JP2015010480A - Variable valve gear of multi-cylinder internal combustion engine and control device for variable valve gear - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a variable valve gear of a multi-cylinder internal combustion engine capable of improving an engine startability due to reduction of friction and the like and improving exhaust emission performance due to production of strong exhaust gas flow and the like.SOLUTION: This invention comprises: a pair of exhaust valves 3a to 3b and a pair of intake valves 71a to 71b arranged for each of #1 cylinder and #2 cylinder; first valve stopping mechanisms 11b and 11c for changing over a valve operating state and a valve stopped state for one exhaust valve of the pair of intake valves of #1 cylinder and #2 cylinder; and a second stop mechanism 11a for changing over a valve operating state and a valve stopped state for the other exhaust valve in the #1 cylinder. The first valve stopping mechanism is constituted such that it becomes the valve operating state when hydraulic pressure is supplied and it becomes the valve stopped state when supplying of the hydraulic pressure is stopped, and the second valve stopping mechanism is constituted such that it becomes a valve stopped state when the hydraulic pressure is supplied and it becomes a valve operating state when supplying of the hydraulic pressure is stopped.

Description

  The present invention relates to a variable valve operating apparatus for a multi-cylinder internal combustion engine capable of stopping an opening / closing operation of an exhaust valve according to an engine operating state, and a control apparatus for the variable valve operating apparatus.

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

  This variable valve operating system includes a valve stop (pause) mechanism that stops (pauses) the opening and closing operations of the intake valves and exhaust valves of some cylinders. The body of the lassia adjuster serving as a swing fulcrum is fixed to the cylinder head by a switching member. Therefore, since it functions as a normal lash adjuster, when the rocker arm is pushed down by a cam, one end of the rocker arm swings around the swing fulcrum, and the intake valve and the exhaust valve move at a predetermined valve lift by the other end. It comes to work.

  On the other hand, in the normal operation range where low fuel consumption is required, the switching member moves into the body of the lash adjuster in some cylinders, and the lash adjuster is lost to stop the operation of the intake valve and the exhaust valve. (Cylinder deactivation) The fuel consumption is improved by performing so-called reduced-cylinder operation in which only the remaining cylinders are operated for combustion.

JP 2007-100585 A (FIGS. 1 and 3)

  However, when the engine is started, the conventional variable valve operating system operates all cylinders without stopping the cylinders to ensure engine torque, and all intake valves and exhaust valves are driven to open and close (valve operation). However, these do not particularly improve startability and exhaust performance.

  Further, when the engine enters a predetermined practical operation region, some cylinders shift to a resting state. In this case, all of the intake valves and exhaust valves of the resting cylinders are stopped together. That is, the switching members corresponding to all the intake valves and exhaust valves of some cylinders must be moved simultaneously by the supply hydraulic pressure (signal hydraulic pressure).

  The present invention has been devised in view of the above technical problems of the conventional variable valve gear, and by obtaining a better startability such as reduction of friction at the time of engine start, and by generating a strong exhaust gas flow, etc. A variable valve operating apparatus for a multi-cylinder internal combustion engine capable of reducing exhaust emission is provided.

According to the first aspect of the present invention, one of the pair of intake valves and the pair of exhaust valves provided for each cylinder and the pair of exhaust valves in a part of the plurality of cylinders is provided with one exhaust valve. A first valve stop mechanism that switches between a valve operating state that opens and closes and a valve stop state that stops the opening and closing operation; a valve operating state that opens and closes the other exhaust valve of the pair of exhaust valves in the some cylinders; A second valve stop mechanism that switches a valve stop state that stops the opening and closing operation,
The first valve stop mechanism is mechanically configured to be in a valve operation state when switching energy is supplied, and to be in a valve stop state when supply of switching energy is stopped, and the second valve stop mechanism is When the switching energy is supplied, the valve is in a stopped state, and when the switching energy supply is stopped, the valve is in a mechanical state.

  According to the present invention, the engine startability can be improved by reducing friction from the initial stage of cranking, and the exhaust emission performance can be improved by generating a strong exhaust gas flow.

1 is a perspective view of an exhaust valve side showing a first embodiment in which a variable valve device according to the present invention is applied to a two-cylinder internal combustion engine. It is a perspective view which shows the intake valve side of the embodiment. It is the sectional view on the AA line of FIG. 1 and FIG. A is a longitudinal sectional view showing a first valve stop mechanism provided on the exhaust side of the present embodiment, B is a longitudinal sectional view showing the operation of the first valve stop mechanism, and C is a sectional view taken along line BB of A. is there. A is a longitudinal sectional view showing a second (third) valve stop mechanism provided on the exhaust valve (intake valve) side of the present embodiment, and B is a longitudinal section showing an operation of the second (third) valve stop mechanism. Drawing C is a CC sectional view taken on the line B. It is a longitudinal cross-sectional view which shows the hydraulic lash adjuster which is not equipped with the valve stop mechanism in this embodiment. It is the schematic which shows the control hydraulic circuit of this embodiment. It is a valve lift amount of an exhaust valve and an intake valve in this embodiment, and an operating angle characteristic view. A is an operation explanatory diagram when the exhaust valve provided with the valve stop mechanism according to the present embodiment is controlled to the maximum lift amount (L4), and B is an operation explanatory diagram when the exhaust valve is closed. It is. A is an operation explanatory diagram when the exhaust valve is controlled to the maximum lift amount (L4) of the exhaust valve not provided with the valve stop mechanism in the present embodiment, and B is an operation explanatory diagram when the exhaust valve is closed. It is. A is an operation explanatory diagram when the exhaust valve provided with the valve stop mechanism according to the present embodiment is controlled to the minimum lift amount (L1), and B is an operation explanatory diagram when each exhaust valve is closed. , C is an explanatory view of the lost motion action by the valve stop mechanism. The operation state of the intake valve on the # 1 cylinder side is shown, A is a peak lift state where the valve lift amount of the opened intake valve is LI, B is an action explanatory diagram showing the state of the closed intake valve, C is an explanatory view of the lost motion action by the valve stop mechanism. It is a figure which shows the valve stop operation area | region and all the cylinder operation area | regions on the map of the engine speed in this embodiment, and an engine torque. It is a figure which shows the operating characteristic of the intake-exhaust valve | bulb of # 1 cylinder and # 2 cylinder at the time of the change transfer to the operation area | region (A)-(D) shown in FIG. It is a perspective view which shows the valve operating apparatus by the side of the exhaust valve in 2nd Embodiment of this invention. It is a perspective view which shows the valve operating apparatus by the side of the intake valve in 2nd Embodiment. It is the schematic which shows the control hydraulic circuit of 2nd Embodiment. It is a figure which shows the valve stop operation area | region and all the cylinder operation area | regions on the map of the engine speed in this embodiment, and an engine torque. It is a figure which shows the operating characteristic of the intake-exhaust valve | bulb of # 1 cylinder and # 2 cylinder at the time of the change transfer to the operation area | region (A)-(D) shown in FIG.

Embodiments of a variable valve apparatus for a multi-cylinder internal combustion engine according to the present invention will be described below with reference to the drawings. In each embodiment, the present invention is applied to an in-line two-cylinder internal combustion engine with gasoline specifications.
[First Embodiment]
FIG. 1 shows the exhaust side valve operating devices of # 1 cylinder and # 2 cylinder, and FIG. 2 shows the intake side valve operating devices of # 1 cylinder and # 2 cylinder.

  In the exhaust side valve system, as shown in FIG. 1, a swing cam 7 of an exhaust VEL which is a variable lift mechanism described later opens and closes the exhaust valves 3a, 3a, 3b and 3b via a swing arm 6. It is like that. On the other hand, in the intake side valve system, as shown in FIG. 2, each rotary cam 73a of the intake cam shaft 73 directly opens and closes each intake valve 71a, 71a, 71b, 71b via each swing arm 74. It has become.

  The front (F) side # 1 cylinder in the right position shown in FIGS. 1 and 2 is a cylinder capable of cylinder deactivation, that is, a cylinder in which all intake valves and exhaust valves can be stopped, and the rear The (R) side # 2 cylinder does not deactivate the cylinder, and is always a normally operated cylinder in which at least one exhaust valve and intake valve operate. However, in the present embodiment, only the exhaust valve 3b on the rear (R) side of the # 2 cylinder of the normally operating cylinder can be stopped.

  In the exhaust side valve system shown in FIG. 1, the R-side exhaust valves 3a, 3b of each cylinder are provided with first valve stop mechanisms 11b, 11c, respectively, on hydraulic lash adjusters 10b, 10d, as will be described later. The exhaust valve 3a on the F side of the # 1 cylinder is provided with a second valve stop mechanism 11a. On the other hand, both intake valves 71a and 71a of the # 1 cylinder of the intake side valve system shown in FIG. 2 are provided with third valve stop mechanisms 11d and 11e in lassia adjusters 75a and 75b, respectively. Except for these, the hydraulic lash adjusters 10c, 75c, 75d of the intake and exhaust valves 3b (F side), 71b, 71b are not provided with valve stop mechanisms.

3 shows a cross section taken along line AA of FIGS. 1 and 2, that is, the intake side and exhaust side valve operating devices in the # 1 cylinder (cylinder that can be deactivated).
[Valve device on the exhaust side]
Explaining specifically the exhaust side valve gears of the # 1 and # 2 cylinders, as shown in FIGS. 1 and 3, one cylinder that opens and closes a pair of exhaust ports 2 and 2 formed in the cylinder head 1. A pair of exhaust valves is provided. That is, the first and second exhaust valves 3a and 3a are provided in the # 1 cylinder, and the first and second exhaust valves 3b and 3b are provided in the # 2 cylinder. Here, in each cylinder, the first exhaust valves 3a and 3b are arranged on the F side, and the second exhaust valves 3a and 3b are arranged on the R side.

  The pair of exhaust ports 2 and 2 of each cylinder are separated from the port opening 2a shown in FIG. 3 by the partition wall 1b shown by a broken line (dashed line) in FIG. 1c, and the downstream side of the front end portion 1c is free of the partition wall 1b, forming a collective exhaust port in which both exhaust ports are gathered, and its cross-sectional area is the exhaust port on the F side. The cross-sectional area of the port 2 and the sum of the cross-sectional areas of the R-side exhaust port 2 are set to substantially coincide with each other. The collective exhaust port of each cylinder is further gathered into one by an exhaust manifold (not shown). Therefore, the exhaust gas discharged from the combustion chambers of the cylinders to the exhaust ports 2 and 2 is gathered in the collective exhaust port on the way, and further, the collective exhaust ports of all the cylinders (two cylinders) in the exhaust manifold. Are gathered into one, sent to the downstream, and then discharged to the outside via the catalyst and further through the exhaust pipe and silencer.

  As shown in FIGS. 1 and 3, the exhaust VEL provided in each of the exhaust valves 3a to 3b is disposed along the front-rear direction of the engine on the upper side of each cylinder and has two drive cams 5a on the outer periphery. A pair of swing cams having a shaft 5 and cam surfaces 7b and 7b that are rotatably supported on the outer peripheral surface of the drive shaft 5 and open and close the exhaust valves 3a to 3b via the swing arms 6, respectively. 7, 7, a transmission mechanism 8 that converts the rotational force of each driving cam 5 a into a swinging force and transmits the swinging force to the swinging cam 7, and the operation of the exhaust valves 3 a to 3 b via the transmission mechanism 8. The control mechanism 9 controls the angle (valve opening period) and the lift amount.

  Further, in the cylinder head 1, the clearance between the swing arms 6 and the exhaust valves 3 a to 3 b and the clearance between the cam circles 7 b of the swing cams 7 are adjusted to zero lash. Four first to fourth hydraulic lashia adjusters 10a, 10b, 10c, and 10d that are fulcrum members (pivots) are held.

  That is, the first and second hydraulic lash adjusters 10a and 10b are disposed on the exhaust valve side of the # 1 cylinder, and the third and fourth hydraulic lash adjusters 10c and 10d are disposed on the exhaust valve side of the # 2 cylinder. It is installed.

