JP6184834B2 - Control device and controller for multi-cylinder internal combustion engine - Google Patents

Description

  The present invention relates to a control device and a controller for a multi-cylinder internal combustion engine that can improve fuel efficiency during actual traveling by combining a cylinder deactivation mechanism and a variable compression ratio mechanism.

  As a conventional control device for a multi-cylinder internal combustion engine, a control device for a variable valve gear described in Patent Document 1 below is known.

  This is a so-called cylinder deactivation (reducing cylinder operation) in which the intake and exhaust valves of half of all cylinders are stopped and only the remaining half of the cylinders are operated (combusted).

  As described above, the cylinder deactivation increases the throttle valve opening and reduces the pump loss, and the load per cylinder increases. Therefore, the high load shift improves the thermal efficiency, and as a result, the driving fuel consumption improves. Can be made.

JP-A-10-82334

  However, in the prior art described in the above publication, since a small number of cylinders are operated (combusted) during cylinder deactivation control, the engine torque may be restricted. In addition, when looking at the operating cylinder, the load (torque per cylinder) at the same engine torque is about twice as high, so that knocking is likely to occur and the engine torque may be further suppressed. There is. Further, if the ignition timing is delayed or the fuel is enriched in order to suppress the knocking, the fuel efficiency is deteriorated, and the improvement of the fuel efficiency, which is originally aimed at stopping the cylinder, is suppressed.

  In other words, it is meaningless to increase the torque slightly until the cylinder deactivation effect (fuel efficiency effect) is lost.

  As described above, with the conventional technology, it has been difficult to increase the torque of the engine by reducing the cylinder operation while obtaining the fuel efficiency effect. As a result, the cylinder deactivation region with good fuel efficiency is limited to the low torque region, and the frequency of traveling with reduced cylinder operation is reduced, so that the fuel consumption performance during actual traveling of the vehicle cannot be sufficiently improved.

  The present invention has been devised in view of the above-described conventional technical problems, and shifts the reduced-cylinder operation region to the higher engine torque side while maintaining the fuel efficiency effect, thereby improving the fuel efficiency performance during actual traveling of the vehicle. The purpose is to further increase.

The present invention has been devised in view of the above-described conventional technical problems, and among a plurality of cylinders, a cylinder deactivation mechanism capable of stopping the operation of an intake valve and an exhaust valve of some cylinders, and a machine of all cylinders A control device for a multi-cylinder internal combustion engine, comprising: a variable compression ratio mechanism capable of changing a compression ratio; and a controller that controls the cylinder deactivation mechanism and the variable compression ratio mechanism ,
The cylinder deactivation mechanism has a first operation region in which the operation of the intake valves and exhaust valves of the partial cylinders is stopped, and a second operation region in which the engine valves of all cylinders are operated. When changing from the first operating region on the side to the second operating region on the engine high torque side, the variable compression ratio mechanism causes the second machine to be relatively high from the relatively low first mechanical compression ratio. When the engine operating state changes from the second operating region to the first operating region, the mechanical compression ratio is reduced by the variable compression mechanism in advance and the cylinder is thereafter The operation of the intake valves and exhaust valves of the partial cylinders is stopped by a stop mechanism .

  According to the present invention, in the first operating region (reduced cylinder operating region), the resistance to knocking can be improved due to a low mechanical compression ratio. Can be expanded to the side. That is, it is possible to increase the driving frequency in the reduced-cylinder region where the fuel efficiency is good, and to improve the fuel efficiency performance during actual traveling of the vehicle.

  In addition, the frequency of operation in the second operation region (all cylinder operation region), which is disadvantageous in fuel consumption, is reduced, while in the second operation region, thermal efficiency can be improved as much as possible by high mechanical compression ratio control. As a result, it is possible to improve the fuel efficiency performance in the total during actual traveling of the vehicle.

