US3727595A - Control device for hydraulically operated tappet valves of internal combustion engines - Google Patents

Control device for hydraulically operated tappet valves of internal combustion engines Download PDF

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US3727595A
US3727595A US00068362A US3727595DA US3727595A US 3727595 A US3727595 A US 3727595A US 00068362 A US00068362 A US 00068362A US 3727595D A US3727595D A US 3727595DA US 3727595 A US3727595 A US 3727595A
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Prior art keywords
valve
bore
hydraulically operated
actuating piston
sphere
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US00068362A
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English (en)
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H Links
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Robert Bosch GmbH
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Robert Bosch GmbH
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Priority claimed from DE19691944177 external-priority patent/DE1944177A1/de
Priority claimed from DE19691962916 external-priority patent/DE1962916A1/de
Priority claimed from DE19702006304 external-priority patent/DE2006304A1/de
Priority claimed from DE19702006844 external-priority patent/DE2006844A1/de
Priority claimed from DE2008668A external-priority patent/DE2008668C3/de
Priority claimed from DE19702010291 external-priority patent/DE2010291A1/de
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L9/00—Valve-gear or valve arrangements actuated non-mechanically
    • F01L9/10—Valve-gear or valve arrangements actuated non-mechanically by fluid means, e.g. hydraulic
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B1/00—Engines characterised by fuel-air mixture compression
    • F02B1/02—Engines characterised by fuel-air mixture compression with positive ignition
    • F02B1/04—Engines characterised by fuel-air mixture compression with positive ignition with fuel-air mixture admission into cylinder
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00—Other engines
    • F02B75/12—Other methods of operation
    • F02B2075/125—Direct injection in the combustion chamber for spark ignition engines, i.e. not in pre-combustion chamber
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00—Road transport of goods or passengers
    • Y02T10/10—Internal combustion engine [ICE] based vehicles
    • Y02T10/12—Improving ICE efficiencies

