EP2275683A1 - Procédé de commande d'une pompe à engrenages et application du procédé - Google Patents

Procédé de commande d'une pompe à engrenages et application du procédé Download PDF

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Publication number
EP2275683A1
EP2275683A1 EP09163048A EP09163048A EP2275683A1 EP 2275683 A1 EP2275683 A1 EP 2275683A1 EP 09163048 A EP09163048 A EP 09163048A EP 09163048 A EP09163048 A EP 09163048A EP 2275683 A1 EP2275683 A1 EP 2275683A1
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EP
European Patent Office
Prior art keywords
gear
gear pump
pump
drive
center
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.)
Granted
Application number
EP09163048A
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German (de)
English (en)
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EP2275683B1 (fr
Inventor
Markus Aregger
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.)
Maag Pump Systems AG
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Maag Pump Systems AG
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 Maag Pump Systems AG filed Critical Maag Pump Systems AG
Priority to EP09163048.3A priority Critical patent/EP2275683B1/fr
Priority to US12/818,615 priority patent/US20100322806A1/en
Priority to US12/818,502 priority patent/US8500414B2/en
Priority to JP2010138975A priority patent/JP2011001958A/ja
Publication of EP2275683A1 publication Critical patent/EP2275683A1/fr
Application granted granted Critical
Publication of EP2275683B1 publication Critical patent/EP2275683B1/fr
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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
    • F04C2/12—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
    • F04C2/18—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 with similar tooth forms
    • 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
    • F04C14/00—Control of, monitoring of, or safety arrangements for, machines, pumps or pumping installations
    • F04C14/08—Control of, monitoring of, or safety arrangements for, machines, pumps or pumping installations characterised by varying the rotational speed
    • 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
    • F04C2240/00—Components
    • F04C2240/40—Electric motor
    • F04C2240/402—Plurality of electronically synchronised motors

