EP3676498B1 - Regelbare kühlmittelpumpe für haupt- und nebenförderkreislauf - Google Patents

Regelbare kühlmittelpumpe für haupt- und nebenförderkreislauf Download PDF

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Publication number
EP3676498B1
EP3676498B1 EP18740810.9A EP18740810A EP3676498B1 EP 3676498 B1 EP3676498 B1 EP 3676498B1 EP 18740810 A EP18740810 A EP 18740810A EP 3676498 B1 EP3676498 B1 EP 3676498B1
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EP
European Patent Office
Prior art keywords
pump
valve
pressure
circuit
hydraulic control
Prior art date
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Active
Application number
EP18740810.9A
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German (de)
English (en)
French (fr)
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EP3676498A1 (de
Inventor
Franz Pawellek
Toni Steiner
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.)
Nidec GPM GmbH
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Nidec GPM GmbH
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.)
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Publication date
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Publication of EP3676498A1 publication Critical patent/EP3676498A1/de
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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D15/00—Control, e.g. regulation, of pumps, pumping installations or systems
    • F04D15/0027—Varying behaviour or the very pump
    • F04D15/0038—Varying behaviour or the very pump by varying the effective cross-sectional area of flow through the rotor
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D15/00—Control, e.g. regulation, of pumps, pumping installations or systems
    • F04D15/0005—Control, e.g. regulation, of pumps, pumping installations or systems by using valves
    • F04D15/0022—Control, e.g. regulation, of pumps, pumping installations or systems by using valves throttling valves or valves varying the pump inlet opening or the outlet opening
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01P—COOLING OF MACHINES OR ENGINES IN GENERAL; COOLING OF INTERNAL-COMBUSTION ENGINES
    • F01P7/00—Controlling of coolant flow
    • F01P7/14—Controlling of coolant flow the coolant being liquid
    • F01P2007/146—Controlling of coolant flow the coolant being liquid using valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01P—COOLING OF MACHINES OR ENGINES IN GENERAL; COOLING OF INTERNAL-COMBUSTION ENGINES
    • F01P5/00—Pumping cooling-air or liquid coolants
    • F01P5/10—Pumping liquid coolant; Arrangements of coolant pumps
    • F01P5/12—Pump-driving arrangements
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
    • F05D—INDEXING SCHEME FOR ASPECTS RELATING TO NON-POSITIVE-DISPLACEMENT MACHINES OR ENGINES, GAS-TURBINES OR JET-PROPULSION PLANTS
    • F05D2250/00—Geometry
    • F05D2250/50—Inlet or outlet
    • F05D2250/52—Outlet
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
    • F05D—INDEXING SCHEME FOR ASPECTS RELATING TO NON-POSITIVE-DISPLACEMENT MACHINES OR ENGINES, GAS-TURBINES OR JET-PROPULSION PLANTS
    • F05D2270/00—Control
    • F05D2270/50—Control logic embodiments
    • F05D2270/54—Control logic embodiments by electronic means, e.g. electronic tubes, transistors or IC's within an electronic circuit
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
    • F05D—INDEXING SCHEME FOR ASPECTS RELATING TO NON-POSITIVE-DISPLACEMENT MACHINES OR ENGINES, GAS-TURBINES OR JET-PROPULSION PLANTS
    • F05D2270/00—Control
    • F05D2270/60—Control system actuates means
    • F05D2270/64—Hydraulic actuators