Definitions

  • SHEET 10 [1F 10 PATENTEB APR 1 H973 CONTROL DEVICE FOR HYDRAULICALLY OPERATED TAPPET VALVES OF INTERNAL COMBUSTION ENGINES BACKGROUND OF THE INVENTION valve in a closed position.
  • the elasticity of the fluid in the system is a determining factor for the limits of possible applications. Due to the relatively long fluid conduits, a large fluid volume is present which, dependent upon the rpm and thus the throttle effect, causes a shift in the motion pattern between the pump piston and the engine valve. Such an occurrence results in an interference with the predetermined closing and opening moments of the valve. Furthermore, temperature variations also have an adverse effect, since they cause changes in the volume of the control liquid. Also, due to the elasticity of the fluid conduits, pressure oscillations may occur which may cause natural resonances of the valves thus resulting in an interruption of the connection between pump piston and cam or valve stem and actuating piston.
  • the tappet valve stem is in engagement with an actuating piston which is intermittently displaced by the liquid delivered under pressure to the control device by a delivery pump in a continuous manner.
  • a periodically energized solenoid valve The latter includes a movable valve member which, dependent upon the energized or the de-energized condition of the valve solenoid, may assume two positions. In one position it admits the pressurized liquid to said actuating piston, while in the other position it causes withdrawal of said liquid therefrom.
  • the actuating piston When communication between the actuating piston and the continuously delivered pressurized liquid exists, the actuating piston, urged by the pressurized liquid, executes its working stroke whereby the tappet valve is opened.
  • FIG. 1 is an axial sectional view of an embodiment of the invention including a schematic representation of g the associated liquid circuit;
  • FIG. 2 is an axial sectional view of a further embodiment of the invention.
  • FIG. 3 is a valve lift diagram pertaining to the operation of the embodiments according to FIGS. 1 or 2;
  • FIG. 4 is a sectional view of another embodiment of the invention including a schematic representation of an associated liquid circuit
  • FIG. 5 is a fragmentary sectional view of the same embodiment showing some components in an alternate position
  • FIG. 6 is a sectional view of still another embodiment of the invention including a schematic representation of an associated liquid circuit
  • FIG. 7 is a fragmentary sectional view of the same embodiment showing some components in an alternate position
  • FIG. 8 is a sectional view of a further embodiment of the invention including a schematic representation of an associated liquid circuit
  • FIG. 9 is a fragmentary sectional view of the same embodiment showing some components in an alternate position
  • FIG. 10 is a sectional view of a further embodiment of the invention including a schematic representation of an associated, electronically controlled liquid cir- 'cuit;
  • FIG. 10a is a block diagram of an electronic control apparatus associated with the embodiment shown in FIG. 10;
  • FIG. 11 is a diagram of valve lift curves illustrating the operation of the embodiment according to FIG. 10;
  • FIG. 11 is a sectional view of another embodiment of the invention including schematic representations of associated fluid circuits; 7 l
  • FIG. 13 is a diagram of valve lift curves illustrating the operation of the embodiment according to FIG. 12;
  • FIG. 14 is a sectional view of still another embodiment of the invention including a diagrammatic representation of associated liquid circuits;
  • FIG. 15 is a diagram of valve lift curves illustrating the operation of the embodiment according to FIG. 14.
  • FIG. 1 illustrates the invention in its simplest form.
  • a partially shown cylinder head 1 of an internal combustion engine there operates a tappet valve 2 having a valve stem 3 axially slidably guided by a bearing sleeve 4 secured in the wall of the cylinder head 1.
  • the outer end of the valve stem 3 carries a spring seat disc 5.
  • a valve closing spring 6 Between the upper face of the cylinder head 1 and the spring seat disc 5 there is disposed in a preloaded condition a valve closing spring 6.
  • the lower terminal face of the actuating piston 9 is urged into contact with the valve stem 3.
  • the actuating piston 9 projects 'into a chamber 11 which is provided in the bushing 10 and which receives a spring 12 urging the actuating piston 9 in the closing direction of the tappet valve 2.
  • a delivery pump 14 draws liquid from a tank 13 and 1 delivers it under a pressure of, for example 100 kg/cm through a pressure conduit 15 towards the solenoid valve 8. From the conduit 15 there extends, downstream of the pump 14, a return conduit 16 in which there is disposed a pressure control valve 17 and which terminates in the tank 13.
  • the liquid first flows into a pressure chamber 19 which is connected with a control chamber 21 by means of a bore 20 serving as a supply channel.
  • the mouth of the bore 20 at the control chamber 21 serves as a valve seat for a sphere 22 which is movably disposed in the control chamber 21.
  • From the control chamber 21 there extends a bore 23 to the chamber 11.
  • From the control chamber 21 there also extends a bore 24 which is in continuous communication with the tank 13 through a discharge channel 25 and a return conduit 26.
  • the 'mouth of the bore 24 at the control chamber 21 serves as a valve seat for the valve sphere 22.
  • An armature 28, having a pin-like extension 28' in contact with the sphere 22, is slidably disposed in the bore 24 and urges, under the action of a spring 27, the sphere 22 into a position in which it closes the bore 20.
  • the control chamber 21, the bore 24 and the armature 28 are contained in a valve support sleeve 29 which is inserted in the housing bracket 7 and which is held in position by a flange of the housing 30 of the electromagnet forming part of the solenoid 8.
  • the electromagnet chamber accommodating the spring 27 and the bore 20 immediately upstream of the solenoid valve, are interconnected by means of a channel 31 so that in both I aforenoted spaces identical pressures prevail.
  • the diameter of the cylindrical surface of armature 28 sliding in the bore 24 in a fluid-tight manner, - is identical to the diameter of both seats for the sphere 22.
  • the force derived from the pressure prevailing in bore 20 and exerted on the sphere in the opening direction is identical to the force of the pressurized liquid exerted on the sphere by the armature 28 in the closing direction.
  • Spring 27 aids the latter force so that the sphere 22 is, as a net result of the opposing forces, urged against the mouth or valve seat provided about the bore 20.
  • the force of the spring 27 is overcome by the magnetic force and the armature 28 is displaced towards the left.
  • the pressurized liquid thus may flow through bore 20, pressing the sphere 22 against its other, opposite seat formed about the opening of bore 24.
  • the spring 27 returns the armature 28 and the sphere 22 into their initial position in which the bore 20 is again closed.
  • the liquid may flow from the chamber 11 through the bore 24, the channel 25 and the return conduit 26 to the liquid tank 13 allowing the actuating piston 9 to return. This, in turn, causes the tappet valve 2 to assume its closed position.
  • the admission of liquid to the actuating piston 9 is effected by a servooperated control means.
  • This embodiment finds application particularly in large internal combustion engines, where greater forces are needed to open the engine valves.
  • the hydraulic servo circuit is controlled by a solenoid valve of the type and in the manner as described in connection with the embodiment depicted in FIG. 1.
  • the liquid admitted through the conduit 15 flows through a channel 37 into an annular groove 38 which is formed in the wall of a bore 39. From the annular groove 38 there extend the pressure chamber 19 and the bore 20. In the bore 39 there operates a control plunger 40 which is in engagement with an axially aligned piston 50.
  • the pressurized liquid flows from bore 20 to control chamber 21, bore 23 and then through a channel 41 to a lower radial face of piston 50.
  • the piston 50 and thus the control plunger 40 are displaced against the force of a spring 42 disposed in bore 39.
  • a circumferential annular groove 43 which is in continuous communication with the annular groove 38.
  • the portion of the bore 39 accommodating the spring 42 is'connected through a channel 44 with the return conduit 26 terminating in the liquid tank 13.
  • annular groove 45 which is in communication through a channel 46 with one side of a bore 47 in which there is reciprocably arranged an actuating piston 48.
  • radial bores 49 which, in the position of the control plunger 40 as shown in FIG. 2, connect the bore 47 with the discharge channel 44.
  • the pressurized liquid displaces the piston 50 and the control plunger 40 against the force of the return spring 42 whereby the annular groove 43 shifts into alignment with the annular groove 45 while, at the same time, the bores 49 hydraulically separate from the annular groove 45.
