JP2003293802A - Valve system control device for engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve fuel economy by reducing the occurrence of a pumping loss in low rotation low load operation including an idling time. <P>SOLUTION: A valve system control device comprises a phase varying mechanism to vary the switching timings of a suction valve and an exhaust valve; and a lift varying mechanism capable of respectively varying lift amounts of the two suction valves and the phase varying mechanisms of each of cylinders. A lift varying mechanism of the second suction valves of the two suction valves of each cylinder is substantially stopped. Control is effected such that in a low rotation low load operation area, the opening timing of a first suction valve is delayed by IO from a suction top dead center, the substantial suction stop of a second suction valve is stopped, and the closing timing of the exhaust valve is advanced by EC from a suction top bead center. Further, a relation of IO and EC is set to IO<EC, an exhaust process is completed slightly in a compression state, and in a suction stroke, a piston is pushed back, the occurrence of a pumping loss is reduced, and combustion is improved by a swirl effect. <P>COPYRIGHT: (C)2004,JPO

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an engine valve control system, and more particularly to an engine valve control system provided with a variable valve control system for varying opening / closing timings and lift amounts of intake valves and exhaust valves. .

[0002]

2. Description of the Related Art Conventionally, it has been necessary to provide a variable valve operating device in an engine that can change lift characteristics such as an opening / closing timing (phase angle with respect to a crankshaft) of an intake valve and an exhaust valve and a lift amount according to an operating state. Has been done.

As an example of valve control of an engine provided with such a variable valve device, for example, Japanese Patent Laid-Open No. 10-26 is known.
As described in Japanese Patent No. 6878, in a low-to-medium-load self-ignition region, a state where there is no overlap in the opening period of the intake and exhaust valves, that is, a minus overlap state, suppresses pumping loss, It has been conventionally known that NOx is reduced by self-ignition, and as described in Japanese Patent Application Laid-Open No. 55-128610, in order to reduce pumping loss, an intake valve is used instead of a throttle valve. It is known that the output is controlled by adjusting the closing timing, and it is also known from Japanese Patent Laid-Open No. 59-85408.
As described in the publication, in order to improve idle stability and fuel efficiency during low load operation, one of the two intake valves with the larger lift amount and the two exhaust valves with the lift amount. It is known that the valve with the larger valve stop is stopped.

[0004]

By the way, in an engine, improvement of fuel consumption is an important issue, and in particular, it is necessary to reduce pumping loss during low rotation and low load operation such as idling to improve fuel consumption. Therefore, there is a demand to achieve that by controlling a variable valve operating device of a variable phase type or a variable lift type.

However, the conventional control of the variable valve operating system of the variable phase type or the variable lift type does not aim to improve the fuel consumption by reducing the pumping loss at the time of low rotation and low load operation such as idling.

The present invention has been made in view of the above problems, and it is an object of the present invention to reduce pumping loss during low rotation and low load operation including idling to improve fuel efficiency.

[0007]

The invention according to claim 1 is
In an engine valve control device equipped with a variable valve device for varying at least the opening / closing timing of each of an intake valve and an exhaust valve, the intake valve is controlled in a low rotation and low load operation range by control of the variable valve device. Is opened by a first predetermined amount after the intake top dead center, and the exhaust valve is closed by a second predetermined amount before the intake top dead center, and
It is an object of the present invention to provide a valve operating control system for an engine, wherein the second predetermined amount is set to be larger than a predetermined amount.

In this case, in the low rotation and low load operation range, the intake valve opens at a timing after the intake top dead center by a first predetermined amount and the exhaust valve opens at a second predetermined amount before the intake top dead center. The intake / exhaust valve opening timing is in a minus overlap state where it does not completely overlap. Then, by closing the exhaust valve before the intake top dead center, a slightly compressed state is reached, the exhaust stroke ends, and the piston is pushed back into the intake stroke, reducing pumping loss. Further, the effect of reducing pumping loss in the low rotation and low load operation range is that the advance amount with respect to the intake top dead center at the exhaust valve closing timing is greater than the first predetermined amount which is the delay amount with respect to the intake top dead center at the intake valve opening timing. Is most significant when the second predetermined amount is Further, in this case, the increase in the residual gas due to the exhaust valve closing before the intake top dead center is within an allowable range, and combustion stability can be secured. Therefore, during low rotation and low load operation such as idling, it is possible to reduce combustion loss and improve fuel efficiency while ensuring combustion stability.

According to a second aspect of the present invention, there is provided an engine valve control system according to the first aspect, wherein the engine has two intake valves for each cylinder, and the variable valve drive system includes at least a first intake valve. A low-rotation, low-load operating range is provided that includes a phase variable mechanism that changes the opening and closing timing of the valve, and a lift variable mechanism that can change the lift characteristic of the second intake valve to a low lift characteristic including at least substantial intake stop. Then, the phase variable mechanism sets the timing of opening the first intake valve to the timing after the first predetermined amount from the intake top dead center, and the lift variable mechanism substantially stops the opening of the second intake valve. It is characterized by having a low lift characteristic including.