Definitions

  • the present invention relates to a method according to the preamble of claim 1 and an application of the method.
  • Gear pumps consist of two intermeshing gears which are mounted on shafts, usually a shaft is connected to a drive unit.
  • the shaft which is not driven by a drive unit, is driven by means of torque transmission from the driven shaft via the tooth flanks.
  • a gear pump is known, in turn, two drive units are provided for individually driving the waves, wherein phase and angular velocity of the intermeshing gears are coordinated such that on the one hand lifting of tooth flanks of intermeshing gears and on the other hand too high excess torque on the tooth flanks Interlocking gears are avoided.
  • the present invention initially relates to a method for controlling a gear pump consisting of two intermeshing gears, in which the two gears are each driven via corresponding shafts with a drive unit.
  • the invention is characterized in that a momentary position of the one gear with respect to a current position of the other Gear is determined and that the current position of the one gear with respect to the current position of the other gear is set according to predefined operating conditions continuously.
  • a variant is characterized in that the determination of the instantaneous position of the one gear with respect to the current position of the other gear is set by a reference value determined before the normal operation of the gear pump or during interruptions of the normal operation of the gear pump.
  • FIG. 1 Further embodiments of the present invention are characterized in that the reference value in the middle between tooth flanks of a tooth gap of a toothed wheel, preferably in the middle between tooth flanks of a tooth gap of a toothed wheel, is located.
  • a method for automatic calibration of the arrangement is indicated with a gear pump.
  • the system can carry out this calibration both before commissioning and during interruptions to operation without further action by the operating personnel.
  • a possible wear of tooth flanks can be determined, because then resp. the difference between the first angular difference and the second angular difference also increases. An excessive occlusion can then be detected by a simple threshold violation.
  • each encoder / sensor unit are arranged centrally between the teeth of the respective gear and a rotor of the respective drive.
  • a central arrangement of the encoder / sensor units has the advantage that an existing angle of rotation due to a non-ideal rigidity of the entire drive train has a reduced influence on the measurement error of the system.
  • the measurement error is halved by the central arrangement.
  • This operating condition is also referred to as an edge change, because the touching tooth flanks change in the course of a Ausquetschvorganges.
  • An exact torque adjustment can be achieved by appropriate control of the speeds or instantaneous positions of the gears to each other.
  • the tooth flanks thus transmit an arbitrarily adjustable torque, but never lift off each other during operation, whenever a defined flank seal is to be achieved.
  • Further embodiments of the present invention are characterized in that the rotational speed of the shafts driven by the drive units is set so synchronously that a delivery medium pressure on a delivery side of the gear pump is substantially constant.
  • Further embodiments of the present invention are characterized in that the delivery medium pressure is measured on the pressure side of the gear pump and that the speed is adjusted in dependence on the measured delivery medium pressure.
  • a gear pump comprising a pump housing, two housed in the pump housing and intermeshing gears and two shafts, which are operatively connected to the gears and are guided by the pump housing, wherein the two waves with are each a drive unit operatively connected, wherein between the gear and the drive unit in each case a coupling unit for equalizing eccentricities between the drive unit and the respective shaft is arranged and wherein between the gear center and the drive center in each case a rotary encoder / sensor unit is arranged.
  • An embodiment variant of the present application is characterized in that the rotary encoder / sensor unit lies in an axial region which is defined by the middle between the center of the gear and the center of the drive plus a deviation on both sides of at most 10% of the distance between the center of the gear and the center of the drive.
  • rotary encoder / sensor units are each arranged in the middle between the respective gear center and the respective drive center.
  • rotary encoder / sensor unit to the axis of rotation of the respective shaft have a radial distance which is larger, preferably at least twice as large as an outer radius of the gears.
  • rotary encoder / sensor units are either optical or magnetic rotary encoder / sensor units.
