EP1961961A2 - Dispositif de contrôle de la pression dans une pompe à huile - Google Patents

Dispositif de contrôle de la pression dans une pompe à huile Download PDF

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Publication number
EP1961961A2
EP1961961A2 EP07122748A EP07122748A EP1961961A2 EP 1961961 A2 EP1961961 A2 EP 1961961A2 EP 07122748 A EP07122748 A EP 07122748A EP 07122748 A EP07122748 A EP 07122748A EP 1961961 A2 EP1961961 A2 EP 1961961A2
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EP
European Patent Office
Prior art keywords
circumferential side
side rotor
discharge passage
passage
discharge
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
EP07122748A
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German (de)
English (en)
Other versions
EP1961961B1 (fr
EP1961961A3 (fr
Inventor
Keiichi Kai
Kenichi Fujiki
Yasunori Ono
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.)
Yamada Manufacturing Co Ltd
Original Assignee
Yamada Manufacturing Co Ltd
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 Yamada Manufacturing Co Ltd filed Critical Yamada Manufacturing Co Ltd
Publication of EP1961961A2 publication Critical patent/EP1961961A2/fr
Publication of EP1961961A3 publication Critical patent/EP1961961A3/fr
Application granted granted Critical
Publication of EP1961961B1 publication Critical patent/EP1961961B1/fr
Not-in-force legal-status Critical Current
Anticipated expiration legal-status Critical