The pumping loss during low rotation and low load operation is
It can be reduced by setting the intake and exhaust valves in the minus overlap state and setting the advance amount with respect to the intake top dead center at the exhaust valve closing timing larger than the delay amount with respect to the intake top dead center at the intake valve opening timing. In this case, the combustion stability is affected by an increase in the residual gas within the allowable range, so that it is preferable to further improve the combustion. According to the invention of claim 2, in the low rotation and low load operation range, the second intake valve is in the open stop state, and only the first intake valve is opened, so that the swirl of intake air occurs in the combustion chamber, and the swirl effect is generated. Improves combustion.

According to a third aspect of the present invention, in the engine valve control system of the second aspect, the engine has an air-fuel ratio set leaner than the stoichiometric air-fuel ratio in the low rotation and low load operation range. It is characterized by being one.

When the combustion is improved by the swirl effect, lean burn is possible and fuel consumption can be further improved.

According to a fourth aspect of the present invention, there is provided an engine valve control system according to the second aspect, wherein in the low rotation medium load operating range, the first intake valve is open and the exhaust valve is open. The second intake valve is made to have a low lift characteristic including the substantial stop of the valve opening, while providing an overlap during the open period.

As described above, in the low rotation medium load operating range,
By increasing the internal EGR rate and reducing pumping loss by providing an overlap between the period when the intake valve is open and the period when the exhaust valve is open, and by making the second intake valve have a low lift characteristic that includes the substantial open stop. It is possible to improve fuel efficiency.

According to a fifth aspect of the present invention, in the engine valve control system according to the second aspect, in the low rotation and high load operation range, the timing for closing the two intake valves is set to the low rotation and low load operation. The feature is that it is faster than the area.

As described above, in the low rotation and high load operating range, the volume efficiency can be increased and the required engine output can be obtained by advancing the timing of closing the intake valve.

According to a sixth aspect of the present invention, in the engine valve control system of the second aspect, in the high rotation and low load operation range, the timing for closing the two intake valves is advanced and the intake valves are increased. It is characterized in that an overlap is provided between the period when the valve is open and the period when the exhaust valve is open.

As described above, in the high-rotation low-load operation range, the volume efficiency is improved by advancing the timing of closing the intake valve and providing the overlap between the period when the intake valve is open and the period when the exhaust valve is open. It is possible to obtain a required engine output, reduce pumping loss, and improve fuel efficiency.

According to a seventh aspect of the present invention, there is provided an engine valve control system according to the fifth aspect, wherein in the high load operation range, the timing of closing the two intake valves is increased as the engine speed increases. It is characterized by delaying.

As described above, in the high load operation range, the timing of closing the intake valve is delayed with the increase of the engine speed, thereby ensuring the volume efficiency in the high speed range,
You can get the required engine power.

[0021]

BEST MODE FOR CARRYING OUT THE INVENTION Embodiments of the present invention will be described below with reference to the drawings.

FIG. 1 is a plan view of a cylinder head of an engine equipped with a valve operating system according to an embodiment of the present invention. The engine according to this embodiment is an in-line 4-cylinder DO
In an HC (double overhead cam) engine, an intake camshaft 2 and an exhaust camshaft 3 are arranged above a cylinder head 1 so as to extend in the cylinder row direction in parallel with each other, and for each cylinder A1 to A4. Is provided with a spark plug 4. This engine has two intake ports Pin for each cylinder A1 to A4, as shown in FIG. 4 described later.
1, Pin2 and two exhaust ports Pex1, Pex2
, Two intake valves (not shown) and two exhaust valves (not shown) are provided, and four tappets 5 (FIG. 1) are provided for each of the cylinders A1 to A4 corresponding to the intake valves and the exhaust valves. Then, only the intake side of the fourth cylinder A4 is shown). Then, the intake camshaft 2 and the exhaust camshaft 3 have three different cams 6, 7, 8 (having different profiles) with respect to the respective tappets 5 corresponding to the two intake valves of each of the cylinders A1 to A4. In FIG. 1, only the intake side of the third cylinder A3 is shown).

A cam carrier 13 is arranged on the upper surface of the base of the cylinder head 1. The cam carrier 13 is provided with a vertical wall portion 14 that supports the lower portions of the intake camshaft 2 and the exhaust camshaft 3 between the left and right and between the cylinders, and slidably accommodates the intake-side and exhaust-side tappets 5. A tappet guide 15 is formed for guiding.
A cam cap (not shown) that supports the upper portions of the intake camshaft 2 and the exhaust camshaft 3 is bolted to the upper surface of the vertical wall portion 14.

