JP2003307139A - Valve train for internal combustion engine - Google Patents

Abstract

(57) [Summary] [Purpose] To achieve a good combustion state in a low-speed, medium-load region, reduce pump loss and improve fuel efficiency. The present invention relates to a valve train including a valve lift control mechanism and a valve timing control mechanism, wherein a control unit first detects an engine operation state based on an information signal input from each sensor in section 1; 2
It is determined whether or not the engine water temperature is equal to or higher than a predetermined value. If it is equal to or more than the predetermined value, it is determined in section 3 whether the engine speed is equal to or less than the predetermined value. If the value is equal to or less than the predetermined value, it is determined in the section 4 that the current time is in the middle load range.
Thereby, the closing timing of the intake valve is controlled to be sufficiently before the bottom dead center.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a valve operating system for an internal combustion engine in which valve lift characteristics are switched while variably controlling valve timing.

[0002]

2. Description of the Related Art Conventionally, an internal combustion engine for an automobile has a lift of an intake valve or an exhaust valve depending on an operating state for the purpose of both improving fuel efficiency during low-medium speed operation and improving output torque during high-speed operation. It is known that a valve lift control mechanism is provided which controls the intake / exhaust timing or the intake / exhaust amount by varying the characteristics (for example, Japanese Patent Laid-Open No. 62-12).
1811, etc.).

This is a common rocker in which a rocking arm for low speed whose rocking tip abuts an intake valve, for example, and a rocker arm for high speed which is adjacent to one side of the low speed rocker arm and has no contact portion with the intake valve. It is swingably supported by the shaft. Further, a low speed cam is in sliding contact with the low speed rocker arm, and a high speed cam having a profile in which the valve opening angle or the valve lift amount is larger than the low speed cam is in sliding contact with the low speed rocker arm.

Further, each rocker arm is provided with a connection switching means composed of a plunger, a guide hole or the like for integrally connecting or disconnecting the rocker arms.

Then, the connection switching means is controlled on the basis of the output signal from the controller according to the current engine operating state, and when the engine is at a low rotation speed, the connection of each rocker arm is released, as shown by the solid line in FIG. The valve lift characteristic (small operating angle) on the low speed side is set, and at the time of high rotation, the rocker arms are integrally connected to selectively switch to the valve lift characteristic (large operating angle) on the high speed side as shown by the alternate long and short dash line in FIG. It is like this. As a result, the valve lift of the intake valve is reduced and the valve closing timing is controlled to be earlier than the bottom dead center when the engine speed is low, and mechanical loss such as engine pump loss and friction is minimized. While improving the fuel efficiency and the like, at the time of high rotation, the intake valve charging efficiency is increased by increasing the valve lift amount of the intake valve and advancing the valve opening timing to ensure sufficient output.

[0006]

However, in the above-mentioned conventional apparatus, the valve lift control mechanism simply converts the valve lift characteristic into a fixed small operating angle when the engine is running at low speed to open the intake valve. The timing is retarded (retarded) and the valve closing timing is advanced (advanced). Therefore, control is not performed in accordance with changes in the load in addition to various engine operating states, that is, the number of revolutions, and it is difficult to sufficiently exhibit engine performance.

That is, for example, in the low rotation and low load region, although the effect of improving the combustion and reducing the pump loss can be obtained by the small operating angle control as described above, when the operating angle is controlled to the large operating angle at the high rotation speed, In comparison, the intake charging efficiencies are remarkably different from each other, so that a large output change, that is, a large torque shock occurs at the time of switching between the large and small operating angles. As a result, driving becomes unstable. Therefore, the small operating angle control by the valve lift control mechanism naturally limits the control for improving the fuel efficiency by simply reducing the overlap with the exhaust valve or advancing the closing timing of the intake valve. Will end up.

