US20130340694A1

US20130340694A1 - Variably operated valve system for internal combustion engine

Info

Publication number: US20130340694A1
Application number: US13/859,369
Authority: US
Inventors: Makoto Nakamura
Original assignee: Hitachi Automotive Systems Ltd
Current assignee: Hitachi Astemo Ltd
Priority date: 2012-06-22
Filing date: 2013-04-09
Publication date: 2013-12-26
Also published as: JP2014005756A; DE102013010507A1; CN103511016A

Abstract

A variably operated valve system for an internal combustion engine, a valve stop mechanism (11) makes at least one of a pair of fulcrum members lost motion to stop open-and-closure drives for one of two engine valves per cylinder and a stop of open-and-closure drives for the one of the engine valves is inhibited, in a case where a lost motion quantity of the valve stop mechanism exceeds a predetermined value (M3).

Description

BACKGROUND OF THE INVENTION

(1) Field of the Invention
The present invention relates to a variably operated valve system for an internal combustion engine which is capable of performing a variable control of a valve lift quantity of an engine valve (one or each of a pair of a pair of engine valves per engine cylinder) and performing a valve stop.
(2) Description of Related Art
A Japanese Patent Application First Publication (tokkai) No. 2010-007636 published on Jan. 14, 2010 exemplifies a previously proposed variably operated valve system for an internal combustion engine.
The previously proposed variable valve system disclosed in the above-described Japanese Patent Application First Publication includes a variable mechanism which continuously varies lift quantities and working angles of two intake valves per cylinder and a stop mechanism which stops open-and-closure operations of one of the two intake valves. In a commonly used driving range, one of the two intake valves is made lost motion to stop the one intake valve so that only the other of the two intake valves is driven to strengthen an intake swirl, thereby a fuel consumption and a combustion performance being improved. On the other hand, in a driving range requiring an engine torque, the two intake valves are driven and the working angles of both of the two intake valves are made substantially equal to each other so that an intake air charging efficiency can be improved.
Then, in a case where the required engine torque is increased in a one-valve driving state, the one-valve driving state is maintained and the working angle (or the lift quantity) is increased if this required engine torque is smaller than a maximum torque in the one-valve driving.

SUMMARY OF THE INVENTION

However, in the previously proposed variably operated valve system, as described before, the one-valve driving state is maintained if the required engine torque is smaller than the maximum torque in the one-valve driving. Therefore, the working angle of one of variably operated valve mechanisms in which the related valve is stopped is forced to be expanded, in other words, the lift quantity and a lost motion quantity are forced to be large.
Therefore, an excessive load is applied onto a posture of the one of the variably operated valve mechanisms in which the related valve is stopped so that a space between one end section of one of swing arms via which the one intake valve is open-and-closure operated and a pivot which provides a swing fulcrum of the one swing arm becomes non-uniform or becomes a local contact. Then, a positional gap between both of the one end section and the pivot occurs and, in some cases, there is a possibility of a dropping out of the one end section of the swing arm from the pivot.
It is, with the above-described problem in the previously proposed variably operated valve system in mind, an object of the present invention to provide a variable operated valve system which is capable of suppressing an irregular behavior such as the positional gap and drop out of the swing arm with respect to the pivot by inhibiting a valve stop control in a case where the lost motion quantity (a stroke quantity) of the valve stop mechanism during the one-valve stop state exceeds a predetermined value.
According to one aspect of the present invention, there is provided a variably operated valve system comprising: a drive shaft to which a rotation driving force is transmitted from an engine crankshaft and on an outer periphery of which drive cams are disposed; two swing cams that operate two engine valves per engine cylinder to open against spring forces of respective valve springs; a transmission mechanism that converts rotation motions of the drive cams into swing motions and transmits the converted swing motions to the respective swing cams; a pair of swing arms interposed between the respective swing cams and the respective engine valves to operate the respective engine valves to open or close; a pair of fulcrum members each of which provides a swing fulcrum of each of the swing arms; a control mechanism that varies a posture of the transmission mechanism to variably control a lift quantity of each of the engine valves; a valve stop mechanism that makes at least one of the pair of fulcrum members lost motion to stop open-and-closure drives for one of the two engine valves; and valve stop inhibit means for inhibiting a stop of open-and-closure drives for the one of the engine valves, in a case where a lost motion quantity of the valve stop mechanism exceeds a predetermined value.
According to another aspect of the present invention, there is provided a variably operated valve system for an internal combustion engine, comprising: drive cams to which a rotational force is transmitted from a crankshaft; engine valves, two of the engine valves being mounted in each of engine cylinders and each engine valve being biased toward a valve closure direction by means of a spring force of a corresponding one of valve springs; swing cams that make the respective engine valves open operation against spring forces of valve springs; a transmission mechanism that converts rotation motions of the drive cams into swing motions and transmits the converted swing motions to the swing cams; a control mechanism that varies a posture of the transmission mechanism to variably control a lift quantity of each of the engine valves; a pair of operational members that swing in accordance with the swing motions of the swing cams to operate the respective engine valves to open or close; a valve stop mechanism that absorbs a swing quantity of one of the pair of operational members to stop open-and-closure operations of one of the two engine valves; and valve stop inhibiting means for inhibiting an absorption of a swing motion of the valve stop mechanism in a case where the swing quantity of each of the pair of the operational members exceeds a predetermined quantity.
According to a still another aspect of the present invention, there is provided a variably operated valve system for an internal combustion engine, comprising: a drive cam rotationally driven through a crankshaft; a pair of engine valves, each engine valve being biased toward a closure direction by means of a spring force of a valve spring; a pair of swing cams that swing to drive the pair of engine valves via a pair of operational members; a transmission mechanism that converts a rotation motion of the drive cam into a swing motion and transmits the converted swing motion to the pair of swing cams; a control mechanism that varies a posture of the transmission mechanism to vary operation characteristics of the pair of the engine valves; and a valve stop mechanism disposed on at least one of the pair of operational members to make a lost motion to stop a drive of one of the pair of engine valves, wherein, in a case where a valve lift quantity of one of the pair of engine valves exceeds a predetermined value, an operation of a valve stop through the valve stop mechanism is inhibited.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a perspective view representing a first preferred embodiment of a variably operated valve system according to the present invention which is applicable to one bank side of a V shaped six cylinder internal combustion engine.

FIG. 2 is a cross sectional view of the variably operated valve system cut away along a line A to A shown in FIG. 1.

FIG. 3 is a cross sectional view of the variable operated valve system cut away along a line B to B shown in FIG. 1.

FIGS. 4A and 4B are a longitudinal cross sectional view representing a first hydraulic pressure rush adjuster in the first preferred embodiment shown in FIG. 1 and a longitudinal cross sectional view representing an action of the first hydraulic pressure rush adjuster.

FIG. 5 is a longitudinal cross sectional view of a second hydraulic pressure rush adjuster in the first preferred embodiment shown in FIG. 1 according to the present invention.

FIG. 6 is a rough configuration view representing a control hydraulic pressure circuit in the first embodiment shown in FIG. 1.

FIG. 7A is an explanatory view for explaining an action of the first hydraulic pressure rush adjuster at a time of a valve closure in a case where a valve lift quantity of an intake valve in the first embodiment is controlled to L2 and FIG. 7B is an explanatory view for explaining an action of the first hydraulic pressure rush adjuster at a time of open of the intake valve in the case where the valve lift quantity of the intake valve in the first embodiment is controlled to L2.

FIG. 8A is an explanatory view for explaining an action of the second hydraulic pressure rush adjuster at a time of a valve closure in a case where a valve lift quantity of an intake valve in the first embodiment is controlled to L2 and FIG. 8B is an explanatory view for explaining an action of the second hydraulic pressure rush adjuster at a time of open of the intake valve in the case where the valve lift quantity of the intake valve in the first embodiment is controlled to L2.

FIG. 9 is an explanatory view for explaining an action of the first hydraulic pressure rush adjuster in a case where a state transition is made from a state in which the lift quantity is controlled to L3 to a state in which a valve stop control is carried out.

FIG. 10A is an explanatory view for explaining an action of the first hydraulic pressure rush adjuster at the time of the valve closure in a case where the lift quantity of the intake valve in the first embodiment is controlled to a maximum lift quantity (L7) and FIG. 10B is an explanatory view for explaining the action of first hydraulic pressure rush adjuster at the time of the valve open in the case where the lift quantity of the intake valve in the first embodiment is controlled to the maximum lift quantity (L7).