  Here, the first hydraulic lash adjuster 10a is disposed on the F side of the # 1 cylinder, the second hydraulic lash adjuster 10b is disposed on the R side, and the third hydraulic lash adjuster 10c is disposed on the # 2 cylinder. The fourth hydraulic lash adjuster 10d is disposed on the R side, and is disposed on the R side.

  The first and second exhaust valves 3a and 3a of the # 1 cylinder are connected to the # 1 cylinder via the first and second hydraulic lash adjusters 10a and 10b on the # 1 cylinder side according to the engine operating state. There are provided second and first valve stop mechanisms (lost motion mechanisms) 11a and 11b for stopping the opening and closing operations of the first and second exhaust valves 3a and 3a, respectively.

  On the R side second exhaust valve 3b side of the # 2 cylinder, another opening / closing operation of the second exhaust valve 3b is stopped via the R side fourth hydraulic lash adjuster 10d on the # 2 cylinder side. A one-valve stop mechanism (lost motion mechanism) 11c is provided.

  Further, on the exhaust side, as shown in FIG. 7, at the F side end of the drive shaft 5, the valve timing control for making the opening and closing timings of the exhaust valves 3a to 3b variable according to the engine operating state. A device (exhaust VTC) is provided. The exhaust VTC may be a normal type in which a vane rotor (not shown) is phase-converted by hydraulic pressure, for example.

  Hereinafter, the components in the # 1 and # 2 cylinders will be described. The four exhaust valves 3a to 3b are slidably held by the cylinder head 1 via the valve guides 4 and each stem end. Each valve retainer 3d provided in the vicinity of 3c is urged in a closing direction by each valve spring 12 elastically contacted between the inner upper surface of the cylinder head 1.

  The drive shaft 5 is rotatably supported by a plurality of bearing portions 13 provided at an upper end portion of the cylinder head 1 via a cam shaft 7a of the swing cam 7, and the exhaust VTC described above provided at one end portion. The rotational force of the crankshaft is transmitted by the timing belt via a timing pulley provided in the housing (not shown). The drive cam 5a provided on the outer periphery of the drive shaft 5 per cylinder has an axis Y that is eccentric in the radial direction from the axis X of the drive shaft 5, and the outer cam profile is circular. Is formed.

  Each of the swing arms 6 has a flat or slightly convex lower surface of one end portion 6a in contact with each stem end 3c of each of the exhaust valves 3a to 3b, while a lower surface concave portion 6c of the other end portion 6b The rollers 14 are in contact with the heads of the hydraulic lash adjusters 10a to 10d, and are rotatably accommodated in the accommodation holes formed in the center via the roller shafts 14a.

  As shown in FIGS. 1 and 3, each of the swing cams 7 has a cam surface 7b formed of a base circle surface, a ramp surface and a lift surface on the lower surface at both ends of a cylindrical cam shaft 7a. The base circle surface, the ramp surface, and the lift surface are in rolling contact with the upper surface of the roller 14 of the swing arm 6 in accordance with the swing position of the swing cam 7.

  In the camshaft 7a, a journal portion formed at a substantially central position in the axial direction of the outer peripheral surface is rotatably supported by the bearing portion 13 with a minute clearance, and the outer peripheral surface of the drive shaft 5 is supported by the inner peripheral surface. It is designed to be freely supported.

  Each transmission mechanism 8 includes a rocker arm 15 disposed above the drive shaft 5, a link arm 16 that links the one end 15a of the rocker arm 15 and the drive cam 5a, and the other end 15b of the rocker arm 15. A link rod 17 that links the swing cam 7.

  The rocker arm 15 has a cylindrical base portion at the center thereof rotatably supported by a control cam, which will be described later, via a support hole, and one end portion 15 a is rotatably connected to the link arm 16 by a pin 18. On the other hand, the other end 15 b is rotatably connected to the upper end of the link rod 17 via a pin 19.

  In the link arm 16, the cam body of the drive cam 5a is rotatably fitted in a fitting hole 16a at the center position of an annular base portion, while the protruding end is connected to the rocker arm one end portion 15a by the pin 18. It is connected.

  The link rod 17 has a lower end portion rotatably connected to a cam nose portion on which one cam surface 7 b of the swing cam 7 is formed via a pin 20.

  An adjustment mechanism 23 is provided between the other end 15b of each rocker arm 15 and the upper end of the link rod 17 to finely adjust the lift amount of each exhaust valve 3a-3b when each component is assembled. ing.

  The control mechanism 9 is slidably fitted in a support hole of the rocker arm 15 on the outer periphery of the control shaft 21, and is rotatably supported on the same bearing portion above the drive shaft 5. Two control cams 22 serving as rocking fulcrums of the rocker arms 15 are fixed.

  The control shaft 21 is arranged in the longitudinal direction of the engine in parallel with the drive shaft 5 and is rotationally controlled by an actuator 50 shown in FIG. On the other hand, the control cam 22 has a cylindrical shape and its axis is offset from the axis of the control shaft 21 by a predetermined amount.

  As shown in FIG. 7, the actuator 50 is provided inside the housing with an electric motor 51 fixed to one end of the housing (not shown), and the rotational driving force of the electric motor 51 is applied to the control shaft 21. As a speed reduction mechanism for transmission, a ball screw mechanism 52 including a ball screw element, a conversion link, and the like is configured.

  The electric motor 51 is composed of a proportional DC motor, and is controlled to rotate forward and backward by a control signal from a control unit 53 that detects an engine operating state.

  As shown in FIGS. 4 to 6, the four hydraulic lash adjusters 10 a to 10 d include a bottomed cylindrical body 24 held in each cylindrical holding hole 1 a of the cylinder head 1, and the body 24. The plunger 27 is accommodated in the lower portion of the body 24 and is formed in the lower portion of the body 24 through a partition wall 25 which is slidable in the vertical direction and is integrally formed in the lower portion. A high-pressure chamber 28 that communicates with the reservoir chamber 26 through the communication hole 25a, and is provided inside the high-pressure chamber 28 to allow the hydraulic oil in the reservoir chamber 26 to flow only in the direction of the high-pressure chamber 28. And a check valve 29. Further, a discharge hole 1b for discharging the hydraulic oil accumulated in the holding hole 1a to the outside is formed in the cylinder head 1.

  The body 24 has a cylindrical first concave groove 24a formed on the outer peripheral surface thereof, and is formed on the peripheral wall of the first concave groove 24a inside the cylinder head 1 so that the downstream end is the first concave groove. A first passage hole 31 that communicates between the oil passage 30 opened in the groove 24a and the inside of the body 24 is formed penetrating in the radial direction.

  Also, the bodies 24 of the first and second hydraulic lash adjusters 10a and 10b (F, R side) of the # 1 cylinder and the fourth hydraulic lash adjuster 10d (R side) of the # 2 cylinder are shown in FIGS. As shown in FIGS. 5A and 5B, the bottom 24b side is lower than the bottom 24c of the body 24 of the third hydraulic lash adjuster 10c (F side) on the # 2 cylinder side where the valve stop mechanism shown in FIG. 6 is not provided. It extends in the direction and is formed in a substantially cylindrical shape.

  As shown in FIG. 3, the oil passage 30 communicates with a main oil gallery 30a for supplying lubricating oil formed in the cylinder head 1. The main oil gallery 30a includes an oil pump shown in FIG. Lubricating oil is pumped from 54 or 64.

  As shown in FIGS. 4 to 6, the plunger 27 has a cylindrical second concave groove 27 a formed on the outer peripheral surface at the substantially central portion in the axial direction, and the second concave groove 27 a has the first groove on the peripheral wall. A second passage hole 32 communicating with the first passage hole 31 and the reservoir chamber 26 is formed penetrating along the radial direction. Further, the distal end surface of the distal end head portion 27b of each plunger 27 is formed in a spherical shape in order to ensure good slidability with the spherical lower surface concave portion 6c of the other end portion 6b of each swing arm 6.

  Each plunger 27 has its maximum protruding amount regulated by an annular stopper member 33 fitted and fixed to the upper end portion of the body 24.

  The second concave groove 27a is formed to have a relatively large width in the axial direction, whereby the first passage hole 31 and the second passage hole 32 are formed at any of the vertically sliding positions of the plunger 27 with respect to the body 24. It always comes to communicate.

  Each check valve 29 includes a check ball 29a that opens and closes a lower opening edge (seat) of the communication hole 25a, a first coil spring 29b that urges the check ball 29a in a closing direction, and the first coil spring 29b. The retainer 29c is held between the inner bottom surface of the bottom wall 24c of the body 24 and the annular upper end of the retainer 29c, and the plunger 27 as a whole is urged toward the partition wall 25. And a second coil spring 29d for energizing the upper part of the coil spring.

  In the base circle section of the cam surface 7b of the rocking cam 7, when the pressure in the high pressure chamber 28 becomes low as the plunger 27 moves forward (upward movement) due to the urging force of the second coil spring 29d, the oil The hydraulic oil supplied from the passage 30 into the holding hole 1a flows into the reservoir chamber 26 from the first concave groove 24a through the first passage hole 31, the second concave groove 27a and the second passage hole 32, and further check. The ball 29 a is pushed open against the spring force of the first coil spring 29 b, and hydraulic oil flows into the high pressure chamber 28.

  As a result, the plunger 27 pushes up the other end 6 b of the swing arm 6 and contacts the roller 14 and the swing cam 7 to contact the swing cam 7, one end 6 a of the swing arm 6, and the stem of each intake valve 3. The gap with the end 3a is adjusted to zero lash.

  In the lift section of the swing cam 7, a downward load acts on the plunger 27, so that the hydraulic pressure in the high pressure chamber 28 rises, and the oil in the high pressure chamber 28 leaks from the gap between the plunger 27 and the body 24. The plunger 27 is slightly lowered (leak down).

  When the base circle section of the cam surface 7b of the oscillating cam 7 is reached again, as described above, the plunger 27 moves forward (upward movement) by the urging force of the second coil spring 29d, so that the gaps of the respective parts are reduced to zero. It is adjusted to.

  All of the first to fourth hydraulic lash adjusters 10a to 10d have such a lash adjustment function.

  The lost motion mechanisms (second and first valve stop mechanisms) 11a, 11b, and 11c are the first and second hydraulic lash adjusters 10a and 10b on the F side and the R side of the # 1 cylinder and the # 2 cylinder. This is provided only on the fourth hydraulic lash adjuster 10d on the R side, and is not provided on the third hydraulic lash adjuster 10c on the F side of the # 2 cylinder, as shown in FIG.

  That is, the second and fourth hydraulic lash adjusters 10b and 11d on the R side of each of the # 1 and # 2 cylinders, and the first hydraulic lash adjuster 10a side on the F side of the # 1 cylinder The second valve stop mechanism 11a is provided so that the valve stop and the valve operation can be switched according to the engine operating state as will be described later. On the other hand, the valve stop mechanism is not provided on the F side of the # 2 cylinder, and therefore has only a normal pivot function and a zero lash adjustment function.

  Further, the first valve stop mechanisms 11b and 11c and the second valve stop mechanism 11a are different in a part of the structure (regulation mechanism) as shown in FIGS.

  As shown in FIGS. 4A and 4B, the first valve stop mechanisms 11b and 11c include a pair of cylindrical sliding holes 34 formed continuously on the bottom side of each holding hole 1a and the sliding holes. A pair of lost motion springs 35 that are elastically mounted between the bottom surface of 34 and the lower surface of the body 24 to urge the second and fourth hydraulic lashia adjusters 10b and 10d upward, respectively. And a pair of regulating mechanisms 36 that regulate the lost motion of the hydraulic lash adjusters 10b and 10d.

  Each sliding hole 34 has an inner diameter set to be the same as the inner diameter of the holding hole 1a, and each body 24 is held so as to be slidable vertically from the holding hole 1a.