1 is an overall schematic diagram showing a first embodiment in which a control device according to the present invention is applied to a two-cylinder internal combustion engine. The variable mechanical compression ratio mechanism provided for this embodiment is shown, A shows a control position where the mechanical compression ratio is maximized, and B shows a control position where it is minimized. It is sectional drawing which shows the valve operating apparatus of # 1 cylinder which is a cylinder which can be stopped. It is a perspective view which shows the valve operating apparatus by the side of intake (exhaust side) ranging from # 1 cylinder to # 2 cylinder of this embodiment. A is a longitudinal sectional view showing first and second valve stop mechanisms provided on the intake side of the present embodiment, B is a longitudinal sectional view showing the operation of the first and second valve stop mechanisms, and C is A of B FIG. 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 a map which shows the relationship between the number of operating cylinders in this embodiment, and a mechanical compression ratio. It is a change characteristic figure of the mechanical compression ratio in this embodiment. FIG. 8 is a mechanical compression ratio at the time of switching to the operation range (1) to (4) shown in FIG. 7, the operation of the # 1 cylinder and the # 2 cylinder, and the valve lift characteristics of the intake and exhaust valves. It is a control flowchart figure on the acceleration side by the control unit of this embodiment. It is a control flowchart figure on the deceleration side. It is the whole control apparatus schematic of a 2nd embodiment. It is a perspective view which shows the valve operating apparatus provided with VEL by the side of the intake valve in this embodiment. It is a map which shows the relationship between the number of operating cylinders in this embodiment, and a mechanical compression ratio. It is a change characteristic figure of the mechanical compression ratio in this embodiment. FIG. 15 is a characteristic diagram of the mechanical compression ratio, the operation of the # 1 cylinder and the # 2 cylinder, the valve lift of the intake and exhaust valves, and the operating angle at the time of switching to the operation range (1) to (4) shown in FIG. It is a control flowchart figure on the acceleration side by the control unit of this embodiment.

Embodiments of a control apparatus for a multi-cylinder internal combustion engine according to the present invention will be described below with reference to the drawings. In this embodiment, the present invention is applied to an in-line two-cylinder internal combustion engine 01 of gasoline specification, and a variable mechanical compression ratio mechanism (VCR) 02 capable of changing the mechanical compression ratio ε of all cylinders is provided. A cylinder deactivation mechanism 03 is provided that can deactivate the cylinder by deactivating the intake valve and the exhaust valve in only one cylinder.
[First Embodiment]
FIG. 1 shows a first embodiment of the present invention. In addition to the VCR 02 and the cylinder deactivation mechanism 03, the internal combustion engine 01 is provided with a supercharger 04, which is a mechanical supercharger, and an intake valve 3 There is provided a phase variable electric VTC05 that can control the opening / closing timing (opening / closing phase).

  The VCR02 shown in FIGS. 1 and 2 is described in Japanese Patent Application Laid-Open No. 2002-276446 filed earlier, and therefore the structure will be briefly described.

The crankshaft 50 has a plurality of journal portions 51 and a crank pin 52, and the journal portion 51 is rotatably supported by the main bearing of the cylinder block 06. The crank pin 52 is a predetermined amount eccentric from journal portion 51, wherein the lower link 53 is connected rotatably.

  The lower link 53 is configured to be split into two left and right members, and the crank pin 52 is rotatably fitted in a substantially central connecting hole.

  The upper link 54 connected to one end of the lower link 53 via a connecting pin 55 is rotatably connected to the piston 57 by a piston pin 56 at the upper end. The piston 57 receives combustion pressure and reciprocates in the cylinder 58 of the cylinder block 06.

  Above the cylinder 58, the intake valve 3 and the exhaust valve 71 supported in an openable and closable manner in the cylinder head 07 are arranged.

  A control link 59 having an upper end connected to the other end of the lower link 53 via a connecting pin 60 is connected to a lower portion of the cylinder block 06 via a control shaft 61 so as to be swingable. . That is, the control shaft 61 is supported by the cylinder block 06 and has an eccentric cam 61a that is eccentric from the rotation center thereof, and a lower end portion of the control link 59 is rotatably connected to the eccentric cam 61a. ing.

  The rotation position of the control shaft 61 is controlled by a compression ratio control actuator 62 using an electric motor based on a control signal from an engine control unit 63 that is a controller.

  Therefore, as shown in FIG. 2A, when the control shaft 61 is rotated in one direction (counterclockwise in FIG. 2) by the compression ratio control actuator 62, the VCR02 has a center position X of the eccentric cam 61a. It is located in the middle left lower part. As a result, the swing support position of the lower end of the control link 59 changes, the stroke position of the piston 57 changes, and the mechanical compression ratio becomes the maximum control position, that is, the piston top dead center position becomes the highest. The position can be changed (ε15 described later).

  On the other hand, when the control shaft 61 is rotated in the other direction (clockwise in FIG. 2), as shown in FIG. 2B, the center position X of the eccentric cam 61a is positioned vertically upward in the drawing. As a result, the swing support position of the lower end of the control link 59 changes, the stroke position of the piston 57 changes, and the mechanical compression ratio becomes the minimum control position, that is, the piston top dead center position becomes the lowest control. The position can be changed (compression ratio ε9 described later).