  • the liquid present under pressure in the annular groove 38 may flow through the annular grooves 43 and 45 to the radial face of the actuating piston 48 and thereby effect an opening of the tappet valve.
  • the control plunger 40 is returned by the spring 42 into its initial position in which the annular grooves 45 and 43 are separated from one another, whereas the radial bores 49 are re-connected with the annulargroove 45.
  • the pressurized liquid present above the actuating piston 48 may flow from the bore 47 through the annular groove 45 to the liquid tank 13.
  • the tappet valve closes.
  • the work pistons 9 and 48 affect directly the stem 3 of the tappet valve 2. It is to be understood that between piston and valve stem there may be provided any other force transmitting means, such as a rocker arm, or the like.
  • FIG. 3 particularly well illustrates the advantages of the afore-described embodiments.
  • the stroke s of the tappet valve is shown as a function. of the angular position of the cam shaft.
  • the stroke s designates the structurally possible maximum stroke of the tappet valve.
  • the valve lift curve I pertains to a tappet valve operated conventionally by the cam of a cam shaft.
  • the relatively flat initial portion of the ascent and relatively flat terminal portion of the descent of the curve is caused by the fact that the flanks of a cam, for starting and terminating the force transmission to the push rod, must not be steep.
  • a transition into a steeper range may occur only when the stroke is already under way.
  • the effective open period x flow passage section (hereinafter designated an open time area) is the areabelow the curve I.
  • the lift curves II, III and IV represent the stroke ofa tappet valve operated by means of a device according to the afore-described embodiments of the invention. These curves all have a relatively steep course since the switching time of the electromagnet is extremely short. Thus, the slope of these curves depends exclusively from. flow resistances and refill'periods. Since the open ing time of the solenoid valve always remains the same, whereas the aforenoted effects change with the rpm (i.e. with the available length of time), to each rpm there corresponds a different curve. Thus, curves II, III and IV correspond to three different rpms. In the stroke portions s s and s s the dampening becomes effective.
  • the area under the curves II, III and IV is substantially larger than that under the curve 1, resulting in the advantages set forth earlier.
  • FIGS. 4 and 5 in a partially shown cylinder head la of an internal combustion engine there operates a tappet valve 2a having a valve stem 3a axially slidably guided by a bearing sleeve 4a secured in the wall of the cylinder head la.
  • the outer end of the valve stem 3a carries a spring seat disc 5a.
  • a valve closing spring 6a Between the upper face of the cylinder head 1a and the spring seat disc 5a there is disposed in a preloaded condition a valve closing spring 6a.
  • a delivery pump 14a draws liquid through a suction conduit 12a from a tank 13a and delivers it under a pressure of, for example kg/cm through a pressure conduit towards the solenoid valve 8a. From the conduit 15a there extends, downstream of -the pump 14a, a return conduit 16a in which there is disposed a pressure control valve 17a and which terminates in the tank 13a.
  • conduit 151 From conduit 151; there extend conduits 18a which lead to the control devices of the other engine valves and which carry liquid under pressure delivered by the same pump 14a.
  • the liquid first flows through a supply channel 19a which is connected with a control chamber 21a by means of a bore 20a serving also as a supply channel.
  • the mouth of the bore 20a at the control chamber 21a serves as a valve seat for a sphere 22a which is movably disposed in the control chamber 21a.
  • From the. control chamber 210 there extends a bore 230 to the chamber 110.
  • From the control chamber 21a there also extends a bore 240 which is in continuous communication with the suction conduit 12a through discharge channels 25a and 26a and a return conduit 26a.
  • the mouth of the bore 24a at the control chamber 21a serves as a valve seat for the valve sphere 220.
  • An armature 28a having a pin-like extension 28a in contact with the sphere 22a, is slidably disposed in the bore 24a and urges, under the action of aspring 27a, the sphere 220 into a position in which it closes the bore 20a.
  • the control chamber 21a, the bore 24a and the ar mature 28a are contained in a valve support sleeve 29a which is inserted in the housing bracket 7a and which is held in position by a flange of the housing 30a. of the electromagnetic forming part of the solenoid 8a.
  • the electromagnet chamber accommodating the spring 27a and the bore 200 immediately upstream of the solenoid valve, are interconnected by means of a channel 310 so that in both aforenoted spaces identical pressures prevail.
  • the diameter of the cylindrical surface of armature 28a sliding in the bore 24a in a fluidtight manner is identical to the diameter of both seats for the sphere 22a.
  • the force derived from the pressure prevailing in bore 20a and exerted on the sphere in' the opening direction is identical to the force of the pressurized liquid exerted on the sphere by the armature 28a in the closing direction.
  • Spring 27a aids the latter force so that the sphere 22a is, as a net result of the opposing forces, urged against the mouth or valve seat provided about the bore 20a.
  • the force of the spring 27a is overcome by the magnetic force and the armature 28a is displaced towards the left.
  • the pressurized liquid thus may flow through bore 20a, pressing the sphere 22a against its other, opposite seat formed about the opening of bore 24a.
  • sphere 22a closes the bore 24a so that the liquid ad mitted under pressure through the bore 23a may flow into the chamber 11a, displacing the actuating piston 9a. This results in the opening of the tappet valve 2a.
  • the spring 27a returns the armature 28a and the sphere 22a into their initial position in which the bore 20a is again closed.
  • the liquid may flow from the chamber 11a through the bore 24a, the channels 25a and 26a and the return conduit 26a to the suction side of pump 14a allowing the actuating piston 9a to return. This, in turn, causes the tappet valve 2a to assume its closed position.
  • a check valve 36a In the suction conduit 12a upstream of the return conduit 26'a there is disposed a check valve 36a, so that the returning liquid generates a pressure build-up at the suction side of the pump 14a resulting in an improved efficiency and power thereof.
  • the motion of the actuating piston 9a is hydraulically braked towards the end of each stroke.
  • a collar 34a which, towards the end of the strokes of actuating piston 9a penetrates into one or the other cavity 35a, 35'a which have approximately the same diameter as the collar 340.
  • the latter displaces the liquid present in that cavity through a radial throttle gap which is defined by the wall of the cavity and the periphery of collar 34a.
  • a dampening of the motion of work piston 9a is achieved.
  • the walls of cavities 35a and 35 'a, as well as collar 34a may have a conical configuration.
  • the return and supply conduits circumvent the solenoid valve and, controlled by a plunger 37a, lead directly to the chamber 11a.
  • plunger 37a has a circumferential annular groove 38a I which is in continuous communication with the supply channel 190 and an annular circumferential groove 39a which, in turn, is in continuous communication with the return channel 26a.
  • one of the circumferential annular grooves 38a or 390 communicates with a channel 40a which, in turn, when the tappet valve 2a is in an open position, communicates with the chamber 11a.
  • the connection between channel 40a and chamber 11a is controlled by the actuating piston 9a. Only after the latter has traveled a predetermined path, does the channel 40a open. For an axial pressure relief of the piston 9a, the latter has, in the range of the mouth of channel 40a, a circumferential annular groove 41a.
  • the control plunger 37a operates in a bore 44a which is connected through a channel 45a with the channel 25a upstream of the throttle 43a.
  • liquid flows under pressure through channel 45a into the bore 44a and displaces the control plunger 37a against the force of a spring 46a until said plunger abuts against a shoulder 47a of the bore 44a (FIG. 5).
  • the supply channel 19a is separated from the channel 40a, whereas through groove 39a, the return channel 26a is connected with the channel 40a.
  • the liquid may flow in an unthrottled manner from the chamber 11a through the bore 40a, the annular groove 39a, the return channel 26a and the return conduit 26a to the suction side of the pump 14a.