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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
    • F04C14/00—Control of, monitoring of, or safety arrangements for, machines, pumps or pumping installations
    • F04C14/24—Control of, monitoring of, or safety arrangements for, machines, pumps or pumping installations characterised by using valves controlling pressure or flow rate, e.g. discharge valves or unloading valves
    • F04C14/26—Control of, monitoring of, or safety arrangements for, machines, pumps or pumping installations characterised by using valves controlling pressure or flow rate, e.g. discharge valves or unloading valves using bypass channels
    • 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
    • F04C14/06—Control of, monitoring of, or safety arrangements for, machines, pumps or pumping installations specially adapted for stopping, starting, idling or no-load operation
    • F04C14/065—Capacity control using a multiplicity of units or pumping capacities, e.g. multiple chambers, individually switchable or controllable
    • 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 present invention relates to an oil pump pressure control device for a variable flow rate oil pump in which two discharge ports are provided and which uses three rotors as means for provision of the two discharge sources thereof that facilitates reduced friction while maintaining pressure (oil pressure) characteristics not in multi stage characteristics but in characteristics identical to the pressure characteristics of a common oil pump based on a newly devised method of switching oil passages.
  • Japanese Unexamined Patent Application No. H11-280666 As disclosed in Japanese Unexamined Patent Application No. H11-280666 , the use of three rotors that themselves serve as means for transporting oil to two locations is well known in the conventional art. Japanese Unexamined Patent Application No. H11-280666 refers only to means for transporting oil to two locations and makes no reference to the subsequent effects thereof.
  • the tooth profile is determined once the flow rate produced by an inner side rotor and outer side rotor is established and, accordingly, the flow rate ratio remains substantially constant no matter what the revolutions. More particularly, a pressure difference is generated between the space between the inner-side teeth and the space between the outer-side teeth due to the characteristics of the inner-side tooth profile and the outer-side tooth profile which, as a result, there is a fear of the middle rotor being pressed against one side and, in turn, of the wear of the teeth surface thereof being accelerated. In other words, a significant drawback inherent to this device pertains to the generation of biased wear on the middle rotor caused by pressure difference.
  • Japanese Unexamined Patent Application No. 2005-140022 describes a device designed with the aim of decreasing superfluous work and increasing efficiency at the low revolution range based on oil being relieved (returned) at a desired revolution range. Referring to FIG. 8 of page 13 thereof, superfluous work is reduced and efficiency is increased as a result of drop in the flow rate at a desired number of revolutions. While the efficiency can be increased by the adoption of this kind of variable flow rate oil pump, the following problems are inherent thereto.
  • the port position and opened area are governed by fixed design restrictions.
  • Another drawbacks pertains to the rotor being a single rotor assembly and, accordingly, to the fact that there is no means available for reducing superfluous work when it occurs.
  • the problem to be solved by the present invention is to produce a three rotor-using oil pump for a variable flow rate oil pump in which two discharge ports are provided and which uses three rotors as means for provision of the two discharge sources thereof that facilitates reduced friction while maintaining pressure (oil pressure) characteristics identical to the pressure characteristics of a common oil pump (The oil pump of Japanese Unexamined Patent Application No. 2002-70756 that exhibits the nonlinear stepped characteristic passing through the broken line as shown in FIG. 10 of page 7 thereof, and comprises a valve with a ON/OFF relief function. In addition, which exhibits approximately one characteristic inflection point.) rather than exhibiting multi-stepped characteristics based on a newly devised method of switching oil passages.
  • the oil pump pressure control device of the invention of claim 1 comprising: an oil pump having three rotors including an outer circumferential side rotor and an inner circumferential side rotor; a first discharge passage for feeding oil from the abovementioned outer circumferential side rotor to an engine; a return passage that returns to an intake side of the abovementioned outer circumferential side rotor; a second discharge passage for feeding oil from the abovementioned inner circumferential side rotor to the engine; a return passage that returns to an intake side of the abovementioned inner circumferential side rotor; and a pressure control valve whose valve main body is provided between the discharge port from the abovementioned inner circumferential side rotor and the abovementioned first discharge passage, wherein the abovementioned first discharge passage and the abovementioned first discharge passage and the abovementioned
  • the oil pump pressure control device of the invention of claim 2 comprising: an oil pump having three rotors including an outer circumferential side rotor and an inner circumferential side rotor; a first discharge passage for feeding oil from the abovementioned outer circumferential side rotor to an engine; a first return passage that returns to an intake side of the abovementioned outer circumferential side rotor; a second discharge passage for feeding oil from the abovementioned inner circumferential side rotor to the engine; a second return passage that returns to an intake side of the abovementioned inner circumferential side rotor; and a pressure control valve whose valve main body is provided between a discharge port from the abovementioned inner circumferential side rotor and the abovementioned first discharge passage, wherein the abovementioned first discharge passage and the abovementioned second discharge passage are coupled, and a flow passage control is executed in each of: a low revolution range in a