FIG. 11A is an explanatory view for explaining an action of the second hydraulic pressure rush adjuster at the time of the valve closure in a case where the lift quantity of the intake valve in the first embodiment is controlled to the maximum lift quantity (L7) and FIG. 11B is an explanatory view for explaining the action of the second hydraulic pressure rush adjuster at the time of the valve open in the case where the lift quantity of the intake valve in the first embodiment is controlled to the maximum lift quantity (L7).

FIG. 12 is a graph representing the lift characteristic of the intake valve in the first embodiment.

FIG. 13 is a characteristic graph representing a relationship between a lost motion quantity of the first hydraulic pressure rush adjuster and a rotation angle of a control shaft, in the first embodiment.

FIG. 14 is a characteristic graph representing a relationship between the lift quantity of the intake valve and the rotation angle of the control shaft in the first embodiment.

FIG. 15 is a control flowchart executed by a control unit used in the first embodiment shown in FIG. 1.

FIG. 16 is a rough configuration view representing the control hydraulic pressure circuit in a second preferred embodiment according to the present invention.

FIG. 17 is a longitudinal cross sectional view representing the first hydraulic pressure rush adjuster in a third preferred embodiment according to the present invention.

FIG. 18 is a longitudinal cross sectional view representing the second hydraulic pressure rush adjuster in a fourth preferred embodiment according to the present invention.

DETAILED DESCRIPTION OF THE INVENTION

Hereinafter, preferred embodiments of a variably operated valve system for an internal combustion engine according to the present invention will be described on a basis of attached drawings in order to facilitate a better understanding of the present invention. In these embodiments, the present invention is applied to a V shaped six cylinder internal combustion engine and the variably operated valve system includes a variable mechanism which variably controls working angles and valve lifts (quantities) of intake valves which are engine valves. A right bank includes a first cylinder #1, a third cylinder #3, and a fifth cylinder #5 and a left bank includes a second cylinder #2, a fourth cylinder #4, and a sixth cylinder #6. However, the right bank and the left bank have mutually the same structure. Hence, hereinafter, only one side bank, namely, the right bank will be described.

