JP2015178799A - Variable valve device for internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a variable valve device for an internal combustion engine which can improve exhaust emission performance by improving combustion in cold start for example.SOLUTION: A variable valve device for an internal combustion engine includes: an intake VVL 3 for making an operating angle of intake valves 1, 1 variable; and an exhaust VVL 4 for making an operating angle of exhaust valves 2, 2 variable. The intake VVL mechanically holds each intake valve with a cam profile of a small lift cam 8 toward a small operating angle and a small lift side, when discharge oil pressure as conversion drive force does not operate, which comes from an oil pump 19 through a first solenoid selector valve 20. On the other hand, the exhaust VVL mechanically holds each exhaust valve according to a cam profile of a large lift cam 31 toward a large operating angle and a large lift side, when discharge oil pressure does not operate, which comes from an oil pump through a second solenoid selector valve 47.

Description

  The present invention relates to a variable valve operating apparatus for an internal combustion engine that can improve the fuel efficiency by improving the combustibility at the time of cold start of an internal combustion engine for an automobile, for example.

  As a conventional variable valve operating apparatus for an internal combustion engine, one described in Patent Document 1 below is known.

  This variable valve operating apparatus includes an intake operation angle variable mechanism that makes the operation angle of the intake valve variable, and an exhaust operation angle variable mechanism that makes the operation angle of the exhaust valve variable.

  The intake operating angle variable mechanism mechanically controls the intake valve to the small operating angle side when the hydraulic pressure that is the conversion driving force does not act when the engine is stopped, for example, and the exhaust operating angle variable mechanism is also converted and driven. When the hydraulic pressure is not applied, the exhaust valve is mechanically controlled to the small operating angle side.

Automotive technology Vol43, No8, P14-P19, published in 1989

  However, this conventional variable valve device is mechanically controlled to a small operating angle when the engine is stopped, so that both the intake valve and the exhaust valve are mechanically controlled. In addition, both the intake valve and the exhaust valve are controlled to a small operating angle.

  Here, considering the combustion cycle at the initial start of the engine, since the exhaust top dead center of the piston is shifted to the intake stroke and the piston starts to descend, the exhaust valve closes early, so the exhaust port It becomes difficult to reversely introduce high-temperature combustion gas (high-temperature EGR gas) from the side into the cylinder (combustion chamber). For this reason, the inside of the cylinder is difficult to warm and combustion becomes unstable, and there is a possibility that a lot of exhaust emissions such as HC occurs.

  The present invention has been devised in view of the above-described conventional technical problems. The intake operating angle variable mechanism mechanically controls the intake valve to the small operating angle side, and the exhaust operating angle variable mechanism is an exhaust valve. A variable valve operating apparatus for an internal combustion engine capable of improving combustion and improving fuel consumption by mechanically controlling the engine to a large operating angle side.

The invention according to claim 1 includes an intake operation angle variable mechanism that makes the operation angle of the intake valve variable, and an exhaust operation angle variable mechanism that makes the operation angle of the exhaust valve variable,
The intake operating angle variable mechanism mechanically holds the intake valve on the small operating angle side when the conversion driving force does not act, while the exhaust operating angle variable mechanism operates the exhaust valve when the conversion driving force does not operate. It is characterized by being mechanically held on the large operating angle side.

  In the present invention, the intake valve is mechanically stable at a small operating angle and the exhaust valve is at a large operating angle from the beginning of the start (default valve timing), exceeds the exhaust top dead center of the piston, enters the intake stroke, Since the exhaust valve does not close immediately even when the engine starts to descend, high-temperature combustion occurs from the exhaust port side through the exhaust valve during a relatively long period (first period) from the exhaust top dead center to the exhaust valve closing timing. A large amount of gas (high temperature EGR gas) can be directly flowed back into the cylinder.

  On the other hand, at this time, the intake valve is held at a small operating angle, and the period from the opening timing of the intake valve to the exhaust top dead center (second period) is relatively short, and therefore, the second half of the exhaust stroke in this second period. The high-temperature combustion gas (high-temperature EGR gas) is once returned to the cold intake port / intake pipe side through the intake valve, so that the cooled combustion gas (low-temperature EGR gas) is reintroduced into the cylinder in the next intake stroke. The amount introduced can be reduced.

  Thus, among the EGR gas introduced into the cylinder, the ratio of the high temperature EGR gas in the first period can be increased while the ratio of the low temperature EGR gas in the second period is kept small.

  In addition, since the intake valve is held at a small operating angle, the intake valve closing timing is advanced to the vicinity of the intake bottom dead center, and the effective compression ratio can be increased, which also contributes to combustion stabilization.

  Due to the effect of increasing the ratio of these high-temperature EGR gases and the effect of increasing the effective compression ratio, the effect of stabilizing start-up combustion and reducing exhaust emissions such as HC can be obtained.

  Furthermore, because the exhaust valve has a large operating angle, the exhaust valve opening time is relatively early, and the exhaust gas is discharged to the exhaust system before the temperature of the combustion gas is sufficiently lowered. The exhaust emission can be reduced from this aspect as well.

  Here, since the intake valve has a small operating angle and the exhaust valve has a large operating angle from the initial stage of cranking and starting, the above-described series of effects can be surely obtained from the initial stage of starting combustion.

  In addition, even if a failure such as disconnection of the electric system occurs, it is mechanically stable at the default valve timing described above, and even at the time of the failure, it has the same start combustion stabilization effect and emission reduction effect. That is, it also has a mechanical fail-safe function.

  According to the present invention, it is possible to improve exhaust emission performance by improving combustion at the time of cold start, for example.

It is the schematic of the variable valve apparatus of the internal combustion engine of 1st Embodiment of this invention. A is an operation explanatory view at the time of small lift control by the intake side lift variable mechanism (intake VVL), and B is an operation explanatory view at the time of large lift control by the same mechanism. A is an operation explanatory diagram at the time of small lift control by the exhaust side lift variable mechanism (exhaust VVL), and B is an operation explanatory diagram at the time of large lift control by the same mechanism. It is a valve lift characteristic view of an intake valve and an exhaust valve by each lift variable mechanism of this embodiment. It is a control flowchart figure by the controller with which this embodiment is provided. It is the sectional view on the AA line of FIG. 7 which shows the phase variable mechanism used for the exhaust side (intake | emission side) in 2nd Embodiment. It is a longitudinal cross-sectional view of the phase variable mechanism of this embodiment. It is a valve lift characteristic view of an intake valve and an exhaust valve in this embodiment. It is a control flowchart figure by the controller with which this embodiment is provided. It is the schematic of the variable valve apparatus of 3rd Embodiment. It is an operation explanatory view of intake VEL in this embodiment, A shows the minimum lift state of an intake valve, and B shows the maximum lift state. The operation state of the electric actuator provided to this embodiment is shown, A shows the operation state which controls the intake valve to the minimum lift, and B shows the operation state which controls the intake valve to the maximum lift. It is a valve lift characteristic view of an intake valve and an exhaust valve in this embodiment. It is a control flowchart figure by the controller with which this embodiment is provided.

Hereinafter, embodiments of a variable valve operating apparatus for an internal combustion engine according to the present invention will be described in detail with reference to the drawings. In each embodiment, the present invention is applied to a gasoline four-cycle multi-cylinder internal combustion engine.
[First Embodiment]
FIG. 1 shows a variable valve operating apparatus according to the first embodiment, and an intake side and an exhaust side are provided with two intake valves 1 and 1 and exhaust valves 2 and 2 per cylinder, respectively.

  Each of the intake valves 1, 1 and the exhaust valves 2, 2 has a variable operating angle and valve lift in stages according to the engine state. An intake VVL3 that is an intake operating angle variable mechanism to be controlled and an exhaust VVL4 that is an exhaust operating angle variable mechanism are provided.

  First, the intake VVL3 will be described. As shown in FIGS. 1 and 2, the intake VVL3 is integrally fixed to the intake camshaft 5 (exhaust camshaft 6) and is arranged at the center of each cylinder. An egg-shaped large lift cam 7 for lift, small lift cams 8 and 8 for small lifts provided on both sides of the large lift cam 7, and a rocker shaft 9 are swingably supported. 8, a pair of follower portions 10a, 10a is disposed, and the lower ends of the tip portions of the follower portions 10a, 10a are in contact with the stem ends of the intake valves 1, 1 (exhaust valves 2, 2). A follower portion 11 a is provided at a position corresponding to the main rocker arm 10 in contact with the large lift cam 7, and a sub rocker arm 11 capable of lost motion, and a sub rocker arm 11 provided in the sub rocker arm 11. The lower rocker arm 11 is swingably supported by a lost motion mechanism 12 for urging the sub rocker arm 11 toward the large lift cam 7 and a support shaft 13 fixed to the main rocker arm 10. The sub-rocker arm 11 and the main rocker arm 10 are interlocked with each other by being engaged / disengaged with the lever 11b, or a lever member 14 for releasing the interlock, a hydraulic plunger 15 for engaging / disengaging the lever member 14, and a main rocker arm 10 Return spring 16.