  Each of the lost motion springs 35 is formed by a coil spring, and urges the bottom surface of the body 24 upward to cause the distal end head portion 27b of the plunger 27 to become a recess 6c on the lower surface of the other end portion 6b of the swing arm 6. It is supposed to be in contact with

  Further, the maximum upward movement position of each body 24 is regulated by a stopper pin 37 inserted and arranged inside the cylinder head 1. That is, the stopper pins 37 are disposed in the cylinder head 1 in the direction perpendicular to the axis toward the body 24, and the distal end portion 37a is slidably disposed in the first concave groove 24a. When the tip portion 37a comes into contact with the lower end edge of the first concave groove 24a in accordance with the upward movement of the body 24, the maximum sliding position of the body 24 is regulated.

  Therefore, each hydraulic lash adjuster 10b, 10d is lost by vertically moving between the holding hole 1a and the sliding hole 34 through the spring force of the lost motion spring 35 as the swing arm 6 swings. By performing the motion, the function as the swing fulcrum of the swing arm 6 is lost, the lift operation of the swing cam 7 is absorbed, and the opening / closing operation of each exhaust valve 3a is stopped.

  As shown in FIGS. 4A to 4C, the first regulating mechanisms 36 of the first valve stop mechanisms 11 b and 11 c include a movement hole 38 formed so as to penetrate in the inner radial direction of the bottom 24 b of the body 24, and the cylinder. A restriction hole 39 formed in a direction perpendicular to the holding hole 1 a in the head 1, a spring support retainer 40 fixed to one end of the inside of the movement hole 38, and a slide inside the movement hole 38. A slidable pin 41 slidably provided in the movement hole 38 and a columnar restriction pin (movable across the movement hole 38 and the restriction hole 39). (First restriction pin) 42, elastically mounted between the rear end of the slide pin 41 and the retainer 40, and urges the restriction pin 42 toward the restriction hole 39 via the slide pin 41. And a return spring 43.

  The restriction hole 39 is adapted to coincide with the movement hole 38 from the axial direction when the body 24 is restricted to the maximum upper position by the stopper pin 37, and the inner diameter thereof is the movement hole 38. The signal hydraulic pressure is introduced from an oil passage hole 44 formed in the cylinder head 1 on one end side.

  Here, the regulation of the rotation direction of the body 24 is to slightly increase the amount of protrusion of the stopper pin 37, and to provide a slit in the longitudinal direction in the first concave groove 24a of the body 24. This can be easily realized by engaging the lower end with the tip of the stopper pin 37. Alternatively, a separate rotation restricting member may be mounted between the cylinder head 1 and the body 24.

  The retainer 40 is formed in a cylindrical shape with a lid, a breathing hole 40a for ensuring smooth movement of the sliding pin 41 is formed in the bottom wall, and a central portion where the breathing hole 40a on the rear end face faces. Although 40b is formed flat, the outer end portions 40c and 40c are formed in a circular arc surface shape having substantially the same curvature as the inner peripheral surface of the sliding hole 34 in order to ensure smooth slidability. . Further, as shown in FIG. 4B, the length of the retainer 40 in the axial direction is such that the rear end edge of the restricting pin 42 is brought into contact with the front end edge before the restricting pin 42 is completely accommodated in the movement hole 38. It is set to a length that restricts further backward movement. It should be noted that a slight amount of hydraulic oil leaking to the moving hole 38 is guided into the sliding hole 34 through the breathing hole 40 a through the outer surface of the bottom wall of the retainer 40 and the inner peripheral surface of the sliding hole 34. It has become.

  As shown in FIGS. 4A and 4C, the sliding pin 41 is formed in a covered cylindrical shape, and has an outer diameter slightly smaller than the inner diameter of the moving hole 38 to ensure smooth slidability. At the same time, the tip surface 41a of the tip portion is formed in an arc surface shape having the same curvature as the inner peripheral surface of the sliding hole 34 in order to ensure smooth slidability.

  The restriction pin 42 is formed in a solid cylindrical shape, and has an axial length substantially the same as the axial length of the restriction hole 39. As shown in FIG. 4A, the spring of the return spring 43 is formed. When the force is moved into the restriction hole 39 via the sliding pin 41 by force, the whole is accommodated in the restriction hole 39. As a result, the R-side hydraulic lash adjusters 10b and 10d of the # 1 and # 2 cylinders are moved in the vertical direction, that is, lost motion is performed.

  The regulating pin 42 is formed with an outer diameter slightly smaller than the inner diameters of the moving hole 38 and the regulating hole 39 so as to ensure smooth slidability with respect to them, and the oil passage. As shown in FIG. 4B, when the hydraulic pressure supplied from the hole 44 to the restriction hole 39 is received by the tip end surface 42a as a flat pressure receiving surface, the spring force of the return spring 43 is resisted in the left direction in the figure. When the sliding pin 41 moves and comes into contact with the retainer 40 from the axial direction, the tip end portion is accommodated from the restriction hole 39 into the movement hole 38, and the second and second cylinders of the # 1, # 2 cylinders are accommodated. The fourth hydraulic lash adjuster 10b, 10d is restricted from moving in the vertical direction, that is, the lost motion is restricted and locked to the cylinder head 1.

  As shown in FIG. 7, the oil pressure fed from the first oil pump 54 is supplied to the oil passage hole 44 (regulation hole 39) as a signal oil pressure via the first electromagnetic switching valve 55. It has become. That is, the first electromagnetic switching valve 55 is a switching energy changing means (first hydraulic pressure changing means) that converts a state of supplying hydraulic pressure as switching energy and a state of stopping supply.

  The first electromagnetic switching valve 55 (first hydraulic pressure supply / supply stop conversion means) is configured such that a spool valve slidably provided inside a valve body (not shown) is connected to an electromagnetic force of a solenoid and a spring force of a coil spring. Thus, it is switched on and off in two stages. The solenoid is energized and de-energized (on and off) from the same control unit (controller) 53 that controls the driving of the electric motor 51, thereby providing a pump discharge passage and a first oil passage hole 44. The first signal oil pressure is supplied to the first restriction pin 42 through communication, or the pump discharge passage is closed and the oil passage hole 44 and the drain passage 45 are communicated to be switched. .

  Therefore, when the engine is stopped, the control unit 53 does not energize the solenoid, and the first electromagnetic switching valve 55 closes the pump discharge passage and connects the oil passage 44 and the drain passage 45. Therefore, the first valve stop mechanisms 11b, 11c Lost motion by is ready to operate. That is, the first valve stop mechanisms 11b and 11c are of a valve stop stable type that mechanically stabilizes in a valve stop state when the supply of hydraulic pressure as switching energy is stopped.

  On the other hand, in the second valve stop mechanism 11a, as shown in FIGS. 5A to 5C, the second restriction mechanism 46 is different in structure from the first restriction mechanism 36 of the first valve stop mechanisms 11b and 11c, and the sliding pin 41 and It is formed integrally with the first restriction pin 42.

  That is, the same constituent members as those of the first restricting mechanism 36 are denoted by the same reference numerals, and will be briefly described. The second restricting mechanism 46 is a moving member formed so as to penetrate in the inner radial direction of the bottom 24b of the body 24. A hole 38, a restriction hole 39 formed in a direction perpendicular to the holding hole 1 a in the cylinder head 1, a retainer 40 fixed to one end of the movement hole 38, and the movement hole 38 A second restriction pin 47 that is slidably provided inside and is movable from the movement hole 38 to the restriction hole 39, and between the rear end of the second restriction pin 47 and the retainer 40. And a return spring 49 that urges the second restricting pin 47 toward the restricting hole 39.

  As shown in FIG. 5B, the retainer 40 has a rear end of the second restricting pin 47 at the leading edge when the second restricting pin 47 is completely accommodated in the movement hole 38 as shown in FIG. 5B. The length is set so as to abut and restrict further backward movement.

  The second restricting pin 47 is formed in a cylindrical shape, and has a solid tip portion 47 a extending in the axial direction. The outer diameter is slightly smaller than the inner diameters of the moving hole 38 and the restricting hole 39. In order to ensure smooth slidability. Further, the second regulating pin 47 receives the hydraulic pressure supplied from the second oil passage hole 48 to the regulating hole 39 as a pressure receiving surface, and the tip surface 47a receives the hydraulic pressure supplied to the regulating hole 39, as shown in FIG. It moves backward against the spring force, the tip part comes out of the restriction hole 39 and the whole is accommodated in the movement hole 38, and the restriction is released. Further, as shown in FIG. 5C, the second restricting pin 47 is formed in an arcuate surface shape having a curvature substantially the same as the inner peripheral surface of the sliding hole 34 in order to ensure good slidability of the tip surface 47a. Has been. Further, the retainer 40 has a flat central portion 40b facing the breathing hole 40a on the rear end surface, but the outer end portions 40c and 40c have the sliding hole 34 in order to ensure smooth slidability. It is formed in the shape of a circular arc surface having substantially the same curvature as the inner peripheral surface.

  As shown in FIG. 7, the hydraulic pressure fed from the second oil pump 64 is supplied to the second oil passage hole 48 as the second signal hydraulic pressure via the second electromagnetic switching valve 65. .

  The second electromagnetic switching valve 65 (second hydraulic pressure supply / supply stop conversion means) is configured such that a spool valve slidably provided inside a valve body (not shown) is connected to an electromagnetic force of a solenoid and a spring force of a coil spring. Thus, the solenoid is switched on and off in two stages, and the control current is supplied to the solenoid from the control unit 53, and the pump discharge passage and the second oil passage hole 48 communicate with each other. The second control oil pressure is supplied to the second restriction pin 47, or the pump discharge passage is closed and the second oil passage hole 48 and the second drain passage 66 are communicated to be controlled. Yes.

  Therefore, when the engine is stopped, the control unit 53 does not energize the solenoid, and the second electromagnetic switching valve 64 closes the pump discharge passage and makes the second oil passage 48 and the second drain passage 45 communicate with each other. The lost motion by the stop mechanism 11 a does not operate, and the first hydraulic lash adjuster 10 a is in a valve operating mode in which it is locked to the cylinder head 1. That is, the second valve stop mechanism 11a is a valve operation stable type that mechanically stabilizes in a valve operation state when the supply of hydraulic pressure as switching energy is stopped.

  The control unit 53 detects an engine operating state based on information signals such as an engine speed, a load, and a throttle valve opening amount from various sensors such as a crank angle sensor, an air flow meter, a water temperature sensor, and a throttle valve angle sensor. At the same time, the electric motor 51 is driven and controlled by the information signal (VEL control shaft actual position signal) from the rotational position sensor (not shown) for detecting the engine operating state and the current rotational position of the control shaft 21 to control the control shaft. The rotational position of 21 is controlled. As a result, the lift amounts and operating angles of the four exhaust valves 3a to 3b of the # 1 and # 2 cylinders are changed.

  That is, as shown in FIG. 8, the lift amount of each of the exhaust valves 3a to 3b is determined so that the corresponding hydraulic lash adjusters 10a, 10b, and 10d are not lost by the second and first valve stop mechanisms 11a to 11c. When the valve is not stopped due to being locked by the cylinder head 1, it is changed in the range from the minimum L1 to the maximum L4.

  The operation of the exhaust VEL when each of the exhaust valves 3a to 3b is controlled to the maximum lift amount L4 is as shown in FIGS. FIG. 9 shows a valve stop mechanism, for example, the R side of the # 1 cylinder, and the same applies to the R side of the # 2 cylinder and the F side of the # 1 cylinder. Stop mechanisms 11b, 11c, and 11a are provided. The state shown in FIG. 9 shows a valve operating state in which, for example, the second hydraulic lash adjuster 10b is locked to the cylinder head without the lost motion by the first valve stop mechanism 11b. The rotation angle of the control shaft 21 is θ4 corresponding to the maximum lift amount L4, the drive cam 5a rotates in the clockwise direction, reaches the maximum lift L4 at the position shown by the peak lift in FIG. 9A, and the position shown in FIG. 9B. The valve is closed. The same applies to the R side of the # 2 cylinder having the first valve stop mechanism 11c, and the F side of the # 1 cylinder has the same effect by simply replacing the second valve stop mechanism 11a. These valves operate (valve opening / closing operation) with similar lift characteristics.