  • rotary encoder / sensor units are arranged such that a perpendicular to the shaft and extending through the corresponding encoder / sensor unit Connecting line with a centrally extending between the two axes of rotation plane suction side an angle in the range of 35 ° to 55 °, preferably 40 ° to 50 °, preferably 45 °, includes.
  • each drive unit comprises a rotor and a stator, wherein the rotor is axially displaceable with respect to the stator.
  • drive units each have a compensating bearing unit on the side facing away from the gear pump, which radially supports the respective rotor of the drive unit.
  • the coupling unit is a diaphragm coupling.
  • FIG. 10 Further embodiments of the present application are characterized in that a flange between the pump housing (10) and the stator of the respective drive unit is arranged, wherein the flange holes has, through which circulates a cooling medium for adjusting the temperature.
  • connection between the drive units and the respective shafts of the gear pump are conical polygonal connections.
  • the one drive unit, the gear pump and the other drive unit are each contained in a temperature zone in which the temperatures are adjustable to predetermined values, wherein between adjacent temperature zones preferably isolation areas are present.
  • a known arrangement is shown with a gear pump 1, the conveying medium F from a suction side S on a pressure side D promotes. It is evident in Fig. 1 a pump housing 10, are guided by the waves 2 and 3 to the outside.
  • the guided outward shaft 3 is connected via a first universal joint 4, an adjustable in length axle 6 and a second universal joint 6 with a drive unit 7. Accordingly, the guided outward shaft 2 via a corresponding first and second universal joint and a corresponding axis section with a further drive unit (in Fig. 1 not shown).
  • gears in Fig. 1 not visible
  • the double universal joint consisting of the first and second universal joint 4 and 5 is provided together with the adjustable axle section 6 for receiving lateral and angular deviations of the drive unit with respect to the shaft 2 and 3, respectively.
  • Due to the double shaft joint in combination with the adjustable Achsabites 6 acts an additional bearing force on a shaft 10 contained in the pump housing shaft bearing. This additional bearing force is due to the dead weight of the double-pivot joint and the axle section 6.
  • the additional bearing force is due to a relatively short bearing distance of the pump bearings, which are in the pump housing 10 for supporting the shafts 2 and 3, with respect to the length of the double Swing joint considerably.
  • Fig. 2 is a section through an inventive arrangement shown with a gear pump 1, wherein the cutting plane in the axes of rotation 13 and 14 of the shafts 2 and 3 and by a sensor 25 is placed, according to the in Fig. 4 Plotted sectional plane AA.
  • Fig. 2 shows the sake of simplicity, only one half of the gear pump 1. Accordingly, only one drive unit 7 is shown.
  • the drive unit 7 is pressed directly via a flange 15, ie without intermediate gear, to the pump housing 10 or its lid. Via a screw 21, the rotating parts of the drive unit 7, such as a hub 16, a diaphragm coupling 22 and a rotor 18, connected to the shaft 3 of the gear pump 1.
  • the screw 21 can be released if necessary, whereby the drive unit 7 can be solved by the gear pump 1 again. After loosening screws 40, which connects the flange 15 with the pump housing 10 or with its lid, and after loosening the screw 21, the complete drive unit 7 can be released from the gear pump 1.
  • the shafts 2, 3 of the gear pump and their bearing units stay within the gear pump and can be disassembled individually.
  • the drive unit 7 is in addition to the flange 15 and the hub 16 further comprises a rotor 18, a stator 17 and a drive cover 19 with an opening 20.
  • the drive cover 19 closes the drive unit 7 on the side facing away from the gear pump 1 and is connected to the stator 17, wherein the opening 20 is arranged centrally on the extended axis of rotation 13 of the shaft 3.
  • Gear pump side, the stator 17 is connected to the flange 15.
  • the gear pump 1 is directly, ie without intermediate gear, connected to the drive unit 7.
  • the screw 21 is provided by means of which the rotor 18 is fixed axially via the hub 16 and the flange 15. The screw 21 is guided during the assembly of the drive unit 7 to the gear pump 1 through the opening 20 in the drive cover 19 along the axis of rotation 13 of the shaft 3 and secured in a corresponding bore in the shaft 3.
  • the hub 16 is connected to the shaft 3 via a so-called conical polygon connection, on the one hand allows a precise axial alignment of the rotor 18 to the shaft 3, on the other hand, a very torsionally rigid connection between the rotor 18 of the drive unit 7 and the driven shaft 3 of the gear pump 1 allows.
  • the membrane coupling 22 and the hub 16 are conceivable, for example, as a single part, as well as from Fig. 2 shows that in the left, drive side half of the individual part of the classic function of a hub which can be coupled to the shaft 3, in the right part of this single part is thin-walled and thus fulfills the functions of a membrane coupling.