Further, the cam carrier 13 is formed with a plug insertion hole 16 for inserting the ignition plug 4 which is mounted in a plug hole (not shown) of each cylinder A1 to A4 of the cylinder head 1. Then, on one end side and the other end side of the cam carrier 13, valve insertion holes for inserting the two hydraulic control valves 101, 102 for the variable lift mechanism at positions aligned with the plug insertion holes 16 on the one end side and the other end side. 103 and 104 are formed. The hydraulic pressure control valves 101 and 102 control the hydraulic pressure for the variable lift mechanism built in each tappet 5. The cam carrier 13 is formed with main oil passages 105, 106 and 107 for supplying working oil pressure to the variable lift mechanism incorporated in each tappet 5.

This engine corresponds to two intake valves and exhaust valves for each cylinder A1 to A4, and each tappet guide 1
The tappet 5 housed in 5 has a built-in lift variable mechanism that changes the lift amount (and opening angle) of the intake valve and the exhaust valve in response to the hydraulic pressure.

In this engine, although not shown, a crank sprocket is fitted and fixed to one end of a crankshaft, and cam sprockets 22 and 23 are provided at one end of the intake camshaft 2 and one end of the exhaust camshaft 3, respectively. , The timing chain 24 between those sprockets
And are covered with a cover member 21. When the crankshaft rotates, both camshafts 2 and 3 rotate via the crank sprocket, the timing chain 24, and the cam sprockets 22 and 23, and the intake valve and the exhaust valve are opened and closed.

At one end of the intake camshaft 2 and the exhaust camshaft 3, the rotational phase angle of the camshafts 2 and 3 with respect to the crankshaft, that is, the phase angle of the intake valve and the exhaust valve with respect to the crankshaft during the opening period of the exhaust valve. Phase changing mechanism 20 for independently changing each
1, 202 are provided respectively, and a hydraulic control valve (not shown) for the phase varying mechanism is arranged at the engine end covered with the cover member 21. .

These valve timing variable mechanisms 201,
Since 202 has almost the same structure, the structure of the intake side phase changing mechanism 201 will be described, and the detailed description of the exhaust side phase changing mechanism 202 will be omitted.

That is, as shown in FIG. 2, the variable phase variable mechanism 201 has a hollow structure and a plurality of protrusions 20 on the inner surface side.
2a, and a housing 203 having a housing 202 that is bolted to the cam pulley 22 together with a lid member (not shown), and a rotor 203 having a plurality of engaging portions 203a protruding in a cross shape in the housing 202. It is fixed to the tip of the intake camshaft 2 by means of bolts 204 and is stored so as to be relatively rotatable within a predetermined range. The rotor 203 has a protrusion 202a of the engaging portion 203a with respect to the housing 202.
Relative rotation is possible within the range of contact with. As a result, of the intake camshaft 3 to which the rotor 203 is fixed,
The phase angle with respect to the cam sprocket 22 to which the housing 202 is fixed can be changed, that is, the phase angle of the intake camshaft 2 with respect to the crankshaft can be changed, and thereby the opening / closing timing of the intake valve can be changed. There is.

Inside the phase variable mechanism 201, the cam sprocket 22, the housing 202, and the rotor 20 are provided.
3 and the lid member define a plurality of oil chambers 205, and the rotor 2
In 03, an oil passage 206 for advancing and an oil passage for retarding (not shown) are provided so as to communicate with the plurality of oil chambers 205 so as to be displaced in the front-rear direction and the circumferential direction. Under the control of an oil control valve (OCV) not shown, the oil chamber 20 is passed through the advance oil passage 206 or the retard oil passage.
By supplying the hydraulic pressure to 5, the rotor 203 is rotated, and the phase angle of the intake camshaft 2 with respect to the crankshaft can be continuously changed and set to an arbitrary phase angle.

Similarly, the phase changing mechanism 202 on the exhaust side can continuously change the phase angle of the exhaust camshaft 3 with respect to the crankshaft to set an arbitrary phase angle.

The tappet 5 has a built-in lift variable mechanism and is cylindrical as a whole, and as shown in FIG. 3, a center tappet 51 which forms a substantially rectangular flat center cam contact surface long in the cam sliding contact direction. And a side tappet 52 which is arranged on both sides of the cam contact surface of the center tappet 51 and constitutes a side cam contact surface having a substantially circular flat shape together with the center cam contact surface of the center tappet 51. The center tappet 51 and the side tappet 52 are configured to be movable relative to each other in the tappet sliding direction, and can be coupled and disengaged from each other by a lock mechanism described later.