In particular, in the low-rotation medium-load range, the valve closing timing of the intake valve can be advanced more than the valve closing timing at the large operating angle by the small operating angle control, but the advance amount is limited. However, it is only slightly advanced from the bottom dead center position, and cannot be fully advanced. Therefore, there is a problem that the effect of reducing the pump loss, that is, the improvement of fuel efficiency becomes insufficient. That is, the pump loss is as shown in FIG.
As shown in (4), since the intake valve closing time is maximized when the valve closing timing is near the bottom dead center (BDC), the pump loss cannot be sufficiently reduced with a limited advance amount.

[0009]

The present invention has been devised in view of the above problems of the prior art. The invention of claim 1 is based on an output signal from a controller for detecting an engine operating state. A valve lift control mechanism that switches the valve lift characteristic of the valve to a small valve operating angle in the low engine speed range and a large valve operating angle in the high engine speed range, and the phase of the valve operating angle based on the output signal from the controller. A valve operating system including a valve timing control mechanism for converting the opening / closing timing of an intake valve to an advance side or a retard side by converting the intake valve to a small operating angle by the valve lift control mechanism at low engine speed. This phase is converted to the advanced side by the valve timing control mechanism when the engine is under medium load, so that the intake valve closing timing is set before the bottom dead center position, and The high load is characterized by converting retarded.

According to a second aspect of the present invention, the closing timing of the intake valve is set to be near the bottom dead center when the valve lift control mechanism controls the operating angle to a small angle during low engine speed. It is characterized by that.

[0011]

According to the first aspect of the present invention, when the engine is operating at low engine speed and low load, the phase of the small operating angle is controlled to the advance side so that the intake valve closing timing comes before the bottom dead center position. Since it is set to, it becomes possible to reduce the pump loss of the engine.
At the same time, the opening timing of the intake valve is also advanced, so that the overlap with the exhaust valve is increased, and the proportion of high-temperature residual gas in the cylinder is increased, so that the intake air-fuel mixture can be warmed. Therefore, as described above, even if the effective compression ratio is lowered by the large advance control of the closing timing of the intake valve, the temperature of the air-fuel mixture at the start of combustion (the end of the compression stroke) rises and good combustion can be obtained.

That is, if the closing timing of the intake valve is set earlier than the bottom dead center in the medium load range, the effective compression ratio is lowered and the temperature of the air-fuel mixture after compression is lowered, which may cause deterioration of combustion. There is. However, in this operating region, the intake air-fuel mixture amount is large and the heat generation amount due to combustion is also high, so that the wall temperature of the combustion chamber, the residual gas temperature, etc. rise and the air-fuel mixture temperature after compression also rises, so that the deterioration of combustion It can be suppressed sufficiently. Therefore, it is possible to reduce the pump loss by positively advancing the closing timing of the intake valve. In other words, the priority is given to advancing the closing timing of the intake valve to improve the fuel consumption by reducing the pump loss.

According to the second aspect of the present invention, the compression ratio is increased because the intake valve closing timing is set near the bottom dead center when the engine is in the vicinity of idling or when the engine is operating at low speed and low load such as during idling. be able to. Therefore, it is possible to raise the combustion temperature at the end of the compression stroke and obtain a good combustion state. In other words, under operating conditions near idling, the flow rate of the air-fuel mixture is throttled, so the amount of heat generated by combustion is reduced, and since the wall temperature of the combustion chamber is also low, the combustion speed is slow and combustion tends to become unstable, Since the compression ratio can be increased as described above, combustion can be improved, and fuel consumption can be improved and rotation can be stabilized.

[0014]

Embodiments of the present invention will now be described in detail with reference to the drawings. In this embodiment, the one applied to an internal combustion engine having two intake valves per cylinder is shown.