  • the abovementioned problems are able to be resolved by the oil pump pressure control device of the invention of claim 2 according to the configuration described above in which the abovementioned pressure control valve is configured as a type comprising a first valve portion and a second valve portion.
  • the abovementioned problems are able to be resolved by the oil pump pressure control device of the invention of claim 3 according to the configuration described above in which the abovementioned pressure control valve is configured as a type comprising a first valve portion, a second valve portion and a third valve portion.
  • a significant merit of the invention of claim 1 is that, even in variable flow rate oil pumps that use three rotors, a pressure ratio determined at the design stage can eliminate the other pressure difference produced by the discharge passages and return passages and can prevent a middle rotor from which the inner circumferential side rotor is configured from being pressed to one side whereupon, in turn, friction on the teeth surface can be prevented and the durability thereof improved.
  • a further effect thereof is that at times of high-speed revolution of the outer circumferential side rotor and the inner circumferential side rotor the second discharge passage of the inner circumferential side rotor is fully closed so as to form the inner circumferential side rotor as an independent circuit whereupon, even in the absence of the generation of a superfluous work pressure by the inner circumferential side rotor, there is no drop in overall pump pressure.
  • work pressure x flow rate
  • the superfluous work can be reduced so long as the pressure is lowered and, because the outer circumferential side rotor pump (main pump) and the inner circumferential side rotor pump (sub pump) are not in communication, the sub pump pressure is significantly lowered.
  • the inner circumferential side pump forms an independent circuit at times of high revolution, if the opened area return passage of the pump is increased more oil can be discharged and the pump pressure further reduced. Because a single discharge port is not partitioned into two for two discharge sources thereof and instead a single discharge port can be established for each, the flow rate is not partitioned. Accordingly, because the rotor diameter of the three rotors can be reduced and the slide area of the rotors decreased comparative to a common single rotor assembly, the friction (torque) can be reduced and the pump efficiency increased.
  • An additional merit of the invention is that, because the three rotors can be configured as a rotor having a dual-rotor assembly, the superfluous work of one rotor can be reduced.
  • the invention of claim 2 affords an identical effect to the effect afforded by the invention of claim 1.
  • the invention of claim 3 based on the provision of a simple pressure control valve, affords a reduction in the number of component parts.
  • the invention of claim 4 based on the provision of a 3-valve-type pressure control valve, facilitates a pressure control by which desired values can be better approximated.
  • the device shown in the diagrams is a three rotor-type oil pump principally configured from an outer rotor 1, a middle rotor 2 and an inner rotor 3.
  • An outer circumferential side intake port 4 and an outer circumferential side discharge port 5 are provided in the abovementioned outer rotor 1 and the abovementioned middle rotor 2, and an inner circumferential side intake port 6 and inner circumferential side discharge port 7 are provided in the abovementioned middle rotor 2 and inner rotor 3.
  • the abovementioned outer rotor 1, the abovementioned middle rotor 2, the outer circumferential side intake port 4 and the outer circumferential side discharge port 5 generically describe an outer circumferential side rotor
  • the abovementioned middle rotor 2 the abovementioned inner rotor 3
  • the inner circumferential side intake port 6 and inner circumferential side intake port 7 generically describe an inner circumferential side rotor.
  • the three rotor-type oil pump is configured from a first discharge passage 11 for feeding oil from the abovementioned outer circumferential side discharge port 5 to an engine E, a first return passage 12 that returns to an intake passage 8 of the outer circumferential side intake port 4, a second discharge passage 13 for feeding oil from the abovementioned inner circumferential side discharge port 7 to the engine E, and a second return passage 14 that returns to an intake passage 9 of the abovementioned inner circumferential side intake port 6, an end portion side of the abovementioned second discharge passage 13 being coupled with the abovementioned first discharge passage 11 at a suitable position therealong.
  • the abovementioned intake passage 8 and the abovementioned intake passage 9 may generically describe an intake body D (see FIG. 4 ).
  • the abovementioned first return passage 12 and the abovementioned second return passage 14 may generically describe a return passage E (see FIG. 4 ).
  • the symbol C denotes a pressure control valve configured from a valve main body 20 and a valve housing 30 which is provided across the abovementioned first discharge passage 11, first return passage 12, second discharge passage 13 and second return passage 14.
  • the abovementioned valve main body 20 is configured from a first valve portion 21, a narrow-diameter coupling portion 23 and a second valve portion 22.
  • a valve comprising the abovementioned first valve portion 21 and second valve portion 22 is referred to as a two-valve pressure control valve C.
  • a long-hole portion 31 slidable in accordance with need with respect to the abovementioned valve main body 20 is formed in the pressure control valve C, the abovementioned valve main body 20 being constantly push-pressured from a cover body 33 fixed in a rear portion side of the second valve portion 22 to the abovementioned first valve portion 21 side by the elastic pressure produced by a compression coil spring 40 within the long-hole portion 31.
  • the symbol 32 denotes a stopper portion formed in one end of the abovementioned long-hole portion 31 and positioned in a suitable position of the abovementioned first discharge passage 11.