  On the other hand, FIG. 10 shows an exhaust valve 3b and a third hydraulic lash adjuster 10c on the # 2 cylinder F side that is not provided with a valve stop mechanism, and no valve stop mechanism is provided here. The valve is operated with the same lift characteristic as the R side of the # 2 cylinder in the case of the valve operation mode shown in FIG.

  FIGS. 11A and 11B are explanatory views of the operation when the intake valve is controlled to the minimum lift. For example, on the R side of the # 1 cylinder on the intake side, the VEL when the exhaust valve 3a is controlled to the minimum lift amount L1. The operation of is shown. When the first valve stop mechanism 11b does not perform the valve stop operation and the second hydraulic lash adjuster 10b does not perform the lost motion and is locked to the cylinder head, the lift amount is controlled to L1. On the other hand, when the first valve stop mechanism 11b is in the valve stop operation (lost motion operation), as shown in FIG. 11C, the second hydraulic lash adjuster 10b performs the lost motion by M1, and the lift amount is zero. This is the valve stop mode (state) that continues this state.

  The operation on the R side of the other # 2 cylinder provided with the first valve stop mechanism is the same as the R side of the # 1 cylinder shown in FIGS. Also. The F-side of the # 1 cylinder provided with the second valve stop mechanism also has the same operation as shown in FIG. Further, the F side of the # 2 cylinder not provided with the valve stop mechanism has the same lift characteristics as the R side of the # 1 cylinder in the valve operation mode shown in FIGS.

Further, the control unit 53 outputs a control signal to an exhaust VTC electromagnetic switching valve (not shown) of the exhaust VTC in accordance with the engine operating state, so that the control unit 53 receives the hydraulic pressure discharged from the oil pump 54 or the oil pump 64. The rotation phase of the drive shaft 5 is made variable by rotating the vane rotor (not shown) relative to the crankshaft relative to the advance side or the retard side. Thereby, the opening / closing timing of each exhaust valve 3a, 3a, 3b, 3b and the phase of the peak lift are controlled.
[Valve on the intake side]
The intake side valve operating device is shown in FIGS. 2, 3 and 7, and does not have a variable lift mechanism (exhaust VEL) as in the exhaust side, but is similar to the exhaust VTC described above. A timing control mechanism (intake VTC) is provided, and only the # 1 cylinder is provided with a valve operation stable type valve stop mechanism.

  That is, as shown in FIGS. 2 and 3, two intake valves 71a, 71a, 71b, 71b are provided per cylinder for opening and closing a pair of intake ports 70, 70 formed in the cylinder head 1 per cylinder. It has been. That is, the first and second intake valves 71a and 71a on the F side and the R side are provided in the # 1 cylinder, and the first and second intake valves 71b and 71b on the F side and the R side are provided in the # 2 cylinder.

  As an intake side valve operating device, an egg-shaped rotation is arranged on the upper side of each cylinder along the longitudinal direction of the engine and opens the intake valves 71a to 71b on the outer periphery against the spring force of the valve springs 72. An intake camshaft 73 having a cam 73a is provided, and the intake valves 71a to 71a are provided via rollers 77 and swing arms 74 interposed between the intake valves 71a to 71b and the rotary cams 73a. 71b is opened and closed with the constant valve lift LI shown in FIGS. 8 and 12A as a peak lift.

  The support member (pivot) is held by the cylinder head 1 and adjusts the gap between each swing arm 74 and each intake valve 71a to 71b and the base circle of each rotary cam 73a to zero lash. Hydraulic lash adjusters 75a to 75d are respectively provided. In other words, there are also four hydraulic lash adjusters 75a to 75d on the intake side, the first and second hydraulic lash adjusters 75a and 75b are provided in the # 1 cylinder, and the third and fourth hydraulic lash adjusters 75c are provided in the # 2 cylinder. , 75d.

  Here, the first hydraulic lash adjuster 75a is disposed on the F side of the # 1 cylinder, the second hydraulic lash adjuster 75b is disposed on the R side, and the third hydraulic lash adjuster 75c is disposed on the # 2 cylinder. The fourth hydraulic lash adjuster 75d is disposed on the F side, and is disposed on the R side.

The first and second hydraulic lash adjusters 75a and 75b on the F-side and R-side intake valves 71a and 71a side of the # 1 cylinder shown in FIG. 2 are respectively provided with a lost motion mechanism (third valve stop mechanism 11d, 11e). The third valve stop mechanisms 11d and 11e have the same configuration as the F-side second valve stop mechanism 11a of the exhaust valve side # 1 cylinder shown in FIG. That is, when the supply of the hydraulic pressure that is the switching energy is stopped, the valve operation is stable so that the valve operation state is mechanically stabilized. While the second valve stop mechanism 11a is on the exhaust valve 3a side, these are used on the intake valves 71a and 71a side, so that they are referred to as third valve stop mechanisms 11d and 11e in order to avoid confusion.
On the other hand, the third and fourth hydraulic lash adjusters 75c and 75d of the intake valves 71b and 71b on the F and R sides of the # 2 cylinder do not have a valve stop mechanism.

  Since the third valve stop mechanisms 11d and 11e have the same structure as the second valve stop mechanism 11a shown in FIG. 5 described above, the same reference numerals in FIG. . That is, it is elastically mounted between a cylindrical sliding hole 34 formed continuously on the bottom side of each holding hole 1a of the cylinder head 1, and a bottom surface of the sliding hole 34 and a lower surface of the body 24, Lost motion springs 35 and 35 for urging the first and second hydraulic lash adjusters 75a and 75b upward, and a third regulating mechanism 76 for regulating the lost motion of the first and second hydraulic lash adjusters 75a and 75b. And is composed of.

  The first and second intake valves 71a and 71a on the # 1 cylinder side provided with the third valve stop mechanisms 11d and 11e are stopped by the lost motion as indicated by the broken line in FIG. When lifted, the lift is zero, and when the valve is not stopped, the peak lift amount is constant LI, which is slightly larger than the exhaust valves 3a to 3b of the # 1 and 2 cylinders shown by the solid line in FIG. It is substantially the same as the lift L3 and is set to be smaller than the maximum lift amount L4.

  FIG. 12 shows the operating state of the # 1 cylinder on the intake valve side, A shows the state where the first and second intake valves 71a, 71a are operating with a slightly large constant peak lift amount LI (≈L3), B indicates the closed state of the first and second intake valves 71a and 71a, and C indicates the lost motion operating state (valve stopped state) of the first and second lash adjusters 75a and 75b by the third valve stop mechanisms 11d and 11e. ing.

  Also, as shown in FIG. 12C, the lost motion amount of the hydraulic lash adjusters 75a and 75b by the third valve stop mechanisms 11d and 11e on the intake valves 71a and 71a side of the # 1 cylinder is relatively large as M3. The angle α3 formed by the arm 74 and the lost motion direction is also a relatively large value. Here, more specifically, the angle α3 is an angle formed by a line connecting the swing fulcrum of the swing arm 74 and the rotation center of the roller and an axis as the lost motion direction of the hydraulic lash adjuster.

  However, if it is about α3, the contact between the heads of the hydraulic lash adjusters 75a and 75b and the recesses of the swing arms 74 does not float even when the rotation speed is high, and a smooth lost motion operation is obtained. It is done. Conversely, the value of M3 (α3) is a value within a range where a smooth lost motion operation can be obtained. Here, if it is assumed that it has further increased to M4 (α4), when the rotation speed is high, the contact portion becomes non-uniform or local contact, and a deviation occurs between them. In some cases, a floating (separation) occurs, and a smooth lost motion operation cannot be obtained, resulting in irregular behavior. Considering this, a predetermined engine rev limit shown in FIG. 13 is provided.

  On the other hand, the third and fourth lash adjusters 75c and 75d on the intake side of the # 2 cylinder that do not have the valve stop mechanism have the same structure as the third lassia adjuster 10c on the exhaust side of the # 2 cylinder shown in FIG. is there.

  The cam profile of the rotating cam 73a is the same so that the fixed valve lift amount of each intake valve 71b, 71b of the # 2 cylinder is the same as that of the intake valve 71a, 71a of the # 1 cylinder. Is set to

  As shown in FIG. 12A, the angle β3 formed between the swing arm 74 in the peak lift state and the lost motion direction is close to the ideal 90 °, and even if the valve jumping or the like occurs in the high rotation range, the swing arm The lateral displacement between the head 74 and the hydraulic lash adjuster head 27 is less likely to occur, and the swing arm 74 is less likely to come off.

  Further, on the exhaust side, the angle formed with the lost motion direction of the swing arm 6 in the peak lift state is close to an ideal 90 ° as shown by β4 in the lift amount L4 control shown in FIGS. 9A and 10. In the same manner, irregular behavior such as swing arm disengagement during valve operation is less likely to occur. That is, the difference between β3 to β4 and 90 ° at the time of valve operation is smaller than the difference between α3 and 90 ° at the time of lost motion, and irregular behavior such as the swing arm coming off at the time of valve operation is less than that at the time of lost motion. It is hard to occur.

The intake VTC described above has the same structure as the exhaust VTC described above, and the control unit 53 outputs a control signal to an intake VTC electromagnetic switching valve (not shown) of the intake VTC according to the engine operating state. The rotational phase of the drive shaft 5 is made variable by rotating the vane rotor (not shown) relative to the crankshaft toward the advance side or the retard side via the hydraulic pressure discharged from the oil pump 54 or the oil pump 64. Thereby, the opening / closing timing (lift phase) of each of the intake valves 71a to 71b is controlled. Here, the oil pump that supplies hydraulic pressure to the intake VTC electromagnetic switching valve may be shared with the oil pump that supplies hydraulic pressure to the exhaust VTC electromagnetic switching valve, or may be provided separately. In the former case, the engine system structure is simplified, and in the latter case, conversion responsiveness of each VTC is improved.
[Operation of variable valve gear]
Hereinafter, the operation of the variable valve operating apparatus in the present embodiment will be described.

  Since the oil pumps 54 and 64 are not in operation when the engine is stopped, the signal oil pressure is inactive or low regardless of the on / off positions of the first and second electromagnetic switching valves 55 and 65. Since the one-valve stop mechanisms 11b and 11c are valve-stop stable types, the valve-stop mode, that is, the state in which the lost-motion operation can be performed, while the second valve stop mechanism 11a and the third valve stop mechanisms 11d and 11e operate Since it is a stable type, it is in a valve operating mode.

  Therefore, as shown in “cylinder operation” in the operation region A of FIG. 14, the two exhaust valves 3a and 3b on one side (R side) of both cylinders # 1 and # 2 are in the valve stop state, and the other two on the F side The exhaust valves 3a and 3b and all the intake valves are in a drive (valve operation) state.

  Even when cranking for starting the engine is started and starting combustion is started, the oil pressure of the oil pumps 54 and 64 does not suddenly rise, and the state (mode) described above is maintained.

  Further, both the signal of the first electromagnetic switching valve 55 (converting the first valve stop mechanism) and the signal of the second electromagnetic switching valve 65 (converting the second and third valve stop mechanisms) are turned off, that is, each signal hydraulic pressure is It communicates with the drain passages 45 and 66 instead of the pump hydraulic pressure, and each signal hydraulic pressure is in a state where only a low pressure can act. Even if the hydraulic pressure of the oil pumps 54 and 64 rises early, the above-mentioned state is maintained. It can be done.

  Here, when the engine is started, the friction of each part of the engine is increased due to the low engine temperature, and at the time of starting, the engine temperature is low and the combustion tends to be poor.