FIG. 1 is a schematic view showing the overall construction of this embodiment, in which A is a valve lift control mechanism provided on the camshaft 11, and B is a valve timing provided on one end of the camshaft 11. The control mechanism C is a hydraulic switching valve D according to the current engine operating state and the oil / water temperature detected based on output signals from the crank angle sensor F, the air flow meter G, the oil / water temperature sensor H, the throttle valve opening sensor, and the like. , E, the hydraulic pressure switching valve D switches the control hydraulic pressure of the valve lift control mechanism A between a low speed side and a high speed side, while the hydraulic pressure switching valve E is
The supply of hydraulic pressure to the pressure chamber described later of the valve timing control mechanism B is switched ON / OFF.

The valve lift control mechanism A is constructed as shown in FIGS. That is, 2 for each cylinder
A single main rocker arm 1 corresponding to the intake valves 3 and 3 of the book is provided, and the main rocker arm 1 swings to the cylinder head via a main rocker shaft 4 common to each cylinder. While being freely supported,
The tip portion 1b is in contact with the stem top portions of the intake valves 3 and 3. The main rocker arm 1 has a substantially rectangular shape in plan view, and a rectangular opening 5 is cut out in the longitudinal direction of one side portion, and the opening 5 is also formed in the longitudinal direction of the other side portion.
A rectangular hole 6 having a substantially rectangular shape having a larger area than that of the rectangular hole 6 is cut out, and a cutout window 7 for exposing the rectangular hole 6 to the outside is formed on the front end side of the outer wall 1C of the rectangular hole 6. .
As shown in FIG. 5, a roller 10 is rotatably provided on the shaft 8 via a needle bearing 9 in the opening 5, and a sub-rocker arm 2 is arranged in the rectangular hole 6. The roller 10 is in contact with a low speed cam 12 provided on a camshaft 11 that rotates in synchronization with a crankshaft (not shown).

As shown in FIG. 4, the sub-rocker arm 2 has its base end supported by the main rocker arm 2 via a sub-rocker shaft 13 so as to be relatively swingable, and contacts the intake valve 3. And a cam follower portion 15 that is in sliding contact with the high-speed cam 14 that is juxtaposed with the low-speed cam 12 is formed at the tip thereof in an arc shape.
A coil-shaped lost motion spring 16 for pressing the cam follower portion 15 against the high speed cam 14 is interposed below the cam follower portion 15. The sub rocker shaft 13
Is slidably inserted into an insertion hole 2a formed inside the base end of the sub-rocker arm 2 and has both end portions 13 thereof.
a and 13b are press-fitted and fixed in press-fitting holes 17 and 17 which are formed in the base end portion 1a at positions facing both sides of the rectangular hole 6.

Further, as shown in FIG. 4, the main rocker arm 1 has a cylindrical recess 1 which is located immediately below the sub rocker arm 2 and which holds the lost motion spring 16.
8 is integrally formed. The lost motion spring 16 has a lower end seated on the bottom plate 18a of the recess 18, and an upper end thereof is integrally formed with the sub-rocker arm 2 via a retainer 19 slidably fitted in the recess 18.
Is pressing.

Reference numeral 21 in the drawing denotes a connection switching means for connecting and disconnecting the main rocker arm 1 and the sub rocker arm 2 as appropriate, and the connection switching means 21 is constructed as shown in FIGS. 2 and 3. That is, in the side portion of the roller 10 of the main rocker arm 1, a bottomed cylindrical first guide hole 23 is formed in the width direction, and a cylindrical short piston 22 is slidably held therein. Together with the piston 2
An oil chamber 24 is defined behind the 2. On the other hand, the sub-rocker arm 1 is formed with a second guide hole 25 that is coaxial with the first guide hole 23 and has the same diameter. The second guide hole 25 has a return end supported at one end by a retainer 28. A plunger 27 for urging the piston 22 toward the oil chamber 24 via a spring 26 is housed.