The side tappet 52 has a structure in which a pair of left and right tubular portions 52a and 52b having side cam contact surfaces are integrally connected by a lower connecting portion 52c which constitutes an abutting portion with a stem end of a valve. A tappet spring 53 is installed on the lower connecting portion 52c, and a center tappet 51 is incorporated between the tappet spring 53 and the left and right cylindrical portions 52a and 52b. In the tappet 50, the lower surface of the lower connecting portion 52c contacts the stem ends of the intake valve and the exhaust valve via a shim (not shown).

The center tappet 51 extends vertically downward from the edges of both long sides along the cam sliding contact direction of the substantially rectangular flat center cam contacting surface, and is parallel to the tappet sliding direction and the tappet sliding direction. It has a pair of side end surfaces 54a, 54b which are symmetrical with respect to the axis, and has a circular arc-shaped cross section which is vertically downward from the edges of both short sides and is parallel to the tappet sliding direction and symmetrical with respect to the tappet axis. It has a pair of outer peripheral surfaces.

The center tappet 51 is formed with a through hole 56 parallel to the cam axis direction and concentric with the tappet axis and penetrating from one side end surface to the other side end surface. Further, in the center tappet 51, the outer sides of the center tappet 51 continuously extend from the outer circumferential surface to the side tappet 52 side at both circumferential edge portions of both outer circumferential surfaces, and the inner side faces 54a. 54b extending substantially in the direction of the cam axis at a substantially right angle from 54b, 54b to form a protruding portion that constitutes a sliding contact surface with the side tappet 52 over substantially the entire length in the direction of the tappet axis.

The side tappet 52 has a cam sliding contact direction and a tappet sliding that extend vertically downward from inner edge portions facing each other along the cam sliding contact direction of the side cam contact surfaces of the pair of left and right cylindrical portions 52a and 52b. The inner end surfaces 58a and 58b are parallel to each other and symmetrical with respect to the tappet axis, and the inner end surfaces 58a and 58b are provided with predetermined side end surfaces 54a and 54b of the center tappet 51 at both ends in the cam sliding contact direction. Projecting portions 59a facing each other in a non-sliding contact state at intervals of
An outer peripheral surface 6 having an arcuate cross-section and having 59b, which extends vertically downward from the edge of the side cam contact surface on the outer end side in the cam axis direction and which is parallel to the tappet sliding direction and symmetrical with respect to the tappet axis.
0a and 60b.

The tubular portion 52 of the side tappet 52
a and 52b are parallel to the cam axis direction and concentric with the tappet axis from the inner end surfaces 58a and 58b to the outer peripheral surface 60a.
The through holes 61a, 61b penetrating the 60b are in direct communication with the through hole 56 of the center tappet 51 when the center cam sliding contact surface of the center tappet 51 and the side cam sliding contact surface of the side tappet 52 are flush with each other. Has been formed.

The tubular portion 52 of the side tappet 52
At the end portions of the inner end surfaces 58a and 58b of the a and 52b in the cam sliding contact direction, the sliding contact portions that are in sliding contact with the sliding contact surfaces of the overhang portions formed at the edges of the center tappet 51 in the cam sliding contact direction. Is formed in a rib shape over substantially the entire length in the axial direction of the tappet.

Center tappet 51 and side tappet 52
Is the side end surfaces 54a, 54b of the center tappet 51.
And projecting portions 59a, 5 located at both ends of the inner end surfaces 58a, 58b of the tubular portions 52a, 52b of the side tappet 52.
9b face each other at a predetermined interval in a non-sliding contact state, and also the protruding portion of the center tappet 51 and the side tappet 5
The two tubular portions 52a and 52b can be slidably contacted with the sliding contact portions.

The three cams 6, 7 and 8 for the intake side tappets 5 are provided on both sides of one of the two intake valves (first intake valve) of each cylinder A1 to A4. Sheet (6,
8) is a low speed cam having the same cam profile with a low lift amount, and one central piece (7) is a high speed cam having a high lift amount cam profile, and two intake valves for each cylinder A1 to A4. For the other of the two (second intake valve),
Two pieces (6, 8) on both sides are stopping cams having the same cam profile with a minute lift amount that substantially stops intake, and one piece (7) in the center is a high speed cam having a high lift amount cam profile. . Also, the three cams 6, 7, 8 for each tappet 5 on the exhaust side are low speed cams, two on both sides having the same cam profile with a low lift amount, and one cam in the center.
The sheet is a high speed cam having a high lift cam profile.

Center tappet 51 and side tappet 52
The center cam sliding contact surface of the center tappet 51 abuts on the base circle portion of the central cam 7, and the side tappet 52
In the state where the side cam sliding contact surface of the side cam is in contact with the base circle portions of the cams 6 and 8 on both sides, the center cam sliding contact surface of the center tappet 51 and the side cam sliding contact surface of the side tappet 52 are substantially flush with each other. Through hole 56 of the center tappet 51 and through holes 61a, 61 of the side tappet 52.
b communicates in a straight line.