On the other hand, in the present embodiment, the exhaust valves 3a to 3b of all the cylinders are in a single valve stop (exhaust valve single valve stop) mode before start cranking, so that the effect of reducing valve friction from the initial stage of cranking. Is obtained. As a result, the start-up friction is reliably reduced, and the combustion gas flows only to one of the exhaust ports. Therefore, the port surface area with which the exhaust gas contacts is reduced by half, and the wall surface cooling is suppressed. Further, the gas flow in the exhaust port 2 is strengthened by stopping the one-valve, so that the reaction of the untwisted component in the exhaust gas is promoted, and the catalyst heats up quickly due to the reaction heat, and the early activation Is promoted. As a result, exhaust emission at the time of starting can also be reduced.
As described above, before cranking, the first valve stop mechanism is mechanically stable in advance in the valve stop mode, and the second valve stop mechanism is mechanically stable in advance in the valve operation mode. Since both cylinders (2 cylinders) are mechanically preliminarily set in the exhaust single valve stop mode, the above-described valve friction reduction effect and the like can be reliably obtained without delay from the beginning of cranking. Further, even if each electromagnetic switching valve that controls each valve stop mechanism has an electrical failure such as disconnection, it is mechanically in the above-described all-cylinder exhaust single valve stop mode, so startability In other words, it has a so-called mechanical fail-safe effect.

  In addition to these effects, in the present embodiment, the drive-side exhaust valves 3a and 3b on the F side of the # 1 and # 2 cylinders also have a small lift amount and a small operating angle (lift amount L1, actuating) by the electric exhaust VEL. Since the angle D1) can be obtained, the friction of the engine sliding portion is further reduced and the startability is further improved.

  Further, the opening timing of the exhaust valves 3a and 3b is delayed to near the bottom dead center due to the reduction of the operating angle, and the closing timing of the exhaust valve is a timing advanced sufficiently from the top dead center.

  That is, the opening time of the exhaust valve is delayed to near the bottom dead center, so that the period until the high-temperature combustion gas in the cylinder is discharged is extended, so that the engine itself can be warmed effectively, Therefore, the engine temperature rise can be promoted.

On the other hand, the exhaust valve closing timing is a timing that is sufficiently advanced from the top dead center (the piston is much before the top dead center), so the exhaust valve remains in a state where a large amount of high-temperature combustion gas (exhaust gas) remains in the cylinder. 3a and 3b are closed. The combustion gas is compressed as the piston rises, and the engine is further warmed as the temperature rises. As described above, the engine itself is quickly warmed up.
[When combustion gas is discharged from the exhaust port]
Next, the case where combustion gas is discharged from the exhaust port 2 will be considered.

  Here, as described above, the partition wall 1b exists between the F-side exhaust port 2 and the R-side exhaust port 2 provided independently. In both the # 1 and # 2 cylinders, the R-side exhaust valves 3a and 3b are in the valve stop state, so that the combustion gas flows out from the F-side exhaust port 2 on the valve operating side at a high flow rate (high gas flow). Due to this high gas flow, oxidation reactions such as unburned HC and PM (particulate matter particulate matter) in the exhaust gas (combustion gas) proceed (so-called afterburning), and these cause the temperature increase effect of the exhaust gas, This suppresses the exhaust gas temperature drop in the downstream.

  Further, immediately after the high temperature exhaust gas is discharged from the cylinder, the exhaust gas contacts only with the exhaust port 2 on the F side, that is, the surface area for cooling the exhaust gas is halved. Temperature drop (cooling) due to heat transfer (heat dissipation) is suppressed. Therefore, the temperature of the exhaust gas is maintained high in terms of both the afterburning effect and the heat transfer cooling suppression.

  Here, looking at the downstream side of the tip 1c of the partition wall 1b shown by the broken line in FIG. 3, the partition wall 1b disappears and both the exhaust ports 2 and 2 gather, and the sectional area of the collective exhaust port is: The cross-sectional area of the F-side exhaust port 2 before assembly and the sum of the cross-sectional areas of the R-side exhaust port 2 are set so as to substantially coincide with each other.

  When the exhaust gas flow exceeds the tip 1c of the partition wall 1b, the high flow velocity flow of the F-side exhaust port 2 is directed toward the center side of the aforementioned collective exhaust port (R-side exhaust port 2 side). It becomes the changed diagonal flow and produces a strong swirl flow, and the turbulent flow component increases.

  As a result, the oxidation reaction of unburned HC and PM (particulate matter) in the exhaust gas (combustion gas) further proceeds (so-called afterburning), and the temperature of the exhaust gas further increases.

  Here, the port direction is abruptly changed by γ at a position slightly downstream of the tip 1c of the partition wall 1b. This also increases the turbulent flow component and enhances the exhaust gas temperature rise effect. it can.

  As described above, in both the region up to the tip 1c of the partition wall 1b where both the exhaust ports 2 and 2 are separated and the region where both the exhaust ports 2 and 2 are gathered beyond the tip 1c of the partition wall 1b, The exhaust gas temperature rises further.

  As a result, even at the position of the downstream catalyst, the exhaust gas temperature can be relatively increased at the stage where the exhaust gas temperature has decreased due to exhaust pipe cooling, thereby activating the catalyst by increasing the catalyst temperature. Thus, the exhaust emission conversion rate (purification rate of harmful components) can be increased at an early stage, thereby suppressing the generation of harmful exhaust emissions such as HC and PM at the start.

  On the other hand, the engine temperature itself is warmed up quickly as described above, so that the heating effect is accelerated and the combustion is improved as the warming up progresses, so exhaust of harmful exhaust emissions from the cylinders. The amount itself decreases, and from this aspect, harmful exhaust emissions (HC, NOx, PM, etc.) after passing through the catalyst can be further reduced.

  Next, when the engine is warmed up, it becomes less necessary to hold a large amount of high temperature residual gas in the cylinder. Therefore, as shown in the upper part of FIG. 14A, the operating angles of the exhaust valves 3a and 3b (opening period) ), And eliminate the negative valve overlap with the intake valve, provide a small positive valve overlap, reduce residual gas and improve combustion stability. Here, since the exhaust single valve is stopped, the effect of reducing the burned gas (combustion gas) taken into the cylinder via the intake / exhaust valve in the positive valve overlap section is also obtained.

  By these, by suppressing the residual gas amount in the cylinder in the idling operation and the low torque operation region that are usually used, it is possible to suppress the rotational fluctuation and improve the sound vibration performance.

  In addition, since the exhaust single valve is stopped and the operating side exhaust valve is slightly small in operating angle, the valve friction is small and the fuel efficiency is improved.

  Further, since the opening timing of the exhaust valves 3a and 3b is slightly advanced from the bottom dead center, the exhaust valves 3a and 3b are opened before the negative pressure of the in-cylinder pressure develops in the expansion stroke in the low torque operation region. The pump loss during the expansion stroke, which tends to be in the low torque operation region, can be suppressed, and the fuel efficiency can be further improved.

  On the other hand, it has the same exhaust gas temperature rise effect as the above-mentioned start by stopping the exhaust single valve, and it can raise the exhaust gas temperature, thereby increasing the conversion rate of the catalyst. Even in this normal range, exhaust emission A reduction effect is obtained.

  Next, let us consider a case where the accelerator is further blown from the area A, and the rotation and torque are accelerated along the arrows on the map of FIG.

  First, when the engine speed or torque becomes slightly larger than the A range, the A / B boundary line is exceeded, and the exhaust single valve state is maintained while only the # 1 cylinder is deactivated (reduced cylinder region, B region). Transition.

  That is, the second switching valve control signal ON signal is sent to the second valve stop mechanism 11a and the third valve stop mechanisms 11d and 11e, and the # 1 cylinder shifts to the cylinder deactivation mode (all four intake / exhaust valves are stopped). To do.

  Of the four intake / exhaust valves, the R-side exhaust valve 3a is originally in a stopped state, and therefore, only three valves (three places) newly shift to the valve stopped state. Specifically, there are only three (three valves), the exhaust valve stop mechanism 11a for the # 1 cylinder and the intake valve stop mechanisms 11d and 11e for the # 1 cylinder.

These three valve stop mechanisms 11a, 11e, and 11d are classified into the stable operation type valve stop mechanism shown in FIG. 5, and when the signal oil pressure becomes high, the three second restriction pins 47 are pushed out by the high oil pressure, -The valve stops and shifts to the state of transition. The volume by which the oil is pushed out when the three second restriction pins 47 move is 3 × A (pin area) × S (moving stroke), and when the hydraulic pressure is P, P × 3 × A × S is the job of oil. The time required to complete the 3PAS work is a response delay.

  Here, considering the normal cylinder deactivation transition in the above-described conventional example, all four valves (four locations) will be decelerated at a time, so the required work increases to 4 PAS, and the cylinder deactivation occurs. The conversion responsiveness will deteriorate. Further, the number of oil passages 48 to the second restricting pin for applying a high pressure is four compared to the number of the present embodiment, and the oil leakage from the oil passage 48 is increased correspondingly. The hydraulic pressure P itself also decreases. From this aspect, the conversion response is further deteriorated.

  However, in this embodiment, since the number of restriction pins operated at high hydraulic pressure is smaller than that in the conventional example, oil work (hydraulic work) required for conversion can be reduced and the hydraulic pressure is reduced less, so that the response to transition to cylinder deactivation is possible. Can be increased. In addition, since the intake valves 71a and 71a and the exhaust valve 3a can be converted simultaneously at the same time, it is possible to suppress the occurrence of a shift in the conversion timing between the intake and exhaust valves.

  In this region B (reduced cylinder operation state), the operating cylinder is maintained in the exhaust valve single valve mode, and the exhaust emission reduction effect similar to that in region A (after warming up) is maintained by stopping the exhaust single valve, Because the exhaust gas is discharged from only one exhaust port 2 of the exhaust valve 3b on the F side of the working cylinder (# 2 cylinder) by the reduced cylinder operation, heat transfer from the exhaust port of the exhaust gas (heat escape) Cooling) is further suppressed, the exhaust gas having a higher temperature is sent to the catalyst position, the catalyst conversion rate is further improved, and the exhaust emission is further reduced.

  On the other hand, since only one cylinder is combusted, the amount of exhaust gas is small, excessive heating of the catalyst is suppressed, and catalyst thermal deterioration is suppressed.

  Next, consider fuel efficiency.

  The engine friction is further reduced by adding an additional stop of the three valves of the # 1 cylinder (further reduction of valve operating friction). Further, due to the reduced-cylinder operation, the cycle high efficiency is improved and the fuel efficiency is further improved by the combustion high load shift in the active # 2 cylinder. That is, even if the engine torque is the same, the in-cylinder surface area in contact with the combustion gas is almost halved by the reduced cylinder operation, so that the cooling loss and the like are also reduced.

  Here, the region B is a practical region that is frequently used, such as traveling on a highway at a constant speed, and also here, a large exhaust emission reduction effect and a fuel consumption reduction (fuel consumption improvement) effect can be obtained.

  As described above, in the B region of the reduced-cylinder operation, the exhaust emission is further reduced and the fuel consumption is further improved. In particular, in this embodiment, the reduced exhaust emission / low fuel consumption reduced cylinder region is extended to the high torque side. It can be expanded and the fuel efficiency in the vehicle total can be improved. This is because, as shown in the intake / exhaust valve operation of FIG. 14B, the lift amount of the exhaust valve 3b is expanded to L3.5 and the operating angle is expanded to D3.5 (intermediate between the L3 lift curve and the L4 lift curve), and the exhaust VTC. As a result, the opening timing of the exhaust valve is almost unchanged and the closing timing of the exhaust valve is delayed. As a result, a moderate valve overlap is formed, a good intake / exhaust operation is performed, and the intake valve closing timing is advanced to near the bottom dead center, ensuring a low / medium-speed intake charging efficiency and high torque It is because it can generate | occur | produce.

  If the engine rotation / torque increases further, the required engine torque cannot be produced in the reduced cylinder operation, and therefore, when the BC boundary line in FIG. 13 is exceeded, the entire cylinder operation (region C) is converted.