First Embodiment

FIGS. 1 through 3 show a first preferred embodiment of the variably operated valve system according to the present invention.
The variably operated valve system includes: first and second intake valves 3, 3 per (one) cylinder which open or closes a pair of intake ports formed within a cylinder head 1; a drive shaft 5 disposed at an upper side of first cylinder #1, third cylinder #3, and fifth cylinder #5 along a forward-or-backward direction of the engine and having three drive cams 5 a on an outer periphery of drive shaft 5; a pair of swing cams 7, 7 rotatably supported on an outer peripheral surface of drive shaft 5 to make the respective intake valves 3, 3 open-and-closure operation via respective swing arms 6, 6 which are operational members; a transmission mechanism 8 which converts a rotational force of respective swing cams into a swing force and transmits the converted swing force to respective swing cams 7, 7; a control mechanism 9 which controls working angles and lift quantities of respective intake valves 3, 3 via transmission mechanism 8; first and second hydraulic pressure rush adjusters 10 a, 10 b which are two fulcrum members (pivots) to make a valve clearance between each intake valve 3, 3 and each swing cam 7, 7 via each swing arm 6 zero-rush; and three valve stop mechanisms 11, each valve stop mechanism 11 stopping the open-and-closure operations of one (first intake valve) of the two intake valves 3 via one side first hydraulic pressure rush adjuster 10 a in accordance with an engine driving condition.
It should be noted that drive shaft 5, swing cam 7, transmission mechanism 8, and control mechanism 9 constitute a variable mechanism.
Hereinafter, for explanation conveniences, each structural member in a single cylinder, for example, in first cylinder #1 will be described below.
Each intake valve 3 is slidably held on a cylinder head 1 via a valve guide 4 and is biased toward a closure direction by means of each valve spring 12 elastically contacted between each spring retainer 3 b disposed adjacently to each stem end 3 a and an upper surface of the inner part of cylinder head 1.
Drive shaft 5 has swing cam 7 which is rotatably supported on a plurality of bearing sections 13 mounted on an upper end section of cylinder head 1 and to which a rotational force of an engine crankshaft is transmitted via a timing belt disposed on one end section of drive shaft 5. A drive cam 5 a per cylinder is disposed on the outer periphery of drive shaft 5 has an axial center X eccentrically disposed in a radial direction from an axial center Y of drive shaft 5 and has a cam profile on an outer periphery thereof formed substantially in an ordinary circular shape.
A recessed lower surface of one end section 6 a of each swing arm 6 is contacted on a stem end 3 a of each intake valve 3 and a lower surface recess section 6 c of the other end section 6 b is contacted on each of first and second hydraulic pressure rush adjusters 10 a, 10 b and a roller 14 is rotatably housed in a housing hole formed on a center section of each swing arm 6 via a roller shaft 14 a.
Each swing cam 7 is, as shown in FIG. 1, integrally installed at both ends of a cylindrical camshaft 7 a and a cam surface 7 b including a base circle surface, a ramp surface, and a lift surface is formed on a lower surface of each swing arm 7 and the base circle surface, the ramp surface, and the lift surface are rollably contacted on an upper surface of roller 14 of each swing arm 6 in accordance with a swing position of swing cam 9.
Camshaft 7 a has a journal section which is formed on an axially substantial center position of an outer peripheral surface of cam shaft 7 a and is rotatably supported on journal sections 13 with a minute clearance and has an inner peripheral surface rotatably supporting an outer peripheral surface of drive shaft 5
Transmission mechanism 8 includes a rocker arm 15 arranged on an upper side of drive axle 5; a link arm 16 which links between one end section 15 a of rocker arm 15 and drive cam 5 a; and a link rod 17 which links between the other end section 15 b of rocker arm 15 and one swing cam 7.
Rocker arm 15 has: a cylindrical base section at a center section of rocker arm 15 which is rotatably supported on a control cam, as will be described later; one end section 15 a rotatably linked to link arm 16 by means of a pin 18; and the other end section 15 b rotatably linked to an upper end section of link rod 17 via a pin 19.
Link arm 16 has a fitting hole 16 a provided on a center position of an annular base section of link arm 16 to which a cam main body of drive cam 5 a is rotatably fitted and has a projection end linked to one end section 15 a of rocker arm by means of pin 18.
It should be noted that a lower end section of link rod 17 is rotatably linked to a cam nose section of swing cam 7 via pin 20.
It should also be noted that an adjuster mechanism 23 which minutely adjusts the lift quantity of each intake valve 3 during an assembly of structural parts is interposed between the other end 15 b of rocker arm 15 and the upper end of link rod 17.
On control mechanism 9, a control shaft 21 rotatably supported on the same journal section at the upper position of drive shaft 5 and a control cam 22 which is provided on the outer periphery of control shaft 21, is slidably fitted into a supporting hole pf rocker arm 15, and provides a swing fulcrum of rocker arm 15 are fixed onto control mechanism 9.
Control shaft 9 is disposed in parallel to drive shaft 5 in the engine forward-or-backward direction and rotatably controls actuator 50 shown in FIG. 6. On the other hand, control cam 22 has an axial center position eccentrically deviated from the axial center of control shaft 21 by a predetermined distance.
Actuator 50, as shown in FIG. 6, includes: an electrically driven motor 51 fixed to one end section of a housing (not shown); and a ball screw mechanism 52 as a speed reduction mechanism which disposed in the inside of the housing and transmits a rotational drive force of electrically driven motor 51 to control shaft 21.
Electrically driven motor 51 is constituted by a DC motor of a proportional type and is controlled to be normally or reversely rotated by a control signal from a control unit 53 as will be described later.
Each of first and second hydraulic pressure rush adjusters 10 a, 10 b, as shown in FIGS. 1 through 5, includes: a bottomed cylindrical body 24 slidably held in an upward-or-downward direction within a retaining hole is of cylinder head 1; a plunger 27 slidably housed within body 24 in the upward-or-downward direction, having a partitioning wall 25 integrally provided at a lower section of plunger 27, and having an inner section constituting a reservoir chamber 26; a high pressure chamber 28 communicated with reservoir chamber 26 via a communication hole 25 a penetrated through partitioning wall 25; and a check valve 29 disposed in the inner section of high pressure chamber 28 to allow the working oil within reservoir chamber 26 only in the direction to high pressure chamber 28. In addition, an exhaust hole 1 b through which working oil reserved within retaining hole 1 a is exhausted externally is formed.
Body 24 has an outer peripheral surface on which a circular first recess groove 24 a is formed and a first passage hole 31 radially formed in the inner section of cylinder head 1. A downstream end of first passage hole 31 communicates oil passage 30 opened to first recess groove 24 a with the inner part of body 24.
In addition, body 24 at first hydraulic pressure rush adjuster 10 a side has a bottom section, as shown in FIGS. 4A and 4B, extended toward a more downward direction than body 24 at second hydraulic pressure rush adjuster 10 b and cylindrically formed.
Oil passage 30 is communicated with a main oil gallery 31 a for a lubricating oil supply formed within cylinder head 1 and lubricating oil is supplied under pressure into main oil galley 31 a from an oil pump 54 shown in FIG. 6.
A cylindrical second recess groove 27 a is formed on an outer peripheral surface at a substantially center section in the axial direction of plunger 27 and a second passage hole 32 which communicates first passage hole 31 with reservoir chamber 26 is penetrated through a peripheral wall of second recess groove 27 a along the radial direction. In addition, a tip surface of a tip head section 27 b is formed in a spherical (surface) shape in order to secure a favorable sliding characteristic on a lower surface recess section 6 c of the other end section 6 b of corresponding one of swing arms 6, 6.
It should be noted that a maximum projection quantity of plunger 27 is limited by means of a stopper member 33 fitted and fixed onto an upper end section of body 24.
An axial length of second recess groove 27 a is relatively largely formed so that first passage hole 31 and second passage hole 32 are, at all times, communicated with each other regardless of whether any upward-or-downward sliding position is taken by plunger 27 with respect to body 24.
Check valve 29 includes: a check ball 29 a which opens or closes a lower section opening edge of communication hole 25 a; a cup-shaped retainer 29 c which retains a first coil spring 29 b; and a second coil spring 29 d which biases whole plunger 27 while biasing retainer 29 c in the direction of partitioning wall 25.
Then, at a base circle interval of swing cam 7, high pressure chamber 28 provides a low pressure along with an advance movement of plunger 27 (upward movement) by a biasing force through second coil spring 29 d. At this time, a working oil supplied from into retaining hole is caused to flow into reservoir chamber 26 via second recess groove 27 a and second passage hole 32, presses check ball 29 a to open against the spring force of first coil spring 29 b so that the working oil is caused to flow within high pressure chamber 28.
Plunger 27, at all times, pushes other end section 6 b of swing arm 6, a gap among swing cam 7, one end section 6 a of swing arm 6, and stem end 3 a of each intake valve 3 is adjusted to zero-rush via a contact between roller 14 and swing cam 7.
Then, at the lift interval of swing cam 7, downward load is acted upon plunger 27. Hence, the hydraulic pressure within high pressure chamber 28 is raised and oil within high pressure chamber 28 leaks out from a gap between plunger 27 and body 24 so that plunger 27 is slightly dropped (leak down).
At the next (again) base circle interval of swing cam 7, as described above, the gap between each part is adjusted to provide a zero rush due to an advance movement (upward motion) of plunger 27 by means of a biasing force of second coil spring 29 d.
The rush adjustment function as described above is provided in both of first and second hydraulic pressure rush adjusters 10 a, 10 b.
Valve stop mechanism 11 is, as shown in FIGS. 4A and 4B, disposed only at first hydraulic pressure rush adjuster 10 a. Valve stop mechanism 11 includes: a cylindrical sliding purpose hole 34 continuously formed at a bottom side of retaining hole 1 a; a lost motion spring 35 which biases first hydraulic pressure rush adjuster 10 a toward the upward direction; and a limitation mechanism 36 which limits a lost motion of first hydraulic pressure rush adjuster 10 a.
Sliding purpose hole 34 has an inner diameter set at the same length as the inner diameter of retaining hole 1 a and is arranged such that body 24 is slidably and continuously held in the upward-or-downward direction from retaining hole 1 a.
Lost motion spring 35 is formed of a coil spring and biases the bottom surface of body 24 in the upward direction so that tip end section 27 a of plunger 27 is elastically contacted on the lower surface of the other end section 6 b of swing arm 6.
In addition, a maximum upward movement position of body 24 is limited by means of stopper pin 17 inserted and arranged through an inside of cylinder head 1. That is to say, stopper pin 37 is disposed in the axial right angle direction toward body 24 within cylinder head 1 and tip section 37 a of stopper pin 37 is slidably disposed in the axial right angle direction within first recess groove 24 a. Then, a maximum upward sliding position of body 24 is limited by a contact of tip section 37 a along with the upward movement of body 24 on the lower end edge of first recess groove 24 a.
Therefore, first hydraulic pressure rush adjuster 10 a carries out the lost motion by stroking in the upward-or-downward direction between retaining hole 1 a and sliding purpose hole 34 via a spring force of lost motion spring 35 along with the swing motion of swing arm 6 so that a function of swing arm as the swing fulcrum is lost and the open-and-closure operations of first intake valve 3 is stopped.
Limitation mechanism 36 mainly includes: a movement purpose hole 38 penetrated in the inner diameter direction of bottom section 24 b of body 24; a limitation purpose hole 39 formed in the axial right angle direction to retaining hole is of cylinder head 1; a retainer 40 fixed to one end side of the internal of movement purpose hole 38; a limitation pin 41 slidably disposed in the inside of movement purpose hole 38 and movable from movement purpose hole 38 to limitation purpose hole 39; and a return spring 42 elastically disposed between a rear end of limitation pin 41 and retainer 40 to bias limitation pin 41 toward limitation purpose hole 39.
Limitation purpose hole 39 is made coincident with movement purpose hole 38 from the axial direction when body 34 is limited to the maximum upward direction by means of stopper pin 37. An inner diameter of limitation purpose hole 39 is formed at substantially the same as movement purpose hole 38 and a signal hydraulic pressure is introduced from an oil passage hole 43 formed within cylinder head 1 at one end side of limitation purpose hole 39.
It should be noted that the limitation of body 24 in the rotational direction can easily be realized by slightly increasing a jump out quantity (projection quantity) of stopper pin 37, providing an elongated slit within first recess groove 24 a of body 24, and by engaging the elongated slit with the tip of stopper pin 37. Or alternatively, a separate rotation limitation member may be equipped between cylinder head 1 and body 24.
Retainer 40 is formed in a bottomed cylindrical shape and has a breathing hole 40 a penetrated through a bottom section of retainer 40 to secure a smooth movement of limitation pin 41. In addition, an axial length of retainer 40 is set to a length such that a rear end of limitation pin 41 is contacted on a tip end edge of retainer 40 and a further retractable movement of limitation pin 41 is limited at a time point at which limitation pin 41 is completely housed within movement purpose hole 38, as shown in FIG. 4B.
Limitation pin 41 is formed in an approximately hollow cylindrical shape, an outer diameter is slightly smaller than movement purpose hole 38 and an inner diameter of a limitation purpose hole 39 so that a smooth slidability (a smooth sliding motion) is secured. In addition, this limitation pin 41 receives the hydraulic pressure supplied from oil passage hole 43 to limitation purpose hole 39 by a pressure receiving surface of tip section 41 a so that a retractable movement against the spring force of return spring 42 is carried out, the tip section of limitation pin 41 is dropped out from limitation purpose hole 39 and housed within movement purpose hole 38 and the limitation is released.
The hydraulic pressure supplied under pressure from oil pump 54 is supplied to oil passage hole 43 (limitation purpose hole 39) 6 via an electromagnetic switching valve 55 as a signal hydraulic pressure as shown in FIG. 6.
Electromagnetic switching valve 55 switches a spool valve slidably disposed in the inside of the valve body (not shown) at two stages in an on-or-off manner by means of an electromagnetic force of a solenoid thereof and a spring force of a coil spring thereof. A control current is supplied or not supplied from a control unit 53 which controls a drive of electrically driven motor 51 to the solenoid so that a switching control such that a pump draining passage and oil passage hole 43 are communicated or the pump draining passage is closed to communicate oil passage hole 43 with a drain passage 44 is carried out. Thus, the signal hydraulic pressure is controlled in two stages of large and small.
Control unit 53 detects an engine driving state (or an engine driving condition) on a basis of information signals from various sensors such as a crank sensor, an airflow meter, coolant temperature sensor, a throttle valve angle sensor, and so forth and controls a rotational position of control axis 21 by drivingly controlling electrically driven motor 51 in response to the information signal from a rotational position sensor (not shown) which detects the present rotational position of control shaft 21. Thus, lift quantities and working angles of respective intake valves 3, 3 are varied.
In addition, this control unit 53 includes a valve stop inhibit circuit which is valve stop inhibiting means for inhibiting the lost motion of valve stop mechanism 11 via electromagnetic switching valve 55. This valve stop inhibit circuit performs such a control that oil passage hole 43 is communicated with drain passage 44 via electromagnetic switching valve 55 on a basis of a rotation angle θ of control shaft 21. This causes limitation pin 41 to move toward a direction of limitation purpose hole 39 by means of the spring force of return spring 42 and tip end section 41 a of limitation pin 41 is engageably inserted into limitation purpose hole 39. Thus, body 24 (first hydraulic pressure rush adjuster 10 a) is locked to cylinder head 1 and the lost motion of first hydraulic pressure rush adjuster 10 a is inhibited.
The rotation angle position of control shaft 21 to control the drive of electromagnetic switching valve 55 by means of the valve stop inhibit circuit can arbitrarily be set in accordance with an engine driving state or so forth. In the first embodiment, the rotation angle position is θ3 for which control shaft 21 provides in such a way that the lift quantity becomes L3 shown in FIG. 12.
In other words, in a case where that the lift quantity of first intake valve 3 exceeds L3, this lift quantity having a proportional relationship to the stroke quantity of the lost motion of first hydraulic pressure rush adjuster 10 a, is detected by rotational angle θ of control shaft 21 and, at this time point, intake valves 3, 3 is in the valve stop state, the control current flowing toward electromagnetic switching valve 55 is interrupted to communicate oil passage hole 43 with drain passage 44 so that limitation pin 41 forcibly locks first hydraulic pressure rush adjuster 10 a. It should be noted that if, at a time point at which the lift quantity of first intake valve 3 exceeds L3, limitation pin 41 is already locked and the state is in a two-valve lift operation state, this state is maintained.