The side tappet 52 has an outer peripheral surface 60.
The a and 60b substantially contact the inner peripheral surface of the tappet guide 15 in the cam axis direction. And the protruding portion 5 of the side tappet 52
A predetermined gap is held between the center tappet 51 and the center tappet 51, and a machining error between the side tappet 52 of the tappet 5 and the center tappet 51 inside the tappet guide 15 is allowed by this gap.

A through hole 56 (through hole portion) of the center tappet 51 and a through hole 6 of the side tappet 52 of the tappet 5.
1a and 61b (through holes), as described above, the center cam sliding contact surface of the center tappet 51 contacts the base circle portion of the central cam 7, and the side cam sliding contact surfaces of the side tappet 52 are cams 6 and 8 on both sides. While being in contact with the base circle portion of, a series of through holes are formed by communicating in a straight line in the cam axis direction. Then, in this through hole, it is possible to project to a through hole portion (through hole 61b) that penetrates one tubular portion 52b of the side tappet 52 and a through hole portion (through hole 56) that penetrates the center tappet 51 by hydraulic pressure. As described above, the hydraulic plunger 71 is built into the through hole portion (through hole 56) penetrating the center tappet 51, and the other tubular portion 52 of the side tappet 52 is hydraulically operated through the hydraulic plunger 71.
A lock pin 72 is built into the through hole portion (through hole 61a) penetrating a so that the lock pin 72 can be projected.
A pressing spring 73 is built in so as to urge the center tappet 51 and the side tappet 52 from each other to a position where they can move relative to each other in the sliding direction of the tappet.

The center tappet 51 and the side tappet 5 are formed by the through holes 56, 61a and 61b, the hydraulic plunger 71, the lock pin 72, the pressing spring 73 and the like.
Locking means for connecting and disconnecting the two with each other is configured.

In the through-hole portion (through-hole 56) on the side of the center tappet 51, bushes 74, which are flush with the side end surfaces 54a, 54b of the center tappet 51, are provided at both ends of the divided portion with the side tappet 52. 75 are inserted and fixed. Among them, the bush 75 fixed to the end portion on the hydraulic plunger 71 side is a collar portion 76 on the outer periphery of the lock pin 72.
To restrict the movement of the lock pin 72 to the hydraulic plunger 71 side. The bush 74 fixed to the other end holds the pressing spring 73 in a compressed state between the bush 74 and the collar portion 76 on the outer periphery of the lock pin 72.

Further, bushes 77, 78 are provided in the through hole portions (through holes 61a, 61b) on the side tappet 52 side.
Are inserted and fixed. Of which, the hydraulic plunger 71
Through-hole portion (through-hole 61
The bush 78 inserted and fixed in b) has a fitting hole 78b, which is closed on the outer peripheral side of the tappet by an end wall 78a, and in which the hydraulic plunger 71 is slidably fitted.

The bush 77 inserted into and fixed to the through hole portion (through hole 61a) of the other tubular portion 52a has a partition wall 77a as a stopper portion for restricting the movement of the lock pin 72 in the axial center portion. Then, the fitting hole 77b for fitting the end of the lock pin 72 inward with the partition wall 77a sandwiched therebetween.
Has a fitting hole 77c on the outer peripheral side of the tappet for fitting a later-described locking member, and the partition wall 77a has both fitting holes 7c.
There is a through hole 77d that connects 7b and 77c.

The bushes 77, 78 on the side tappet 52 side have the end portions on the side of the center tappet 51 projected from the through-hole portions (through-holes 61a, 61b) by a predetermined dimension, so that the through-hole portions (through-holes of the center tappet 51 The bushes 74, 75 inserted and fixed in the holes 56) face each other at predetermined intervals.

In the lock means having such a structure, the hydraulic plunger 71 is provided with the lock pin 72 when the hydraulic pressure is applied.
The lock pin 72 is moved to the side and the end portion is inserted into the through hole portion (through hole 56) on the side of the center tappet 51, and the lock pin 72 is moved against the pressing spring 73 to move the lock pin 72.
End portion of the side tappet 52 through hole portion (through hole 61
Introduce a). At this time, the hydraulic plunger 71 straddles the split portion of the center tappet 51 and the tubular portion 52b of the side tappet 52, and the lock pin 72
Straddles the divided portion of the center tappet 51 and the other tubular portion 52a of the side tappet 52, and they cooperate to join the center tappet 51 and the side tappet 52 into a locked state. When the hydraulic pressure is released, the pressing spring 73 moves the lock pin 72 to the hydraulic plunger 71 side, and the hydraulic plunger 71 returns to the original position,
Both end surfaces of the lock pin 72 are flush with the split portion of the center tappet 51 and the side tappet 52 in the through hole 56 of the center tappet 51, and the center tappet 51 is disengaged from the side tappet 52 and is unlocked (locked). It will be released).