  The sub rocker arm 11 is swingably supported by a support shaft 23 provided at the rear end portion of the main rocker arm 10.

  The lost motion mechanism 12 is formed in the lower part of the sub rocker arm 11 and abuts against a convex part 10b formed in the central lower part of the main rocker arm 10, and urges the plunger 25 toward the convex part 10b. And a lost motion spring 24. When the lever member 14 is disengaged from the lower end step portion 11 b of the sub rocker arm 11, the sub rocker arm 11 is lost through the plunger 25 and the lost motion spring 24.

  The hydraulic plunger 15 moves forward and backward when oil pressure (conversion operating force) from the oil pump 19 shown in FIG. 1 is supplied to and discharged from an oil chamber 17 formed on the rear end side via the first electromagnetic switching valve 20. It is supposed to be.

  That is, the oil pressure is supplied to and discharged from the oil chamber 17 through the internal axial direction of the rocker shaft 9 and the hydraulic passages 18a and 18b formed in the main rocker arm 10. As shown in FIG. 1, the switching valve 20 switches the conduction between the hydraulic passages 18 a and 18 b and the drain passage 21 or the discharge passage 19 a of the oil pump 19.

  Therefore, when an ON signal is output (energized) to the first electromagnetic switching valve 20 and hydraulic pressure is supplied from the oil pump 19 to the oil chamber 17, the hydraulic plunger 15 moves forward as shown in FIG. 2B. Thus, the tip end portion of the lever member 14 engages with the step portion 11b. Thereby, each intake valve 1, 1 is controlled to a large lift as shown by LI <b> 2 in FIG. 4 according to the cam profile of the large lift cam 7 through the main rocker arm 10 by the swing of the sub rocker arm 11.

  On the other hand, when an off signal is output (non-energized) to the first electromagnetic switching valve 20 and no pump hydraulic pressure is supplied to the oil chamber 17, the hydraulic plunger 15 is moved to the spring of the return spring 16 as shown in FIG. The lever member 14 is moved backward by force, and the distal end portion of the lever member 14 is separated from the stepped portion 11b to be disengaged. As a result, the sub rocker arm 11 is brought into a lost motion state by the lost motion mechanism 12, so that the intake valves 1 and 1 are driven by the main rocker arm 10 according to the cam profiles having the same shape of both the small lift cams 8 and 8. It is controlled to a small lift as indicated by LI1.

Here, considering the case where the oil pressure (conversion driving force) of the oil pump 19 is not generated, such as when the engine is stopped, each intake valve 1, 1 is a machine regardless of the off position on position of the first electromagnetic switching valve 20. It is stable to small lift control (this is called the default position). In other words, the small lift operation is the default mode for each of the intake valves 1 and 1.
[Exhaust VVL]
Next, as described above, the exhaust VVL4 has the same basic configuration as the intake VVL3, but differs in the direction of action of the hydraulic pressure on the hydraulic plunger. In FIGS. Will be explained.

  A large lift cam 31 for large lift, which is integrally fixed to the exhaust camshaft 6 extending in the longitudinal direction of the engine and disposed in the center of the cylinder, and small lift small lifts provided on both sides of the large lift cam 31. A pair of follower parts 34a, 34a are disposed at positions corresponding to the lift cams 32, 32 and the rocker shaft 33 so as to be swingable, and corresponding to the small lift cams 32, 32, and the tip ends of the follower parts 34a, 34a. The main rocker arm 34 whose lower end is in contact with the stem ends of the exhaust valves 2 and 2, the sub rocker arm 35 that is provided at a position corresponding to the large lift cam 31 and capable of lost motion, and the sub rocker arm A lost motion mechanism which is provided in the inner surface 35 and biases the follower portion 35a of the sub rocker arm 35 toward the large lift cam 31; 6 and a support shaft 41 fixed to the main rocker arm 34 so as to be swingable. The sub rocker arm 35 and the main rocker arm 34 are synchronized with each other by being engaged with and disengaged from a step 35a at the lower end of the sub rocker arm 35. A lever member 38 for interlocking or releasing the interlock, a hydraulic plunger 39 for engaging and disengaging the lever member 38, and a return spring 40 provided in the main rocker arm 34 are provided.

  The sub rocker arm 35 is swingably supported by a support shaft 41 provided at the rear end portion of the main rocker arm 34.

  The lost motion mechanism 36 includes a plunger 42 that abuts the main rocker arm convex portion 34b, and a lost motion spring 43 that biases the plunger 42 toward the convex portion 34b.

  The hydraulic plunger 39 moves backward when oil pressure (conversion operating force) from the oil pump 19 is supplied to the oil chamber 44 formed on the outer periphery on the front end side via the second electromagnetic switching valve 47. 2 When the electromagnetic switching valve 47 is turned off (not energized) and no hydraulic pressure is supplied from the oil pump 19, it moves forward by the spring force of the coil spring 46 mounted inside.

  That is, hydraulic pressure is supplied to and discharged from the oil chamber 44 via the internal axial direction of the rocker shaft 33 and hydraulic passages 45 a and 45 b formed in the main rocker arm 34. As shown in FIG. 1, the electromagnetic switching valve 47 switches between the hydraulic passages 45 a and 45 b and the drain passage 48 or the discharge passage 19 a of the oil pump 19.

  Therefore, when the second electromagnetic switching valve 47 is turned on (energized) and hydraulic pressure is supplied to the oil chamber 17 from the oil pump 19, the hydraulic plunger 39 resists the spring force of the coil spring 46, as shown in FIG. As a result, the lever member 38 is disengaged from the stepped portion 35a. As a result, the sub rocker arm 35 is brought into a lost motion state by the lost motion mechanism 36, so that the exhaust valves 2 and 2 are driven by the main rocker arm 34 according to the cam profiles of the same shape of both the small lift cams 32 and 32 in FIG. It is controlled to a small lift as indicated by LE1.

  On the other hand, when an off signal is output (non-energized) to the second electromagnetic switching valve 47 and no hydraulic pressure is supplied from the oil pump 19 to the oil chamber 44, the hydraulic plunger 39 is turned into a coil spring as shown in FIG. 3B. The spring member 46 moves forward to press the lower end of the lever member 38 and engage the tip of the lever member 38 with the step 35a. Thus, the exhaust valves 2 and 2 are controlled to a large lift as shown by LE2 in FIG. 4 according to the cam profile of the large lift cam 31 via the main rocker arm 34 by the swing of the sub rocker arm 35.

  Here, as described above, when the hydraulic pressure of the oil pump 19 is not generated due to engine stop or the like, the exhaust valves 2 and 2 are mechanically large regardless of the OFF position ON position of the second electromagnetic switching valve 47. Stable to lift control (this is called the default position). That is, the large lift operation is the default mode for each of the exhaust valves 2 and 2, which is the reverse of the intake VVL3.

The controller (ECU) 22 is based on various sensors such as a crank angle sensor, an air flow meter, an engine water temperature sensor, an oil temperature sensor, and a throttle opening sensor that detects the opening of a throttle valve. An operation state is detected, and a control current is output to the first and second electromagnetic switching valves 20 and 47 in an on-off manner.
[Effects of this embodiment]
As described above, as shown in FIG. 4, the default mode is that each intake valve 1, 1 has a small lift (LI 1) and a small operating angle, and the exhaust valves 2, 2 have a large lift (LE 2), Large operating angle. Therefore, when the oil pump hydraulic pressure is hardly generated, such as when the engine is stopped or cranking, it is mechanically stable in this default mode.

  This default valve timing can stabilize the combustion during start-up (particularly in the case of cold start) that tends to be unstable, and can obtain the effect of reducing start-up emissions such as HC.

  Moreover, since this default valve timing has already been reached when the engine is stopped, this effect can be obtained from the very beginning of starting combustion.

  Since the exhaust valves 2 and 2 have a large operating angle DE2 (lift LE2), the exhaust valves 2 and 2 are not immediately closed even when the piston starts to descend after exceeding the exhaust top dead center of the piston. Therefore, in the relatively long period from the exhaust top dead center (TDC) to the exhaust valve closing timing (EVC2) (the period indicated by (1) -2 in FIG. 4 which is the first period), the exhaust is discharged from the exhaust port side. A large amount of high-temperature combustion gas (high-temperature EGR gas) can be directly introduced back into the cylinder through the valves 2 and 2. On the other hand, the intake valves 1 and 1 have a small operating angle (DI1) and a period from the intake valve opening timing (IVO1) to the exhaust top dead center (TDC) (second period (2) -1 in FIG. 4). Therefore, the high-temperature combustion gas (high-temperature EGR gas) in the latter half of the exhaust stroke is once returned to the cold intake port / intake pipe side through the intake valves 1 and 1 during this period. The amount by which the cooled combustion gas (low temperature EGR gas) is reintroduced into the cylinder in the next intake stroke can be reduced.