  • the control of the pressure control valve C also requires that various conditions dependent on change in the discharge pressure of the abovementioned first discharge passage 11 be satisfied. More specifically, a flow rate control must be executed in each of a low revolution range which constitutes a state in which only the first discharge passage 11 and the second discharge passage 13 are opened as shown in FIG. 1 , an intermediate revolution range which constitutes a state in which the first discharge passage 11 and the second discharge passage 13 are opened and the abovementioned first return passage 12 is closed to open the second return passage 14 as shown in FIG. 2 and, in addition, in a high revolution range which constitutes a state in which the second discharge passage 13 is closed to open the first discharge passage 11 and the first return passage 12 and the second return passage 14 are open as shown in FIG. 3 .
  • each of the return passages of the outer circumferential side rotor A and the inner circumferential side rotor B are closed by the first valve portion 21 and the second valve portion 22 of the pressure control valve C, and all oil discharged from the first discharge passage 11 and the second discharge passage 13 is discharged to the engine.
  • the first discharge passage 11 of the outer circumferential side rotor A and the second discharge passage 13 of the inner circumferential side rotor B are in communication and, accordingly, an equalization of pressure occurs.
  • the overall discharge flow rate of the oil pump is equivalent to a sum of the flow rates of the outer circumferential side rotor A and the inner circumferential side rotor B.
  • the characteristics produced in the low revolution range are shown in a characteristics graph of number of revolutions and discharge pressure (see FIG. 8A ) and a characteristics graph of number of revolutions and discharge flow rate (see FIG. 8B ).
  • a state in which the engine revolutions have risen further is taken as the intermediate revolution range.
  • this state which constitutes the state of FIG. 2
  • an opening portion 141 of the second return passage 14 starts to open, and an opening portion 131 of the second discharge passage 13 starts to close.
  • the first discharge passage 11 of the outer circumferential side rotor A and the second discharge passage 13 of the inner circumferential side rotor B remain in communication. Because the opening portion 141 of the second return passage 14 of the inner circumferential side rotor B starts to open, first, the rise in pressure in the inner circumferential side rotor B stops.
  • the opening portion 131 of the second discharge passage 13 of the inner circumferential side rotor B gradually closes and the opening portion 141 of the second return passage 14 of the inner circumferential side rotor B gradually opens consequent to a rise in the number of revolutions in the intermediate revolution range, the overall increase in the flow rate when the number of revolutions rises is negligible.
  • the pressure not expressed in the true surface of the discharge of the inner circumferential side rotor B gradually drops due to the opening portion 141 of the second return passage 14 of the inner circumferential side rotor B being gradually opened.
  • the discharge flow rate of the outer circumferential side rotor A increases together with the number of revolutions.
  • the discharge flow rate of the inner circumferential side rotor B decreases together with the number of revolutions and the opening portion 141 of the second return passage 14 of the inner circumferential side rotor B being opened.
  • the backflow rate from the discharge of the outer circumferential side rotor A exceeds the discharge flow rate of the inner circumferential side rotor B subsequent to a certain number of revolutions being attained and, accordingly, the resultant discharge flow rate of the inner circumferential side rotor B is negative.
  • the generation of a negative flow rate in this way means that a flow rate equivalent to a sum of the flow rate of two oil pumps can be produced and that a flow rate equivalent to less than a flow rate of a single pump can be produced and, accordingly, that a broad variation in flow rate is possible.
  • the characteristics at the intermediate revolution range are expressed in the pressure characteristics graphs of revolutions with respect to discharge pressure and discharge flow rate (see FIG.
  • a state in which the engine revolutions have increased further is taken as the high revolution range.
  • this state which is the state shown in FIG. 3 , the opening portion 121 of the first return passage 12 starts to open and the opening portion 131 of the second discharge passage 13 has finished closing.
  • a more specific description thereof will be hereinafter provided. Because the discharge of the inner circumferential side rotor B is fully closed, the discharge of the outer circumferential side rotor A and the discharge of the inner circumferential side rotor B are no longer in communication. That is to say, the inner circumferential side rotor B is formed as an oil circuit independent of the outer circumferential side rotor A.
  • the pressure from the discharge of the outer circumferential side rotor A is unable to reach the inner circumferential side rotor B and instead is simply returned through the second return passage 14 of the inner circumferential side rotor B, and this results in an instant drop in the pressure of the inner circumferential side rotor B. Because backflow to the inner circumferential side rotor B also stops and all the oil discharged from the inner circumferential side rotor B is returned by way of the second return passage 14, a zero flow rate from the inner circumferential side rotor B to the engine E is established.
  • the friction (torque) can be caused to drop instantly and superfluous work can be eliminated due to the zero flow rate from the inner circumferential side rotor B and the discharge of the inner circumferential side rotor B performing no work at all, the overall efficiency of the pump is increased.
  • the characteristics at the high revolution range are expressed in the pressure characteristics graphs of revolutions with respect to discharge pressure and discharge flow rate (see FIG. 8 ) and, while there is a gradual increase in the outer circumferential side rotor A, the inner circumferential side rotor B is in a closed state and a pressure linking line obtained as a sum of the outer circumferential side rotor A and inner circumferential side rotor B is equivalent to the outer circumferential side rotor A alone. Because of the decrease in friction (torque) due to the drop in the pressure of the inner circumferential side rotor B in this way, the efficiency is increased.