Definitions

  • the present invention relates to a mechanically driven coolant pump with a controllable delivery rate for a main delivery circuit from a first outlet and for a secondary delivery circuit from a second outlet of the coolant pump.
  • auxiliary devices such as exhaust gas recirculation, a turbocharger, charge air cooling or the like, as well as so-called split cooling, i.e. separate cooling of the engine block and cylinder heads of the internal combustion engine, are used in vehicles.
  • split cooling i.e. separate cooling of the engine block and cylinder heads of the internal combustion engine.
  • both systems are known in the prior art that include one or more additional water pumps to enable independent promotion of individual circulations, as well as systems with water valves, which enable a needs-based distribution of a coolant flow delivered by a pump to different branches.
  • the provision of additional water pumps and water valves with actuators for valve adjustment in a branched pipeline network again requires the corresponding installation and susceptibility of cabling to interference for a power supply and control signal transmission between decentralized actuators or pump motors, a central control device and a battery.
  • a failure of a drive or a defect in a cable can have effects on other areas of the coolant circulation that do not comply with a uniform fail-safe mode to prevent consequential damage.
  • control slide In a closed state, the control slide covers the pump impeller to form a spiral housing, whereby the pump outlet is shut down. In this case, an opening to a bypass is released in a rear wall of the pump chamber behind the pump impeller, which allows coolant to be discharged from the pump chamber separately from the pump outlet.
  • an open position of the control slide in which a flow through the pump outlet is completely released, the opening of the bypass to the pump chamber is closed by a part of the control slide.
  • the disclosed coolant pump thus provides a function for switching a large delivery rate through the pump outlet or a small delivery rate through the bypass.
  • intermediate states of a dividing ratio of the delivery flow occur, the course of which cannot be controlled separately in the desired manner, but is established as a function of a pressure difference between the individual volume flows, which in turn results from a fixed flow geometry of the pump.
  • WO 2013/034126 A1 discloses a controllable coolant pump according to the preamble of claim 1.
  • one object of the present invention is to create a compact actuator for a coolant system with two delivery circuits.
  • Another aspect of the invention is also to create a constructive relationship for a fail-safe mode implemented jointly in the conveying circuits.
  • the controllable mechanical coolant pump with a first outlet for a main delivery circuit and a second outlet for a secondary delivery circuit has, inter alia, a hydraulic control circuit derived from the coolant with an inlet-side auxiliary pump, an outlet-side proportional valve and a control slide as a hydraulic actuator for limiting the flow of the main delivery circuit, and is characterized in particular in that a control valve is connected to the hydraulic control circuit as a hydraulic actuator for limiting the flow of the secondary delivery circuit, with actuations of the control slide and the control valve being assigned to respective pressure ranges in the hydraulic control circuit.
  • the invention thus provides for the first time a coolant pump with two hydraulic actuators, in particular for regulating two different pump outlets or delivery circuits.
  • the invention provides for the first time to connect two hydraulic actuators to a hydraulic control circuit, which is derived in particular from the coolant, i.e. to operate them with the same control pressure.
  • the actuators By designing the actuation of the actuators for different pressure ranges, they respond to a control of an assigned pressure in the hydraulic control circuit, at least in some areas, independently of one another, so that different valve positions can be made in the two conveying circuits.
  • two new basic states can be implemented in comparison to the state of the art of an ECF pump with bypass, i.e. the main conveyor circuit and secondary conveyor circuit are completely closed, or the main conveyor circuit and secondary conveyor circuit are fully open, as well as two adjustment ranges, in which For example, the main conveyor circuit remains closed and a flow rate of the secondary conveyor circuit can be set.
  • control valve can be connected to the hydraulic control circuit as a branched hydraulic actuator between the auxiliary pump and the proportional valve, and can be closed by means of the pressure in the hydraulic control circuit against an elastic bias.
  • This configuration of the hydraulic control means that the hydraulic actuators or the control slide and the control valve have the same control pressure at.