  Thus, among the EGR gas introduced into the cylinder, the ratio of the high temperature EGR gas in the first period can be increased while the ratio of the low temperature EGR gas in the second period is kept small.

  Further, since the exhaust valves 2 and 2 have a large operating angle, the opening timing of the exhaust valves 2 and 2 is relatively early (EVO2), and the exhaust gas is discharged to the exhaust system before the temperature is sufficiently lowered. Therefore, the downstream catalyst can be effectively warmed, the catalyst activity (conversion rate improvement) can be promoted, and the exhaust emission can be reduced from this aspect.

  Further, since each intake valve 1, 1 has a small operating angle DI1 (lift LI1), the closing timing (IVC1) of the intake valve 1, 1 has advanced to the vicinity of the intake bottom dead center (BDC), The effective compression ratio can be increased, which also contributes to combustion stabilization.

  The effect of increasing the ratio of the high temperature EGR gas and the effect of increasing the effective compression ratio stabilize the start-up combustion, and also add the effect of promoting the activity of the catalyst as described above, thereby further reducing exhaust emissions such as HC. Is obtained.

  Furthermore, since the intake valves 1 and 1 have a small operating angle and the exhaust valves 2 and 2 have a large operating angle from the very beginning of cranking and starting, the above-described effects can be reliably obtained from the very beginning of starting combustion. It is done.

  FIG. 5 shows a control flowchart by the controller 22, which includes a flow including engine stop transition.

  First, in step 1, it is determined whether or not the engine stop condition (key-off) is satisfied. If it is determined that the engine stop condition is satisfied, an engine stop (fuel cut) signal is output in step 2 and the engine speed (Ne) decreases. To go.

  Along with this, the discharge hydraulic pressure of the oil pump 19 decreases, and the converted hydraulic pressure (converting operating force) acts regardless of the control positions of the first and second electromagnetic switching valves 20 and 47 of the variable lift mechanisms 3 and 4. Therefore, in step 3, it shifts to the default position (intake valve 1, 1 small operating angle, exhaust valve 2, 2 large operating angle), where it is mechanically stable. Thereafter, in step 4, the engine is stopped.

  Then, in step 5, it is determined whether or not the engine start condition (key-on) has been reached again. If it is determined that the engine start condition is satisfied, even in this engine state, it is mechanically stable at this default position. In step 6, a control signal for controlling to a position corresponding to the default position may be output to both electromagnetic switching valves 20 and 47 just in case. This is because the pump hydraulic pressure fluctuates unexpectedly during the cranking of the next stroke, or even if there is a rotational fluctuation due to cranking, there is an effect of stabilizing the default position with a slight pump hydraulic pressure (this The process can be omitted).

  After that, when cranking is output in step 7 and cranking is started, it is determined in step 8 whether or not a predetermined cranking rotation has been reached, and it is determined that the cranking rotation speed has reached a complete explosion. Then, in step 9, fuel injection / ignition (complete explosion control) for starting combustion is performed.

  Since the start valve combustion is stable from the very beginning to the default valve timing, the start combustion can be reliably stabilized and the start emission can be reduced.

  Thereafter, in step 10, it is determined whether or not a predetermined time has elapsed after starting by a timer. If it is determined that the predetermined time has elapsed, the engine speed T is equal to or lower than T0 in step 11 assuming that the rotational fluctuation has stabilized. It is determined whether or not. If it is determined that the temperature does not exceed the predetermined temperature T0, the process returns. If it exceeds, the warm-up is finished and the process proceeds to step 12.

  In step 12, a signal for reducing the operating angle of the exhaust valves 2 and 2 to the small operating angle LD1 (small lift LE1) is output to the exhaust VVL4. This is because it is considered that the warm-up is sufficiently advanced when the temperature exceeds the predetermined temperature T0, and in this state, there is a possibility that abnormal combustion such as knocking or pre-ignition will occur on the contrary. At the default valve timing, a large amount of high-temperature EGR gas is introduced into the cylinder, so that the cold engine combustion is improved. However, when the engine temperature rises, this high-temperature EGR gas facilitates the above-described abnormal combustion.

  Therefore, when the operating angle (lift) of the exhaust valves 2 and 2 is decreased, the first period from the exhaust top dead center (TDC) to the exhaust valve closing timing (EVC1) decreases ((1) -1 in FIG. 4). ), The amount of high-temperature combustion gas (high-temperature EGR gas) flowing back directly into the cylinder through the exhaust valves 2 and 2 from the exhaust port side decreases. Thereby, excessive heating in the cylinder is suppressed, and abnormal combustion such as knocking and pre-ignition is avoided.

  Further, in step 13, it is determined whether or not the engine speed Ne has reached a predetermined speed N0. If it is determined that the engine speed Ne has reached the predetermined speed N0, then in step 14, each intake valve 1, 1 is actuated by the intake VVL3. Control to convert to corner (DI2) is performed.

  This is because if the operating angle remains small, the lift is low and the closing timing IVC1 of the intake valves 1 and 1 is near the bottom dead center. It is because it will be suppressed.

  Therefore, the operating angle of each intake valve 1 and 1 is expanded to DI2, the lift is increased to LI2, delayed to the closing timing IVC2, the charging efficiency is increased, and the torque / output is increased.

  Thus, when the torque / output increases, the thermal load also increases, and abnormal combustion such as knocking and pre-ignition is likely to occur. However, the opening timing IVO is advanced at the large operating angle of the intake valves 1 and 1. That is, since the second period from the opening timing IVO2 to the exhaust top dead center (TDC) is expanded, the ratio of the low temperature EGR re-intaked into the cylinder is relatively increased, thereby suppressing abnormal combustion. As a result, the knocking and pre-ignition can be suppressed while securing the torque and output in the high rotation range.

It should be noted that the start combustion stabilization and emission reduction effect at the start by the above-described default valve timing has a mechanical fail-safe function that is effective even when an electrical failure such as disconnection occurs in each electromagnetic switching valve 20, 47. Yes. This is because when an OFF signal (non-energized) is constantly output to each electromagnetic switching valve 20, 47 due to disconnection and the OFF position is fixed, or an ON signal is erroneously output due to an abnormality in the electronic control system, etc. Even in the case of starting, when the pump hydraulic pressure which is the conversion operating force is hardly supplied, since it is mechanically stable at the above-mentioned default valve timing, the starting combustion can be stabilized and the exhaust emission can be stabilized. Can also be reduced.
[Second Embodiment]
6 and 7 show a second embodiment of the present invention, in which an intake VTC 49 that is an intake phase variable mechanism and an exhaust VTC 50 that is an exhaust phase variable mechanism are provided in addition to the intake VVL 3 and the exhaust VVL 4 of the first embodiment, respectively. It is. Both VTCs 49 and 50 are formed in substantially the same structure.

  That is, the intake VTC 49 and the exhaust VTC 50 have the same basic structure except that the conversion angle of each camshaft (drive shaft) phase is different from θI and θE (only the stopper position is different). Therefore, only the structure of the exhaust side VTC 50 will be described below based on FIGS. 6 and 7.

  The exhaust VTC 50 is of a so-called vane type, and is rotationally driven by a crankshaft of the engine and transmits a rotational driving force to the exhaust camshaft 6 at the end of the exhaust camshaft 6. A vane member 52 that is fixed and rotatably accommodated in the timing sprocket 51 and a hydraulic circuit 53 that rotates the vane member 52 forward and backward by hydraulic pressure are provided.

  The timing sprocket 51 includes a housing 54 that rotatably accommodates the vane member 52, a disc-shaped front cover 55 that closes the front end opening of the housing 54, and a substantially disc that closes the rear end opening of the housing 54. The housing 54, the front cover 55, and the rear cover 56 are integrally fastened together by four small-diameter bolts 79 from the axial direction of the exhaust camshaft 6.

  The housing 54 has a cylindrical shape in which both front and rear ends are formed, and shoes 54a, which are four partition walls, project inwardly at a position of about 90 ° in the circumferential direction of the inner peripheral surface.

  Each of the shoes 54a has a substantially trapezoidal cross section, and four bolt insertion holes 54b through which the shaft portions of the respective bolts 37 are inserted are formed at substantially central positions so as to penetrate in the axial direction. A U-shaped seal member 58 and a leaf spring (not shown) that presses the seal member 58 inward are fitted and held in a holding groove that is cut out along the direction.