[Operation of Variably Valve Operated System]

Hereinafter, an operation of the variably operated valve system in the above-described first embodiment will be explained.
For example, when the engine is driven in a range from an idling state to a low rotation region, electrically driven motor 51 is rotationally driven by means of a control current outputted from control unit 53. This rotational torque is transmitted to control shaft 21 via ball screw mechanism 52. When this control shaft 21 is rotationally driven in one direction, as shown in FIGS. 7A, 7B, 8A, and 8B, control cam 22 is uni-directionally pivoted so that an axial center of control cam 22 is rotated around the axial center of control shaft 21 at the same radius and a wall thickness section of control cam 22 is spaced apart from drive shaft 5 and moved toward a right upper direction as shown in FIGS. 7A through 8B. Thus, the other end section 15 b of rocker arm 15 and a pivot point (linkage pin 19) of link rod 17 are moved in the upper direction with respect to drive shaft 5 and, thus, the cam nose section side of each swing cam 7 is forcibly lifted via link rod 17.
Thus, when drive cam 5 a is rotated and pushes up one end section 15 a of rocker arm 15 via link arm 16, the lift quantity of rocker arm 15 is transmitted to each swing cam 7 and each swing arm 6 via link rod 17 so that each intake valve 3 is open against the basing spring of valve spring 12 and the lift quantity of each intake valve 3 becomes sufficiently small as L1 through L3 shown in FIG. 12.
For example, in a case where the engine driving state is transferred from the low engine speed range to a middle or high engine speed range, the control current causes electrically driven motor 51 to reversely be rotated so that ball screw mechanism 52 is rotated in the same direction. As shown in FIGS. 10A, 10B, 11A, and 11B, control shaft 21 causes control cam 22 to be rotated in the other direction and the axial center of control cam 22 is moved toward the downward direction.
Thus, whole rocker arm 15 is, at this time, moved toward the direction of drive axle 5 so that the other end section 15 b of rocker arm 15 presses a cam nose section of swing cam 7 in the downward direction via link rod 17 so that a whole of each swing cam 7 is pivoted in a counterclockwise direction by a predetermined quantity from a position shown in FIGS. 7A through 8B. Hence, as shown in FIGS. 10A through 11B, a contact position of cam surface 7 b of each swing cam 7 to an outer peripheral surface of roller 14 is moved toward the cam nose side (lift section side).
Therefore, when drive cam 5 a is rotated at a time of valve open of intake valve 3 so as to push up one end section 15 a of rocker arm 15 via link arm 16, each intake valve 3 is opened against the spring force of each valve spring 12 via each swing arm 6. The valve lift quantity of each intake valve 3 is continuously varied up to a maximum of L7 shown in FIG. 12 and becomes large from L4 to L7 as the rotation of drive cam 5 a is increased. Therefore, an intake air charging efficiency is improved and an improvement in the output can be achieved.

[Operation of Valve Stop Mechanism]