As described above, the locking means can be switched between the locked state and the unlocked state by the hydraulic pressure. In the locked state, the side tappet 52 and the center tappet 51 are locked.
And the side tappet 52 and the center tappet 51 driven by the central cam 7 cooperate to lift the intake valve and the exhaust valve in a high-speed manner, and the center tappet 51 moves the side tappet 52 in the unlocked state. The first intake valve and the exhaust valve are lifted in a low speed mode by being slid relative to each other and driven by both sides 6 and 8 to substantially stop the operation of the second intake valve.

The tappet 5 has a hydraulic plunger 7
In a state in which a spherical member 86 as a tappet stopping member partially protrudes from the outer circumference of the tappet in a fitting hole 77c on the outer circumference side of the tappet of the bush 77 fixed to the through hole 61a of the tubular portion 52a in which one does not include 1 Holds freely rolling. The spherical member 86 has a ring-shaped holding member 87 having a bent claw portion 87a inserted from the outside along the inner circumference of the through hole 61a, and has a reduced diameter portion provided on the outer surface of the end portion of the bush 77. It is rotatably held by being fitted and fixed in between.

The tappet guide 15 provided on the cam carrier 13 is engaged with a spherical member 86, which serves as a tappet stopping member, held on the tappet 5 side so as to be movable only in the tappet sliding direction. As described above, a groove 88 extending in the sliding direction of the tappet is provided. Spherical member 8
6 cooperates with the groove 88 on the inner circumference of the tappet guide hole 25 to regulate the rotation of the tappet 5.

The end faces 5 on both sides of the center tappet 51, which face the tubular parts 52a and 52b of the side tappet 52, face each other.
4a and 54b project toward the inner end surfaces 58a and 58b of the tubular portions 52a and 52b of the side tappet 52 in the cam axis direction below the through hole 56, and the side tappet 5
Locking portions 89a and 89b that are locked to the ends of the bushes 77 and 78 that protrude inward from the through holes 61a and 61b of the two tubular portions 52a and 52b are formed. By locking the locking portions 89a and 89b on both sides to the bushes 77 and 78 on the side tappet 52 side, the center tappet 51 is restricted from moving upward in the tappet sliding direction.

As shown in FIGS. 1 and 4, an oil passage 108 extending from one end surface is formed in the cam carrier 13, and this oil passage 108 is a first hydraulic control valve (OCV) 101 for a variable lift mechanism. Connected to. When the OCV 101 is off, the hydraulic oil supplied to the oil passage 108 is cut off, and when the OCV 101 is turned on, the oil passage 108 extends to the intake side.
Through the first main oil passage 105.

The first main oil passage 105 is extended along the longitudinal direction of the cam carrier 13, and the tappet guide 15 is provided.
Between the cylinders A1 to A4, the branch passage 109 extends toward the intake side tappet guide 15 of each of the cylinders A1 to A4.
Are formed in the same number as. This branch passage 109 is
Each of the cylinders A1 to A4 communicates with the tappet guide 15 corresponding to one of the intake valves, and the OCV 101 is turned on to supply the hydraulic pressure to the variable lift mechanism incorporated in the tappet 5.

As shown in FIG. 3, for the fourth cylinder A4, the intake ports Pin1 and Pin2 are arranged in the other cylinder A.
It is the reverse of 1-A3. And the fourth cylinder A4
The first main oil passage 105 is connected to the second hydraulic control valve (OCV2) 102 via an intermediate oil passage 110 continuous to the branch passage 109 corresponding to. When the second OCV 102 is off, the first main oil passage 1
The hydraulic oil in 05 has its path cut off, and the second OCV 102
When turned on, the first main oil passage 105 passes through the intermediate oil passage 111, and the pair of intermediate oil passages 112, 11 extending from the intermediate oil passage 111 to the intake side and the exhaust side, respectively.
3 to communicate with the second and third main oil passages 106 and 107, respectively.

Second and third main oil passages 106 and 107
Also, as with the first main oil passage 105, the cam carrier 1
3 is extended in the longitudinal direction. And the second on the intake side
In the main oil passage 106, the laterally extending branch passage 114 communicates with the tappet guide 15 corresponding to another intake valve for each of the cylinders A1 to A4, and the second OC
When the V102 is turned on, the operating hydraulic pressure is supplied to the hydraulic plunger 71 of the lift variable mechanism incorporated in the tappet 5.

On the other hand, in the third main oil passage 107 on the exhaust side, between the tappet guides 15, the branch passages 115 extending laterally are provided in the tappet guides 15 corresponding to the exhaust valves of the cylinders A1 to A4, respectively. Communication, second O
When the CV 102 is turned on, the hydraulic pressure of the lift variable mechanism 71 of the lift variable mechanism incorporated in the tappet 5 is changed to the hydraulic pressure.
Supply to.