  That is, the second switching valve control signal to the second and third valve stop mechanisms 11a, 11d, and 11e is turned off, the second restriction pin 47 is moved to the right by the pin spring, and the lassia adjuster can perform the lost motion operation. It is switched again from the state to the state fixed to the cylinder head 1. As a result, as shown in “Cylinder operation” in the region C in FIG. 14, the # 1 cylinder also starts to operate again. On the other hand, the exhaust single valve state can be maintained in both cylinders, and the emission reduction effect and low fuel consumption effect (low friction caused by exhaust single valve stop) can be reduced and reduced in all cylinders from A to C. It is obtained regardless of cylinder operation.

  Here, when the BC boundary line is exceeded, the operation is switched to the all-cylinder operation. Therefore, if the intake / exhaust valve characteristics shown in FIG. 14B are maintained, the engine torque increases rapidly.

  Therefore, the closing timing of the intake valves 71a and 71b is sufficiently delayed by the retard control by the intake VTC. Then, the exhaust VEL is converted into the maximum lift L4 and the maximum operating angle D4, and is slightly retarded by the exhaust VTC. This makes it possible to sufficiently delay the intake valve closing timing while suppressing changes in the opening timing and valve overlap of the exhaust valve, suppressing the intake charging efficiency, suppressing pump loss, and torque for all cylinder operation. The fuel consumption in all-cylinder operation is improved while suppressing the rapid increase.

Furthermore, as the accelerator pedal is depressed and accelerated, the torque demand increases, so the exhaust valve operating angle is directed to D3.5 (lift L3.5) while suppressing changes in the exhaust valve opening timing and valve overlap. The engine torque is reduced and the closing timing of the intake valve is advanced to the bottom dead center side to increase the engine torque.
However, even when all cylinders are operated, the engine torque cannot be sufficiently increased. This is because stopping the exhaust single valve has contributed to reducing exhaust emissions due to the exhaust gas flow effect, but the exhaust resistance increases due to this exhaust gas flow in a high load region (exhaust gas large region) close to the full load. . Therefore, the exhaust single valve stop is stopped and the operation is switched to the dual exhaust valve operation. That is, when the first switching valve control signal ON signal of the first valve stop mechanisms 11b and 11c, which are valve stop stable type valve stop mechanisms, is sent, it is the cylinder head that the hydraulic lash adjusters 10b and 10d are operating in the lost motion. Since it is fixed to 1, it becomes a valve operation mode. The conversion responsiveness here is good because it is conversion in only two places.

  In addition, as the accelerator pedal is depressurized, the intake valve closing timing is retarded again so that the maximum charging efficiency corresponding to the increase in the engine speed can be obtained. Torque can be improved.

  When retarding the closing timing of the intake valve, if the operating angle of the exhaust valve is increased while maintaining the valve overlap, the exhaust valve is opened earlier to increase the exhaust efficiency and further increase the torque. You can also.

Note that in the high load D region, which is close to the full load, the amount of exhaust gas is large, so conversely, if the exhaust single valve is stopped, the temperature of the exhaust gas flowing into the catalyst will rise excessively and the catalyst may be thermally deteriorated. There is sex. However, in this embodiment, the D region is operated by both exhaust valves of all the cylinders, the gas flow of the exhaust gas is suppressed, the afterburning is suppressed, and the exhaust gas in both the cylinders 2 , 2, the cooling effect due to heat transfer is maximized, so that an excessive increase in the exhaust gas temperature can be suppressed, and the catalyst thermal deterioration which is a concern at high loads close to the full load can be prevented.
[Second Embodiment]
15 to 19 show a second embodiment. As shown in FIG. 15, unlike the first embodiment, an exhaust VEL is not attached to the valve operating system on the exhaust valve side.

  That is, FIG. 15 shows a valve system on the exhaust side, and on the outer peripheral surface of the exhaust camshaft 80, a normal rotating cam 80a that opens the four exhaust valves 3a to 3b via the spring force of the valve spring 12 is shown. Are provided for each cylinder. Other configurations such as an exhaust VTC provided at one end of the exhaust camshaft 80 are the same as in the first embodiment.

  FIG. 16 shows the valve system on the intake side, and the third valve stop mechanism 11d of the valve operation stable type is different from the first embodiment in that it is provided only on the R-side intake valve 71a of the # 1 cylinder. The intake valve 71a on the # 1 cylinder F side is provided with a valve stop stable fourth valve stop mechanism 11g. The same fourth valve stop mechanism 11h is also provided on the intake valve 71b on the F side of the # 2 cylinder.

  Therefore, as shown in FIGS. 18 and 19, the exhaust valve 3a, 3b is in a single valve stop over all the cylinders (two cylinders), and the intake valves 71a, 71b are as shown in FIGS. Is also one valve stop.

  Although the system diagram is shown in FIG. 17, the oil passage 44 downstream of the first electromagnetic switching valve 55 is branched into four, and the two first valve stop mechanisms 11b and 11c and the two fourth valve stop mechanisms 11g and 11h. Communicating with Therefore, the first electromagnetic switching valve 55 switches the four valve stop stable type valve stop mechanisms.

  Then, the oil passage 48 downstream of the second electromagnetic switching valve 65 is branched into two and communicates with one second valve stop mechanism 11a and one third valve stop mechanism 11d. Therefore, the second electromagnetic switching valve 65 switches between the two valve operation stable type valve stop mechanisms.

  Next, the startability effect of this embodiment will be described.

  When the engine is stopped, the oil pressure of the oil pumps 54 and 64 is not activated, so that the signal oil pressure is inactive or low regardless of the on / off positions of the first and second electromagnetic switching valves 55 and 65. The first and fourth valve stop mechanisms 11b, 11c, 11g, and 11h are in a valve stop mode, while the second and third valve stop mechanisms 11a and 11d are in a valve operation mode.

  Accordingly, as shown in “Cylinder operation” in the operation region A of FIG. 19, the exhaust valves 3a and 3b on one side (R side) of both the # 1 and # 2 cylinders are in the valve stop mode, and the # 1 and # 2 both cylinders are on. The intake valves 71a and 71b on one side (F side) are in a valve stop mode, and the other intake and exhaust valves 3a, 3b, 71a and 71b are in a valve operation mode.

  Even when the cranking for the start is started and the start combustion is started, the oil pressure of the oil pumps 54 and 64 does not suddenly rise, and the above-described mode is maintained.

  Further, both the first electromagnetic switching valve control signal and the second electromagnetic switching valve control signal are controlled to be off, that is, each signal oil pressure communicates with the drain passages 45 and 66 instead of the pump oil pressure, and each signal oil pressure can only act at a low pressure. Thus, even when the oil pressure of the oil pumps 54 and 64 rises early, the above-described mode can be maintained.

  Here, at the time of start-up, the engine friction is increased due to the low engine temperature, and at the time of start-up, the engine temperature is low and combustion is poor, and the exhaust emission is also large.

  On the other hand, in the present embodiment, the exhaust valves of both cylinders before the start cranking are in a single valve stop (single valve operation) mode, and the intake valves of both cylinders are also single valve stopped. That is, since half of the intake and exhaust valves are in the valve stop state, the friction is further reduced compared to the first embodiment, and the effect of reducing the friction is obtained from the initial stage of cranking. As a result, combustion is improved and stabilized, and the starting performance can be further improved from that of the first embodiment. In addition, the engine warm-up performance is improved and the warm-up time is shortened by these combustion improvement and stabilization effects.

Here, since the exhaust gas is discharged from one exhaust port by stopping the exhaust single valve, the exhaust gas temperature is prevented from lowering (heat transfer suppression), and the gas flow in the exhaust port 2 is enhanced. Therefore, the reaction of the unburned components in the exhaust gas is promoted, the reaction heat is added, the catalyst is activated early, and the exhaust emission at the start can be reduced as in the first embodiment. . Furthermore, fuel atomization is promoted by the intake swirl by stopping the intake single valve, and PM generated when minute fuel droplets are burned can be further reduced.
As described above, before cranking, the first and fourth valve stop mechanisms are mechanically stabilized in advance in the valve stop mode, and the second and third valve stop mechanisms are mechanically stabilized in advance in the valve operation mode. Yes. That is, since all cylinders (two cylinders) are mechanically preliminarily set in the exhaust single valve stop mode and the intake single valve stop mode, the above-mentioned valve friction reduction effect and exhaust emission reduction effect from the initial stage of cranking and the initial stage of startup, etc. Is surely obtained without delay. Furthermore, even if each electromagnetic switching valve that controls each valve stop mechanism has an electrical failure such as disconnection, the above-described all-cylinder exhaust single valve stop mode and all-cylinder intake single valve stop mode are mechanically Therefore, the startability can be ensured, that is, it has a so-called mechanical fail-safe effect.

Furthermore, in addition to these effects, in the present embodiment, in particular, the valve overlap can be doubled by controlling the advance angle of the intake valve by the intake VTC and the retard control of the exhaust valve by the exhaust VTC. A large amount of combustion gas (exhaust gas, EGR) can be introduced, and warming up of the engine can be promoted. Further, unburned HC and PM generated by combustion are discharged into the intake system during the large valve overlap section, re-intaked into the cylinder during the intake stroke, and re-combusted, thereby reducing these exhaust emissions.
By the way, if a large amount of inert exhaust gas with a large overlap is introduced into the cylinder, the combustion tends to deteriorate, but the combustion is improved by the intake swirl effect due to the stop of the intake single valve as in this embodiment. It is.

  Next, considering the case where the combustion gas is discharged from the exhaust port 2, the same effect as in the first embodiment can be obtained. Supplementally, a partition wall 1b indicated by a broken line (dashed line) in FIG. 15 exists between the F-side exhaust port 2 and the R-side exhaust port 2 provided independently.

  In both cylinders, the exhaust valves 3a and 3b on the R side are stopped, so that the combustion gas flows out from the exhaust ports 2 and 2 of the exhaust valves 3a and 3b on the F side on the valve operation side at a high flow rate (high gas flow). To do. Due to this high gas flow, oxidation reactions such as unburned HC and PM (particulate matter) in the exhaust gas (combustion gas) proceed (so-called afterburning), and the temperature of the exhaust gas increases. Further, immediately after the high temperature exhaust gas is discharged from the cylinder, the exhaust gas contacts only with the F side exhaust 2 and 2 ports, that is, the surface area for cooling the exhaust gas is halved. Temperature drop (cooling) due to heat transfer (heat dissipation) of gas is suppressed.

  Therefore, the temperature of the exhaust gas is maintained high in terms of both the afterburning effect and the heat transfer cooling suppression.

  Here, looking at the downstream side from the tip of the partition wall 1b shown by the broken line in FIG. 15, the partition wall 1b disappears at the front end portion 1c in the middle as in the first embodiment, and the collective exhaust in which both exhaust ports gather. It becomes a port, and its cross-sectional area is set to substantially coincide with the sum of the cross-sectional area of the F-side exhaust port 2 and the cross-sectional area of the R-side exhaust port 2. When the exhaust gas flow exceeds the tip 1c of the partition wall 1b, the high flow velocity flow of the F side exhaust port 2 is directed toward the central side of the collective exhaust port (R side exhaust port 2 side). It becomes the changed diagonal flow and produces a strong swirl flow, and the turbulent flow component increases. As a result, the oxidation reaction of unburned HC and PM (particulate matter) in the exhaust gas (combustion gas) further proceeds (so-called afterburning), and the exhaust gas temperature rise (suppression of temperature decrease) further proceeds. .

  As described above, in both the region up to the tip 1c of the partition wall 1b where both the exhaust ports 2 and 2 are separated and the collective exhaust port region where the both exhaust ports 2 and 2 are gathered, The relative temperature rise further proceeds.

  As a result, even at the position of the downstream catalyst, the exhaust gas temperature can be maintained high, the exhaust emission conversion rate of the catalyst can be maintained high, and the generation of harmful emissions such as HC and PM at the start can be suppressed.