  The front cover 55 is formed in the shape of a disk plate, and a support hole 55a having a relatively large diameter is formed in the center thereof, and at a position corresponding to each bolt insertion hole 54b of each shoe 54a on the outer periphery. Four bolt holes (not shown) are formed.

  The rear cover 56 is integrally provided with a gear portion 56a meshing with the timing chain on the rear end side, and a large-diameter bearing hole 56b is formed in the axial direction so as to penetrate therethrough.

  The vane member 52 includes an annular vane rotor 52a having a bolt insertion hole in the center, and four vanes 52b integrally provided at a position approximately 90 ° in the circumferential direction of the outer peripheral surface of the vane rotor 52a. .

  In the vane rotor 52a, a small-diameter cylindrical portion on the front end side is rotatably supported in the support hole 55a of the front cover 55, while a small-diameter cylindrical portion on the rear end side is freely rotatable in the bearing hole 56b of the rear cover 56. It is supported.

  The vane member 52 is fixed to the front end portion of the exhaust camshaft 6 from the axial direction by a fixing bolt 57 inserted through the bolt insertion hole of the vane rotor 52a from the axial direction.

  Each of the vanes 52b is formed in a relatively long and narrow rectangular shape, and one of the other vanes 52b is formed in a trapezoidal shape having a large width. While the vane 52b is set to be substantially the same, the width of the vane 52b is set to be larger than that of the three vanes 52, so that the weight balance of the entire vane member 52 is achieved.

  Each vane 52b is disposed between the shoes 54a, and has a U-shaped seal member 60 slidably contacting the inner peripheral surface of the housing 54 in an elongated holding groove formed in the axial direction of each outer surface. Leaf springs that press the seal member 60 in the direction of the inner peripheral surface of the housing 54 are respectively fitted and held. Further, two substantially circular concave grooves 52c are formed on one side surface of each vane 52b opposite to the rotation direction of the exhaust camshaft 6, respectively.

  Further, four advance-side hydraulic chambers 61 and retard-side hydraulic chambers 62 are respectively formed between both sides of each vane 52b and both sides of each shoe 54a.

  As shown in FIG. 7, the hydraulic circuit operates with respect to the first hydraulic passages 63 for supplying and discharging the hydraulic oil pressure to and from the respective advance side hydraulic chambers 61 and the respective retard side hydraulic chambers 62. There are two systems of hydraulic passages, a second hydraulic passage 64 for supplying and discharging oil pressure, and a supply passage 65 and a drain passage 66 communicating with the discharge passage 19a are respectively provided in both the hydraulic passages 63 and 64. It is connected via a third electromagnetic switching valve 67 for switching the passage. The supply passage 65 is provided with a one-way oil pump 19 for pumping oil in the oil pan 01, while the downstream end of the drain passage 66 communicates with the oil pan 01.

  The first and second hydraulic passages 63 and 64 are formed in a cylindrical passage constituting portion 59, and one end of the passage constituting portion 59 extends from a small diameter cylindrical portion of the vane rotor 52a to an internal support hole. The other end is connected to the third electromagnetic switching valve 67 while being inserted into 52 d.

  Further, between the outer peripheral surface of one end portion of the passage constituting portion 59 and the inner peripheral surface of the support hole 52d, three annular seal members 70 for separating and sealing one end side of each of the hydraulic passages 63 and 64 are fitted. It is fixed.

  The first hydraulic passage 63 is formed in an oil chamber 63a formed at the end of the support hole 62d on the exhaust camshaft 6 side, and substantially radially inside the vane rotor 52a. And four branch paths 63 b communicating with the chamber 61.

  On the other hand, the second hydraulic passage 64 is stopped in one end portion of the passage constituting portion 59, and is formed into an annular chamber 64a formed on the outer peripheral surface of the one end portion and a substantially L-shape bent inside the vane rotor 52a. The annular chamber 64a and a second oil passage 64b communicating with each retarded-side hydraulic chamber 62 are provided.

  The third electromagnetic switching valve 67 is a four-port three-position type, and an internal valve body switches and controls each of the hydraulic passages 63 and 64, the supply passage 65, and the drain passage 66 relatively. At the same time, it is switched by a control signal from the controller 22.

The third electromagnetic switching valve 67 of the exhaust VTC 50 is configured so that the supply passage 65 communicates with the second hydraulic passage 64 communicating with the retard side hydraulic chamber 62 when the control current does not act, and the drain passage 66 communicates with the advance side hydraulic pressure. The first hydraulic passage 63 communicated with the chamber 61 communicates with the first hydraulic passage 63. Further, the position is mechanically applied by a coil spring in the third electromagnetic switching valve 67.
Further, a fourth electromagnetic switching valve 68 is provided for the intake VTC 49, but since it is the same as the above-described third electromagnetic switching valve 67 for the exhaust VTC 50, detailed description thereof is omitted.

  The controller 22 is the same as that used for the intake VVL3 (first electromagnetic switching valve 20) and the exhaust VVL4 (second electromagnetic switching valve 47) of the first embodiment. Further, the intake / exhaust VTCs 49 and 50 (the fourth and third electromagnetic switching valves 68 and 67) are used in common, and the engine operating state is detected, and the crank angle sensor and the drive shaft angles of the intake and exhaust respectively. The relative rotational positions of the intake and exhaust timing sprockets and the intake and exhaust camshafts 5 and 6 are detected by signals from the sensors.

  Further, in the intake / exhaust VTCs 49 and 50, a lock mechanism is provided between the vane member 52 and the housing 54 to lock and unlock the rotation of the vane member 52 with respect to the housing 54. This locking mechanism is provided between the one vane 52b having a large width and the rear cover 56, and has a sliding hole 72 formed along the axial direction of the exhaust camshaft 6 inside the vane 52b. A lid-shaped cylindrical lock pin 73 slidably provided inside the sliding hole 72 and a cross-sectional cup-shaped engagement hole constituting portion 74 fixed in a fixing hole provided in the rear cover 56 are provided. The lock pin 73 is held by a lock hole 74a in which the tapered tip 73a of the lock pin 73 engages and disengages, and a spring retainer 75 fixed to the bottom surface side of the slide hole 72, and the lock pin 73 is moved toward the engagement hole 74a. And a spring member 76 that biases the spring.

  The lock hole 74a is directly supplied with the hydraulic pressure in the retard side hydraulic chamber 62 or the hydraulic pressure of the oil pump 19 through an oil hole (not shown).

The lock pin 73 is engaged with the lock hole 74a by the spring force of the spring member 76 at the position (default position) where the vane member 52 is rotated to the most retarded angle side. And the relative rotation of the exhaust camshaft 6 are locked. Further, the lock pin 73 is moved backward by the hydraulic pressure supplied from the retard side hydraulic chamber 62 into the lock hole 74a or the hydraulic pressure of the oil pump 19, and the engagement with the lock hole 74a is released. Yes. That is, the lock mechanism has a function of holding and fixing at the default position.
Here, the lock mechanism may not be provided.

  Further, a biasing member that rotationally biases the vane member 52 to the retard side is provided between the bottom surface of each concave groove 52c on one side of each vane 52b and the opposing surface of each shoe 54a facing the bottom surface. A pair of coil springs 77 and 78 are arranged.

  The coil springs 77 and 78 are arranged side by side with an inter-axis distance that does not contact each other even during the maximum compression deformation, and each one end portion is interposed through a thin plate-like retainer (not shown) that fits into the groove 52c. Are connected.

Hereinafter, the basic operation will be described taking the exhaust VTC 50 as an example, but the same applies to the intake VTC 49.
First, when the engine is stopped, the output of the control current from the controller 22 to the third electromagnetic switching valve 67 is stopped, and the valve body is mechanically moved to the retard control position shown in FIG. The corner-side second hydraulic passage 64 communicates with the drain passage 66 and the advance-side first hydraulic passage 63 communicates. Further, when the engine is stopped, the oil pressure of the oil pump 19 does not act and the supply oil pressure becomes zero.

  Therefore, as shown in FIG. 6A, the vane member 52 is rotationally biased to the most retarded angle side by the spring force of each of the coil springs 77 and 78, and one shoe 54a facing one end face of one wide vane 52b. At the same time, the tip 73a of the lock pin 73 of the lock mechanism is engaged in the lock hole 74a, and the vane member 52 is stably held at the most retarded position. That is, the exhaust VTC 50 is a default position where the most retarded angle position is mechanically stable.