  Here, the stopper position (not shown) that restricts the maximum counterclockwise rotation of the control shaft 61 is the maximum control position (mechanical compression ratio ε15), and the stopper position that restricts the maximum clockwise rotation is the minimum control position (mechanical compression ratio). It may be set as ε9).

  Here, the mechanical compression ratio ε is a value obtained by dividing the cylinder internal volume at the bottom dead center (BDC) of the piston 57 by the cylinder internal volume at the top dead center (TDC) of the piston 57.

The cylinder deactivation mechanism 03 is configured as shown in FIG. 3 and FIG. 4, and FIG. 3 shows the intake side and exhaust side valve operating devices in the # 1 cylinder (cylinder deactivation possible cylinder). And a valve operating device on the intake side (or exhaust side) of the # 2 cylinder. The front (F) side # 1 cylinder shown in FIG. 4 is a cylinder capable of cylinder deactivation, that is, a cylinder in which all intake valves and exhaust valves can be stopped, and a rear (R) side cylinder. The # 2 cylinder does not deactivate the cylinder, and is a normally operating cylinder in which at least one intake valve and exhaust valve are always operated.
[Valve on the intake side]
The intake side valve operating mechanism of the cylinders # 1 and # 2 will be described in detail. As shown in FIGS. 3 and 4, one cylinder that opens and closes a pair of intake ports 2 and 2 formed in the cylinder head 07. A pair of hitting intake valves are provided. That is, the first and second intake valves 3a and 3a are provided in the # 1 cylinder, and the first and second intake valves 3b and 3b are provided in the # 2 cylinder. Here, in each cylinder, the first intake valves 3a and 3b are arranged on the F side, and the second intake valves 3a and 3b are arranged on the R side, respectively.

  Each of the intake valves 3a to 3b includes a rotational force and a valve of two rotary cams 5a provided per cylinder provided on the intake camshaft 5 supported between the upper end portion of the cylinder head 07 and the bearing bracket 13. The open end of each intake port 2 is opened / closed via each swing arm 6 by the spring force of the spring 12.

  Further, held by the cylinder head 07, the gap between each swing arm 6 and each intake valve 3a-3b and the gap between each cam surface of each rotary cam 5a are adjusted to zero lash. Four first to fourth hydraulic lash adjusters 10a, 10b, 10c, and 10d that are fulcrum members (pivots) are disposed.

  That is, the first and second hydraulic lash adjusters 10a and 10b are arranged on the intake valve side of the # 1 cylinder, and the third and fourth hydraulic lash adjusters 10c and 10d are arranged on the intake 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.

  Further, the first and second intake valves 3a and 3a of the # 1 cylinder are connected to the # 1 via the first and second hydraulic lash adjusters 10a and 10b on the # 1 cylinder according to the engine operating state. First and second valve stop mechanisms (lost motion mechanisms) 11a and 11b are provided as cylinder deactivation mechanisms 03 for stopping the opening and closing operations of the first and second intake valves 3a and 3a of the cylinders.

  Further, as described above, the intake side of the intake camshaft 5 is provided with a phase change type valve that makes the opening and closing timings of the intake valves 3a to 3b variable according to the engine operating state. A timing control device (intake electric VTC) 05 is provided. As the intake electric VTC 05, one that controls the relative rotation angle between the crankshaft 50 and the intake camshaft 5 by an electric motor described in, for example, Japanese Patent Application Laid-Open No. 2012-145036 is used.

  Hereinafter, the components in the # 1 and # 2 cylinders will be described. The four intake valves 3a to 3b are slidably held by the cylinder head 07 via the valve guides 4 and each stem end. Each valve retainer 3d provided in the vicinity of 3c is urged in the closing direction by each valve spring 12 elastically contacted between the inner upper surface of the cylinder head 07.

  The intake camshaft 5 is configured such that the rotational force of the crankshaft is transmitted by a timing belt via a timing pulley provided in a housing of the intake VTC05 provided at one end. The rotating cam 5a has an outer peripheral cam profile formed in an egg shape.

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

  As shown in FIGS. 5 and 6, the four hydraulic lash adjusters 10 a to 10 d include a bottomed cylindrical body 24 held in each cylindrical holding hole 07 a of the cylinder head 07, 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. A discharge hole 07b is formed inside the cylinder head 07 to discharge the hydraulic oil accumulated in the holding hole 07a to the outside.