  In particular, in the present embodiment, unburned HC and unburned PM at the time when the exhaust is discharged from the cylinder to the exhaust ports 2 and 2 by the improvement of combustion by the intake swirl by stopping the intake single valve or the improvement of fuel atomization. Since these are reduced, these harmful components are further reduced when viewed in the final exhaust gas that passes through the downstream catalyst and is released to the atmosphere.

  On the other hand, as described above, the engine temperature itself is warmed up quickly, so that the efficiency of heating is improved, and combustion is also improved with the progress of warming up, so the emission amount of harmful exhaust emissions from the cylinder itself is reduced. From this aspect, it is the same as in the first embodiment that harmful exhaust emissions (HC, NOx, PM, etc.) after passing through the catalyst can be reduced.

  Next, when the warming-up proceeds, as shown in the upper part of region A in FIG. 19, the exhaust valve is controlled to the advance side by the exhaust VTC, and the intake valve is controlled to the retard side by the intake VTC. By this double control, the valve overlap is reduced, and the inert residual gas can be reduced to improve the combustion stability. Here, since both the intake valve and the exhaust valve are single valve stop, even in the same valve overlap section, the residual gas introduced through the intake / exhaust valve can be sufficiently reduced, and the combustion stability effect is great. .

  Thereby, rotation fluctuation can be suppressed and sound vibration performance can be improved in a commonly used idling operation or low torque operation region.

  In particular, since the closing timing of the intake valve is delayed, the compression and the effective compression ratio are reduced. From this aspect as well, the rotational fluctuation suppressing effect and the sound vibration improving effect are great.

  In addition to the exhaust valve, the intake valve is also stopped one-sided, so the valve friction is further reduced and fuel efficiency is improved. In addition, since the opening timing of the exhaust valve is controlled in a direction to advance from the bottom dead center, the exhaust valve can be opened before the in-cylinder pressure develops to a negative pressure during the expansion stroke, which tends to be in a low torque region. It is possible to suppress the pump loss in the expansion stroke and further improve the fuel consumption. On the other hand, it has the above-mentioned effect of increasing the exhaust gas temperature by stopping the exhaust single valve, so that the conversion rate of the catalyst can be increased, and the effect of reducing exhaust emission can also be obtained in this normal range.

  Next, let us consider a case where the accelerator pedal is further depressed from the A region, and the engine rotation and torque are accelerated along the arrows on the map of FIG.

First, when the rotational speed or torque becomes slightly larger than the A region, only the # 1 cylinder shifts to cylinder deactivation while maintaining the intake / exhaust single valve state of the # 2 cylinder exceeding the A / B boundary line. (Reduced cylinder operation; B area)
That is, the second electromagnetic switching valve control signal ON signal is output to the second valve stop mechanism 11a and the third valve stop mechanism 11d, and the # 1 cylinder shifts to the cylinder deactivation mode (all four intake / exhaust valves are stopped). To do. Of the four intake / exhaust valves, the R-side exhaust valve 3b and the F-side intake valve 3a are originally in a stopped state, so that only two valves (two locations) are newly shifted to the valve stop state. . Specifically, there are only two, an exhaust valve stop mechanism 11a for the # 1 cylinder and an intake valve stop mechanism 11d for the # 1 cylinder.

  In the present embodiment, the total number of the second restriction pins 47 to be operated at high hydraulic pressure is two, which is further smaller than the three in the first embodiment, so that oil work (hydraulic work) required for conversion can be further reduced. Moreover, since the decrease in hydraulic pressure is further reduced, the response to transition to cylinder deactivation can be further enhanced. As a result, smoother drivability at the initial stage of acceleration can be realized, and the fuel consumption can be further reduced by quick cylinder deactivation. Further, since the intake valve and the exhaust valve can be converted at the same time, there is no shift in the conversion timing between the intake and exhaust valves.

  In this B region (reduced cylinder operation state), the operating cylinder is maintained in the intake / exhaust valve single valve mode, and the exhaust emission reduction effect similar to that in the A region is maintained by stopping the exhaust single valve. As a result, exhaust gas is discharged only from the exhaust port 2 on the F side of the operating cylinder (# 2 cylinder), so that heat escape from the exhaust port 2 is further suppressed, and higher temperature exhaust gas is sent to the catalyst. As a result, the catalyst conversion rate is further improved, and the exhaust emission is further reduced.

  On the other hand, since only one cylinder is burned, the amount of exhaust gas is small, the catalyst is not excessively heated, and catalyst deterioration is suppressed.

  Further, due to the intake swirl effect (combustion improvement) by stopping the intake single valve, HC and PM in the exhaust gas flowing out from the cylinder to the exhaust port 2 are originally reduced. HC, PM, NOx, etc. are further reduced by the high conversion efficiency resulting from the exhaust temperature.

  Here, considering the fuel efficiency, the cycle efficiency is improved and the fuel efficiency is further improved by the reduced cylinder operation and the combustion high load shift in the operating # 2 cylinder as in the first embodiment. That is, the cylinder surface area in contact with the combustion gas is almost halved by the cylinder reduction operation, so that the cooling loss and the like are also reduced.

  On the other hand, for engine friction, in addition to cylinder # 1 cylinder deactivation (all intake / exhaust valve stop) and operation # 2 cylinder exhaust single valve stop, in this embodiment, operation # 2 cylinder intake single valve stop is also added. Therefore, the fuel consumption is further reduced compared to the first embodiment, and the fuel efficiency is improved accordingly.

  Further, the fuel consumption is further improved by the progress of the combustion improvement by the intake swirl due to the stop of the intake single valve.

  As described above, the fuel efficiency is further improved in the B region of the reduced-cylinder operation, but in this embodiment, in particular, the reduced-cylinder region with good fuel efficiency can be expanded to the high torque side, and the fuel efficiency at the vehicle total is increased. Can be improved.

  This is because, as shown in the intake / exhaust valve operation of FIG. 19B, the exhaust valve advances the intake valve phase while maintaining the advance angle phase. As a result, the opening timing of the exhaust valve is almost unchanged (advanced position) and a moderate valve overlap is formed, so that a good intake / exhaust operation is performed, and the closing timing of the intake valve advances to near the bottom dead center. This is because the intake and charging efficiency of low and medium rotation is ensured and high torque can be generated. Here, since the opening timing of the exhaust valve is held at the advanced position, exhaust push-out loss due to the high combustion load shift of the operating cylinder is reduced, torque is further generated, and fuel efficiency is improved.

  If the rotation and torque are further increased, the required engine torque cannot be produced in the reduced cylinder operation. Therefore, when the BC boundary line is exceeded, the entire cylinder operation (region C) is converted. That is, the second electromagnetic switching valve control signal to the second and third valve stop mechanisms 11a and 11d is turned off, the regulating pin 47 is moved to the right by the spring force of the return spring 49, and the lassia adjuster is activated. It is switched from a state where it can be fixed to a state where it is fixed to the cylinder head 1. As a result, as shown in “Cylinder operation” in region C in FIG. 19, the # 1 cylinder also starts to operate again. On the other hand, the intake / exhaust single valve state can be maintained in both cylinders, and the emission reduction effect and fuel consumption effect due to the stop of the intake / exhaust single valve can be obtained in the range from A to C regardless of all cylinder operation or reduced cylinder operation. It is.

  Here, when the BC boundary line is exceeded, the operation is switched to the all-cylinder operation. Therefore, if the intake / exhaust valve characteristic shown in FIG. 19B is maintained, the engine torque increases rapidly.

  Accordingly, the closing timing of the intake valve is sufficiently delayed by performing the retard control by the intake VTC. Then, the retardation is controlled by the same amount by the exhaust VTC.

  As a result, the intake valve closing timing can be sufficiently delayed while suppressing the change in valve overlap, the intake charging efficiency is suppressed, the pump loss is also suppressed, and the sudden increase in torque in all cylinder operation is also suppressed. This improves fuel efficiency in all-cylinder operation.

  In addition, the exhaust push-out loss is reduced corresponding to the reduction of the combustion load of the operating cylinders as all cylinders are operated, so that the opening timing of the exhaust valve is shifted toward the bottom dead center and the expansion work is shifted accordingly. When the accelerator pedal is depressed and accelerated, the torque demand increases.Therefore, the intake valve closing timing is controlled to advance toward the bottom dead center while suppressing changes in valve overlap. As a result, the opening angle of the exhaust valve is also advanced and the engine torque is increased.

  However, even when all cylinders are operated, the engine torque cannot be sufficiently increased. This is because although the exhaust single valve stop has contributed to emission reduction due to the exhaust gas flow effect, the exhaust resistance increases due to this exhaust gas flow in a high load region (exhaust gas large region).

Although the intake single valve stop also has a combustion improvement effect (fuel consumption effect) by the intake swirl, the intake resistance increases. Therefore, the intake / exhaust single valve is stopped and the operation is switched to the intake / exhaust double valve operation.
That is, when the first electromagnetic switching valve control signal ON signal is sent to the exhaust side first valve stop mechanisms 11b and 11c and the intake side fourth valve stop mechanisms 11g and 11h which are valve stop stable type valve stop mechanisms, Since the rassia adjusters 10b, 10d, 75a, and 75c are operating in the lost motion are fixed to the cylinder head 1, the valve operating mode is achieved.
Here, since the operation of the four first restriction pins 42 is performed in total, the responsiveness may be deteriorated, but since the high engine torque range has already been reached, there is almost no actual damage on the driving feeling.

  In addition, when the accelerator pedal is depressed, the closing timing of the intake valve is retarded as the rotation rises, so that the maximum charging efficiency can be obtained according to the rotation speed, so that the torque in the entire rotation range can be obtained. It can be improved.

  When retarding the closing timing of the intake valve, if the lift phase of the exhaust valve is maintained, the valve overlap gradually decreases, and the exhaust gas increases due to the exhaust pressure that increases as the engine rotates at a higher speed. Can be prevented from flowing back into the cylinder, and a reduction in torque and output can be suppressed.

Further, in this high load region D, since the amount of exhaust gas is large, when the exhaust single valve is stopped, the temperature of the exhaust gas flowing into the catalyst becomes excessive, and the catalyst may be thermally deteriorated. However, in this embodiment as well, as in the first embodiment, the D region is operated as a double exhaust valve, the exhaust gas flow is suppressed and the afterburning is suppressed, and the exhaust gas is in both exhaust ports. Since it passes through 2 and 2, a cooling effect due to heat transfer can be obtained, so that the exhaust gas temperature can be suppressed, and catalyst thermal deterioration which is a concern at high loads can be prevented.
[Industrial applicability]
In each of the embodiments described above, the switching energy is energy by hydraulic pressure, but according to this, a plurality of valve stop mechanisms can be converted at a time by providing a plurality of hydraulic paths downstream of the electromagnetic switching valves 55 and 65. Has the advantage.

  On the other hand, it is also conceivable to directly or indirectly control the movement of each regulating pin of each valve stop mechanism directly or indirectly by, for example, an on / off operation of an electromagnetic solenoid valve, using energy by electromagnetic force.

  In this case, as for the difference between the stable stop type and the stable operation type, whether the lashia adjuster is in the lost motion state or the cylinder head fixed state when the return spring presses the restriction pin, as in the case of hydraulic pressure. Therefore, this electromagnetic type can be easily realized. For example, when electromagnetic force is used as conversion energy, it is difficult to obtain hydraulic energy from the hydraulic pump. Smooth conversion operation even during extremely low engine rotation (including engine stoppage) or extremely cold machine. Is obtained.

In addition, these valve stop lost motion mechanisms are provided on the rascia adjuster which is the fulcrum of the swing arm, and the lassia adjuster is slid directly on the cylinder head. However, an iron-based material is used for the sliding portion. An existing color may be interposed. Then, even if the cylinder head is made of a material such as aluminum or magnesium, the wear resistance can be improved. In addition, a valve stop (lost motion) mechanism can be provided in addition to the lassia adjuster that serves as the fulcrum of the swing arm. it can. For example, the lost motion mechanism may be provided in the swing arm itself.