The engine valve operating system of this embodiment is
As described above, the phase changing mechanism that changes the opening / closing timing of the intake valve and the exhaust valve (the phase with respect to the crankshaft during the opening / closing period) is provided, and two intake valves and two exhaust valves for each of the cylinders A1 to A4 are provided. Equipped with a variable lift mechanism capable of changing the lift amount, and each cylinder A1
The variable lift mechanism for the second intake valve of the two intake valves for each A4 is capable of changing the lift characteristic of the second intake valve to a low lift characteristic including at least a substantial intake stop. Then, the phase variable mechanism and the lift variable mechanism are controlled as follows according to the operating state of the engine.

In other words, in the low rotation and low load operation range including the idling operation, the timing of opening the first intake valve is set to the first predetermined amount (IO) from the intake top dead center by the phase variable mechanism.
Just after that, with the variable lift mechanism, the second
The intake valve is made to have a low lift characteristic that includes the substantial stop of opening, and the timing of closing the exhaust valve is set to the second position above the intake top dead center.
Control so that the timing is a predetermined amount (EC) earlier,
Further, the second predetermined amount (E) is calculated from the first predetermined amount (IO).
C) is set to a large value (IO <EC). Further, in this low rotation and low load operation range, the air-fuel ratio is set leaner than the stoichiometric air-fuel ratio.

In this case, the intake / exhaust valve opening timing is in a minus overlap state where they do not completely overlap,
By closing the exhaust valve before the intake top dead center, the exhaust stroke ends slightly, the exhaust stroke ends, and the piston is pushed back into the intake stroke, reducing pumping loss. Then, the effect of reducing the pumping loss is from the first predetermined amount (IO) which is the delay amount with respect to the intake top dead center at the intake valve opening timing to the second amount which is the advance amount with respect to the intake top dead center at the exhaust valve closing timing. Large quantification (EC) (IO <
In the case of EC), the increase in the residual gas is within the allowable range, and combustion stability can be ensured. In addition, in this low rotation and low load operation range, the second intake valve is in the open stop state,
The swirl effect improves combustion and improves fuel economy.

FIGS. 5 to 8 show simulation data of changes in volume efficiency, pumping loss, combustion chamber temperature and residual gas amount with respect to the valve timing retard amount during idle operation. In FIGS. 5 to 8, the data of the square mark plots is the case of IO = EC, the data of the triangle mark plots is the case of IO <EC, and the data of the x mark plots is the case of IO> EC. The data in the circle plots are for IO <EC, and the intake valve is in the one-valve stopped state. These figures show engine speed 1500 rpm
After fixing the IO to the top dead center 40 °, 50 °, 60 °, 7
The simulation result when changing to 0 ° is shown.
The difference between IO and EC was set to 10 °. This difference may be set to an optimum value according to the specifications of the engine. Further, the alternate long and short dash line is data in the case of a normal engine with overlap (reference valve timing). From these data,
It can be seen that the pumping loss is reduced by setting the minus overlap during the idling operation, and particularly when IO <EC, the pumping loss reduction effect is the largest. Even when comparing each data at the valve timing that can obtain the volumetric efficiency equivalent to that of the reference valve timing, I
When O <EC, the pumping loss is the smallest.

FIG. 9 is a P-V diagram at the reference valve timing, and is a P-V diagram in the case of this embodiment (IO <EC) of FIG. In the case of this embodiment, the opening timing of the intake / exhaust valve becomes minus overlap, and the exhaust valve closes before the intake top dead center, so that the compression stroke becomes slightly compressed and the exhaust stroke ends and the intake stroke begins. Since the piston is pushed back, there is little pumping loss.

Further, in the control of this embodiment, an overlap is provided between the period in which the first intake valve is open and the period in which the exhaust valve is open in the low rotation medium load operation range, and
The second intake valve is substantially stopped. Thereby, the internal E
It is possible to improve the fuel efficiency by increasing the GR (exhaust gas recirculation) rate and reducing the pumping loss.

In the low rotation and high load operation range, the timing for closing the two intake valves is set earlier than in the low rotation and low load operation range. Thereby, the volumetric efficiency is increased and the required engine output is obtained.

Further, in the high rotation and low load operation range, the timing of closing the two intake valves is advanced and an overlap is provided during the period when the intake valves are open and the period when the exhaust valves are open. As a result, the volumetric efficiency is increased to obtain the required engine output, the pumping loss is reduced, and the fuel consumption is improved.

Further, in the high load operation range, the timing of closing the two intake valves is delayed as the engine speed increases. As a result, the volume efficiency in the high rotation speed range is secured and the required engine output is obtained.

[0068]

As is apparent from the above description, according to the engine valve control system of the present invention, the pumping loss is reduced while ensuring the combustion stability at the time of low rotation and low load including the idling time. Fuel consumption can be improved.