  • the actuating piston 9a Shortly before the tappet valve 2a closes, the actuating piston 9a shuts off the channel 40a. During the entire period of the return motion of actuating piston 9a, effected by the valve spring 6a, in the channel 25a, there prevails a pressure (caused by the throttle 43a) which is sufficient to maintain the control plunger 37a in its terminal position shown in FIG. 5. As soon as the actuating piston 90, however, returns into its initial position, the liquid pressure in the channel 25a decreases, so that the spring 46a may return the control plunger 37a into its initial position in which, as shown in FIG. 4, the supply channel 19a is in communication with the channel 40a.
  • FIGS. 6 AND 7 the embodiment no play or clearance between the valve stem 3a and the actuating piston 9a.
  • the control plunger 37a is provided with a further annular circumferential groove 49a which, by means of radial bores a and an axial bore 51a, is connected with the annular circumferential groove 39a.
  • the latter is provided with a collar 52a which is slidable in a fluidtight manner in a bore 530. From one end of the bore 53a there extends a channel 48a which, dependent upon the position of the plunger 37a, connects said end of bore 53a either with groove 49a (FIG. 6) or with groove 38a (FIG. 7) of the plunger 37a.
  • the other end of the bore 53b is in continuous communication with the annular circumferential groove 390 through a channel 540.
  • a throttle 43a in the channel25a there is provided, in this embodiment, in the return conduit 26'a, a check valve 56a the opening pressure of which is designed in such a manner that in the entire return channel system a sufficiently high pressure is generated and maintained.
  • the supply channel 19a is in hydraulic communication with the channel 48a through the annular circumferential groove 38a, so that the radial face of the collar 52a of the piston 90 is exposed to a pressure in the closing DESCRIPTION OF THE EMBODIMENT SHOWN IN FIGS.
  • FIGS. 8 and 9 in a partially shown cylinder head lb of an inter nal combustion engine there operates a tappet valve 2b having a valve stem 3b axially slidably guided by a bearing sleeve 4b secured in the wall of the cylinder head lb.
  • the outer end of the valve stem 3b carries a spring seat disc 5b.
  • a valve closing spring 6b Between the upper face of the cylinder head lb and the spring seat disc 5b there is disposed in a preloaded condition a valve closing spring 6b.
  • the latter is axially displaceably guided in a fluid-tight manner in a bushing 10b secured to the bracket 7b.
  • a bushing 10b secured to the bracket 7b.
  • the lower terminal face of the actuating piston 9b is urged into contact with the valve stem 3b.
  • the upper end of actuating piston 9b projects into a chamber 11b.
  • a delivery pump 14b draws liquid from a tank 13b and delivers it under a pressure of, for example 100 kg/cm, through a pressure conduit 15b towards the solenoid valve 8b. From the conduit 15b there extends,
  • conduit 15b there extend conduits 18b which lead to the control devices of the other engine valves and which carry liquid under pressure delivered by the same pump 14b.
  • the liquid first flows into a bore 20b which is connected with a control chamber 21b.
  • the mouth of the bore 20b at the control chamber 21b serves as a 10 valve seat for a sphere 22b which is movably disposed in the control chamber 21b.
  • From the control chamber 21b there extends a bore 23b to the chamber 1 1b.
  • From the control chamber 21b there also extends a bore 24b which is in continuous communication with the tank 13b through a discharge channel 25b and a return conduit 26b.
  • the mouth of the bore 24b at the control chamber 21b serves as a valve seat for the valve sphere 22b.
  • An armature 28b having a pin-like extension 28'b in contact with the sphere 22b, is slidably disposed in the bore 24b and urges, under the action of a spring 27b, the sphere 22b into a position. in which it closes the bore 20b.
  • the control chamber 21b, the bore 24b and the armature 28b are contained in a valve support sleeve 29b which is inserted in the housing bracket 7b and which is held in position by a flange of the housing 30b of the electromagnet forming part of the solenoid 8b.
  • the electromagnet chamber accommodating the spring 27b and the bore 20b immediately upstream of the solenoid valve, are interconnected by means of a channel 31b so that in both aforenoted spaces identical pressures prevail.
  • the diameter of the cylindrical surface of armature 28b sliding in the bore 24b in a fluidtight manner is identical to the diameter of both seats for the sphere 22b.
  • the force derived from the pressure prevailing in bore 20b and exerted on the sphere in the opening direction is identical to the force of the pressurized liquid exerted on the sphere by the armature 28b in the closing direction.
  • Spring 27b aids the latter force so that the sphere 22b is, as a net result of the opposing forces, urged against the mouth or valve seat provided about the bore 20b.
  • the force of the spring 27b is overcome by the magnetic force and the armature 28b is displaced towards the left.
  • the pressurized liquid thus may flow through bore 20b, pressing the sphere 22b against its other, opposite seat formed about the opening of bore 24b.
  • sphere 22b closes the bore 24b so that the liquid admitted under pressure through the bore 23b may flow into the chamber 11b, displacing the actuating piston 9b. This results in the opening of the tappet valve 2b.
  • the spring 27b returns the armature 28b and the sphere 22b into their initial position in which the bore 20b is again closed.
  • the liquid may flow from the chamber 11b through the bore 24b, the channel 25b and the return conduit 26b to the liquid tank 13b allowing the actuating piston 9b to return. This, in turn, causes the tappet valve 2b to assume its closed position.
  • actuating piston 9b is, similarly to the third embodiment hydraulically braked towards the end of its stroke.
  • the lateral face of the work piston 9b is provided with a collar 34b which, at the end of each return stroke of piston 9b, penetrates into a cavity 35b which has a diameter approximately identical to that of the collar 34b.
  • the piston 9b displaces the liquid through a radial throttle gap formed between the wall of the cavity 35b and the periphery of the collar 34b.
  • the collar 34b opens a bore 36b through which lubricating oil may flow from a lubricating system (not shown) into the cavity 35b.
  • the lubricating oil may flow into the spring chamber of the tappet valve 217.
  • the throttle gap defined between the collar 34b and the cavity 35b narrows towards the end of the return stroke in such a manner that a predetermined minimum volume remains locked between the collar 34b and the cavity 35b.
  • Said minimum volume by means of its elasticity, functions as an equalizer of clearance between the actuating piston 9b and the valve stem 3b. Such an equalization is advantageous since the dimensions of the valve stem 3b and the piston 9b vary as the temperature changes.
  • FIGS. 10 and 11 similarly to the afore-described embodiments, in an only partially shown cylinder head 1c ofan internal combustion engine there operates a tappet valve 20 having a valve stem-3c axially slidably guided by a bearing sleeve 4c secured in the wall of the cylinder head 1c.
  • the outer end of the valve stem 30 carries a spring seat disc 50.
  • a valve closing spring 6c Between the upper face of the cylinder head 16 and the spring seat disc 5c there is disposed in a preloaded condition a valve closing spring 6c.
  • a housing bracket 7c containing a solenoid valve generally indicated at 8c and an actuating piston 9c.
  • the latter is axially displaceably guided in a fluid-tight manner in a bushing 100 secured to the bracket 7c.
  • the lower terminal'face of the actuating piston 96 is urged into contact with the valve stem 3c.
  • the upper end of actuating piston 90 projects into a chamber 11c.
  • a delivery pump 14c draws liquid from a tank 13c and delivers it under a pressure of, for example 100 kg/cm through a pressure conduit 15c towardsthe solenoid valve 8c. From the conduit 15c there extends, downstream of the pump 140, a return conduit 160 in which there is disposed a pressure control valve 17c and which terminates in the tank 13c.
  • conduit 150 From conduit 150 there extend conduits 180 which lead to the control devices of the other engine valves and which carry liquid under pressure delivered by the same pump 140.
  • the liquid first flows through a nipple 190 which is connected with a control chamber 210 by means of a bore 20c.
  • the mouth of the bore 200 at the control chamber 21c serves as a valve seat for a sphere 22c which is movably disposed in the control chamber 210.
  • From the control chamber 21c there extends a bore 230 to the chamber 11c.
  • From the control chamber 210 there also extends a bore 240 which is in continuous communication with the tank 130 through a discharge channel 256 and a return conduit 260.
  • the mouth of the bore 240 at the control chamber 21c serves as a valve seat for the valve sphere 220.
  • An armature 28c having a pin-like extension 28's in contact with the sphere 220, is slidably disposed in the bore 24c and urges, under the action of a spring 270, the sphere 22c into a position in which it closes the bore 20c.