  • a fail-safe mode is achieved for the secondary conveyor circuit, as will be described later.
  • control valve can be designed as a seat valve valve which is acted upon by a spring in the opening direction.
  • the spring-loaded seat valve ensures smooth adjustment of the valve body in relation to the actuating force of the spring, even when the load is absorbed by the delivery pressure.
  • a piston area for receiving a hydraulic actuating force of the control valve in the hydraulic control circuit can be smaller than a piston area of the control slide in the hydraulic control circuit.
  • the piston area of the control valve to the piston area of the control slide can have an area ratio of approximately 1: 3.
  • This hydraulically effective area ratio between the two actuators in conjunction with the restoring forces of the respective spring preload, achieves a preferred spread of the two associated control pressure ranges, which is reflected in a defined response behavior between the two actuators.
  • control valve can be arranged in the second outlet on the pump housing.
  • a pressure valve can be provided between the main delivery flow and the secondary delivery flow which opens above a predetermined pressure difference between a higher pressure in the main delivery flow and a lower pressure in the secondary delivery flow.
  • the pressure valve thus counteracts the drying up of the small secondary conveying circuit during the transient pressure difference described, since part of the main conveying circuit flows into the secondary conveying circuit.
  • the pressure valve can be designed as a check valve which is acted upon by a spring in the closing direction.
  • a spring-loaded check valve forms the preferred means of providing a pressure valve which gradually opens for a flow from the main conveyor circuit to the secondary conveyor circuit as the pressure difference increases.
  • the pressure valve can open downstream from the control slide into the main delivery circuit and upstream from the control valve into the secondary delivery circuit.
  • This arrangement of the pressure valve achieves a preferred response behavior in the manner of operation described and enables a highly integrated, compact pump structure.
  • Fig. 1 shows a longitudinal section through the pump without complete outer contours of a pump housing 1.
  • a pump shaft 3 extends from a belt pulley 4, through a shaft bearing into a pump chamber 10 of the pump housing 1 and drives a pump impeller 2.
  • the pump impeller 2 and the pump chamber 10, which is not shown in full, are designed in the form of a radial pump assembly, in which a pump inlet 13 (not shown) flows axially towards the pump impeller 2, and a first pump outlet 11 for a main delivery circuit connected to the internal combustion engine a radially outer volute casing section leads out tangentially from the pump chamber 10.
  • the pump assembly of the coolant pump has a hydraulically adjustable control slide 8, which is known from a so-called ECF pump type.
  • a flow-effective radial area around the pump impeller 2 can be variably covered by the control slide 8 with a cylindrical section formed coaxially to the pump shaft 3 along an adjustment path running parallel to the pump shaft 3.
  • the control slide 8 is in a closed position, in which the flow area of the pump impeller 2 is completely covered and thus no delivery flow to the first pump outlet 11 is brought about.
  • an axial piston pump 6 (shown schematically) is also arranged within the radius of the pump impeller 2 and parallel to the pump shaft 3, the piston of which is actuated via a sliding shoe (not shown) which is mounted on a swash plate (not rotatably arranged with the pump shaft 3) shown) slides.
  • the axial piston pump 6 serves as an auxiliary pump of a coolant-operated hydraulic control circuit 5 (shown schematically) in which a control pressure independent of the flow rate is generated and set to control the control slide 8 and a control valve 9 described later.
  • the axial piston pump 6 sucks in coolant from the flow area between the pump impeller 2 and the control slide 9 and discharges the pressurized coolant into the hydraulic control circuit 5, which is formed in the pump housing 1.
  • the hydraulic control circuit 5 comprises an electromagnetically actuated proportional valve 7 (shown schematically), which limits a return of the coolant into the delivered coolant flow and thus sets a pressure of the hydraulic control circuit 5 in a path between the axial piston pump 6 and the proportional valve 7.
  • a hydraulic branch feeds the pressure of the hydraulic control circuit 5 to an annular piston 18, which is arranged coaxially with the pump shaft 3 and takes over the function of a hydraulic actuator along the displacement path of the control slide 8.
  • a return spring acts on the annular piston 18 in the opposite direction to the pressure of the hydraulic control circuit 5, ie away from the pump impeller 2.