  Next, when the engine is started, that is, when the ignition switch is turned on and the crankshaft is cranked by the starter motor, a control signal is output from the controller 22 to the third electromagnetic switching valve 67. However, at the time immediately after the start of the crank, the discharge hydraulic pressure of the oil pump 19 has not yet sufficiently increased, so that the vane member 52 is moved to the most retarded angle side by the lock mechanism and the spring force of the coil springs 77 and 78. Is held in.

  At this time, the third electromagnetic switching valve 67 causes the supply passage 65 (discharge passage 19a) and the second hydraulic passage 64 to communicate with each other and the drain passage 66 and the first hydraulic passage 63 communicate with each other according to the control signal output from the controller 22. I am letting. The cranking advances, and the hydraulic pressure pumped from the oil pump 19 is supplied to the retarded hydraulic chamber 62 through the second hydraulic passage 64 as the hydraulic pressure is increased. The hydraulic pressure is not supplied and the hydraulic pressure is released from the drain passage 66 into the oil pan 01 to maintain the low pressure state.

  Here, after the cranking rotation is increased and the hydraulic pressure is further increased, free vane position control by the third electromagnetic switching valve 67 can be performed. That is, as the pump hydraulic pressure and the hydraulic pressure in the retard side hydraulic chamber 62 increase, the hydraulic pressure in the lock hole 74a of the lock mechanism also increases, the lock pin 73 moves backward, and the distal end portion 73a comes out of the lock hole 74a and moves to the housing 54. Since the relative rotation of the vane member 52 with respect to is allowed, the vane position can be freely controlled.

  For example, the third electromagnetic switching valve 67 is operated by a control signal from the controller 22 to connect the supply passage 65 and the first hydraulic passage 63, while connecting the drain passage 66 and the second hydraulic passage 64.

  Accordingly, the hydraulic pressure in the retard side hydraulic chamber 62 is returned to the oil pan 01 from the drain passage 66 through the second hydraulic passage 64 this time, while the inside of the retard side hydraulic chamber 62 becomes low pressure, while the advance angle is increased. The hydraulic pressure is supplied into the side hydraulic chamber 61 and becomes high pressure.

  Therefore, the vane member 52 rotates in the clockwise direction in the figure against the spring force of each of the coil springs 77 and 78 due to the high pressure in the advance side hydraulic chamber 61, and relatively rotates toward the position shown in FIG. 6B. Then, the relative rotational phase of the exhaust camshaft 6 with respect to the timing sprocket 51 is converted to the advance side. Further, by setting the position of the third electromagnetic switching valve 67 to the neutral position, it can be held at an arbitrary relative rotation phase.

  Further, the relative rotation phase is continuously changed from the most retarded angle (FIG. 6A) to the most advanced angle (FIG. 6B) according to the engine operating state after the warm-up is completed.

The basic operation has been described with the exhaust VTC 50 as an example, but the same applies to the intake VTC 49.
[Effects of Second Embodiment]
FIG. 8 shows lift characteristics of the intake valves 1 and 1 and the exhaust valves 2 and 2 by the intake VVL3 and the exhaust VVL4 and the intake VTC49 and the exhaust VTC50.

  As in the first embodiment, the default position by the exhaust VVL4 is the large lift LE2 (large operating angle DE2), and the default position by the intake VVL3 is the small lift LI1 (small operating angle DI1 '). Here, the default operating angle DI1 'of the intake VVL3 is set to a slightly narrower operating angle than DI1 of the first embodiment.

  The default valve timings of the intake / exhaust VVLs 3 and 4 and the intake / exhaust VTCs 49 and 50 (both VTCs are the most retarded default) are as shown in the lift curve shown by the solid line in FIG.

  That is, the exhaust valves 2 and 2 are a large lift LE2 (large operating angle DE2) of the exhaust VVL4, the exhaust VTC50 is an exhaust lift curve (solid line) at the most retarded angle, and the intake valves 1 and 1 are small lifts of the intake VVL3. At LI1 (small operating angle DI1 ′), the intake VTC 49 is the intake lift curve (solid line) at the most retarded angle.

  The effect of this is greater than that of the first embodiment. Because the default exhaust valve closing timing (EVC) is delayed until EVC2 ′, the above-mentioned first period becomes (1) -2 ′ wider than (1) -2 ′ of the first embodiment, and therefore the high temperature EGR gas is reduced. This is because a larger amount can be taken into the cylinder. Further, since the opening timing (IVO) of the default intake valves 1 and 1 is delayed until IVO1 ′, the second period described above becomes narrower than (2) -1 of the first embodiment, and almost 0 (IVO1 ′ is near TDC). ) Or, as shown in FIG. 8, IVO1 ′ may be further retarded by α on the contrary to TDC to make it a negative period.

  In this way, even if the piston reaches the top dead center in the exhaust stroke, the intake valves 1 and 1 are not opened, so that the combustion gas is once returned to the intake side via the intake valve, where it is cooled, and the low temperature EGR The amount re-inhaled as gas (a factor that deteriorates combustion stability) is almost eliminated.

  Also, during this period α during the intake stroke, the intake valves 1 and 1 are closed and only the exhaust valves 2 and 2 are open, so no fresh air is sucked when the piston descends. In addition, since only the high temperature EGR gas is introduced into the cylinder, the high temperature EGR introduction effect in the first period can be further enhanced.

  In this way, while introducing a larger amount of high temperature EGR into the cylinder, almost no low temperature EGR is introduced into the cylinder, further improving the combustion stability at start-up and reducing exhaust emissions compared to the first embodiment. Can be achieved.

  On the other hand, since the retarding control is performed by the exhaust VTC 50, the opening timing (EVO2 ') of the exhaust valves 2 and 2 is further delayed by that amount, so that it takes a long time to warm the cylinder with the combustion gas. It is also possible to shorten (speed up) the time required for (temperature rise), and to realize early combustion stabilization and early exhaust emission reduction due to the warm-up effect over time.

  FIG. 9 shows a control flowchart of each mechanism by the controller 22.

  In step 21, it is determined whether or not the engine stop condition (key-off) is satisfied. If it is determined that the engine stop condition (key-off) is satisfied, an engine stop (fuel cut) signal is output in step 22. . As a result, the engine speed (Ne) decreases and the discharge hydraulic pressure of the oil pump 19 decreases accordingly. Therefore, the converted hydraulic pressure (converted driving force) does not act regardless of the control positions of the electromagnetic switching valves 20 and 47 of the intake and exhaust VVLs 3 and 4, so that the default position (the intake valves 1 and 1 have a small operating angle, The exhaust valves 2 and 2 are shifted to a large operating angle) where they are mechanically stabilized.

  In addition, since the intake and exhaust VTCs 49 and 50 do not act on the converted hydraulic pressure (converted driving force) regardless of the control positions of the fourth electromagnetic switching valve 68 and the third electromagnetic switching valve 67, the coil springs 77 and 78 Due to the spring force and the valve drive reaction force, the intake and exhaust VTCs 49 and 50 shift to the most retarded angle which is the default, and are mechanically stabilized there.

  And even if the next starting condition (key-on) is reached, it is mechanically stable at this default position. In step 23, both the intake and exhaust VVL3, 4 electromagnetic switching valves 20, 47, Control signals for controlling the positions corresponding to the respective default positions may be output to the electromagnetic switching valves 68 and 67 of the intake and exhaust VTCs 49 and 50. This is because the pump hydraulic pressure fluctuates unexpectedly during the cranking of the next stroke, or even if there is a rotational fluctuation due to cranking, it is stabilized at the default position with a slight pump hydraulic pressure.

  Next, in step 24, it is confirmed that the engine has stopped. In step 25, it is determined whether the engine start condition (key-on) is satisfied. If it is determined that the engine start condition is satisfied, in step 26, the intake air, exhaust VVL3, 4 and intake air The conversion signals to the respective default positions are output to the electromagnetic switching valves 20, 47, 68, 67 of the exhaust VTCs 49, 50, respectively.

  In step 27, a signal indicating that cranking has been started is output, and in step 28, it is determined whether or not a predetermined cranking rotation has been reached. Injection and ignition (complete explosion control) is performed.

  Since the start valve combustion is stable from the very beginning to the default valve timing, the start combustion can be reliably stabilized and the start emission can be reduced. Also, it has a mechanical fail-safe effect as described in the first embodiment.

  If it is determined in step 30 whether or not a predetermined time has elapsed by the timer, that is, if it is determined that a predetermined time for stabilizing the rotational fluctuation has elapsed, in step 31, the engine temperature is set to a block wall temperature not shown. The engine temperature T is detected by a sensor or the like.

  In step 32, it is determined whether or not the predetermined temperature T0 has been reached. If it is determined that the predetermined temperature T0 has been exceeded, it is determined that the warm-up has already ended, and in step 33 the engine speed Ne and load The intake / exhaust VVLs 3 and 4 and the intake / exhaust VTCs 49 and 50 are controlled based on the map.