Then, in a case where the lift quantity of each intake valve 3, 3 in the engine driving state of the idling state and the low engine speed area becomes a small lift quantity range of L1 through L3 shown in FIG. 12, especially, in a particular engine driving state in which an improvement of fuel economy is achieved, the control current is outputted to electromagnetic switching valve 55 from control unit 53 and, thus, a large discharge hydraulic pressure is introduced from oil pump 54 to limitation purpose hole 39 via oil passage hole 43 as the signal hydraulic pressure.
Limitation pin 41 on which this large signal hydraulic pressure is moved toward the backward direction (or retracted) against the spring force of return spring 42 so that tip end section 41 a of return spring 41 is slipped out from limitation purpose hole 39 so that a lock of first hydraulic pressure rush adjuster 10 a to cylinder head 1 is released.
Hence, first hydraulic pressure rush adjuster 10 a, as shown in FIG. 4B, the whole of first rush adjuster 10 a can make the lost motion and the upward-or-downward movement within retaining hole 1 a and sliding purpose hole 34 is repeated via the spring force of lost motion spring 35 and first hydraulic pressure rush adjuster 10 a enters the lost motion state. Hence, first intake valve 3 is in the valve closure state (valve stop state).
That is to say, until the valve is in the valve stop state, swing cam 7 is varied from a zero-lift position (valve closure) shown in FIG. 7A to a maximum valve open lift position shown in FIG. 7B.
Suppose that each intake valve 3, 3 is opened by a lift quantity of L2. Even if, in the valve stop state, swing cam 7 is maximally swung, first hydraulic pressure rush adjuster 10 a makes the lost motion by stroke quantity M2 shown in FIG. 7B and, actually, the state of the engine is transferred to the valve stop state in which not valve lift is carried out. An instantaneous opening angle formed between first swing arm 6 and first hydraulic pressure rush adjuster 10 a is α (refer to FIG. 7B). This α indicates β7 shown in FIG. 10B when swing cam 7 is placed at position indicating a peak lift (maximum lift control) but this opening angle does not give an excessive opening angle.
Hence, even if swing cam 7 indicates the peak lift (maximum valve open operation), a smooth valve stop operation can be obtained.
On the other hand, second hydraulic pressure rush adjuster 10 b functions as an ordinary swing fulcrum to second swing arm 6 as shown in FIGS. 8A and 8B. Hence, second intake valve 3 still carries out the open-and-closure operations at lift quantity of L2. Thus, the intake air swirl is reinforced and the improvements in the fuel economy and combustion can be achieved.
Next, consider such a case where the engine speed is, furthermore, increased, the required torque is increased, the engine valve is again transferred to the two-valve lift operation state, and the lift quantity is increased so that control shaft 21 is rotated in the clockwise direction to provide θ3 of rotation angle, namely, a case where the lift quantities of both of intake valves 3, 3 indicate L3 shown in FIG. 12. Suppose a case when the requirement (request) of fuel economy is again increased from this state and the engine driving state is transferred to the valve stop state. In this case shown in FIG. 9, the opening angle formed between first swing arm 6 and first hydraulic pressure rush adjuster 10 a is considerable open as α3.
Therefore, a contact of tip end head section 27 b of first hydraulic pressure rush adjuster 10 a on lower surface recess section 6 c of first hydraulic pressure rush adjuster 10 a becomes uniform.
That is to say, tip end head section 27 b of first hydraulic pressure rush adjuster 10 a, ordinarily, stably holds lower surface recess section 6 c of the other end section 6 b of first swing arm 6 while the contact of tip end head section 27 b on a spherical surface section at the lost motion 4 side and the contact of tip end head section 27 b on the spherical surface section at an anti-roller 14 are balanced. However, when opening angle of θ3 becomes wide (large), the contact on the spherical surface section of roller 14 side is moved toward the upward direction and the contact section of the spherical surface section at the anti-roller side is moved in the downward direction.
If a weight (load) from roller 14 acted upon first swing arm 6, the load received at the contact section of the spherical surface section of roller side 14 moved in the upward direction is extremely increased. Thus, the balance is broken and it easily becomes a local contact.
Consequently, in addition to the contact section to move in the upward direction, a positional gap on the contact section due to the load of roller 14 is easily generated. Hence, a phenomenon such that first swing arm 6 is deviated toward an anti-valve side (toward away from the intake valves 3, 3) and, dependent upon the situation, there is a possibility that lower surface recess section 6 c of the other end section 6 b of first swing arm 6 becomes out of place from tip end head section 27 b. However, at the level of this α3, the fact that lower surface recess section 6 c of the other end section 6 b which becomes out of place (deviated from) tip head section 27 b falls one way or another an allowable range but if it exceeds α3, there is increased possibility that actually an deviation-out phenomenon described above is developed.
Therefore, at a time point at which the lift quantity is in excess of lift quantity of L3, the valve stop inhibit circuit of control unit 53 interrupts the cut off of the control current to electromagnetic switching valve 55 so that oil passage hole 43 is communicated with drain passage 44 so that the hydraulic pressure within limitation purpose hole 39 and oil passage hole 43 is exhausted toward an inside of an oil pan 45 (refer to FIG. 6) so that the pressure within limitation purpose hole 39 is under a low pressure state.
Thus, limitation pin 41 is moved in the direction of limitation purpose hole 39 by means of the spring force of return spring 42 so that first hydraulic pressure rush adjuster 10 a at the base circle range of swing cam 7 to move in the upward direction, as shown in FIG. 4A, and stopper pin 37 and a further upward movement is limited by stopper pin 27. At a time point at which movement purpose hole 38 and limitation purpose hole 39 are made coincident with each other, tip end section 41 a of limitation pin 41 is inserted and engaged with limitation purpose hole 39 so that first hydraulic pressure rush adjuster 10 a is locked to cylinder head 1.
Hence, the lost motion of first hydraulic pressure rush adjuster 10 a is limited at this time point.
A dot-and-dash line shown in FIG. 13 represents a correlation between rotation angle θ of control shaft 21 and a lost motion quantity (stroke length M) in a case where the valve is stopped and a broken line shown in FIG. 14 represents a correlation between rotation angle θ of control shaft 21 and lift quantity L of intake valve 3 in a case where the lift operation is carried out.
That is to say, as shown in FIG. 13, the correlation is provided between lost motion quantity (stroke, length M) in a case where the valve is stopped and rotation angle θ of control shaft 21. As shown in FIG. 14, the correlation is provided between rotation angle θ of control shaft 21 and lift quantity L of intake valve 3. Hence, control unit 53 forcibly interrupts the power supply to electromagnetic switching valve 55 at a time point at which rotation angle of control shaft 21 exceeds θ3 on a basis of the information signal from the rotation angle sensor.
The above-described series of operation is depicted as a solid line of each of FIGS. 13 and 14. At a time point at which the lost motion quantity exceeds M3, namely, at a time point at which the lift quantity of first intake valve 3 exceeds L3, lost motion operation of first hydraulic pressure rush adjuster 10 a is mechanically inhibited. Thus, together with second intake valve 3, first intake valve 3 is open-and-closure operated and the engine driving by means of both intake valves 3, 3 (two-valve driving) is carried out.
Hence, at a time point at which the lift quantity of intake valve 3 exceeds L3, the lost motion of first hydraulic pressure rush adjuster 10 a becomes eliminated (nullified or is not carried out). Hence, the uniform and local contact between lower surface recess section 6 c at the other end section 6 b of first swing arm 6 and tip end section 27 b of plunger 27 of first hydraulic pressure rush adjuster 10 a and tip end head section 27 b of plunger 27 of first hydraulic pressure rush adjuster 10 a can be avoided. Therefore, lower surface recess section 6 c of first swing arm 6 is, for example, not dropped out from tip head section 27 b of plunger 27 so that the smooth operation state can, at all times, be obtained.
In addition, in a case where the engine speed is, for example, further increased and fall in a region in which the lost motion quantity of first hydraulic pressure rush adjuster 10 a further exceeds M3 (considerably larger than M3) (in a case where the lift quantity of each of intake valves 3, 3 exceeds further L3 shown in FIG. 12, the non-power supply state from control unit 53 to electromagnetic switching valve 55 is continued so that the signal hydraulic pressure is not introduced into limitation purpose hole 39. Hence, the state in which no lost motion state of first hydraulic pressure rush adjuster 10 a is continued so that first hydraulic pressure rush adjuster 10 a exhibits the function as the swing fulcrum in the same manner as second hydraulic pressure rush adjuster 10 b.
In other words, until the lost motion quantity of first hydraulic pressure rush adjuster 10 a indicates M3 (rotation angle θ of control shaft 21 is θ3), first hydraulic pressure rush adjuster 10 a is allowed to be made the lost motion but, at the time pint at which the lost motion quantity exceeds M3, the valve stop is inhibited, the lost motion is limited and fixed, and first hydraulic pressure rush adjuster 10 a functions as the ordinary swing fulcrum.
FIG. 15 shows a specific control flowchart executed by the valve stop inhibit circuit of control unit 53.
That is to say, at a step S1, control unit 53 reads the present engine driving state on a basis of the information signal from the various kinds of sensors described above, at a step S2, control unit 53 reads target lift quantities of first and second intake valves 3, 3 in accordance with the engine driving state from a preset control map, for example, a control map of the engine speed and an engine load. Suppose now that the target lift quantity of second intake valve 3 is L2, the target lift quantity of first intake valve 3 is zero (the valve stop), and the present lift quantities of first and second intake valves 3, 3 are the proximity of L3.
At a step S3, control unit 53 determines whether target lift quantity of first intake valve 3 of first hydraulic pressure rush adjuster 10 a side is zero or not. If zero, namely, the determination of the valve stop is made, the routine goes to a step S4. At this step S4, control unit 53 determines whether the actual lift quantities of respective intake valves 3, 3 are equal to or below L3 from rotation angle θ derived from the rotation angle sensor of control shaft 21 and, in a case where the actual lift quantities of respective intake valves 3, 3 are determined to be equal to or below L3, the routine goes to a step S5.
At this step S5, control unit 53 carries out an output of an on signal or continuation of the on signal to electromagnetic switching valve 55. In other words, since the lift quantity is equal to or below L3, the signal hydraulic pressure is supplied to limitation purpose hole 39 and the valve state is transferred to the lost motion of first hydraulic pressure rush adjuster 10 a or the lost motion is continued so that the valve is in one valve stop state. (in the present case, the on signal is outputted to electromagnetic switching valve 55 and the valve state is transferred to the lost motion state.)
Thereafter, at a step S6, electrically driven motor 51 is controlled toward the target lift quantity and a process in which rotation angle θ of control shaft 21 is controlled is carried out and the routine returns.
If, at step S3, target lift quantity of first intake valve 3 is determined not to be zero, or if, at step S4, the actual lift quantities of respective intake valves 3, 3 are determined to exceed L3, the routine goes to a step S7.
At step S7, control unit 53 carries out the process such that the on signal is outputted to electromagnetic switching valve 55 or the on signal is continued (at the present case, the on signal is continued).
In details, in a case where the actual lift quantities of respective intake valves 3, 3 exceed L3, the supply of the signal hydraulic pressure within limitation purpose hole 39 is interrupted or the interruption is continued and limitation pin 41 is inserted and engaged with limitation purpose hole 39 by means of the spring force of return spring 42 or the insertion and engagement are continued so that first hydraulic pressure rush adjuster 10 a is locked to cylinder head 1 or this lock is continued. Thus, for first intake valve 3, the valve stop state is inhibited and the open-and-closure operations (a lift operation) by means of first swing arm 6 is carried out with first hydraulic pressure rush adjuster 10 a as the swing fulcrum. Hence, as described above, the excessive lost motion of first hydraulic pressure rush adjuster 10 a is inhibited so that the irregular motion due to the local contact of lower surface recess section 6 c of the other end section 6 b of first swing arm on the tip head section 27 b or so forth can be avoided and a smooth operation state can be obtained.
It should be noted that, although, in the first embodiment, a condition to forcibly inhibit the lost motion of first hydraulic pressure rush adjuster 10 a is when the lost motion exceeds M3 (when the target lift of first intake valve 3 exceeds L3). However, it is natural that, on the control map, the valve stop transition of first intake valve 3 is not carried out even in the case of the lower lift quantity than L3 and both of first and second intake valves 3, 3 may be under a lift operation control.
For example, in a case where the engine is in a cold start state in which a combustion torque overcoming an engine friction is required even in a case of small working angle (for example, small lift quantity L2), it is preferable to perform the open-and-closure operations for two intake valves 3, 3 without the lost motion. Hence, it is possible to inhibit the lost motion even if the list quantity is below L3.
The time point at which lift quantity L3 or lost motion M3 is exceeded merely a reference to forcibly inhibit the lost motion. However, the valve stop transition (transition to lost motion operation) at the smaller lift quantity than L3 in the control map is not carried out depending upon the driving condition but may be carried out at the two-valve operation.
On the other hand, the setting of values of L3 or M3 which provide a criterion to forcibly inhibit the lost motion described above may be modified. For example, in the engine high speed range, a slight separation between each component of the variably operated valve mechanism is apt to occur. Hence, the irregular behavior such as the positional gap or drop out of swing arm 6 with respect to the pivot may furthermore be generated. Hence, the values of L3 and M3 which provide the criterion described above may further be set to smaller.
Hereinafter, merits of the first embodiment to the technique of the previously proposed variably operated valve system described in the Japanese Patent Application First Publication No. 2010-007636 described in the BACKGROUND OF THE INVENTION will supplementary be described.
That is to say, in the previously proposed variably operated valve system, as shown in FIG. 4 of the above-described Japanese Patent Application first Publication, body (41) is housed and fixed by means of cylinder head (11) and plunger itself (42) is in the lost motion so that the valve stop is carried out.
Specifically, the lost motion quantity of plunger (42) is increased by reducing the working oil acted upon the body to perform the valve stop. In a midway through the reduction in the working hydraulic pressure, the lost motion quantity is not instantaneously increased and an instant at which an unstable intermediate lost motion quantity is provided is present. In this case, a lift curve is unstable, the motion of the variably operated valve system is irregular, and it is not unprofitable for the suppression of the irregular motion, and there is an anxiety of an occurrence of a phenomenon such that the engine performance becomes unstable.
Whereas, in the first embodiment, either of the lift operation state or the valve stop state (lost motion state) is selectively selected according to the engagement or the engagement release by means of the limitation pin and no intermediate lift curve is present. Hence, there is no anxiety described above.
In addition, the body (41) in the previously proposed variable operated valve system is fixed to the cylinder head (11) and the plunger itself (42) is made the lost motion. Hence, return spring (45) to push up the plunger (42) serves as a spring function (a minute stroke) to perform a rush adjustment of the hydraulic pressure rush adjuster and a spring function (a large stroke) for the plunger to make the lost motion. Hence, if the importance is placed on the rush adjustment and the spring load is reversely increased, the rush adjustment is not well performed an a pump up phenomenon such that the plunger (42) excessively rises is developed.
It should be noted that reference numerals described in respective brackets denote those described in the above-identified Japanese Patent Application First Publication.
Whereas, in the first embodiment, spring 29 d to make the rush adjustment and spring 35 for the lost motion can separately and independently be set. Hence, there is no anxiety described above.
In addition, in the previously proposed variable operated valve system, when the working oil pressure (the working hydraulic pressure) from the oil pump is low, as described above, the valve state is in the valve stop state. Hence, at the time of an engine cranking or engine start at which the hydraulic pressure from the oil pump is not expected, the engine valve is in the valve stop state. Therefore, an intake air quantity at the time of the engine start becomes insufficient and a startability becomes worsened due to an insufficient torque.
Whereas, in the first embodiment, as described before, in a case where the signal hydraulic pressure is not acted, the valve state is the two-valve operations. Thus, there is no anxiety of worsening of the startability. A favorable startability can be obtained.