  • the annular piston 18 is connected to the control slide 8 and moves it with increasing pressure of the hydraulic control circuit 5 in the direction of the pump impeller 2, whereby the cylindrical section of the control slide 6 is increasingly brought into axial overlap with the pump impeller 2.
  • the electromagnetic proportional valve 7 is opened without the supply of a control current, so that the coolant sucked in by the axial piston pump 6 flows back essentially without pressure via the hydraulic control circuit 5 through the proportional valve 7 and back into the pumped coolant.
  • the electromagnetic proportional valve 7 is closed temporarily or intermittently by supplying a control current controlled with pulse width modulation, the pressure generated by the axial piston pump 6 spreads via the hydraulic control circuit 5 to the annular piston 18. If the proportional valve 7 remains open by switching off the control current, no more pressure builds up in the hydraulic control circuit 5 and the annular piston 18 returns to the unactuated basic position under the action of the return spring.
  • a maximum delivery flow is brought about as a function of the pump speed without shielding a flow-effective region of the pump impeller 2 in the main delivery circuit.
  • This state also represents a fail-safe mode, since in the event of a power supply failure, ie a currentless electromagnetic proportional valve 7, a maximum volume flow and the greatest possible heat dissipation from the internal combustion engine through the main conveyor circuit are automatically ensured.
  • the pump housing 1 comprises a second pump outlet 12 for a secondary delivery circuit, to which a cooling system for an exhaust gas recirculation valve (EGR valve) is connected in the present exemplary embodiment.
  • the second pump outlet 12 opens into the pump chamber 10 on a rear side of the pump impeller 2.
  • the opening of the second pump outlet 12 is accessible through openings on the front of the control slide 8, regardless of its position, so that part of the flow from the pump chamber 10 always enters the second pump outlet 12 penetrates.
  • control valve 9 In the second pump outlet 12, the control valve 9 is arranged, which blocks, limits or opens a flow of the secondary delivery circuit.
  • the control valve 9 is also connected to the hydraulic control circuit 5 by a hydraulic branch.
  • a valve body of the control valve 9 is displaced by the pressure in the hydraulic control circuit 5 approximately perpendicular to the flow direction against the restoring force of a spring and gradually closes the flow in the second pump outlet 12.
  • the valve body of the control valve 9 is pushed back by the spring and the flow of the second pump outlet 12 is released.
  • the pressure in the hydraulic control circuit 5 is controlled by switching on and off times for opening and closing the proportional valve 7.
  • the pressure is controlled in such a way that a balance is achieved between the hydraulic pressure and a restoring force of the pretensioned spring in the regulating valve 9 and a position of the valve body in the regulating valve 9 is maintained.
  • the positions of the valve body of the control valve 9, like a position of the annular piston 18 of the control valve 8, can be detected by a displacement sensor (not shown) and used to control the proportional valve 7.
  • the main conveying circuit and the secondary conveying circuit are throttled in relation to a predetermined machine speed on the basis of a control current for opening and closing the electromagnetically actuated proportional valve 7.
  • the hydraulic configuration was selected in such a way that the control valve 9 for the secondary delivery circuit has a higher hydraulic value Pressure to close is required as the control slide 8 for the main conveyor circuit.
  • the assignment of the pressure ranges in which the hydraulic actuators respond is set on the basis of a hydraulically active piston area, which each actuator has for taking up pressure from the hydraulic control circuit 5, and the selected characteristic curve of the return springs.
  • the response behavior of the two hydraulic actuators is preferably selected such that an adjustment range of the control valve 9 can be controlled by a pressure that begins above a pressure at which the control slide 8 closes completely.
  • a suitable separation between the pressure for closing one hydraulic actuator and the lower pressure at the beginning of the adjustment range of the other actuator is set by a hydraulically effective area ratio.
  • the area ratio between the actuator closing at higher pressure and the actuator closing at lower pressure is 1: 3.
  • the in Fig. 1 The operating state of the controllable coolant pump shown is intended for a cold start situation of a vehicle in which the internal combustion engine or other devices are not yet required to cool.