  If it does not exceed the predetermined temperature T0, it is determined that the warm-up has not ended, and the process proceeds to step 34. In step 34, the target phases of the intake valves 1 and 1 and the exhaust valves 2 and 2 for each engine temperature are calculated.

  If the engine temperature is still very low, it is controlled at the default valve timing described above. That is, the intake / exhaust VTCs 49 and 50 are controlled at the most retarded angle.

  On the other hand, when the engine temperature becomes high to some extent, if a large amount of high-temperature EGR gas is introduced into the cylinder, high-temperature abnormal combustion such as knocking or pre-ignition tends to occur. Therefore, in step 35, control is performed to advance the exhaust VTC 50 from the most retarded position (default position) according to the engine temperature T.

  As a result, the exhaust VTC 50 performs the minimum phase advance angle that can prevent high-temperature abnormal combustion, so that combustion stability and low emission can be maximized while preventing high-temperature abnormal combustion. For example, when the exhaust VTC 50 is advanced to the exhaust valve lift characteristic corresponding to the first embodiment shown by the broken line in FIG. 8 and the engine temperature further rises, the exhaust valve lift characteristic shown by the one-dot chain line on the advance side is further increased. The retard angle is changed.

  Here, the same effect can be obtained by advancing the intake VTC 49 based on the engine temperature T from the most retarded phase instead of the exhaust VTC 50.

As described above, if the engine temperature exceeds the predetermined temperature T0, it is determined that the warm-up has already been completed. In step 33, the intake / exhaust VVL3, 4 based on the engine speed Ne map and the load map are determined. Then, the control shifts to normal control of the intake / exhaust VTCs 49 and 50.
[Third Embodiment]
FIG. 10 shows a third embodiment. On the exhaust side, an exhaust VVL4 having the same hydraulic pressure as that of the first embodiment as a conversion driving force is provided, and the maximum lift LE2 (maximum operating angle DE2) is similarly set to the default position. It has become.

  Unlike the first embodiment, the intake side is provided with an electric intake VEL80 that is an intake-side lift variable mechanism and an electric intake VTC81 that is an intake-side phase variable mechanism.

  The intake VEL 80 makes the operating angle and lift amount of the intake valves 1 and 1 continuously variable. For example, the intake VEL 80 is the same as that described in Japanese Patent Application Laid-Open No. 2012-225287 filed earlier by the present applicant. Since it is a structure, it demonstrates easily based on FIGS. The exhaust VVL4 is assigned the same reference numeral as in the first embodiment, and a detailed description thereof is omitted.

  The minimum lift (minimum operating angle) of the intake valves 1 and 1 by the intake VEL 80 is LI0 (DI0), and the maximum lift (maximum operating angle) is LI2 (DI2).

  The intake VEL 80 includes a hollow drive shaft 82 that is rotatably supported by a bearing provided in the upper part of the cylinder head, a rotary cam 83 that is fixed to the outer peripheral surface of the drive shaft 82 by press-fitting, and the outer periphery of the drive shaft 82. Two swing cams 84 that are swingably supported on the surface and slide in contact with the upper surface of each valve lifter 1a disposed at the upper end of each intake valve 1, 1 to open each intake valve 1, 1. And a transmission mechanism that is interposed between the rotary cam 83 and each swing cam 84 and converts the rotational force of the rotary cam 83 into a swing motion and transmits it to the swing cam 84 as a swing force. ing.

  The drive shaft 82 receives a rotational force from the crankshaft via a timing sprocket 85 provided at one end, and this rotational direction is set in the clockwise direction (arrow direction) in FIG.

  The rotary cam 83 has a substantially ring shape, and the axis Y of the cam body is offset from the axis X of the drive shaft 82 by a predetermined amount in the radial direction.

  Both the swing cams 84 are integrally provided at both end portions of a cylindrical cam shaft, and the cam shafts are rotatably supported by the drive shaft 82 via an inner peripheral surface. Further, a cam surface including a base circle surface, a ramp surface, and a lift surface is formed on the lower surface, and the base circle surface, the ramp surface, and the lift surface are arranged on each valve lifter 1a according to the swing position of the swing cam 84. It comes in contact with a predetermined position on the upper surface.

  The transmission mechanism includes a rocker arm 86, a link arm 87, and a link rod 88. One end of the rocker arm 86 is rotatably connected to the link arm 87, and the other end is one end of the link rod 88. Is rotatably connected to.

  In the link arm 87, the cam body of the rotary cam 83 is rotatably fitted in a central fitting hole, and the protruding end is connected to one end of the rocker arm by a pin. The other end of the link rod 88 is rotatably connected to the cam nose of the swing cam 84 via a pin.

  A control shaft 89 is rotatably supported above the drive shaft 82, and a control cam 90 serving as a rocking fulcrum of the rocker arm 86 is fixed to the outer periphery of the control shaft 89. The rotation of the control shaft 89 is controlled by the electric actuator 91, while the position of the axis P2 of the control cam 90 is offset from the axis P1 of the control shaft 89 by a predetermined amount.

  As shown in FIGS. 12A and 12B, the electric actuator 91 includes an electric motor 92 fixed to one end portion of the casing 91a, and a rotational driving force of the electric motor 92 provided to the control shaft 89. It is comprised from the ball screw transmission mechanism 93 which is a reduction mechanism which transmits.

  The electric motor 92 is composed of a proportional DC motor, and is driven by a control signal from the controller 22 that detects an engine operating state and the like.

  The ball screw transmission mechanism 93 includes a ball screw shaft 93a disposed substantially coaxially with the drive shaft of the electric motor 92, a ball nut 93b screwed onto the outer periphery of the ball screw shaft 93a, and one end of the control shaft 89. It is mainly comprised from the linkage arm 93c connected with the part along a diameter direction, and linked with the ball nut 93b.

  The ball screw shaft 93a is connected to one end of the ball screw shaft 93a via a motor drive shaft, and is rotated by an electric motor 92. The ball nut 93b is formed in a substantially cylindrical shape, and a guide groove for continuously holding a plurality of balls is formed in a spiral manner in cooperation with the ball circulation groove on the inner peripheral surface. An axial moving force is applied through the ball while converting the rotational motion of the ball screw shaft 93a into a linear motion. The ball nut 93b is urged toward the electric motor 92 (minimum lift side) by the spring force of the coil spring 94, which is urging means. Therefore, when the engine is stopped, the ball nut 93b is moved to the minimum lift position along the axial direction of the ball screw shaft 93a by the spring force of the coil spring 94. That is, the default position of the intake VEL 80 is the minimum lift LI0 (minimum operating angle DI0).

  The controller 22 detects the current engine operating state from various information signals as in the embodiments described above, but the controller 22 in this embodiment detects a rotation angle of the drive shaft 82. And a detection signal from a potentiometer 95 that detects the rotational position of the control shaft 89 are input, and the relative rotation angle with respect to the crank angle of the drive shaft 82 by the electric intake VTC 81 and the electric intake VEL 80 are used. The valve lift amount and operating angle of each intake valve 1, 1 are detected.

  Accordingly, when the ball screw shaft 93a is rotated in one direction by the rotational torque of the electric motor 92 that is driven to rotate in one direction by the control current from the controller 22 in a predetermined operation region, the ball nut 93b is moved as shown in FIG. 12A. Then, while being assisted by the spring force of the coil spring 94, it moves linearly in a maximum direction (direction approaching the electric motor 92), whereby the control shaft 89 rotates in one direction via the linkage arm 93c.

  Therefore, as shown in FIG. 11A (front view), the control cam 90 rotates about the axis of the control shaft 89 with the same radius, and the thick portion moves away from the drive shaft 82 upward. To do. As a result, the other end of the rocker arm 86 and the pivot point of the link rod 88 move upward with respect to the drive shaft 82, so that each swing cam 84 is forced by the cam nose portion via the link rod 88. As a result, the whole is rotated clockwise as shown in FIG. 11A.

  Therefore, when the rotary cam 83 rotates and pushes up one end of the rocker arm 86 via the link arm 87, the lift amount is transmitted to the swing cam 84 and the valve lifter 1a via the link rod 88. As shown by the valve lift curve in FIG. 13, the intake valves 1 and 1 have the minimum lift (LI0) as shown by the valve lift curve in FIG. 13, and the operating angle DI0 is also minimum.

  Further, for example, when shifting to the high rotation / high load region, the electric motor 92 is rotated in the other direction by the control signal from the controller 22 to move the ball nut 93b to the maximum right direction as shown in FIG. 12B. . As a result, the control shaft 89 rotates the control cam 90 in the clockwise direction in FIG. 11 to rotate the shaft center P2 downward. Therefore, as shown in FIG. 11B, the entire rocker arm 86 moves toward the drive shaft 82 and the other end presses the cam nose portion of the swing cam 84 downward via the link rod 88. The entire swing cam 84 is rotated counterclockwise by a predetermined amount.