Second Embodiment

FIG. 16 shows a second preferred embodiment of the variably operated valve system in which, as the valve stop inhibiting means, a mechanical structure is added.
A pair of oil holes 43 a, 43 b penetrated in the axial right angle direction through control shaft 21 of bearing section 13 which journals (serves as a bearing for) control shaft 21 and are formed on part of oil passage hole 43. On the other hand, a communication hole 46 appropriately communicated with respective oil holes 43 a, 43 b is penetrated and formed in the inside of control shaft 21 in the axial right angle direction. Arc shaped oil grooves 46 a, 46 b are formed on both end sections of this communication hole 46.
This communication hole 46 serves to communicate both of oil holes 43 a, 43 b of oil passage hole 43 via oil grooves 46 a, 46 b as denoted by a solid line in FIG. 16 in a case where the rotation angle position of control shaft 21 is θ2 (L2). In a case where the rotation angle position of control shaft 21 indicates the angle position exceeding θ3 (L3), as shown in the broken line of FIG. 16, both of oil grooves 46 a, 46 b provide the positional gap to interrupt the communication between both oil holes 43 a, 43 b.
Hence, in a case where the engine rotation (the engine speed) falls in the low speed range and the revolution angle of control shaft 21 falls in a range of θ1 through θ3 (L1 through L3 in each intake valve 3, 3), as denoted by the solid line, oil passage 43 and communication hole 46 are communicated and the signal hydraulic pressure is supplied within limitation purpose hole 39. However, when the rotation angle gives a value exceeding θ3 (L3), as denoted by a broken line in FIG. 16, closes both oil holes 43 a, 43 b by the outer peripheral surface of control shaft 21 to interrupt mechanically the communication with oil passage hole 43.
Therefore, the same effect and action as the first embodiment can be obtained. Especially, in this embodiment, if control unit 53 fails, electromagnetic switching valve 55 is abnormally operated, and an abnormal signal hydraulic pressure is supplied to the upstream side of oil passage hole 43, the passage is interrupted by means of control shaft 21. Hence, if rotation angle (lift quantity) exceeds L3 (L3), the lost motion of first hydraulic pressure rush adjuster 10 a is limited and the valve stop is inhibited.

Third Embodiment

FIG. 17 shows a third preferred embodiment of the variably operated valve system according to the present invention. A bottomed cylindrical holding member 47 which is arranged for first hydraulic pressure rush adjuster 10 a to be slidable in the upward-or-downward direction is press-fitted within retaining hole 1 a of cylinder head 1 at first hydraulic pressure rush adjuster 10 a side and a projection section 47 b is integrally disposed at a center position of an inner surface of bottom wall 47 a of holding member 47.
A height of projection section 47 b is set such that the bottom surface of body 24 of first hydraulic pressure rush adjuster 10 a is contacted on projection section 47 b to limit the further lost motion stroke, in a case where the lost motion quantity shown in FIG. 13, in a slight degree, exceeds M3 when the first hydraulic pressure rush adjuster 10 a is made the lost motion.
It should be noted that a communication passage 47 c which communicates first recess groove 24 a with oil passage 30 and a breathing hole 47 d at a bottom wall 47 a side of holding member 47 are formed.
As described above, in the third embodiment, the lost motion stroke of first hydraulic pressure rush adjuster 10 a is mechanically limited so that the excessive lost motion can be suppressed. Hence, the local contact of first swing arm 6 on tip head section 27 b of plunger 27 can more accurately be avoided.
In addition, this holding member 47 is made of an iron series material different from cylinder head 1 (ordinarily, aluminum material). Thus, a wear and abrasion resistance such as limitation purpose hole 39 which slides the limitation pin, sliding purpose hole 34 which slides body 24 of first hydraulic pressure rush adjuster 10 a can be improved. As described above, holding member 47 except projection member 47 b can be applied to the first embodiment.