  • the "pressure" of the pump main body (sum of the outer circumferential side rotor A and inner circumferential side rotor B) is equivalent to the pressure of the outer circumferential side rotor A alone. While the change in the pressure of the outer circumferential side rotor A is negligible due to the opening portion 121 of the first return passage 12 being open, strictly speaking, only very small increases in pressure occur consequent to increases in the number of revolutions.
  • the "flow rate" of the pump main body because the opening portion 131 of the second discharge passage 13 of the inner circumferential side rotor B is fully closed, the "flow rate" of the outer circumferential side rotor A constitutes the overall pump flow rate. While the change in the pressure of the outer circumferential side rotor A is negligible due to the opening portion 121 of the first return passage 12 being open, strictly speaking, only very small rises in pressure occur consequent to increases in the number of revolutions.
  • the pressure control valve C is configured from a valve body 20 and valve housing 30 and is provided between the abovementioned first discharge passage 11, the first return passage 12, the second discharge passage 13 and the second return passage 14.
  • the abovementioned valve main body 20 is configured from a first valve portion 21, a narrow-diameter coupling portion 23, a second valve portion 22, a third valve portion 24 and a narrow-diameter coupling portion 25.
  • the remaining configuration thereof is identical to the configuration shown in FIG. 1 to FIG. 3 .
  • a valve comprising the abovementioned first valve portion 21, second valve portion 22 and third valve portion 24 is referred to as a three valve-type pressure control valve C.
  • a state in which the engine revolutions have risen further is taken as the intermediate revolution range.
  • the opening portion 141 of the second return passage 14 starts to open and the opening portion 131 of the second discharge passage 13 starts to close. A description thereof has been omitted.
  • a state resulting from further increase in the engine revolutions is taken as the high revolution range.
  • the opening portion 121 of the first return passage 12 starts to open and the opening portion 131 of the second discharge passage 13 has finished closing. Because the discharge of the inner circumferential side rotor B is fully closed, the discharge of the outer circumferential side rotor A and the discharge of the inner circumferential side rotor B are no longer in communication.
  • the inner circumferential side rotor B forms an oil circuit independent of the outer circumferential side rotor A.
  • the pressure from the discharge of the outer circumferential side rotor A is unable to reach the inner circumferential side rotor B and is instead simply returned through the second return passage 14 of the inner circumferential side rotor B resulting in an instant drop in the pressure of the outer circumferential side rotor B.
  • backflow to the inner circumferential side rotor B also stops and all the oil discharged from the inner circumferential side rotor B is returned by way of the second return passage 14, a zero flow rate from the inner circumferential side rotor B to the engine E is established.
  • the friction (torque) can be caused to drop instantly and superfluous work eliminated due to the zero flow rate of the inner circumferential side rotor B and the discharge of the inner circumferential side rotor B performing no work at all, the overall efficiency of the pump is increased.
  • the characteristics at this high revolution range are expressed in the pressure characteristics graphs of revolutions with respect to discharge pressure and discharge flow rate (see FIG. 8 ) and, while the increase in the outer circumferential side rotor A is gradual, the inner circumferential side rotor B is in a closed state and a pressure linking line obtained as a sum of the outer circumferential side rotor A and inner circumferential side rotor B is equivalent to the outer circumferential side rotor A alone. Because of the decrease in friction (torque) due to the drop in the pressure of the inner circumferential side rotor B in this way, the efficiency is increased.
  • the "pressure" of the pump main body (sum of the outer circumferential side rotor A and outer circumferential side rotor B) is equivalent to the pressure of the outer circumferential side rotor A alone. While the change in the pressure of the outer circumferential side rotor A is negligible due to the opening portion 121 of the first return passage 12 being open, strictly speaking, only very small rises in pressure occur consequent to increases in the number of revolutions.
  • the "flow rate" of the pump main body because the opening portion 131 of the second discharge passage 13 of the outer circumferential side rotor B is fully closed, the "flow rate" of the outer circumferential side rotor A constitutes the overall pump flow rate. While the change in the pressure of the outer circumferential side rotor A is negligible due to the opening portion 121 of the first return passage 12 being open, strictly speaking, only very small rises in pressure occurs consequent to increases in the number of revolutions.
  • While the invention of the subject application constitutes an oil pump pressure control device as described above, it may also constitute a variable flow rate oil pump.
  • This is an oil pump comprising two discharge ports that uses three rotors as means for as means for provision of the two discharge sources thereof.
  • a discharge port 130 or the second discharge passage 13 of the outer circumferential side rotor B are closed, the outer circumferential side rotor A and the outer circumferential side rotor B are disengaged.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Fluid Mechanics (AREA)
  • Rotary Pumps (AREA)
  • Details And Applications Of Rotary Liquid Pumps (AREA)
EP07122748A 2007-02-20 2007-12-10 Dispositif de contrôle de la pression dans une pompe à huile Not-in-force EP1961961B1 (fr)