  • a so-called torque motor is used, which is a high-pole permanent-magnet three-phase synchronous motor with hollow shaft rotor for the direct, above-mentioned coupling to the gear pump.
  • Torque motors are characterized in particular by a short compact design and a low torsional backlash (high torsional rigidity).
  • a rotary encoder 24 is arranged, which cooperates with a connected to the stator 17 sensor unit 25.
  • a grid is applied to the hub 16, which is read by the sensor unit 25.
  • corresponding magnetic measuring devices or other methods for position determination can also be used.
  • the rotary encoder 24 In order to minimize any measurement errors due to eccentricity of the rotary encoder 24 to the toothing, the rotary encoder 24 is made as large as possible in diameter. The eccentricity of the encoder 24 itself is minimized by the integration of the inclusion of the encoder 24 in the hub 16. Since the hub 16 is in one piece, very close manufacturing tolerances can be adhered to the inclusion of the encoder 24.
  • the sensor unit 25 is preferably selected between the center of the rotor 18 or stator 17 and the center of the driven gear 11 of the gear pump 1. In a uniform Stiffness distribution over the drive train (ie between the center of the rotor 18 and stator 17 and the center of the driven gear 11 of the gear pump 1) is the encoder 24, respectively.
  • the sensor unit 25 is preferably arranged in the middle between the center of the rotor 18 and stator 17 and the center of the driven gear 11 of the gear pump 1.
  • a possible field of application of the arrangement with a gear pump is the pressure build-up downstream of an extruder in the conveyance of plastic melts in an extrusion line.
  • the polymer melts are conveyed at temperatures of up to 300 ° C. against high discharge pressures (eg 300 bar).
  • high discharge pressures eg 300 bar
  • high drive power and thus high torques are necessary.
  • the gear pump or the pump housing is heated to a temperature of, for example, 300 ° C, caused by the conveying medium, while the temperature of the drive units 7 and 8, especially for the necessary electronic circuits in these, should not exceed 60 ° C.
  • insulating dividing wall 30 and 31 are required, respectively between the temperature zones 32 and 33 resp. between the temperature zones 33 and 34 are present.
  • additional measures are required as needed, so that the temperature in the cold temperature zones 32 and 34 does not reach unacceptable levels.
  • An additional measure for example, is that an active cooling (for example, an active water cooling) is provided.
  • the rotor 18 ( Fig. 2 ) to protect from over temperature by a cooling of the flange 15 is connected between the hub 16 and the gear pump 1.
  • the cooling is realized for example by star-shaped holes in the flange 15. This achieves very good cooling properties, since the deflections produce high turbulences.
  • the hub 16 is cooled on the entire surface flange side by radiation and forced convection.
  • Fig. 4 shows a possible positioning of the sensor unit 25, which is used to determine the current position of the one Gear is used in relation to the current position of the other gear, wherein Fig. 4 a section transverse to the axes of rotation 11 and 13 of the shafts 2 and 3 shows.
  • the fluid is transported in the direction of arrow from the suction side S with the gear pump on the pressure side D.
  • a force component is generated in the direction of the arrows P, P ', which act on the shaft bearing of the gear pump and to a slight displacement of the shafts 2 and 3 (FIGS. Fig. 2 ) to lead.
  • the sensor unit 25 will now move in the direction of displacement, i. in the direction of deflection of the shaft, attached.
  • the attachment takes place, for example, below 45 ° and is thus on average of the possible displacement angle, which is differential pressure dependent and viscosity-dependent.
  • z. B. a schwradbreiten- and game size-dependent arrangement of the sensor unit 25 are made.
  • Fig. 5 shows a section transverse to the axes of rotation 13 and 14 in the region of the gears 11 and 12.
  • delivery medium F is received on the suction side S of the tooth gaps and then along the pump housing to the pressure side D transported where the medium F is squeezed out by the meshing gears 11, 12.
  • a "trapped volume” is created in the toothed area between the tooth base and the tooth tip of the gears, which is sealed off by the almost touching tooth flanks in front of and behind this volume.
  • a flow gap can be generated in a targeted manner at the locations where a large flow gap is desired for tribological reasons (optimum gap thickness relative to the tooth flanks). Due to the existing position control of the shafts, the ratio of these two sealing gaps can be actively controlled. Once, the gap leading to the "trapped volume” can be minimized, once the gap following the trapped volume. This makes it possible for the squeezing process to be actively influenced from this "trapped volume", thus optimizing the uniformity of the flow.