The piston 22 is fitted over the first and second guide holes 22 and 25 by the hydraulic pressure introduced into the oil chamber 24 so that the main rocker arm 1 and the sub rocker arm 2 are integrally connected. Has become. On the other hand, when the operating oil pressure is not introduced into the oil chamber 24, the piston 22 is pushed toward the oil chamber 24 via the plunger 27 by the spring force of the return spring 26 and is housed in the first guide hole 23. The rocker arms 1 and 2 are disconnected from each other.

The hydraulic circuit 29 for guiding the working hydraulic pressure to the oil chamber 24 has a main rocker shaft 4 as shown in FIG.
The oil gallery 30 is formed in the inner axial direction of the above, and an oil passage 31 that communicates the oil chamber 24 and the oil gallery 30 through the radial direction of the main rocker shaft 4 and the inside of the main rocker arm 1.

The oil pressure discharged from the oil pump is guided to the oil gallery 30 through the oil pressure switching valve D during a predetermined high speed operation. The hydraulic switching valve D is a 3-port 2-position solenoid valve, and its operation is controlled by a control signal from the control unit C.

The low-speed cam 12 and the high-speed cam 14 adjacent to the low-speed cam 12 are integrally formed on a common cam shaft 11 so as to satisfy the valve lift characteristics required at low engine speed and high engine speed. Are formed in different shapes (including similar shapes having different sizes). That is, the high speed cam 14 has a profile that makes both the valve lift amount and the valve opening period larger than the low speed cam 12.

Therefore, this valve lift control mechanism A
According to the above, during low engine speed operation, the main rocker arm 1 swings according to the profile of the low speed cam 12 to drive each intake valve 3 to open and close, and the valve opening angle and lift amount are both set as shown by X1 in FIG. Get smaller.

At this time, although the sub-rocker arm 2 is swung by the high speed cam 14, the pistons 22 and the plungers 27 are housed in the guide holes 23 and 25, respectively, by the urging force of the return spring 26, and the main rocker arm 2 is moved. It does not prevent the movement of 1.

On the other hand, when the engine is operating at high speed (approximately 4
000 to 5000 rpm), the operating oil pressure fed from the oil pump 32 is applied to the oil gallery 30 and the oil passage 3.
When guided to the oil chamber 24 via 1, the piston 22 and the plunger 27 move against the return spring 26, and the piston 22 is fitted over the guide holes 23 and 25. As a result, both rocker arms 1 and 2 swing together. Here, since the high-speed cam 14 is formed so that both the valve opening angle and the lift amount are larger than the low-speed cam 12, the sub-rocker arm 2
Roller 1 of the main rocker arm 1 when swinging together with
0 floats from the low speed cam 12, each intake valve 3 is opened and closed according to the profile of the high speed cam 14, and the valve opening angle and the lift amount both increase as indicated by X2 in FIG.

On the other hand, when the operating state of the engine shifts from the high speed region to the low speed region again, the hydraulic pressure switching valve D is actuated to cause the oil chamber 2 to move.
The hydraulic pressure guided to 4 decreases, the elastic restoring force of the return spring 26 moves the piston 22 and the plunger 27 to their original positions, and the restraint of the main rocker arm 1 is released.

As a result, as shown in FIG. 9, a valve lift characteristic (small operating angle) X1 based on the profile of the low speed cam 12 and a valve lift characteristic (large operating angle) X2 based on the profile of the high speed cam 14 are obtained. Is synthesized. The closing timing of each intake valve 3 in the low-speed valve lift characteristic X1 is set near the bottom dead center.

On the other hand, the valve timing control mechanism B
As shown in FIG. 6, the bracket 4 is attached to the cylinder head.
It is provided between a cam shaft 11 which is rotatably supported by a shaft 0 and a sprocket 41 to which a driving force is transmitted from a crank shaft, and a front end portion 11 of the cam shaft 11 is provided.
A sleeve 42 is provided on a by a mounting bolt 43. Outer teeth 42a are formed on the outer periphery of the sleeve 42, and the outer peripheral surface of the flange portion 44 at the end portion rotatably supports the sprocket 41.