  • the control chamber 21c, the bore 240 and the armature 28c are contained in a valve support sleeve 29c which is inserted in the housing bracket 70 and which is held in position by a flange of the housing 300 of the electromagnet forming part of the solenoid 8c.
  • the electromagnet chamber accommodating the spring 270 and the bore 20c immediately upstream of the solenoid valve,are interconnected by means of a channel 31c so that in both aforenoted spaces identical pressures prevail.
  • the diameter of the cylindrical surface of armature 28c sliding in bore 24c in a fluid-tight manner is identical to the diameter of both seats for the sphere 220.
  • the force derived from the pressure prevailing in bore 200 and exerted on the sphere inthe opening direction is identical to the force of the pressurized liquid exerted on the sphere by the armature 280 in the closing direction.
  • Spring 27c aids the latter force so that the sphere 220 is, as a net result of the opposing forces, urged against the mouth or valve seat provided about the bore 200.
  • the force of the spring 27c is overcome by the magnetic force and the armature 28c is displaced towards the left.
  • the pressurized liquid thus may flow through bore 20c, pressing the sphere 220 against its other, opposite seat formed about the opening of bore 24c.
  • sphere 22c closes the bore 24c, so that the liquid admitted under pressure through the bore 23c may flow into the chamber 11c, displacing the actuating piston 9c. This results in the opening of the tappet valve 2c.
  • the spring 27 creturns the armature 28c and the sphere 220 into their initial position in which the bore 200 is again closed.
  • the liquid may flow from the chamber through the bore 240, the channel 250 and the return conduit 26c to the liquid tank allowing the actuating piston 90 to return. This, in turn, causes the tappet valve 2c to assume its closed position.
  • the motion of the actuating piston 90 may be hydraulically braked towards the end of each stroke.
  • a collar 340 which, towards the end of each stroke, penetrates into one or the other cavity 350, 35c, which have approximately the same diameter as the collar 340. As soon as the latter penetrates into one of the cavities 35c, 35's, it displaces the liquid present in that cavity through a radial throttle gap defined by the wall of the cavity and the periphery of the collar 340.
  • the power of the delivery pump is variable by means of a setting device 370 which may be formed of a hydraulic setting piston, but it may be an electric setting motor or any other similar device.
  • the setting device 370 receives control signals from an electronic control apparatus 38c which converts data relating to actual conditions (particularly characteristics of the engine operation, such as the position of the accelerator 400 or the position of the brake pedal, as well as the pressure in the hydraulic system which is applied to apparatus 380 through a conduit 410) into a desired value for the power of the pump 140. This value is then applied to the setting apparatus 37c.
  • the purpose of varying the power of pump 140 is to alter the pressure of the liquid that cause actuating piston 9c and thus valve 2c to execute its opening stroke. By varying said pressure, the amplitude of the opening stroke of the tappet valve 2c may be varied, because the equilibrium between the opposing forces of hydraulic pressure and valve spring 6c occurs at different positions of the tappet valve 2c as the hydraulic pressure is changed.
  • the electronic control apparatus 380 includes the regulator 38R and the converter 38F which is in the feed-back circuit to the regulator 38R.
  • the regulator 38R is designed as an amplifier, the output volume of which controls the electromagnetically operating volume regulator of the pump 14c. So as to obtain the precise value as indicated by the accelerator 40c for the output pressure of the pump 140, the feed-back circuit 410 forms with the converter 38F a closed pressure regulating circuit.
  • the converter 38F generates an electric signal from the pressure appearing at the output of the pump 140; the magnitude of said electric signal is comparable to the output volume or control value of the accelerator.
  • control apparatus 380 may also serve to energize and de-energize the solenoid valve SC for initiating the opening and closing of the tappet valve 2c and thus determine the timing of valve operation.
  • FIG. 11 illustrates the ad vantage of the embodiment shown in FIG. 10.
  • the stroke s of the tappet valve 20 is shown as a function of the angle of rotation at of the engine cam shaft.
  • the stroke s indicates the structurally possible largest stroke of the tappet valve.
  • Thecurve I encloses an open time area obtained in case the valve 20 executes an opening stroke of maximum amplitude.
  • the power of the pump 14c is lowered, then the amplitude of the opening stroke drops to s
  • the resulting valve lift curve II encloses a smaller open time area.
  • the open time area may be further altered.
  • valve stem 3d axially slidably guided by a bearing sleeve 4d secured in the wall of the cylinder head 1d.
  • the outer end ofthe valve stem 3d carries a spring seat disc 5d.
  • a valve closing spring 6d Between the upper face of the cylinder head Id and the spring seat disc 5d there is disposed in a preloaded condition a valve closing spring 6d.
  • a housing bracket 7d containing a solenoid valve generally indicated at 8d and an actuating piston 9d.
  • the latter is axially displaceably guided in a fluid-tight manner in a bushing 10d secured to the bracket 7d.
  • the lower terminal face of the actuating piston 9d is urged into contact with the valve stem 3d.
  • the upper end of actuating piston 9d projects into a chamber 1 1d.
  • a delivery pump 14d draws liquid from a tank 13d and delivers it under a pressure of, for example 100 kg/cm through a pressure conduit 15d towards the solenoid valve 8d. From the conduit 15d there extends, downstream of the pump 14d, a return conduit 16d in which there is disposed a pressure control valve 17d and which terminates in the tank 13d.
  • conduit 15d there extend conduits 18d which lead to the control devices of the other engine valves and which carry liquid under pressure delivered by the same pump 14d.
  • the liquid first flows into a bore 20d which is connected with a control chamber 21d of the solenoid valve 8d.
  • the mouth of the bore 20d at the control chamber 21d serves as a valve seat for a sphere 22d which is movably disposed in the control chamber 21d.
  • From the control chamber 21d there extends a bore 23d to the chamber 1 1d.
  • From the control chamber 21d there also extends a bore 24d which is in continuous communication with the tank 13d through a discharge channel 25d and a return conduit 26d.
  • the mouth of the bore 240! at the control chamber 21d serves as a valve seat for the valve sphere 22d.
  • An armature 28d having a pin-like extension 28'd in contact with the sphere 22d, is slidably disposed in the bore 24d and urges, under the action of a spring 27d, the sphere 22d into a position in which it closes the bore 20d.
  • control chamber 21d, the bore 24d and the armature 28d are contained in a valve support sleeve 29d which is inserted in the housing bracket 7d and which is held in position by a flange of the housing 30d of the electromagnet forming part of the solenoid 8d.
  • electromagnet chamber accommodating the spring 27d and the bore 20d immediately upstream of the solenoid valve, are interconnected by means of a channel 31d so that in both aforenoted spaces identical pressures prevail.
  • the diameter of the cylindrical surface of armature 28d sliding in the bore 24d in a fluidtight manner is identical to the diameter of both seats for the sphere 22d.
  • the spring 27d returns the armature 28a and the sphere 22d into their initial position in which the bore 20d is again closed.
  • the liquid may flow from the chamber 11d through the bore 24d, the channel 25d and the return conduit 26d to the liquid tank 13d allowing the actuating piston 9d to return.
  • This causes the tappet valve 2d to assume its closed position.
  • the flow resistance of the channels ensures an operation free from play between the actuating piston 9d and the valve stem 3d.
  • the closing motion of the actuating piston 9d is, towards the end of each stroke, hydraulically braked.
  • a collar 34d which, towards the end of each stroke, penetrates into one or the other cavity 35d, 35d, which have approximately the same diameter as the collar 34d. From the cavities 35d, 35d, the collar 34d displaces the liquid through an annular gap defined by the wall of the cavity and the periphery of the collar, resulting in a dampening of the piston stroke.
  • the throttle gap has to be of such minimum dimension that even at high rpms, the tappet valve 2d closes entirely.
  • the moment of opening and closing of the tappet valve 2d may be controlled independently from one another.
  • the desired values of magnitude set by the electronic control apparatus 33d are determined from the evaluation of actual sensed data, particularly those relating to operational magnitudes of the e'ngine, such as the position of the accelerator (i.e. load), rpm, external pressure, engine temperature, etc.