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
JP2007039135A JP4521005B2 (ja) 2007-02-20 2007-02-20 オイルポンプにおける圧力制御装置

Publications (3)

Publication Number Publication Date
EP1961961A2 true EP1961961A2 (fr) 2008-08-27
EP1961961A3 EP1961961A3 (fr) 2009-12-16
EP1961961B1 EP1961961B1 (fr) 2011-01-26

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ID=39446211

Family Applications (1)

Application Number Title Priority Date Filing Date
EP07122748A Not-in-force EP1961961B1 (fr) 2007-02-20 2007-12-10 Dispositif de contrôle de la pression dans une pompe à huile

Country Status (6)

Country Link
US (1) US20080253904A1 (fr)
EP (1) EP1961961B1 (fr)
JP (1) JP4521005B2 (fr)
CN (1) CN101251108B (fr)
DE (1) DE602007012206D1 (fr)
ES (1) ES2358286T3 (fr)

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP2600004A3 (fr) * 2011-12-02 2016-04-27 Myunghwa Ind. Co., Ltd. Pompe à huile variable

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CN101639063B (zh) * 2009-08-21 2011-01-26 深圳益宝实业有限公司 齿轮式油泵
JP5278775B2 (ja) * 2010-12-06 2013-09-04 アイシン精機株式会社 油供給装置
JP5950583B2 (ja) * 2011-03-27 2016-07-13 株式会社山田製作所 ポンプ装置
JP5938901B2 (ja) * 2011-12-28 2016-06-22 株式会社ジェイテクト モータ制御装置および電動ポンプユニット
CN102900665A (zh) * 2012-10-16 2013-01-30 李庆中 一种多层结构的内啮合齿轮泵或齿轮马达装置
JP6083708B2 (ja) * 2013-09-20 2017-02-22 アイシン精機株式会社 電動オイルポンプ
DE102014222396A1 (de) 2014-11-03 2016-05-04 Continental Automotive Gmbh Verdrängerpumpe
US10392977B2 (en) * 2016-02-11 2019-08-27 Slw Automotive Inc. Automotive lubricant pumping system with two piece relief valve
CN106640576B (zh) * 2016-12-09 2019-01-15 柳州职业技术学院 一种卸载式自变量径向柱塞泵
CN108223357B (zh) * 2017-11-24 2019-11-08 河南航天液压气动技术有限公司 一种内泄式齿轮泵
CN109827059A (zh) * 2019-01-23 2019-05-31 浙江零跑科技有限公司 一种双向供油的摆线齿轮泵

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EP0811765A2 (fr) 1996-06-05 1997-12-10 Bayerische Motoren Werke Aktiengesellschaft, Patentabteilung AJ-3 Moteur à combustion interne avec pompe à engrenages internes à deux étages
JP2002070756A (ja) 2000-08-28 2002-03-08 Toyota Motor Corp 可変容量型オイルポンプ

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Patent Citations (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0811765A2 (fr) 1996-06-05 1997-12-10 Bayerische Motoren Werke Aktiengesellschaft, Patentabteilung AJ-3 Moteur à combustion interne avec pompe à engrenages internes à deux étages
JP2002070756A (ja) 2000-08-28 2002-03-08 Toyota Motor Corp 可変容量型オイルポンプ

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP2600004A3 (fr) * 2011-12-02 2016-04-27 Myunghwa Ind. Co., Ltd. Pompe à huile variable

Also Published As

Publication number Publication date
JP4521005B2 (ja) 2010-08-11
JP2008202488A (ja) 2008-09-04
CN101251108A (zh) 2008-08-27
EP1961961B1 (fr) 2011-01-26
ES2358286T3 (es) 2011-05-09
CN101251108B (zh) 2011-09-14
EP1961961A3 (fr) 2009-12-16
DE602007012206D1 (de) 2011-03-10
US20080253904A1 (en) 2008-10-16

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