  In this case, for example, a roller element that can be displaced (lost motion) is provided on the main swing arm as shown in the special table 2009-503345, and the roller element and the main swing arm may be switched between fastening and non-fastening. That is, the present invention can be applied to various configurations without departing from the gist of the present invention.

  For example, the present invention can also be applied to a lifter type valve gear that does not have a hydraulic lash adjuster as disclosed in JP 2010-270633 A. In this case, a valve stop mechanism built in the lifter as shown in JP-A-63-16112 may be used.

  In each embodiment, an example of in-line 2-cylinders has been described. However, the number of components can be doubled and expanded to in-line 4-cylinders. May be. Furthermore, it is possible to divide the constituent elements of the inline 6 cylinders into two banks and develop them into V type 6 cylinders, which are not particularly limited to inline 2 cylinders and can be widely applied.

  That is, as described above, the specific configuration is not limited and can be applied to various configurations without departing from the gist of the present invention.

The technical ideas of the invention other than the claims ascertained from the embodiment will be described below.
(A) A variable valve operating apparatus for a multi-cylinder internal combustion engine according to claim 1,
A first operation region including a cranking time and a starting time in which the first valve stop mechanism is in a valve stop state and the second valve stop mechanism is in a valve operation state;
A second operation region in which the rotational speed or the engine torque is higher than that in the first operation region, and both the first valve stop mechanism and the second valve stop mechanism are in a valve stop state;
A variable valve operating apparatus for a multi-cylinder internal combustion engine.

According to the present invention, in the first operation region including the cranking time and the start time, the startability can be improved and the exhaust emission reduction effect can be obtained by stopping the exhaust valve on one side. In addition, since both exhaust valves are stopped for some cylinders, not only fuel consumption can be reduced, but the number of active combustion cylinders (cylinders other than some cylinders) decreases, so the number of exhaust ports through which exhaust gas passes is also reduced. It decreases and the cooling of the catalyst by exhaust gas is suppressed. In addition, by reducing the number of exhaust ports through which exhaust gas passes, the gas flow in the exhaust ports is further strengthened, and the exhaust gas temperature rises due to promotion of afterburning and the catalyst conversion rate (hazardous substance purification rate) is improved. Therefore, exhaust emissions in the practical range can be reduced.
(B) A variable valve operating apparatus for a multi-cylinder internal combustion engine according to claim 1,
First energy changing means for supplying or stopping switching energy to the first valve stop mechanism;
Second energy changing means for supplying or stopping the switching energy to the second valve stop mechanism,
In a predetermined operation state in which one exhaust valve of each cylinder is stopped by the first valve stop mechanism and the first energy changing unit,
A variable valve operating apparatus for a multi-cylinder internal combustion engine, wherein the operation of the other exhaust valve of the partial cylinder is stopped by the second valve stop mechanism and the second energy changing means.

According to this invention, when both the exhaust valves are stopped, one of the two exhaust valves is already in the valve stopped state, so that the other exhaust valve is stopped by the second energy changing means. Since only the conversion is performed, the response of conversion of both exhaust valves to the valve stop state is improved.
[Claim c] The variable valve operating apparatus for a multi-cylinder internal combustion engine according to claim 1,
A third valve stop mechanism that switches between a valve stop state and a valve operation state of the pair of intake valves in the partial cylinder;
The third valve stop mechanism is configured to be in a valve operating state when the switching energy does not act,
The variable valve operating apparatus for a multi-cylinder internal combustion engine, wherein the second valve stop mechanism and the third valve stop mechanism are simultaneously stopped by the same second switching energy changing means.

According to the present invention, when some of the cylinders are changed to the cylinder deactivation state, the total number of valves per cylinder that shifts to the valve stop is three, which is smaller than the conventional four. Will improve. Further, since the changing operation is performed by the same switching energy changing means, it is possible to suppress the occurrence of time lag in the conversion on the intake valve side and the exhaust valve side, and the device can be simplified.
(D) A variable valve operating apparatus for a multi-cylinder internal combustion engine according to claim 1,
A third valve stop mechanism for switching the other valve operating state and the valve stop state of the pair of intake valves in the partial cylinder;
A fourth valve stop mechanism for switching between one valve operating state and the valve stop state of the pair of intake valves over all cylinders;
The fourth valve stop mechanism is configured to be in a valve stop state when the switching energy does not act, and the third valve stop mechanism is in a valve operating state when the switching energy does not act. A variable valve operating apparatus for a multi-cylinder internal combustion engine, characterized in that:

According to the present invention, the intake and exhaust valves of all the cylinders can be in a single valve stop state from the initial stage of cranking and the start of starting, so that in addition to various effects by stopping the exhaust single valve, further friction reduction effect, combustion improvement by intake swirl The effect can be obtained over all cylinders, and better startability can be realized. Furthermore, since fuel atomization by the intake swirl due to the single valve stop of the intake valve is promoted, an exhaust emission reduction effect (particularly, an exhaust particulate matter PM reduction effect) can be obtained. PM is likely to be generated when the fine droplets are burned, and the atomization of the fine droplets is promoted by the intake swirl effect, so that the generation of PM is suppressed.
[Claim e] In the variable valve operating apparatus for a multi-cylinder internal combustion engine according to claim d,
The variable valve operating system for a multi-cylinder internal combustion engine, wherein the first valve stop mechanism and the fourth valve stop mechanism are simultaneously converted into a valve stop state or a valve operation state by the same first switching energy changing means. .

According to the present invention, it is possible to suppress the occurrence of a time lag in the change between the intake side and the exhaust side, and the device is simplified.
[Claim f] In the variable valve operating apparatus for a multi-cylinder internal combustion engine according to any one of claims 1 to e,
A variable valve operating apparatus for a multi-cylinder internal combustion engine, wherein the switching energy is hydraulic energy.

According to the present invention, by providing a plurality of hydraulic paths branched downstream of the switching energy supply changing means, the stop mechanism of the plurality of valves can be converted at a time, so that the structure is not complicated, and between the valve stop mechanisms. It is difficult for the timing deviation to occur.
[Claim g] In the variable valve operating apparatus for a multi-cylinder internal combustion engine according to any one of claims 1 to f,
A variable valve operating apparatus for a multi-cylinder internal combustion engine having a variable valve operating apparatus capable of changing a phase or lift amount in lift characteristics of an intake valve and an exhaust valve.

  In the present invention, the valve timing control device VTC and the variable lift mechanism VEL are used, and the effects of improving various performances by combining valve stop or valve operation can be further enhanced. For example, at the time of starting, the engine body can be stopped by stopping one exhaust valve and reducing the operating angle of the other exhaust valve, thereby confining high-temperature combustion gas in the cylinder in addition to reducing friction. Warm-up time is promoted and the warm-up time can be reduced. As a result, the total exhaust emission amount can be reduced. Or, for example, in the reduced-cylinder operation, the intake valve phase is advanced and the intake valve closing timing is advanced to near the bottom dead center, thereby improving the charging efficiency and reducing the fuel consumption and reducing the emission. It can be expanded to the high engine torque side.

DESCRIPTION OF SYMBOLS 1 ... Cylinder head 1a ... Holding hole 3a, 3a ... 1st, 2nd exhaust valve 3b, 3b ... # 1 cylinder side 1st, 2nd exhaust valve 5 # drive shaft 5a ... Drive cam 6 ... Swing arm 6a ... one end 6b ... other end 7 ... rocking cam 8 ... transmission mechanism 9 ... control mechanism 10a, 10b ... # 1 cylinder exhaust side first and second hydraulic lash adjusters (fulcrum members)
10c, 10d... # 2 cylinder exhaust side third and fourth hydraulic lash adjusters (fulcrum members)
11a ... second valve stop mechanism 11b, 11c ... first valve stop mechanism 11d, 11e ... third valve stop mechanism 11g, 11h ... fourth valve stop mechanism 12 ... valve spring of exhaust valve 13 ... bearing portion 14 ... roller 24 ... Body 27 ... Plunger 27b ... Tip head 34 ... Sliding hole 35 ... Lost motion spring (biasing member)
36 ... Restriction mechanism 38 ... Movement hole 39 ... Restriction hole 40 ... Retainer 41 ... Sliding pin 42 ... First restriction pin 43 ... Return spring 44 ... Oil passage hole 45, 66 ... Drain passage 47 ... Second restriction pin 54 , 64 ... Oil pump 55 ... First electromagnetic switching valve 65 ... Second electromagnetic switching valve 71a, 71a ... First and second intake valves 71b, 71b on the # 1 cylinder side First and second intake valves on the # 2 cylinder side Valve 72 ... Valve spring of intake valve 73 ... Intake side camshaft 73a ... Rotating cam 74 ... Intake side swing arm 75a, 75b ... # 1 cylinder intake side first and second hydraulic lash adjusters (fulcrum members)
75c, 75d ... # 2 cylinder intake side third and fourth hydraulic lash adjusters (fulcrum members)

Claims (3)

In a multi-cylinder internal combustion engine,
A pair of intake valves and a pair of exhaust valves respectively provided for each cylinder;
A first valve stop mechanism that switches between a valve operating state that opens and closes one of the pair of exhaust valves in the cylinders of the plurality of cylinders and a valve stop state that stops the opening and closing operation;
A second valve stop mechanism that switches between a valve operating state that opens and closes the other exhaust valve of the pair of exhaust valves in the some cylinders and a valve stop state that stops the opening and closing operation;
The first valve stop mechanism is mechanically configured to be in a valve operating state when switching energy is supplied, and to be in a valve stop state when supply of switching energy is stopped,
The multi-cylinder internal combustion engine characterized in that the second valve stop mechanism is mechanically configured so that the valve is stopped when switching energy is supplied and the valve is operated when supply of switching energy is stopped. Variable valve gear for engine. In an internal combustion engine having a plurality of cylinders,
A pair of intake valves and a pair of exhaust valves respectively provided for each cylinder;
A swing arm that swings using a lash adjuster as a swing fulcrum and opens and closes the intake valve and the exhaust valve;
A first valve that is actuated by hydraulic pressure and is in a valve stop state by causing a lost motion of a swing amount of the swing arm corresponding to one of the pair of exhaust valves in the cylinders of the plurality of cylinders. A stop mechanism;
A second valve stop mechanism that is actuated by hydraulic pressure and that makes the valve stop state by causing a lost motion of a swing amount of the swing arm corresponding to the other exhaust valve of the pair of exhaust valves in the some cylinders;
With
The first valve stop mechanism is configured to be in a valve operating state when hydraulic pressure is supplied, and to be in a valve stopped state when supply of hydraulic pressure is restricted,
The variable valve for a multi-cylinder internal combustion engine, wherein the second valve stop mechanism is configured to be in a valve stop state when hydraulic pressure is supplied and to be in a valve operation state when supply of the hydraulic pressure is limited. apparatus. A pair of intake valves and a pair of exhaust valves respectively provided for each cylinder of the multi-cylinder internal combustion engine;
A first valve stop mechanism that switches between a valve operating state that opens and closes one exhaust valve of the pair of exhaust valves in a part of the plurality of cylinders and a valve stop state that stops the opening and closing operation;
A variable valve apparatus comprising: a second valve stop mechanism that switches between a valve operating state that opens and closes the other exhaust valve of the pair of exhaust valves in the partial cylinder and a valve stop state that stops the opening and closing operation A control device of
The first valve stop mechanism is mechanically configured to operate the exhaust valve when switching energy is supplied, and to stop the operation of the exhaust valve when supply of switching energy stops,
The second valve stop mechanism is mechanically configured to stop the operation of the exhaust valve when switching energy is supplied, and to operate the exhaust valve when supply of switching energy is stopped,
A control apparatus for a variable valve operating apparatus, wherein supply of switching energy to the first valve stop mechanism and the second valve stop mechanism is separately controlled.

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