  • the first shaft 2 drives the second shaft 3 with a defined torque.
  • a first absolute rotation angle difference with the aid of the illustrated rotary encoder 24 in combination with the sensor unit 25 (FIG. Fig. 2 ) at both shafts 2 and 3 is determined by determining a difference between a measured value of one sensor unit 25 and a measured value of the other sensor units 25 '.
  • the second shaft 3 drives the first shaft 3 with the same defined torque as in the first step.
  • a second absolute rotation angle difference is in turn determined by means of the described rotary encoder 24 in combination with the sensor unit 25 in both shafts 2 and 3, again determining the difference between a measured value of one sensor unit 25 and a measured value of the other sensor units 25 ' ,
  • a difference between the first absolute rotation difference and the second absolute rotation difference is formed.
  • This difference is the actual range in which the gears can move to each other, provided that the defined torque that was used in the first and second step in the measurement is not exceeded.
  • a reference value can now be defined, in relation to which the current positions of the gears are specified.
  • the reference value is then a zero point of a defined coordinate system. For example, the reference value lies in the middle between tooth flanks of a tooth gap, so that the absolute values of the maximum deflections are identical.
  • the operating conditions can now be selected in a first setting, for example, such that one gear transmits half plus a defined percentage of the total torque. Accordingly, then transmits the other gear half minus the defined percentage of total torque.
  • the backlash between the flanks of two intermeshing teeth is selectable, namely for example, in 10% increments from the contact of the flanks (no backlash) over a central alignment (ie, the tooth entering a tooth gap lies exactly in the middle of the gap) until the tooth flanks touch again, this time around the trailing tooth flanks is.
  • Fig. 6 illustrates the operating conditions just explained in turn in a section transverse to the axes of rotation 13 and 14 in the region of the gears 11 and 12. Again, in the meshing region of the gears 11 and 12, a detail X shown enlarged as detail, in which also an adjusted backlash 26th is shown highlighted.
  • the edge change operation is based on Fig. 7 explained, in turn, sections transverse to the axes of rotation 13, 14 in a region of the meshing gears 11, 12 show.
  • a state is shown in which the tooth Z 1 'of the toothed wheel 11, which engages in a tooth gap of the toothed wheel 12, contacts the tooth Z 1 .
  • a state shown later in time in which the engaging in a tooth gap of the gear 11 tooth Z 2 of the gear 12, the tooth Z 1 'touches.
  • the operating conditions according to the mentioned edge change operation are used, for example, in highly viscous media in which the squeezing pressure is so great that a very large torque is required to produce the squeeze pressure energy, as these represent a pure energy loss.
  • pressure fluctuations are eliminated or at least greatly reduced by actively influencing the rotational speeds of the two gear shafts.
  • the inventive arrangement or the inventive method is capable of the speed curve per Ausquetschvorgang to vary, in such a way that the pressure on the pressure side is within narrow limits resp. that the pressure on the pressure side is constant.
  • the Ausquetschvorgang of the pumped medium from the tooth base is controlled specifically on the current position of the one gear with respect to the current position of the other gear.
  • a degree of coverage of 1 must be selected. If an overlap degree of 1 is selected, then only one pair of teeth is involved in the displacement, ie the squeezing (see Vogel subuch Jarosla and Monika Ivantysyn: “Hydrostatic Pumps and Motors", 1993, p 319). In this case, a sinusoidal course of the displacement volume flow results. This can be easily and efficiently corrected via a sinusoidal compensation table (for example, a so-called "look-up" table).
  • the speed curve 90 Due to the periodicity, it is possible to store the speed curve 90 in a memory unit (Look-up Table). The values for the speed to be set are then read out in a predetermined cycle, the predetermined clock resulting from the pressure to be set on the pressure side.
  • the selective influencing of the position control can also be used for shear-sensitive materials to reduce a total shear stress.
  • care is taken to ensure that a maximum permissible shear load is not exceeded.
  • the present invention makes it possible for the first time to specifically influence the effects of pulsation, crushing pressures and tribological behavior.
  • the settings may take into account all effects that are relevant to the specific case, or individual operating conditions may be considered as a priority. By this is meant that these operating conditions should have a more significant influence on the behavior of the overall system.