In the sprocket 41, inner teeth 45a are formed on the inner circumference of a cylindrical main body 45 fitted on the outer circumference of a sleeve 42, and the outer end opening of the cylindrical main body 45 has an annular cover portion. It is blocked by 46. Also,
A tubular gear 47, which is movable in the axial direction, is interposed between the tubular body 45 and the sleeve 42.

This cylindrical gear 47 is composed of two front and rear gear components, and the inner and outer teeth 47 of the outer and outer teeth 42a and 45a of the inner and outer surfaces thereof are both helical and inner and outer teeth 47, respectively.
a and 47b are formed. Furthermore, this cylindrical gear 4
7 is biased forward by the spring force of the compression spring 48 elastically held between the rear gear forming portion and the flange portion 44 until the front gear forming portion abuts the cover portion 46, and the cover portion By the hydraulic pressure in the pressure chamber 49 formed between 46 and the front gear component, the rear gear component is moved rearward until it abuts the flange portion 44.

The hydraulic pressure from the oil pump 32 is supplied to and discharged from the pressure chamber 49 by the hydraulic pressure switching valve E via the hydraulic circuit 50. The hydraulic pressure switching valve E
Is the same as the hydraulic switching valve D of the valve lift control mechanism A.
It is composed of a port 2 position type solenoid valve, and its operation is controlled by an output signal from the control unit C.

The control unit C controls ON / OFF of the hydraulic pressure switching valve E by using not only the engine speed but also the load and cooling water temperature as control elements, and the valve lift control in the engine low speed range. The hydraulic pressure switching valve E is controlled when the mechanism A is performing the small operating angle X1 control.

The control of the hydraulic pressure switching valve E by the control unit C will be described below with reference to the flowchart of FIG.

First, in Section 1, the current engine operating state is detected based on the information signal from each sensor as described above. In section 2, it is determined whether or not the current cooling water temperature T is equal to or higher than the predetermined value TW. In this section 3, it is determined based on the signal from the crank angle sensor F whether or not the current engine speed is equal to or lower than a predetermined speed (for example, about 3,500 rpm). Transition. Here, as shown in the characteristic diagram of FIG. 8, it is determined whether or not the current engine load is in the medium load range based on the opening amount of the throttle valve. If it is determined that the load is in the medium load range, the process proceeds to section 5. In this section 5,
An ON signal (energization) is output to the solenoid of the hydraulic pressure switching valve E to open the supply passage of the hydraulic circuit 50. In this way, advance control of the opening / closing timing of each intake valve 3 described later is performed. As shown by X3 (broken line) in FIG. 9, the advance amount is set such that the closing timing of the intake valve 3 is set sufficiently before the bottom dead center position, and the opening timing is the opening time when the high-speed side valve lift characteristic is X2. It is set to a position advanced by α from the valve timing.

On the other hand, when the cooling water temperature is judged to be below a predetermined value (for example, 60 ° C. or less) in the section 2, the process proceeds to section 6, in which the solenoid of the hydraulic switching valve E is turned on.
An FF signal (non-energized) is output to close the supply passage of the hydraulic circuit 50. When it is determined in Section 3 that the engine speed is equal to or higher than the predetermined engine speed, and when it is determined in Section 4 that the engine is in a low load or a high load range other than the medium load, the process proceeds to Section 6 and the hydraulic switching valve is operated. An OFF signal is output to E to close the supply passage. As a result, the retard control is performed without performing the advance control of the opening / closing timing of the intake valve 3.

That is, during idling or in the low speed and low load region near idling, the supply passage is closed, so that the hydraulic pressure is not supplied from the oil pump 32 to the pressure chamber 49. Therefore, as shown in FIG. 6, the tubular gear 47 is biased to the maximum forward position (illustrated position) by the spring force of the compression spring 48, and moves the camshaft 11 in the direction opposite to the rotational direction with respect to the sprocket 41. Rotate relatively. Accordingly, the valve lift characteristic X3 controlled to the small operating angle by the valve lift control mechanism A is controlled to the delay side as shown in FIG. 9X1.