[Brief description of drawings]

FIG. 1 is a plan view of a cylinder head of an engine equipped with a valve operating system according to an embodiment of the present invention.

FIG. 2 is a cross-sectional view showing a configuration of a phase varying mechanism of the variable valve operating device according to the embodiment of the present invention.

FIG. 3 is a cross-sectional view showing a configuration of a variable lift mechanism of the variable valve operating device according to the embodiment of the present invention.

FIG. 4 is a layout diagram of a hydraulic oil supply passage for the variable valve operating device according to the embodiment of the present invention.

FIG. 5 is a diagram showing simulation data of a change in volume efficiency with respect to a valve timing retard amount during idle operation.

FIG. 6 is a diagram showing simulation data showing a change in pumping loss with respect to a valve timing retard amount during idle operation.

FIG. 7 is a diagram showing simulation data of changes in the combustion chamber temperature with respect to the valve timing retard amount during idle operation.

FIG. 8 is a diagram showing simulation data of changes in the residual gas amount with respect to the valve timing retard amount during idle operation.

[Fig. 9] P during idle operation with reference valve timing
It is a -V diagram.

FIG. 10 is a P-V diagram during idle operation of the engine according to the embodiment of the present invention.

[Explanation of symbols]

2 intake camshaft 3 exhaust camshaft 5 tappets (lift variable mechanism built-in) 6,7,8 cam 101,102 Hydraulic control valve 201,202 Phase variable mechanism A1, A2, A3, A4 cylinder Pin1, Pin2 intake port Pex1, Pex2 exhaust port

─────────────────────────────────────────────────── ─── Continuation of front page (51) Int.Cl. ⁷ Identification code FI theme code (reference) F02D 41/02 301 F02D 41/02 301H 320 320 320 41/04 305 41/04 305C 45/00 312 45/00 312H F Term (Reference) 3G016 AA08 AA19 BA03 BA06 BA19 BA28 BB04 BB31 BB40 DA06 DA18 DA22 GA00 3G018 AB07 AB17 BA22 BA33 CA18 DA17 DA28 DA57 DA73 DA83 EA03 EA04 EA12 EA13 FA02 GA23 CA02 FA04 GA23 CA02 FA04 GA23 CA02 FA04 GA23 CA02 FA04 GA23 CA02 FA04 AA11 BA01 CB02 DA01 DA02 DA03 DA10 DA11 DA12 DF04 DG05 FA24 GA04 GA05 GA06 GA17 GA18 HA11Z HE01Z 3G301 HA19 JA02 KA07 KA08 KA09 KA24 KA25 LA07 PA17Z PE01Z

Claims (7)

[Claims] 1. A valve operating system for an engine, comprising a variable valve operating device for varying at least opening and closing timings of an intake valve and an exhaust valve, wherein a low rotation and low load operating range is controlled by the variable valve operating device. Then, the timing of opening the intake valve is set to a timing that is a first predetermined amount after the intake top dead center, and the timing of closing the exhaust valve is a timing that is set to a second predetermined amount before the intake top dead center.
Further, the engine valve control device is characterized in that the second predetermined amount is set to be larger than the first predetermined amount. 2. Each cylinder is provided with two intake valves, and the variable valve operating device is provided with a phase variable mechanism for varying at least the opening / closing timing of the first intake valve and at least the lift characteristic of the second intake valve. A lift variable mechanism capable of changing to a low lift characteristic including a substantial intake stop is provided, and in the low rotation / low load operation range, the phase variable mechanism allows the first variable
The timing of opening the intake valve is the timing after the first predetermined amount from the intake top dead center, and the second intake valve is made to have a low lift characteristic including the substantial opening stop by the lift variable mechanism. The valve operating control system for the engine according to claim 1. 3. The valve operating system for an engine according to claim 2, wherein the air-fuel ratio of the engine is set leaner than the stoichiometric air-fuel ratio in the low rotation and low load operation range. 4. In the low-rotation medium-load operating range, an overlap is provided between the period in which the first intake valve is open and the period in which the exhaust valve is open, and the second intake valve is stopped substantially by the open valve. 3. The engine valve control device according to claim 2, wherein the engine valve control device includes a low lift characteristic. 5. The valve operating control system for the engine according to claim 2, wherein the timing of closing the two intake valves in the low rotation and high load operation range is set earlier than in the low rotation and low load operation range. 6. In a high-speed low-load operating range, the timing of closing the two intake valves is set earlier than in the low-speed low-load operating range, and the intake valve is open and the exhaust valve is open. The valve operating control system for the engine according to claim 2, wherein the valve is provided with an overlap. 7. A valve operating control system for an engine according to claim 5, wherein in the high load operation range, the timing of closing the two intake valves is delayed with an increase in the engine speed.