  Therefore, when the rotary cam 83 rotates and pushes up one end of the rocker arm 86 via the link arm 87, the lift amount is transmitted to the swing cam 84 and the valve lifter 1a8 via the link rod 88. As shown in FIG. 13, the lift amount increases continuously with LI1, LI1.5, and LI2. As a result, the exhaust efficiency in the high rotation range can be increased and the output can be improved.

  That is, the lift amount of the intake valves 1 and 1 is continuously changed from the minimum lift LI0 to the large lift LI2 in accordance with the operating state of the engine. It varies continuously from the minimum lift DI0 to the large lift DI2.

  When the engine is stopped, as described above, the ball nut 93b is automatically urged toward the electric motor 92 by the spring force of the coil spring 94, so that the minimum operating angle DI0 and the minimum lift LI0 position ( The default position is held stable.

  The electric intake VTC 81 has a structure as shown in, for example, Japanese Patent Application Laid-Open No. 2012-145036 filed by the applicant of the present application, so that a detailed description thereof will be omitted. However, the rotational force of the electric motor is reduced by a reduction mechanism. The rotational phase of the drive shaft 82 is changed via the. Further, an urging spring (not shown) urges the drive shaft 82 in the retarding direction, so that the default position of the intake VTC 81 is the most retarded angle as in the intake VTC 49 of the second embodiment.

  Therefore, the lift curve at the default position of the intake valves 1 and 1 by the intake VEL 80 and the intake VTC 81 is as shown by a thick solid line (LI0) in FIG.

  The effect of the combustion stability at start-up and the emission reduction by this is larger than those of the first embodiment and the second embodiment.

  Because the opening timing (IVO) of the intake valves 1 and 1 at the default position is greatly delayed until IVO0 (α is larger than that in the second embodiment), the second period described above is completely eliminated, and the piston is moved in the exhaust stroke. Even when the top dead center is reached, the intake valves 1 and 1 are not open. For this reason, the amount of the combustion gas once returned to the intake side via the intake valves 1 and 1 is cooled and there is almost no amount re-inhaled as low-temperature EGR gas (cause of deterioration in combustion stability).

  Further, in the period α which is larger than that in the second embodiment, the intake valves 1 and 1 are closed and only the exhaust valves 2 and 2 are opened, so that fresh air is sucked in when the piston descends. Therefore, since only the high temperature EGR gas is introduced into the cylinder, the effect of introducing the high temperature EGR in the first period can be further enhanced, and the effect of the second embodiment or more can be obtained.

  In this embodiment, the lift (operating angle) of the intake valves 1 and 1 at the default position is set sufficiently small, and the closing timing (IVC0) of the intake valves 1 and 1 is set near the bottom dead center. Therefore, α is sufficiently large to further enhance the above-described effect.

  Further, in the present embodiment, the lift amount of the intake valves 1 and 1 is sufficiently low as LI0, so that the intake air flow rate is increased and the in-cylinder agitation effect is increased, thereby further improving the combustion.

  Furthermore, since there is a β period between the exhaust valve closing timing (EVC2) and the intake valve opening timing (IVO0), both the intake valves 1 and 1 / exhaust valves 2 and 2 are closed. There is a negative pressure development, which also increases the intake air flow rate when the intake valves 1 and 1 are opened, thereby further improving the combustion.

  Further, the closing timing of the intake valves 1 and 1 is near the bottom dead center, and the combustion is improved by improving the effective compression ratio, as in the other embodiments.

  Furthermore, the combustion improvement effect by the large operating angle DE2 (large lift LE2) of the exhaust valves 2 and 2 is the same as in the other embodiments.

  In the present embodiment, knocking and pre-ignition are further concerned when the engine temperature rises because the combustion improvement effect is high.

Therefore, as the engine temperature rises, control is performed to finely increase the operating angle of the intake valves 1 and 1 by the intake VEL 80 and finely advance the intake VTC 81.
[Control flow chart]
FIG. 14 shows a control flowchart of the intake VEL 80, the intake VTC 81, and the exhaust VVL4 by the controller 22.

  First, in step 41, it is determined whether or not an engine stop condition (key-off) has been reached. If it is determined that the engine stop condition has been reached, the routine proceeds to step 42, where an engine stop (fuel cut) signal is output. Is done. As a result, the engine speed (Ne) decreases, and the discharge hydraulic pressure of the oil pump 19 decreases accordingly.

  Accordingly, in step 43, the converted hydraulic pressure (converted driving force) does not act regardless of the switching valve position of the electromagnetic switching valve 47 of the hydraulic exhaust VVL4. Therefore, the default positions (large lift LE2, large operation) of the exhaust valves 2 and 2 are eliminated. Transition to the corner DE2), where it is mechanically stable. Also, since the electric intake VEL 80 and the electric intake VTC 81 are not supplied with electric power (conversion driving force), the intake valves 1 and 1 are shifted to the minimum lift LI0 and the most retarded angle which are the default positions, respectively. Stable. In step 44, the engine is stopped.

  Thereafter, in step 45, it is determined whether or not the next restart condition (key-on) is satisfied. If it is determined that the start condition is satisfied, at this point in time, it is mechanically stable at this default position. However, in step 46, as a precaution, control signals for controlling the respective default positions may be output to the exhaust VVL4, the intake VEL 80, and the intake VTC 81. This is because, even when there is a rotational fluctuation due to cranking during cranking of the next stroke, stabilization is performed at the default position by the supplied power and a slight pump hydraulic pressure.

  Next, cranking is started in step 47, and in step 48, it is determined whether or not a predetermined cranking rotation has been reached. Fuel injection / ignition for combustion is performed. Since the start valve combustion is stable from the very beginning to the default valve timing, the start combustion can be reliably stabilized and the start emission can be reduced. Further, similarly to the first and second embodiments, it has a mechanical fail-safe effect.

  In step 50, it is determined whether or not a predetermined time necessary for stabilizing the rotational fluctuation has elapsed. If it is determined that the time has elapsed, in step 51, the engine temperature is determined by the cylinder block wall temperature sensor or the like. The temperature T is detected.

  In step 52, it is determined whether or not the engine temperature T exceeds a predetermined temperature T0. If it is determined that the engine temperature T has exceeded, it is determined that the warm-up has already ended, and in step 53 the engine speed Ne and load operation are determined. Based on the map, intake VEL 80, exhaust VVL4, and intake VTC 81 are controlled.

  As a result of the temperature detection in step 52, if the predetermined temperature T0 is not exceeded, it is determined that the warm-up has not been completed, and in step 54, target values of the intake VEL 80 and the intake VTC 81 are calculated for each engine temperature. Based on the calculated target value, the intake VEL 80 and the intake VTC 81 are finely and continuously controlled.

  If the engine temperature is still very low, it is controlled at the default valve timing described above.

  On the other hand, when the engine temperature is increased to some extent, if a large amount of high-temperature EGR gas is introduced into the cylinder, abnormal combustion at a high temperature such as knocking or pre-ignition tends to occur. Therefore, according to the engine temperature T, the intake VEL 80 is increased from the minimum lift LI0 (minimum operating angle DI0) position to LI1, LI1.5, while the closing timing of the intake valves 1, 1 is maintained near the bottom dead center. In this way, the intake VTC 81 is gradually advanced. Thus, combustion stability and low emission can be maximized while preventing abnormal combustion at high temperatures.

  As described above, when the engine temperature exceeds the predetermined temperature T0, it is assumed that the warm-up has already been completed, and normal control of the intake VEL80, the exhaust VVL4, and the intake VTC81 is performed based on the rotational speed Ne / load map. Transition.

  For example, when shifting to a high rotation / high load region, as shown by the broken line in FIG. 13, the intake VEL 80 converts to the maximum lift LI2, the intake VTC 81 controls to the retard side, and the intake valve 1, 1 opening timing IVO2 While suppressing the change of (≈ IVO1.5), the closing timing of the intake valves 1 and 1 is retarded to IVC2, and the change in the EGR gas amount is suppressed and the charging efficiency in the high rotation range is increased and stably The output is increased smoothly.

  In the present embodiment, when the engine temperature is T0 or lower and the engine is not warmed up (including the cooler), priority is given mainly to these variable mechanisms such as the intake VEL80 and the intake VTC81, whose conversion driving force is electric. Therefore, good conversion responsiveness and conversion stability can be obtained.