Therefore, since the closing timing Y1 of each intake valve 3 is closer to the bottom dead center than the conventional one, the compression ratio can be increased. Therefore, it is possible to raise the combustion temperature at the end of the compression stroke and obtain a good combustion state. In other words, under operating conditions near idling, the flow rate of the air-fuel mixture is reduced, so the amount of heat generated by combustion is reduced, and
Since the wall temperature of the combustion chamber is also low, the combustion speed is slow and combustion tends to become unstable.However, since the compression ratio can be increased as described above, combustion is improved, fuel consumption is improved and rotation is stabilized. Can be achieved.

Further, since the opening timing of the intake valves 3 and 3 is delayed, the overlap with the exhaust valve is reduced, and the amount of burned gas that flows back from the exhaust port into the combustion chamber and remains can be reduced. Therefore, also in this respect, the combustion can be improved.

On the other hand, in the low-rotation medium-load range, the supply passage is opened as described above, so that the working hydraulic pressure is supplied to the pressure chamber 49 and becomes high. Therefore, the cylindrical gear 47
Moves to the maximum rearward position (rightward in the figure) against the spring force of the compression spring 48, and relatively rotates the camshaft 11 in the same direction as the rotation direction. Therefore, as described above, the valve lift is controlled to the advance side while maintaining the small operating angle control state as indicated by X3 (broken line) in FIG. 9, and the valve closing timing and the valve opening timing of each intake valve 3 are controlled. You can get it fast enough. Therefore, it is possible to reduce the pump loss while ensuring a good combustion state.

That is, if the closing timing of each intake valve 3 is advanced in such a medium load range, the effective compression ratio is lowered and the temperature of the air-fuel mixture after compression is lowered, possibly causing deterioration of combustion. . However, in this operating region, the intake air-fuel mixture amount is large and the heat generation amount due to combustion is also high, so that the wall temperature of the combustion chamber, the residual gas temperature, etc. rise and the air-fuel mixture temperature after compression also rises, so that the deterioration of combustion It can be suppressed sufficiently. Therefore, each intake valve 3
It is possible to reduce the pump loss by advancing the valve closing timing of. In other words, priority is given to advancing the closing timing of each intake valve 3 to improve fuel efficiency by reducing pump loss.

However, if the closing timing of each intake valve 3 is advanced too early, the compression temperature will eventually drop and combustion will be deteriorated.
Therefore, as described above, the opening timing of the intake valve 3 is also advanced,
Control was performed to increase the overlap with the exhaust valve.
As a result, the proportion of residual gas in the combustion chamber can be increased and the air-fuel mixture can be warmed by the heat of the residual gas, so even if the effective compression ratio decreases due to the earlier valve closing timing, the combustion state at the start of combustion Is improved, and it is possible to reduce pump loss while preventing deterioration of combustion.

Further, since the residual gas is increased, the negative pressure during the intake stroke in the cylinder is reduced (the pressure is increased) correspondingly, and the pump loss can be reduced also from this point.

Next, in the low rotation and high load region, the same control as in the low rotation and no load region is performed, and the closing timing of the intake valve 3 is delayed so as to be near the bottom dead center, so that the effective intake stroke is lengthened. As a result, the intake charging efficiency can be improved and the output torque can be increased.

Further, since the opening timing of each intake valve 3 is also delayed, the overlap with the exhaust valve is reduced and knocking can be prevented by reducing the residual gas.

Moreover, in the present embodiment, as described above, the advance angle control in the low rotation medium load range is not performed when the cooling water temperature TW is equal to or lower than a predetermined value (for example, 60 ° C.), so that the combustion is deteriorated. Can be prevented.