  • the proportional valve 7 is controlled by a control unit (not shown) by a sampling ratio of a pulse width modulation with a high proportion of switch-on times in order to set a high pressure in the hydraulic control circuit 5.
  • the proportional valve 7 severely limits a return flow of the coolant behind the axial piston pump 6 and a backflow in front of the proportional valve 7 increases the pressure in the hydraulic control circuit 5 to the branched actuators until the control slide 8 and then the control valve 9 close.
  • both flows of the main conveying circuit and the secondary conveying circuit are consequently maximally limited or closed.
  • the in Fig. 2 The operating state of the controllable coolant pump shown is intended, for example, for a warm-up situation of a vehicle in which the internal combustion engine is not yet at operating temperature, but so-called hotspots have already formed on devices such as exhaust gas recirculation, so that there is already a need for cooling to protect components such as an EGR valve is present.
  • the proportional valve 7 is controlled by a sampling ratio of a pulse width modulation with a lower proportion of switch-on times in order to lower the pressure in the hydraulic control circuit 5.
  • a return flow from the hydraulic control circuit 5 through the proportional valve 7 increases and the pressure on the actuators decreases.
  • the control valve 9 initially moves back into the open position via gradual limiting positions, while the control slide 8 remains closed. If a pressure in the hydraulic control circuit 5 is maintained after this process, the flow of the main conveying circuit consequently remains closed and the flow in the secondary conveying circuit 5 remains open.
  • a gradual limitation of the secondary conveyor circuit can be set when the main conveyor circuit is closed.
  • the pressure valve 15 remains closed, since it is still exposed to a pressure of the secondary delivery circuit in the closing direction while the other side is not exposed to any delivery pressure.
  • the in Fig. 3 The operating state of the controllable coolant pump shown is intended for a load situation of a vehicle in which there is a need for cooling both for the internal combustion engine and for one or more other devices that are connected to the secondary delivery circuit.
  • Fig. 3 If the proportional valve 7 is not activated or is activated by a sampling ratio of a pulse width modulation with a small proportion of switch-on times, so that no pressure is generated in the hydraulic control circuit 5.
  • the control slide 8 then moves back into the open position via gradual limiting positions, while the control valve 9, which is already open, remains open.
  • both the flow of the main conveying circuit and the flow of the secondary conveying circuit 5 consequently remain open to the maximum.
  • a gradual limitation of the main conveying circuit can be set when the secondary conveying circuit is open.
  • the pressure valve 15 is opened by a pressure difference while the control slide 8 is opening or while the main conveying circuit is open to the maximum.
  • the pressure difference arises from a low pressure loss of the part of the conveying flow that flows into the main conveying circuit and a high pressure loss of the part of the conveying flow that flows into the secondary conveying circuit.
  • the pressure valve 15 opens and enables a flow of the large delivery rate in the main delivery circuit to compensate the insufficient delivery rate in the secondary delivery circuit. The flow behavior is thus improved during a transient state of the division or a relatively large division ratio between the delivery rates.
EP18740810.9A 2017-09-01 2018-07-12 Regelbare kühlmittelpumpe für haupt- und nebenförderkreislauf Active EP3676498B1 (de)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE102017120191.2A DE102017120191B3 (de) 2017-09-01 2017-09-01 Regelbare Kühlmittelpumpe für Haupt- und Nebenförderkreislauf
PCT/EP2018/068958 WO2019042644A1 (de) 2017-09-01 2018-07-12 Regelbare kühlmittelpumpe für haupt- und nebenförderkreislauf

Publications (2)

Publication Number Publication Date
EP3676498A1 EP3676498A1 (de) 2020-07-08
EP3676498B1 true EP3676498B1 (de) 2021-06-09

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Application Number Title Priority Date Filing Date
EP18740810.9A Active EP3676498B1 (de) 2017-09-01 2018-07-12 Regelbare kühlmittelpumpe für haupt- und nebenförderkreislauf

Country Status (6)

Country Link
US (1) US11002281B2 (zh)
EP (1) EP3676498B1 (zh)
CN (1) CN111051702B (zh)
BR (1) BR112019028100A2 (zh)
DE (1) DE102017120191B3 (zh)
WO (1) WO2019042644A1 (zh)

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Publication number Publication date
CN111051702B (zh) 2021-09-14
WO2019042644A1 (de) 2019-03-07
US11002281B2 (en) 2021-05-11
EP3676498A1 (de) 2020-07-08
BR112019028100A2 (pt) 2020-07-28
CN111051702A (zh) 2020-04-21
DE102017120191B3 (de) 2018-12-06
US20200340482A1 (en) 2020-10-29

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