  Further, as can be seen from FIG. 10, in this embodiment, the conversion driving force of each variable mechanism is electric on the intake side and hydraulic on the exhaust side. Here, assuming that the intake side and the exhaust side are hydraulic, it is normal to use a single oil pump, and the oil passage increases to two systems or more. In the case of increased leakage (decrease in hydraulic pressure) or simultaneous conversion on the intake side and exhaust side, the oil supply speed for each conversion decreases, and conversion responsiveness deteriorates.

  In the present embodiment, as the conversion energy of the variable mechanism, only one of the intake side and the exhaust side is set to hydraulic pressure, so that responsiveness can be ensured even if hydraulic energy is used.

  Here, it is conceivable that both the intake-side and exhaust-side variable mechanisms are electrically operated. However, in addition to an increase in battery load, problems in layout occur. In other words, the electric actuator (electric conversion device) of the intake VEL 80 takes a large space by the electric motor 92 and the speed reduction mechanism, but if this is mounted on both the intake side and the exhaust side, the space occupied in the engine. As a result, the size of the engine becomes extremely large, and there is a concern about its ability to be mounted on an engine room. On the other hand, if one exhaust VVL4 is hydraulic, an electromagnetic switching valve can be easily mounted at a position different from the other intake-side electric actuator (for example, the front of the engine or the side of the engine). Thus, the space efficiency can be increased.

  The present invention is not limited to the configuration of each of the embodiments described above, and can be applied to various configurations and structures without departing from the gist of the present invention.

  For example, the lift variable mechanism and the phase variable mechanism may be step conversion (step conversion) such as two-stage variable, or may be continuous conversion. Further, the conversion driving force may be hydraulic, electric, or air pressure.

  Furthermore, although an example using an urging spring has been shown as means for holding the conversion driving force in a mechanically stable position (default position), it is possible to mechanically stabilize without such an urging spring. Also good. For example, it may be stabilized at the default position using a valve train driving load, or may be directly fixed at the default position using a lock pin or the like.

The opening / closing timing of the intake / exhaust valves (IVO, IVC, EVO, EVC, etc.) may be exactly the timing when the lift starts and ends, but excludes a slight ramp period provided at the beginning and end of lift. It is good as a time. In the latter case, since the gas exchange starts substantially effectively and ends effectively, the effect of the present invention can be further enhanced.
In the present invention, the effect can be further enhanced by combining different types of variable mechanisms. For example, a variable mechanical compression ratio mechanism capable of adjusting the piston top dead center height as disclosed in JP-A-2002-276446 can be used in combination. In this case, for example, as in each of the above-described embodiments, not only the intake valve closing timing is set near the bottom dead center at the start, but the effective compression ratio is increased, and the piston compression top dead center position itself is slightly set by the variable mechanical compression ratio mechanism. By raising the geometric compression ratio, the effective compression ratio can be further increased. As a result, it is possible to further enhance the combustion improvement effect at the start, which is the effect of the present invention.

The technical ideas of the invention other than the claims ascertained from the embodiment will be described below.
[Claim a]
The variable valve operating apparatus for an internal combustion engine according to claim 1,
A variable valve operating apparatus for an internal combustion engine, wherein the conversion driving force of at least one of the intake operating angle variable mechanism and the exhaust operating angle variable mechanism is hydraulic.
[Claim b]
The variable valve operating apparatus for an internal combustion engine according to claim 1,
A variable valve operating apparatus for an internal combustion engine, wherein the conversion driving force of at least one of the intake operating angle variable mechanism and the exhaust operating angle variable mechanism is electric.
[Claim c]
The variable valve operating apparatus for an internal combustion engine according to claim 1,
The variable intake valve for an internal combustion engine, wherein when the conversion drive force is applied, the intake operation angle variable mechanism is converted to an operation angle larger than a small operation angle when the conversion drive force is not applied. apparatus.
[Claim d]
The variable valve operating apparatus for an internal combustion engine according to claim 1,
The variable valve operating device for an internal combustion engine, wherein the exhaust operating angle variable mechanism is converted into an operating angle smaller than an operating angle when the conversion operating angle is not applied when the conversion driving force is applied.
[Claim e]
The variable valve operating apparatus for an internal combustion engine according to claim 1,
The variable valve operating apparatus for an internal combustion engine, wherein the intake operating angle variable mechanism and the exhaust operating angle variable mechanism selectively convert the operating angle according to the presence or absence of an action of a conversion driving force.
[Claim f]
The variable valve operating apparatus for an internal combustion engine according to claim 1,
The variable valve operating apparatus for an internal combustion engine, wherein the intake operating angle variable mechanism and the exhaust operating angle variable mechanism continuously vary the operating angle according to the operating amount of the conversion driving force.
[Claim g]
The variable valve operating apparatus for an internal combustion engine according to claim 6,
An intake phase variable mechanism that changes a peak lift phase of the intake valve, and an exhaust phase variable mechanism that changes a peak lift phase of the exhaust valve,
The variable valve operating apparatus for an internal combustion engine, wherein the intake phase variable mechanism and the exhaust phase variable mechanism are each mechanically held at a retarded phase when no conversion driving force is applied.

DESCRIPTION OF SYMBOLS 1 ... Intake valve 2 ... Exhaust valve 3 ... Intake VVL (intake lift variable mechanism)
4. Exhaust VVL (Exhaust lift variable mechanism)
5 ... intake camshaft 6 ... exhaust camshaft 7 ... large lift cam 8 ... small lift cam 19 ... oil pump 20 ... first electromagnetic switching valve 22 ... controller 31 ... large lift cam 32 ... small lift cam 47 ... second electromagnetic switching valve 49 ... hydraulic pressure Intake VTC
67 ... 3rd electromagnetic switching valve 50 ... Hydraulic exhaust VTC
80 ... Electric intake VEL
81 ... Electric intake VTC

Claims (8)

An intake operating angle variable mechanism that makes the operating angle of the intake valve variable;
An exhaust operating angle variable mechanism that makes the operating angle of the exhaust valve variable,
While the intake operating angle variable mechanism mechanically holds the intake valve on the small operating angle side when the conversion driving force does not act,
The variable exhaust valve operating mechanism for an internal combustion engine, wherein the exhaust operating angle variable mechanism mechanically holds the exhaust valve on the large operating angle side when the conversion driving force does not act. The variable valve operating apparatus for an internal combustion engine according to claim 1,
An exhaust phase variable mechanism for changing a peak lift phase of the exhaust valve;
The variable exhaust valve mechanism for an internal combustion engine, wherein the exhaust phase variable mechanism mechanically holds the exhaust valve at a peak lift phase toward the retarded angle when the conversion driving force does not act. The variable valve operating apparatus for an internal combustion engine according to claim 1,
An intake phase variable mechanism for changing a peak lift phase of the intake valve;
The variable valve operating apparatus for an internal combustion engine, wherein the intake phase variable mechanism mechanically holds the intake valve at a peak lift phase toward the retard side when the conversion driving force does not act. The variable valve operating apparatus for an internal combustion engine according to claim 1,
An intake phase variable mechanism that changes a peak lift phase of the intake valve, and an exhaust phase variable mechanism that changes a peak lift phase of the exhaust valve,
The variable valve operating apparatus for an internal combustion engine, wherein the intake phase variable mechanism and the exhaust phase variable mechanism are each mechanically held on the retard side when a conversion driving force does not act. The variable valve operating apparatus for an internal combustion engine according to claim 1,
A variable valve operating apparatus for an internal combustion engine, wherein each of the intake operating angle variable mechanism and the exhaust operating angle variable mechanism is configured such that one conversion driving force is hydraulic and the other conversion driving force is electric. An intake operating angle variable mechanism in which the conversion driving force for changing the operating angle of the intake valve is hydraulic or electric; and
An exhaust operation angle variable mechanism that makes the conversion drive force that makes the operation angle of the exhaust valve variable hydraulic or electric, and
The intake operating angle variable mechanism mechanically holds the intake valve on the relatively small operating angle side when the conversion driving force does not act, compared to when the conversion driving force acts,
The internal combustion engine characterized in that the exhaust operating angle variable mechanism mechanically holds the exhaust valve at a relatively large operating angle side when the conversion driving force does not act than when the conversion driving force acts. Variable valve gear. The variable valve operating apparatus for an internal combustion engine according to claim 6,
An exhaust phase variable mechanism for changing a peak lift phase of the exhaust valve;
The variable exhaust valve mechanism for an internal combustion engine, wherein the exhaust phase variable mechanism mechanically holds the exhaust valve at a peak lift phase toward the retarded angle when the conversion driving force does not act. The variable valve operating apparatus for an internal combustion engine according to claim 6,
An intake phase variable mechanism for changing a peak lift phase of the intake valve;
The variable valve operating apparatus for an internal combustion engine, wherein the intake phase variable mechanism mechanically holds the intake valve at a peak lift phase toward the retard side when the conversion driving force does not act.

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