Further, since the opening timing of each intake valve 3 in the advance control is advanced more than the opening timing of the high speed valve lift, the effect of the early closing of each intake valve 3 and the large overlap is larger than that of the conventional one. Can be even larger. That is, in the related art, the structural point (the main rocker arm 1 and the sub-rocker arm 2 of the valve lift control mechanism A rotate at high speed,
It is not possible to extend the low-speed valve lift to the front and back of the high-speed valve lift from the structure that constantly swings regardless of the low-speed operation), but in this embodiment, switching between the high-low valve lift and phase conversion is performed. Can be controlled separately, the advance angle control as described above is possible,
It is possible to obtain a higher level of effect.

In the low-rotation medium-load region, the pump is controlled by controlling the high-speed valve lift characteristic X2 shown in FIG. 9 to delay the closing timing of each intake valve 3 further than the intake valve closing timing of the normal engine. In addition to the effect of reducing loss, the valve opening timing of the intake valve 3 is on the advance side of the valve closing timing of the normal engine, and the amount of valve overlap with the exhaust valve becomes large and the heating effect of the air-fuel mixture due to residual gas Although it is further improved, since the valve lift amount is increased, the valve drive loss, that is, the sliding friction resistance is increased, which may reduce the output torque.

On the other hand, in this embodiment, since the low speed valve lift characteristic X3 is selected, the air-fuel mixture is throttled by the small openings of each intake valve 3, the swirl in the combustion chamber is strengthened, and combustion is good. At the same time, a small valve lift reduces the valve drive loss and improves the output torque. Further, advancing the closing timing of each intake valve 3 has various advantages that the vaporization of the fuel is promoted and the combustion is improved because the negative pressure is temporarily increased due to the adiabatic expansion effect in the cylinder.

FIG. 10 shows a valve lift characteristic in which the present invention is applied to an exhaust valve. A one-dot chain line shows a high-speed valve lift characteristic XE3, and a solid line shows a low-speed valve lift characteristic XE.
4, the two-dot chain line indicates the low valve lift characteristic X on the advance side which controls the closing timing of the valve lift characteristic for low speed to the advance side (advance side).
E1, the broken line is the retarding side low valve lift characteristic XE in which the valve closing timing of the low speed valve lift characteristic is controlled to the retarding side (delay side).
2 are shown respectively. That is, the low speed and low load range (valve lift control) of the engine near idling or idling.
Then, the mechanism A switches to the low speed valve lift characteristic XE4 based on the profile of the low speed cam, and the valve timing control mechanism B further changes the XE4 to X.
As shown by E1, control is performed on the advance side (advance side). Therefore, since the exhaust valve opens near the bottom dead center position and closes near the top dead center position, the expansion stroke becomes longer and the overlap amount with the intake valve 3 becomes smaller, which prevents deterioration of combustion. As a result, fuel efficiency can be improved.

Next, in the low engine speed and low load region of the engine, the low speed valve lift characteristic XE4 is set to the valve timing control mechanism B.
Is controlled to the retard side (lag side) as indicated by XE2. Therefore, the valve closing timing of the exhaust valve is determined by the valve lift characteristic XE3 for high speed based on the cam profile of the high speed cam.
Since the valve closing timing can be delayed, the action of sucking back the burned gas (residual gas) that has flowed back into the intake pipe into the combustion chamber becomes large. Therefore, the air-fuel mixture is warmed by the heating action of the residual gas, and the combustion efficiency is improved.

By advancing the closing timing of the exhaust valve to the point before the top dead center (advancing side), the heating operation of the residual gas can be performed. That is, in this case, burned gas (residual gas) is confined in the combustion chamber (see FIG. 10).
XE1). With the containment action and the suction return action described above, exhaust gas having a high HC concentration at the end of the exhaust stroke of the engine can be contained or sucked back into the combustion chamber and re-combusted in the next stroke, so HC can be reduced. In addition, the small valve lift reduces the valve drive loss,
The output torque can be improved.