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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)
EP09163048.3A 2009-06-18 2009-06-18 Procédé de commande d'une pompe à engrenages Active EP2275683B1 (fr)

Priority Applications (4)

Application Number Priority Date Filing Date Title
EP09163048.3A EP2275683B1 (fr) 2009-06-18 2009-06-18 Procédé de commande d'une pompe à engrenages
US12/818,615 US20100322806A1 (en) 2009-06-18 2010-06-18 Arrangement including a gear pump
US12/818,502 US8500414B2 (en) 2009-06-18 2010-06-18 Method of controlling a gear pump as well as an application of the method
JP2010138975A JP2011001958A (ja) 2009-06-18 2010-06-18 歯車ポンプを制御する方法及び方法の使用

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
EP09163048.3A EP2275683B1 (fr) 2009-06-18 2009-06-18 Procédé de commande d'une pompe à engrenages

Publications (2)

Publication Number Publication Date
EP2275683A1 true EP2275683A1 (fr) 2011-01-19
EP2275683B1 EP2275683B1 (fr) 2017-01-11

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US (1) US8500414B2 (fr)
EP (1) EP2275683B1 (fr)
JP (1) JP2011001958A (fr)

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US10801499B2 (en) 2017-08-28 2020-10-13 Jtekt Corporation External gear pump

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CN206206150U (zh) * 2014-02-28 2017-05-31 凤凰计划股份有限公司 与两个独立驱动的原动机成一体的泵
EP3123029B1 (fr) 2014-03-25 2024-03-20 Project Phoenix, LLC Système de pompage de fluide et commande associée
US10294936B2 (en) 2014-04-22 2019-05-21 Project Phoenix, Llc. Fluid delivery system with a shaft having a through-passage
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US10072676B2 (en) 2014-09-23 2018-09-11 Project Phoenix, LLC System to pump fluid and control thereof
WO2016057321A1 (fr) 2014-10-06 2016-04-14 Afshari Thomas Ensemble actionneur linéaire et système associé
US10677352B2 (en) * 2014-10-20 2020-06-09 Project Phoenix, LLC Hydrostatic transmission assembly and system
TWI777234B (zh) * 2015-09-02 2022-09-11 美商鳳凰計劃股份有限公司 泵送流體之系統及其控制
US10865788B2 (en) 2015-09-02 2020-12-15 Project Phoenix, LLC System to pump fluid and control thereof
DE102016113366A1 (de) * 2016-07-20 2018-01-25 Weber-Hydraulik Gmbh Hydraulikaggregat
CN109854498A (zh) * 2019-03-06 2019-06-07 郑州沃华机械有限公司 一种双驱动轴熔体泵及其控制方法
KR102298877B1 (ko) * 2020-08-11 2021-09-06 현대자동차 주식회사 부등 피치 모사 제어를 통한 기어 펌프 소음 저감 제어 장치 및 방법

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CH659290A5 (de) 1982-07-08 1987-01-15 Maag Zahnraeder & Maschinen Ag Zahnradpumpe.
EP0382029A1 (fr) * 1989-01-30 1990-08-16 Matsushita Electric Industrial Co., Ltd. Unité d'entraînement synchrone à deux axes et machine à tailler les engrenages utilisant celle-ci
US5314312A (en) * 1992-01-31 1994-05-24 Matsushita Electric Industrial Co., Ltd. Fluid-rotating apparatus
EP0697523A2 (fr) * 1994-08-19 1996-02-21 Diavac Limited Machine à vis pour fluide
DE19522515A1 (de) * 1995-06-21 1997-01-02 Sihi Ind Consult Gmbh Verfahren zum Stillsetzen einer Verdrängermaschine
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* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP2505335A2 (fr) 2011-03-31 2012-10-03 Frank Reineke Dispositif d'extrusion pour la production d'un profil en forme de bande ou tube en matière plastique ou de caoutchouc
US10801499B2 (en) 2017-08-28 2020-10-13 Jtekt Corporation External gear pump

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US8500414B2 (en) 2013-08-06
JP2011001958A (ja) 2011-01-06
US20100322805A1 (en) 2010-12-23
EP2275683B1 (fr) 2017-01-11

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