  • the signals relating to the sensed actual magnitudes are applied through conductors 38d .to the electronic control apparatus 33d.
  • Such input signal may be derived, for example, from the position of an accelerator pedal 39d, sensed by a device 40d not shown in detail.
  • Such electronic control apparatus 33d as mentioned is described in the German Pat. No. 1,100,377 (Bendix).
  • FIG. 13 illustrates the advantages of the embodiment shown in FIG. 12.
  • the stroke s (ordinate) of the tappet valve is illustrated as a function of the angular position a (abscissa) of the engine crankshaft (KW).
  • the valve lift curves associated with an exhaust valve and an intake valve are shown side by side.
  • the opening stroke is constant and is designated with s
  • the maximum open period of the exhaust valve constant opening and closing velocity of the tappet valves,these curves I have a relatively flat slope.
  • the two lift curves I overlap, that is, the intake valve begins to open before the exhaust valve is completely closed.
  • the angle of overlap is the same.
  • the curves II represent a valve operation with minimum rpm under full load conditions.
  • the slope of the lift curves II for the opening and closing strokes is substantially steeper than that of the lift curves I. It is seen that the duration of the opening stroke expressed in a is shorter here because i of the lower rpm. The result is a greater open time area.
  • the opening moment of the exhaust valve may be the same under full load conditions for maximum rpm and minimum rpm, whereas the closing moment for maximum rpm occurs later (to cause overlap) than in case of minimum rpm.
  • the intake valve the converse ap plies.
  • the curves III correspond to the partial load range.
  • the moment of closing the exhaust valve and the moment of opening the intake valve may be shifted with respect to one another.
  • the positive or negative slop of the curve may vary.
  • the valve operation may be altered, on the one'hand, as a function of the load and, on the other hand, as a function of the rpm.
  • the said variations as a function of load and rpm are independent from one another, because one curve shift is effected as a function of the accelerator position, whereas the other is dependent upon the engine rpm.
  • this type of control has the significant advantage that the mean pressure or the torque substantially increases with decreasing rpm which results in a decrease of the fuel consumption in case of a continuous injection and further results in the decrease of the pollutants in the exhaust gases.
  • a substantial improvement of the engine operation and particularly a quiet run in the idling range is achieved.
  • Curve IV pertains to an exhaust valve and encloses a relatively small area, so that beyond a determined rpm a braking effect appears.
  • the opening or closing moment of the exhaust valve is varied as a function of the position of the brake pedal and the engine rpm.
  • the open period of the exhaust valve is variable as a function of the brake pedal position; the open time area of the exhaust valve decreases as the brake pedal changes its position in the direction of increased braking effect.
  • the braking moment of the engine aids the mechanical braking by forcing the engine piston to perform increased work during its exhaust stroke.
  • the fuel admission is shut off, so that only air is compressed and displaced.
  • the load-and-rpm-dependent change of the opening and closing moments of the intake and exhaust valves may contribute, in either Otto or diesel engines, to additional power increases.
  • the lift curve of the intake valve may much more substantially overlap the exhaust valve curve I than does the inlet valve curve I. Consequently, the open time area of the intake valve is increased.
  • the portion corresponding to the opening course of the intake valve is, in curve V, as compared to curve I, shifted parallel in the direction of the exhaust valve curve.
  • curve VI For the exhaust valve, this curve VI, which corresponds to the partial load range, appears duringthe braking of the engine by means of the exhaust valve.
  • variable overlap or separation of the valve opening also has an effect on the course of combustion and thus, on the composition of exhaust gases. It is noted that the CO, CH and NO components appear mostly in the low to middle load and rpm ranges.
  • the CO content is reduced by means-of an overlap which decreases with the rpm, in that the residual gas quantities are decreased and a leaner mixture adaptation is possible. In addition, a miss-free operation may be achieved.
  • the CH content may be reduced in this range by a large overlap possibly coordinated with the injection periods. Particularly in case of delays when coasting in gear, the CH content may be substantially decreased by a decrease or elimination of the overlap.
  • This valve control may be effective to such an extent that the heretofore necessary fuel shut-off device operative while coasting in gear, may be dispensed with.
  • the NO content is lowered by reintroducing the exhaust gas into the intake suction system; the variable valve time overlap is insofar advantageous since the metering device and shut-off apparatus of an extem al exhaust reintroducing system may be omitted.
  • a tappet valve 2e having a valve stem 3e axially slidably guided by a bearing sleeve 42 secured in the wall of the cylinder head 12.
  • the outer end of the valve stem 32 carries a spring seat disc 52. Between the upper face of the cylinder head 12 and the spring seat disc 52 there is disposed in a preloaded condition a valve closing'spring 62.
  • a housing bracket 72 containing a solenoid valve generally indicated at 82 anda hydraulically operated actuating piston 9e. The latter is axially displaceably guided in a fluid'tight manner in a bushing 102 secured to the bracket 7e.
  • the lower terminal face of the actuating piston 9e is urged into contact with the valve stem 32.
  • the upper end of actuating piston 92 projects into a chamber 112 which leads to solenoid valve 82.
  • a delivery pump 142 draws liquid from a tank 132 and delivers it under a pressure of, for example 100 kg/cm through a pressure conduit 152 towards the solenoid valve 82. From the conduit 152 there extends, downstream of the pump 142, a return conduit 162 in which there is disposed a pressure control valve 172 and which terminates in the tank 132.
  • conduit 152 From conduit 152 there extend conduits 182 which lead to the control devices of the other engine valves and which carry liquid under pressure delivered by the same pump 142.
  • the liquid first flows into abore 202 which is connected with a control chamber 212.
  • the mouth of the bore 202 at the control chamber 212 serves as a valve seat for a sphere 222 which is movably disposed in the control chamber 212.
  • From the control chamber 212 there extends a bore 232 to the chamber 112.
  • From the control chamber 212 there also extends a bore 242 which is in continuous communication with the tank 132 through a discharge channel 252 and a return conduit 262.
  • the mouth of 'the bore 242 at the control chamber 212 serves as a valve seat for the valve sphere 222.
  • An armature 28e having a pin-like extension 282 in contact with the sphere 222, is slidably disposed in the bore 242 and urges, under the action of a spring 272, the sphere 222 into a position in which it closes the bore 202.
  • the control chamber 212, the bore 242 and the armature 282 are contained in a valve support sleeve 292 which is inserted in the housing bracket7e and which is held in position by a flange of the housing 302 of the electromagnet forming part of the solenoid 82.
  • the electromagnet chamber accommodating the spring 272 and the bore 202 immediately upstream of the solenoid valve, are interconnected by means of a channel 312 so that in both aforenoted spaces identical pressures prevail.
  • the diameter of the cylindrical surface of armature 282 sliding in the bore 242 in a fluid-tight manner is identical to the diameter of both seats for the sphere 222.
  • the force derived from the pressure prevailing .in bore 202 and exerted on the sphere in the opening direction is identical to the force of the pressurized liquid exerted on the sphere by the armature 282 in the closing direction.
  • Spring 27e aids the latter'force so that the sphere 222 is, as a net result of the opposing forces, urged against the mouth or valve seat provided about the bore 202.
  • the force of the spring 272 is overcome by the magnetic force and the armature 282 is displaced towards the left.
  • the pressurized liquid thus may flow through bore 202, pressing the sphere seat formed about the opening of bore 242.
  • sphere 222 closes the bore 242, so that the liquid adinto the chamber 112, displacing the actuating piston 92. This results in the opening of the tappet valve 2e.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Magnetically Actuated Valves (AREA)
  • Valve Device For Special Equipments (AREA)
US00068362A 1969-08-30 1970-08-31 Control device for hydraulically operated tappet valves of internal combustion engines Expired - Lifetime US3727595A (en)