Fourth Embodiment

FIG. 18 shows a fourth preferred embodiment of the variably operated valve system according to the present invention.
A basic structure is the same as the first embodiment. The whole lift quantity variable mechanism including respective swing arms 7, 7 is arranged like a reflecting mirror in a reverse direction as each embodiment.
Thus, swing cam 7 swings and lifts in a clockwise direction shown in FIG. 18 so that swing arms 6, 6 and intake valves 3, 3 are valve open lifted.
In this embodiment, as compared with the structure shown in FIG. 9 in the first embodiment, the swing lift direction of swing cam 7 is the same as the lost motion direction of first hydraulic pressure rush adjuster 10 a. Hence, the cam nose section of swing cam 7 and swing arm 6 become difficult to be interfered during the operation.
In addition, a contact point between swing cam 7 and roller 14 of swing arm 6 comes near to first hydraulic pressure rush adjuster 10 a side and just presses under pressure the vicinity to the center of swing arm 6. A contact characteristic between first hydraulic pressure rush adjuster 10 a and swing arm 6 becomes favorable and becomes difficult to remove. This is because, in tip head section 27 b of first hydraulic pressure rush adjuster 10 a, the contact at the spherical surface section at the roller section side and the contact at the spherical surface section at the anti-roller section side are balanced.
As described above, in the fourth embodiment, a favorable contact state between first hydraulic pressure rush adjuster 10 a and swing arm 6 can be obtained and an interference between respective components (parts) can be suppressed.
The present invention is not limited to this structure of each of the preferred embodiments described above. For example, the present invention is applicable to a V shaped eight cylinder engine not only the V shaped six cylinder engine. The present invention is also applicable to an inline four cylinder engine corresponding to one bank of the V shaped cylinder engine, and another type engine.
In addition, one of the two swing cams 7, 7 may have different cam profiles and this achieves a combustion improvement due to a slight swirl while the suction (intake) air charging efficiency is maintained in a high engine load region in which both of the (intake) valves are operated. In addition to intake valves 3, 3 as the engine valves, the present invention is applicable to an exhaust valve side. In this alternative case, since the swirl of exhaust gas can be reinforced, an exhaust emission transformation performance at catalyst can be improved.
In addition, in each of the above-described embodiments, from among the pair of engine valves, one of the pair of engine valves is in the valve stop. However, the present invention is applicable to a case where the two (pair of) engine valves are valve stopped.
Furthermore, as the member which makes the lost motion, a member having no rush function may be applied in addition to the respective hydraulic pressure rush adjusters.
In addition, it is possible to dispose the valve stop mechanism on first swing arm 6. In this case, a roller element which can displace (make the lost motion) to a main swing arm, for example, as disclosed in a Japanese Patent Application First Publication (tokuhyou) No. 2010-270633 published on Dec. 2, 2010 (which corresponds to a U.S. Pat. No. 7,712,443 issued on May 11, 2010) and this roller element and main swing arm may be switched between an engagement and a non-engagement. Even in this case, an unusual posture such that the contact between the roller element and the swing cam becomes out of place or becomes interfered or a bottoming occurs at the time of the lost motion due to the excessive lost motion is suppressed and the smooth operation cam be realized.
Furthermore, the present invention is applicable to the variably operated valve system of a lift type having no hydraulic pressure rush adjuster described in a Japanese Patent Application First Publication (tokkai) No. 2010-270633 published on Dec. 2, 2010. In this case, the valve stop mechanism incorporated in a valve lifter as shown in a Japanese Patent Application First Publication (tokkai) No. showa 63-016112 published on Jan. 23, 1988 may be used. Furthermore, the present invention is applicable to a two-valve stop mechanism. It should be noted that a reference sign T shown in, for example, FIG. 9 denotes a radial clearance formed between an inner peripheral surface of cylindrical member 7 a of swing cam 7 and an outer peripheral surface of drive shaft 5.
Technical ideas graspable from the above-described embodiments except the claims 1, 2, and 3 will be described below.
(a) The variably operated valve system for the internal combustion engine as claimed in claim 1, wherein the valve stop mechanism includes: a retaining hole that holds one of the pair of fulcrum members movable; and a basing member that biases the one of the pair of fulcrum members toward a direction of a corresponding one of the pair of swing arms and the lost motion is made by moving the one of the pair of fulcrum members against the biasing member.
(b) The variably operated valve system for the internal combustion engine as claimed in claim 4, wherein the pair of fulcrum members are hydraulic pressure rush adjusters.
(c) The variably operated valve system for the internal combustion engine as claimed in claim 1, wherein the valve stop mechanism is disposed only in one of the pair of fulcrum members and, when the lost motion quantity is equal to or below the predetermined value, only one of the two engine valves corresponding to the one of the pair of fulcrum members is allowed to be stopped through the valve stop mechanism.
According to the present invention described in item (c), the one-valve stop is carried out and a swirl effect becomes large so that the improvement in fuel economy can be achieved.
(d) The variably operated valve system for the internal combustion engine as claimed in claim 1, wherein the valve stop mechanism is disposed in both of the pair of fulcrum members and, when the lost motion quantity is equal to or below the predetermined quantity, both of the two engine valves are allowed to be stopped through the valve stop mechanism.
According to the present invention described in item (d), both of the engine valves are stopped so that the corresponding engine cylinder is in the cylinder stopped state, a throttle valve of each of the remaining others of the engine cylinders which is not stopped can largely be opened, and a pumping loss can be reduced.
(e) According to the present invention described in item (c), wherein, when the valve stop inhibiting means mechanically inhibits the lost motion of the one of the pair of fulcrum members when the lost motion quantity exceeds the predetermined value.
(f) The variably operated valve system for the internal combustion engine as set forth in item (a), wherein the swing fulcrum of each of the pair of each of the pair of swing arms is formed by a spherical recess section disposed on a corresponding one of the pair of swing arms and a spherical convex section engaged with the recess section.
(g) The variably operated valve system for the internal combustion engine as set forth in item (a), wherein, when the valve stop inhibiting means mechanically inhibits the lost motion of the one of the pair of fulcrum members when the lost motion quantity exceeds the predetermined value.
(h) The variably operated valve system for the internal combustion engine as set forth in item (a), wherein the control shaft includes a control cam that is in an eccentricity state to an axial center of the control shaft, the control cam is inserted through the transmission mechanism, and the posture of the transmission mechanism is varied according to a rotation of the control cam.
(i) The variably operated valve system for the internal combustion engine as set forth in item (g), wherein a limitation of the lost motion is made by a limitation of a movement of the one of the pair of fulcrum members by means of a projection section in a movable range of the one of the pair of fulcrum members.
(j) The variably operated valve system for the internal combustion engine as set forth in item (g), wherein the valve stop inhibit means is constituted by a movement purpose hole and a limitation purpose hole, both of the movement purpose and limitation purpose holes being disposed in the one of the pair of fulcrum members and in an inner wall of the retaining hole, and a limitation pin movably disposed across the movement purpose hole and the limitation purpose hole and a state in which the one of the pair of fulcrum members is in a locked state by disposing the limitation pin across the movement purpose hole and the limitation purpose hole and another state in which the limitation pin is housed in the movement purpose hole to enable the one of the pair of fulcrum members to make the lost motion are controlled.
(k) The variably operated valve system for the internal combustion engine as set forth in item (j), wherein the valve stop inhibiting means includes another biasing member disposed in the movement purpose hole of the one of the pair of fulcrum members to bias the limitation pin toward a direction of the limitation purpose hole and a hydraulic pressure circuit that supplies a hydraulic pressure to press under pressure the limitation pin in another direction of the movement purpose hole of the one of the pair of fulcrum members against the biasing force of the other biasing member.
(l) The variably operated valve system for the internal combustion engine as set forth in item (k), wherein the hydraulic pressure circuit includes: oil passage holes disposed at a bearing section of the control shaft and a communication hole penetrated through a diameter direction of the control shaft to appropriately communicate with the oil passage holes and wherein, when the control shaft is rotated through an angle equal to or below a predetermined rotation angle, both of the oil passage holes and the communication hole are communicated with each other and, when the control shaft exceeds the predetermined rotation angle, the communication between the oil passage holes and communication hole is interrupted.
(m) The variably operated valve system for the internal combustion engine as set forth in item (k), wherein, when the lost motion quantity is smaller than the predetermined value of the lost motion quantity of the one of the pair of fulcrum members, a hydraulic pressure is supplied to move the limitation pin toward the direction of the movement purpose hole against a biasing force of the other biasing member and, when the lost motion quantity exceeds the predetermined value, the biasing force of the other biasing member causes the limitation pin toward the direction of the limitation purpose hole.
(n) The variably operated valve system for the internal combustion engine as set forth in item (a), wherein an angle which is a subtraction of 90° from an angle formed between the corresponding one of the pair of swing arms and the one of the pair of fulcrum members when the lost motion quantity of the fulcrum member is at the predetermined value is smaller than an angle formed between the corresponding one of the pair of swing arms and the one of the pair of fulcrum members when the engine valves are open at a time of a maximum lift control.
According to the present invention described in item (n), a degree of worsening of a posture of the corresponding one of the pair of swing arms becomes more favorable that the worsening of the posture of the corresponding swing arm at the valve open position at the time of the maximum lift control so that the valve stop operation can more smoothly be made.
(o) The variably operated valve system for the internal combustion engine as set forth in item (a), wherein the valve stop inhibiting means includes a limitation member that limits the lost motion by biasing the one of the pair of fulcrum members toward a direction of the corresponding one of the pair of swing arms, when the lost motion quantity of the one of the pair of fulcrum members exceeds the predetermined value.
(p) The variably operated valve system for the internal combustion engine as claimed in claim 14, wherein the valve stop inhibiting means includes a valve stop inhibit circuit constituted by a control unit that interrupts the supply of the hydraulic pressure via an electromagnetic switching valve of the hydraulic pressure circuit.
(q) The variably operated valve system for the internal combustion engine as set forth in item (p), wherein the predetermined value of the lost motion quantity can arbitrarily be set according to a driving condition of the engine.
According to the present invention described in item (q), the inhibited lost motion quantity can arbitrarily be set. Thus, for example, an effective suppression of the irregular behavior at the time of the valve stop can be made at the high engine speed region (area).
This application is based on a prior Japanese Patent Application No. 2012-140574 filed in Japan on Jun. 22, 2012. The entire contents of this Japanese Patent Application No. 2012-140574 are hereby incorporated by reference. Although the invention has been described above by reference to certain embodiments of the invention, the invention is not limited to the embodiment described above. Modifications and variations of the embodiments described above will occur to those skilled in the art in light of the above teachings. The scope of the invention is defined with reference to the following claims.

