DE

REG Reference to a national code

Ref country code: DE

Ref legal event code: R071

Ref document number: 59203844

Country of ref document: DE

REG Reference to a national code

Ref country code: CH

Ref legal event code: PL

REG Reference to a national code

Ref country code: GB

Ref legal event code: PE20

Expiry date: 20121129

REG Reference to a national code

Ref country code: SE

Ref legal event code: EUG

PG25 Lapsed in a contracting state [announced via postgrant information from national office to epo]

Ref country code: GB

Free format text: LAPSE BECAUSE OF EXPIRATION OF PROTECTION

Effective date: 20121129