Applications Claiming Priority (6)

Application Number Priority Date Filing Date Title
DE19691944177 DE1944177A1 (de) 1969-08-30 1969-08-30 Steuerung von Ein- und Auslassventilen von Brennkraftmaschinen durch Fluessigkeit
DE19691962916 DE1962916A1 (de) 1969-12-16 1969-12-16 Steuerung von Ein- und Auslassventilen bei Brennkraftmaschinen durch Fluessigkeit
DE19702006304 DE2006304A1 (de) 1970-02-12 1970-02-12 Steuerung von Ein und Auslaßventilen bei Brennkraftmaschinen
DE19702006844 DE2006844A1 (de) 1970-02-14 1970-02-14 Steuerung von Ein und Auslaßventilen bei Brennkraftmaschinen durch Flüssigkeit
DE2008668A DE2008668C3 (de) 1970-02-25 1970-02-25 Vorrichtung zum Steuern eines Einlaß- oder Auslaßventils einer Brennkraftmaschine
DE19702010291 DE2010291A1 (de) 1970-03-05 1970-03-05 Steuerung von Ein- und Auslaßventilen bei Brennkraftmaschinen durch Flüssigkeit

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US3727595A true US3727595A (en) 1973-04-17

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US00068362A Expired - Lifetime US3727595A (en) 1969-08-30 1970-08-31 Control device for hydraulically operated tappet valves of internal combustion engines

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US (1) US3727595A (enExample)
BE (1) BE755478A (enExample)
CH (1) CH503892A (enExample)
FR (1) FR2059174A5 (enExample)
GB (1) GB1294217A (enExample)
NL (1) NL7012856A (enExample)
PL (1) PL75950B1 (enExample)
RO (1) RO60817A (enExample)

Cited By (33)

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Publication number Priority date Publication date Assignee Title
US3978826A (en) * 1975-04-14 1976-09-07 Curtiss-Wright Corporation Rotary engine with intake valve having a variable open period for power control
US4009695A (en) * 1972-11-14 1977-03-01 Ule Louis A Programmed valve system for internal combustion engine
US4153016A (en) * 1977-04-28 1979-05-08 Hausknecht Louis A Valve control system
US4200067A (en) * 1978-05-01 1980-04-29 General Motors Corporation Hydraulic valve actuator and fuel injection system
US4206728A (en) * 1978-05-01 1980-06-10 General Motors Corporation Hydraulic valve actuator system
US4312494A (en) * 1979-07-03 1982-01-26 Nissan Motor Co., Ltd. Valve device using an on-off functioning type electromagnetic actuator
US4347812A (en) * 1978-04-28 1982-09-07 Nippon Soken, Inc. Hydraulic valve lift device
US4483283A (en) * 1983-05-13 1984-11-20 Hausknecht Louis A Variable valve control system with dampener assembly
US4656976A (en) * 1984-04-01 1987-04-14 Rhoads Gary E Hydraulic rocker arm
US4716863A (en) * 1985-11-15 1988-01-05 Pruzan Daniel A Internal combustion engine valve actuation system
US4821689A (en) * 1987-02-10 1989-04-18 Interatom Gmbh Valve drive with a hydraulic transmission and a characteristic variable by means of a link control
US4872425A (en) * 1989-01-06 1989-10-10 Magnavox Government And Industrial Electronics Company Air powered valve actuator
US4875441A (en) * 1989-01-06 1989-10-24 Magnavox Government And Industrial Electronics Company Enhanced efficiency valve actuator
US4982706A (en) * 1989-09-01 1991-01-08 Robert Bosch Gmbh Valve control apparatus having a magnet valve for internal combustion engines
US5058857A (en) * 1990-02-22 1991-10-22 Mark Hudson Solenoid operated valve assembly
DE4132891A1 (de) * 1991-10-04 1993-04-08 Audi Ag Pneumatisch betaetigte ventilsteuerung fuer brennkraftmaschinen
US5233951A (en) * 1992-09-25 1993-08-10 Hausknecht Louis A Flow restriction controlled variable engine valve system
US5327858A (en) * 1992-09-25 1994-07-12 Hausknecht Louis A Flow restriction controlled variable engine valve system
US5619965A (en) * 1995-03-24 1997-04-15 Diesel Engine Retarders, Inc. Camless engines with compression release braking
US5829396A (en) * 1996-07-16 1998-11-03 Sturman Industries Hydraulically controlled intake/exhaust valve
RU2153082C2 (ru) * 1997-12-29 2000-07-20 Бурячко Владимир Романович Гидравлический привод клапанов двигателя внутреннего сгорания
EP1001143A3 (de) * 1998-11-12 2000-12-06 Hydraulik Ring GmbH Ventilsteuerung für Ein- und Auslassventile von Verbrennungsmotoren
WO2001049981A1 (de) * 1999-12-30 2001-07-12 Robert Bosch Gmbh Ventilsteuerung für einen verbrennungsmotor
RU2171898C2 (ru) * 1999-07-12 2001-08-10 Военный автомобильный институт Электрогидравлическая система управления клапанами двигателя внутреннего сгорания
CN1096538C (zh) * 2000-03-27 2002-12-18 武汉理工大学 用电子控制的共管(轨)液压驱动的内燃机的进排气系统
US20030015155A1 (en) * 2000-12-04 2003-01-23 Turner Christopher Wayne Hydraulic valve actuation systems and methods
US20030150415A1 (en) * 2000-07-10 2003-08-14 Mats Hedman Pressure pulse generator
US20080271705A1 (en) * 2006-05-16 2008-11-06 Sims John T Variable compression engine
EP1843013A3 (en) * 2006-03-30 2009-09-09 Dell'orto S.P.A. Variable-actuation, electro-hydraulic system and device controlling the valves of internal combustion engines
WO2012166035A1 (en) * 2011-06-03 2012-12-06 Alternative Solar Energy Engine Ab Pressure pulse generator
US8381693B2 (en) 2007-11-23 2013-02-26 Empa Eidgenossische Materialprufungs-Und Forschungsanstalt Hydraulically operated valve actuation and internal combustion engine with such a valve actuation
WO2016167715A1 (en) * 2015-04-16 2016-10-20 Freevalve Ab Actuator for axial displacement of an object
US20200173378A1 (en) * 2018-12-04 2020-06-04 Toyota Jidosha Kabushiki Kaisha Internal combustion engine system

Families Citing this family (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE2940112A1 (de) * 1979-10-03 1981-04-16 Daimler-Benz Ag, 7000 Stuttgart Druckbetaetigte steuerungsanordnung
FR2552492B1 (fr) * 1983-09-23 1988-01-15 Alsacienne Constr Meca Bloc electro-hydraulique de commande des soupapes pour moteur a combustion interne
DE102011102008B4 (de) 2011-05-19 2025-05-08 Man Truck & Bus Se Verfahren und Vorrichtung zur Überprüfung der Funktionsfähigkeit, insbesondere der Oxidationsfähigkeit, eines in der Abgasleitung einer mit Luftüberschuss betriebenen Brennkraftmaschine verbauten NO-Oxidationskatalysators
CN105804826B (zh) * 2016-05-04 2018-05-18 哈尔滨工程大学 双压电液压驱动增压式配气系统
CN106703928B (zh) * 2016-12-28 2022-07-15 沪东重机有限公司 由伺服油直接驱动的排气阀控制执行系统

Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2392207A (en) * 1943-01-15 1946-01-01 Weiss Abraham Method and means for saving fuel in internal-combustion engines
US3209737A (en) * 1962-06-27 1965-10-05 Mitsubishi Shipbuilding & Eng Valve operating device for internal combustion engine
US3220392A (en) * 1962-06-04 1965-11-30 Clessie L Cummins Vehicle engine braking and fuel control system
US3548793A (en) * 1968-10-31 1970-12-22 James S Richardson Valve actuating mechanism for internal combustion engines