The present invention is not limited to the above embodiment, and the engine temperature as a control element may be a lubricating oil temperature other than the cooling water temperature. Further, the engine speed upper limit value of the control element can be arbitrarily set according to the specifications of the engine and the like.

[0054]

As is apparent from the above description, claim 1
According to the invention, the closing timing of the intake valve is set to be before the bottom dead center in the low rotation medium load range, so that the pump loss can be reduced while ensuring a good combustion state. As a result, it is possible to stabilize drivability and greatly improve fuel efficiency.

Further, according to the second aspect of the present invention, the compression ratio can be increased by setting the closing timing of the intake valve near the bottom dead center in the low rotation and low load region, and therefore, at the end of the compression stroke. It becomes possible to raise the combustion temperature and obtain a good combustion state. As a result, it is possible to improve fuel efficiency and stabilize engine rotation.

[Brief description of drawings]

FIG. 1 is an overall configuration diagram showing an embodiment of the present invention.

FIG. 2 is a plan view showing a valve lift control mechanism of this embodiment.

FIG. 3 is a sectional view taken along line I-I of FIG.

FIG. 4 is a sectional view taken along line JJ of FIG.

5 is a sectional view taken along line KK of FIG.

FIG. 6 is a sectional view showing a valve timing control mechanism of the present embodiment.

FIG. 7 is a flowchart showing a control mode of this embodiment.

FIG. 8 is a characteristic diagram showing the relationship between throttle opening and load.

FIG. 9 is a diagram showing valve lift and valve timing characteristics in the present embodiment.

FIG. 10 is a diagram showing valve lift and valve timing characteristics in the embodiment applied to the exhaust valve side.

FIG. 11 is a conventional valve lift characteristic diagram.

FIG. 12 is an explanatory diagram showing the relationship between crank angle and pump loss.

[Explanation of symbols]

A: Valve lift control mechanism B: Valve timing control mechanism C ... Control unit (controller) D / E ... hydraulic switching valve 3 ... Intake valve 11 ... Cam shaft

─────────────────────────────────────────────────── ─── Continuation of front page (51) Int.Cl. ⁷ Identification code FI theme code (reference) F02D 41/04 320 F02D 41/04 320 45/00 312 45/00 312H F term (reference) 3G018 AA01 AB04 BA12 CB02 CB06 DA14 DA18 DA28 DA83 EA00 EA03 EA04 EA14 EA35 FA01 FA03 FA06 FA07 GA07 3G084 BA23 CA03 CA04 CA09 DA02 FA18 FA33 3G092 AA11 BA01 DA01 DA04 KA03 KA04 JA25 KA04 JA01 GA02 GA01 GA01 GA17 GA18 GA11 HA11 HA18 GA11 HA18 HA11 LA07 LC08 NC02 NE11 NE12 PA17Z PE01Z

Claims (2)

[Claims] 1. A valve lift for switching a valve lift characteristic of an intake valve to a small valve operating angle in a low engine speed range and a large valve operating angle in a high engine speed range based on an output signal from a controller that detects an engine operating state. A valve operating device comprising a control mechanism and a valve timing control mechanism for converting the phase of the valve operating angle based on an output signal from the controller to switch the opening / closing timing of the intake valve to an advance side or a retard side. Therefore, by changing the phase controlled to a small operating angle by the valve lift control mechanism at the time of low engine speed to the advance side by the valve timing control mechanism at the time of medium load of the engine, the closing timing of the intake valve can be changed. A valve operating system for an internal combustion engine, which is set before the bottom dead center position and is converted to a retard side when the engine is under high load. 2. The valve closing timing of the intake valve is set to be near the bottom dead center when the valve lift control mechanism controls the operating angle to a small angle when the engine is running at low speed. 1. A valve train for an internal combustion engine according to item 1.