Claims

What is claimed is:

1. A variably operated valve system comprising:

a drive shaft to which a rotation driving force is transmitted from an engine crankshaft and on an outer periphery of which drive cams are disposed;

two swing cams that operate two engine valves per engine cylinder to open against spring forces of respective valve springs;

a transmission mechanism that converts rotation motions of the drive cams into swing motions and transmits the converted swing motions to the respective swing cams;

a pair of swing arms interposed between the respective swing cams and the respective engine valves to operate the respective engine valves to open or close;

a pair of fulcrum members each of which provides a swing fulcrum of each of the swing arms;

a control mechanism that varies a posture of the transmission mechanism to variably control a lift quantity of each of the engine valves;

a valve stop mechanism that makes at least one of the pair of fulcrum members lost motion to stop open-and-closure drives for one of the two engine valves; and

valve stop inhibit means for inhibiting a stop of open-and-closure drives for the one of the engine valves, in a case where a lost motion quantity of the valve stop mechanism exceeds a predetermined value.

2. A variably operated valve system for an internal combustion engine, comprising:

drive cams to which a rotational force is transmitted from a crankshaft;

engine valves, two of the engine valves being mounted in each of engine cylinders and each engine valve being biased toward a valve closure direction by means of a spring force of a corresponding one of valve springs;

swing cams that make the respective engine valves open operation against spring forces of valve springs;

a transmission mechanism that converts rotation motions of the drive cams into swing motions and transmits the converted swing motions to the swing cams;

a pair of operational members that swing in accordance with the swing motions of the swing cams to operate the respective engine valves to open or close;

a valve stop mechanism that absorbs a swing quantity of one of the pair of operational members to stop open-and-closure operations of one of the two engine valves; and

valve stop inhibiting means for inhibiting an absorption of a swing motion of the valve stop mechanism in a case where the swing quantity of each of the pair of the operational members exceeds a predetermined quantity.

3. A variably operated valve system for an internal combustion engine, comprising:

a drive cam rotationally driven through a crankshaft;

a pair of engine valves, each engine valve being biased toward a closure direction by means of a spring force of a valve spring;

a pair of swing cams that swing to drive the pair of engine valves via a pair of operational members;

a transmission mechanism that converts a rotation motion of the drive cam into a swing motion and transmits the converted swing motion to the pair of swing cams;

a control mechanism that varies a posture of the transmission mechanism to vary operation characteristics of the pair of the engine valves; and

a valve stop mechanism disposed on at least one of the pair of operational members to make a lost motion to stop a drive of one of the pair of engine valves, wherein, in a case where a valve lift quantity of one of the pair of engine valves exceeds a predetermined value, an operation of a valve stop through the valve stop mechanism is inhibited.

4. The variably operated valve system for the internal combustion engine as claimed in claim 1, wherein the valve stop mechanism includes: a retaining hole that holds one of the pair of fulcrum members movable; and a basing member that biases the one of the pair of fulcrum members toward a direction of a corresponding one of the pair of swing arms and the lost motion is made by moving the one of the pair of fulcrum members against the biasing member.

5. The variably operated valve system for the internal combustion engine as claimed in claim 4, wherein the pair of fulcrum members are hydraulic pressure rush adjusters.

6. The variably operated valve system for the internal combustion engine as claimed in claim 1, wherein the valve stop mechanism is disposed only in one of the pair of fulcrum members and, when the lost motion quantity is equal to or below the predetermined value, only one of the two engine valves corresponding to the one of the pair of fulcrum members is allowed to be stopped through the valve stop mechanism.

7. The variably operated valve system for the internal combustion engine as claimed in claim 1, wherein the valve stop mechanism is disposed in both of the pair of fulcrum members and, when the lost motion quantity is equal to or below the predetermined quantity, both of the two engine valves are allowed to be stopped through the valve stop mechanism.

8. The variably operated valve system for the internal combustion engine as claimed in claim 6, wherein the pair of swing cams are integrally formed.

9. The variably operated valve system for the internal combustion engine as claimed in claim 4, wherein the swing fulcrum of each of the pair of each of the pair of swing arms is formed by a spherical recess section disposed on a corresponding one of the pair of swing arms and a spherical convex section engaged with the recess section.

10. The variably operated valve system for the internal combustion engine as claimed in claim 4, wherein, when the valve stop inhibiting means mechanically inhibits the lost motion of the one of the pair of fulcrum members when the lost motion quantity exceeds the predetermined value.

11. The variably operated valve system for the internal combustion engine as claimed in claim 4, wherein the control shaft includes a control cam that is in an eccentricity state to an axial center of the control shaft, the control cam is inserted through the transmission mechanism, and the posture of the transmission mechanism is varied according to a rotation of the control cam.

12. The variably operated valve system for the internal combustion engine as claimed in claim 10, wherein a limitation of the lost motion is made by a limitation of a movement of the one of the pair of fulcrum members by means of a projection section in a movable range of the one of the pair of fulcrum members.

13. The variably operated valve system for the internal combustion engine as claimed in claim 10, wherein the valve stop inhibit means is constituted by a movement purpose hole and a limitation purpose hole, both of the movement purpose and limitation purpose holes being disposed in the one of the pair of fulcrum members and in an inner wall of the retaining hole, and a limitation pin movably disposed across the movement purpose hole and the limitation purpose hole and a state in which the one of the pair of fulcrum members is in a locked state by disposing the limitation pin across the movement purpose hole and the limitation purpose hole and another state in which the limitation pin is housed in the movement purpose hole to enable the one of the pair of fulcrum members to make the lost motion are controlled.

14. The variably operated valve system for the internal combustion engine as claimed in claim 13, wherein the valve stop inhibiting means includes another biasing member disposed in the movement purpose hole of the one of the pair of fulcrum members to bias the limitation pin toward a direction of the limitation purpose hole and a hydraulic pressure circuit that supplies a hydraulic pressure to press under pressure the limitation pin in another direction of the movement purpose hole of the one of the pair of fulcrum members against the biasing force of the other biasing member.

15. The variably operated valve system for the internal combustion engine as claimed in claim 14, wherein the hydraulic pressure circuit includes: oil passage holes disposed at a bearing section of the control shaft and a communication hole penetrated through a diameter direction of the control shaft to appropriately communicate with the oil passage holes and wherein, when the control shaft is rotated through an angle equal to or below a predetermined rotation angle, both of the oil passage holes and the communication hole are communicated with each other and, when the control shaft exceeds the predetermined rotation angle, the communication between the oil passage holes and communication hole is interrupted.

16. The variably operated valve system for the internal combustion engine as claimed in claim 14, wherein, when the lost motion quantity is smaller than the predetermined value of the lost motion quantity of the one of the pair of fulcrum members, a hydraulic pressure is supplied to move the limitation pin toward the direction of the movement purpose hole against a biasing force of the other biasing member and, when the lost motion quantity exceeds the predetermined value, the biasing force of the other biasing member causes the limitation pin toward the direction of the limitation purpose hole.

17. The variably operated valve system for the internal combustion engine as claimed in claim 4, wherein an angle which is a subtraction of 90° from an angle formed between the corresponding one of the pair of swing arms and the one of the pair of fulcrum members when the lost motion quantity of the fulcrum member is at the predetermined value is smaller than an angle formed between the corresponding one of the pair of swing arms and the one of the pair of fulcrum members when the engine valves are open at a time of a maximum lift control.

18. The variably operated valve system for the internal combustion engine as claimed in claim 4, wherein the valve stop inhibiting means includes a limitation member that limits the lost motion by biasing the one of the pair of fulcrum members toward a direction of the corresponding one of the pair of swing arms, when the lost motion quantity of the one of the pair of fulcrum members exceeds the predetermined value.

19. The variably operated valve system for the internal combustion engine as claimed in claim 14, wherein the valve stop inhibiting means includes a valve stop inhibit circuit constituted by a control unit that interrupts the supply of the hydraulic pressure via an electromagnetic switching valve of the hydraulic pressure circuit.

20. The variably operated valve system for the internal combustion engine as claimed in claim 19, wherein the predetermined value of the lost motion quantity can arbitrarily be set according to a driving condition of the engine.