Patent Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2392207A (en) * 1943-01-15 1946-01-01 Weiss Abraham Method and means for saving fuel in internal-combustion engines
US3220392A (en) * 1962-06-04 1965-11-30 Clessie L Cummins Vehicle engine braking and fuel control system
US3209737A (en) * 1962-06-27 1965-10-05 Mitsubishi Shipbuilding & Eng Valve operating device for internal combustion engine
US3548793A (en) * 1968-10-31 1970-12-22 James S Richardson Valve actuating mechanism for internal combustion engines

Cited By (44)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4009695A (en) * 1972-11-14 1977-03-01 Ule Louis A Programmed valve system for internal combustion engine
US3978826A (en) * 1975-04-14 1976-09-07 Curtiss-Wright Corporation Rotary engine with intake valve having a variable open period for power control
US4153016A (en) * 1977-04-28 1979-05-08 Hausknecht Louis A Valve control system
US4347812A (en) * 1978-04-28 1982-09-07 Nippon Soken, Inc. Hydraulic valve lift device
US4200067A (en) * 1978-05-01 1980-04-29 General Motors Corporation Hydraulic valve actuator and fuel injection system
US4206728A (en) * 1978-05-01 1980-06-10 General Motors Corporation Hydraulic valve actuator system
US4312494A (en) * 1979-07-03 1982-01-26 Nissan Motor Co., Ltd. Valve device using an on-off functioning type electromagnetic actuator
US4483283A (en) * 1983-05-13 1984-11-20 Hausknecht Louis A Variable valve control system with dampener assembly
US4656976A (en) * 1984-04-01 1987-04-14 Rhoads Gary E Hydraulic rocker arm
US4716863A (en) * 1985-11-15 1988-01-05 Pruzan Daniel A Internal combustion engine valve actuation system
US4821689A (en) * 1987-02-10 1989-04-18 Interatom Gmbh Valve drive with a hydraulic transmission and a characteristic variable by means of a link control
US4875441A (en) * 1989-01-06 1989-10-24 Magnavox Government And Industrial Electronics Company Enhanced efficiency valve actuator
US4872425A (en) * 1989-01-06 1989-10-10 Magnavox Government And Industrial Electronics Company Air powered valve actuator
US4982706A (en) * 1989-09-01 1991-01-08 Robert Bosch Gmbh Valve control apparatus having a magnet valve for internal combustion engines
US5058857A (en) * 1990-02-22 1991-10-22 Mark Hudson Solenoid operated valve assembly
DE4132891A1 (de) * 1991-10-04 1993-04-08 Audi Ag Pneumatisch betaetigte ventilsteuerung fuer brennkraftmaschinen
US5233951A (en) * 1992-09-25 1993-08-10 Hausknecht Louis A Flow restriction controlled variable engine valve system
US5327858A (en) * 1992-09-25 1994-07-12 Hausknecht Louis A Flow restriction controlled variable engine valve system
US5619965A (en) * 1995-03-24 1997-04-15 Diesel Engine Retarders, Inc. Camless engines with compression release braking
CN1085774C (zh) * 1996-07-16 2002-05-29 斯特曼工业公司 液压控制的进/排气阀组件和内燃机
US5829396A (en) * 1996-07-16 1998-11-03 Sturman Industries Hydraulically controlled intake/exhaust valve
RU2153082C2 (ru) * 1997-12-29 2000-07-20 Бурячко Владимир Романович Гидравлический привод клапанов двигателя внутреннего сгорания
EP1001143A3 (de) * 1998-11-12 2000-12-06 Hydraulik Ring GmbH Ventilsteuerung für Ein- und Auslassventile von Verbrennungsmotoren
RU2171898C2 (ru) * 1999-07-12 2001-08-10 Военный автомобильный институт Электрогидравлическая система управления клапанами двигателя внутреннего сгорания
WO2001049981A1 (de) * 1999-12-30 2001-07-12 Robert Bosch Gmbh Ventilsteuerung für einen verbrennungsmotor
US6581557B1 (en) 1999-12-30 2003-06-24 Robert Bosch Gmbh Valve control for an internal combustion engine
KR100745391B1 (ko) * 1999-12-30 2007-08-03 로베르트 보쉬 게엠베하 내연기관용 밸브 제어 방법 및 장치
CN1096538C (zh) * 2000-03-27 2002-12-18 武汉理工大学 用电子控制的共管(轨)液压驱动的内燃机的进排气系统
US20030150415A1 (en) * 2000-07-10 2003-08-14 Mats Hedman Pressure pulse generator
US6752106B2 (en) * 2000-07-10 2004-06-22 Cargine Engineering Ab Pressure pulse generator
US20030015155A1 (en) * 2000-12-04 2003-01-23 Turner Christopher Wayne Hydraulic valve actuation systems and methods
US6739293B2 (en) 2000-12-04 2004-05-25 Sturman Industries, Inc. Hydraulic valve actuation systems and methods
EP1843013A3 (en) * 2006-03-30 2009-09-09 Dell'orto S.P.A. Variable-actuation, electro-hydraulic system and device controlling the valves of internal combustion engines
US20080271705A1 (en) * 2006-05-16 2008-11-06 Sims John T Variable compression engine
US8381693B2 (en) 2007-11-23 2013-02-26 Empa Eidgenossische Materialprufungs-Und Forschungsanstalt Hydraulically operated valve actuation and internal combustion engine with such a valve actuation
WO2012166035A1 (en) * 2011-06-03 2012-12-06 Alternative Solar Energy Engine Ab Pressure pulse generator
US8973541B2 (en) 2011-06-03 2015-03-10 Perfecter Ab Pressure pulse generator
EP2715077A4 (en) * 2011-06-03 2015-07-15 Alternative Solar Energy Engine Ab PRESSURE PULSE GENERATOR
KR101557337B1 (ko) 2011-06-03 2015-10-05 얼터네이티브 솔라 에너지 엔진 에이비 압력 펄스 생성장치
WO2016167715A1 (en) * 2015-04-16 2016-10-20 Freevalve Ab Actuator for axial displacement of an object
CN107567535A (zh) * 2015-04-16 2018-01-09 弗瑞瓦勒夫股份公司 用于物体的轴向移动的致动器
CN107567535B (zh) * 2015-04-16 2020-01-10 弗瑞瓦勒夫股份公司 用于物体的轴向移动的致动器
US10577988B2 (en) 2015-04-16 2020-03-03 Freevalve Ab Actuator for axial displacement of an object
US20200173378A1 (en) * 2018-12-04 2020-06-04 Toyota Jidosha Kabushiki Kaisha Internal combustion engine system

Also Published As

Publication number Publication date
GB1294217A (en) 1972-10-25
BE755478A (fr) 1971-02-01
RO60817A (enExample) 1976-10-15
CH503892A (de) 1971-02-28
FR2059174A5 (enExample) 1971-05-28
PL75950B1 (enExample) 1975-02-28
NL7012856A (enExample) 1971-03-02

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