  The body 24 is formed with a cylindrical first groove 24a on the outer peripheral surface, and is formed in the cylinder head 07 on the peripheral wall of the first groove 24a. A first passage hole 31 that communicates between the oil passage 30 opened in the concave groove 24a and the inside of the body 24 is formed to penetrate in the radial direction.

  Further, as shown in FIGS. 5A and 5B, the first and second hydraulic lash adjusters 10a and 10b (F and R sides) of the # 1 cylinder are provided with the valve stop mechanism shown in FIG. 6 on the bottom 24b side. The third and fourth hydraulic lash adjusters 10c, 10d on the # 2 cylinder side that are not provided extend downward from the bottom 24c of the body 24 and are 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 07. The main oil gallery 30a includes an oil pump shown in FIG. The lubricating oil is pumped from 64 through the electromagnetic switching valve 65.

  As shown in FIGS. 5 and 6, the plunger 27 has a cylindrical second concave groove 27a formed on the outer peripheral surface of the substantially central portion in the axial direction, and the peripheral wall of the second concave groove 27a includes the first concave groove 27a. 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 rotating cam 5a, when the pressure in the high pressure chamber 28 becomes low as the plunger 27 moves forward (upward movement) by the biasing force of the second coil spring 29d, the oil passage 30 holds it. The hydraulic fluid supplied into the 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 causes the check ball 29a to pass through the first concave groove 24a. The hydraulic oil is pushed open against the spring force of the one 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 between the end 6a is adjusted to zero lash.

  In the lift section of the rotating cam 5a, a downward load is applied to the plunger 27. Therefore, the check valve is closed, the hydraulic pressure in the high pressure chamber 28 is increased, and the oil in the high pressure chamber 28 becomes a gap between the plunger 27 and the body 24. The plunger 27 slightly leaks out (leak down).

  In the base circle section of the rotating cam 5a again, as described above, the clearance of each part is adjusted to zero lash by the advancement movement (upward movement) of the plunger 27 by the urging force of the second coil spring 29d. is there.

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

  The first and second valve stop mechanisms 11a and 11b are provided only on the first and second hydraulic lash adjusters 10a and 10b on the F side and the R side of the # 1 cylinder. As shown in FIG. It is not provided in the third and fourth hydraulic lash adjusters 10c, 10d on the F side and the R side of the cylinder.

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

  As shown in FIGS. 5A and 5B, the first and second valve stop mechanisms 11a and 11b include a pair of cylindrical sliding holes 34 formed continuously on the bottom side of each holding hole 07a, A pair of lost motion springs 35, which are elastically mounted between the bottom surface of the sliding hole 34 and the lower surface of the body 24 and urge the first and second hydraulic lash adjusters 10a and 10b upward, respectively, , And a restriction mechanism 36 for restricting the lost motion of the second hydraulic lash adjusters 10a and 10b.

  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 designed to be in contact with the ball.

  Further, the maximum upward movement position of each body 24 is regulated by a stopper pin 37 inserted and arranged inside the cylinder head 07. That is, the stopper pins 37 are arranged in the cylinder head 07 in the direction perpendicular to the axis toward the body 24, and the tip end portions 37a are slidably arranged in the first concave grooves 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 10a, 10b is lost by vertically moving between the holding hole 07a and the sliding hole 34 via 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 rotating cam 5a is absorbed, and the opening / closing operation of each intake valve 3a is stopped.

  As shown in FIGS. 5A to 5C, the restriction mechanism 36 of the first and second valve stop mechanisms 11 a and 11 b includes a movement hole 38 formed through the bottom 24 b of the body 24 in the inner radial direction, and the cylinder A restriction hole 39 formed in a direction perpendicular to the holding hole 07a in the head 07, a retainer 40 for supporting a spring fixed to one end of the movement hole 38, and a restriction hole from the movement hole 38 A cylindrical regulation pin 41 slidably provided across the interior of the 39, and is elastically mounted between the rear end of the regulation pin 41 and the retainer 40, so that the regulation pin 41 is placed in the regulation hole 39. And a return spring 42 urging in the direction.

  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 07 on one end side.

  Here, the restriction of the rotation direction of the body 24 slightly increases the amount of protrusion of the stopper pin 37, and a slit in the longitudinal direction of the shaft is provided in the first concave groove 24 a of the body 24. It can be easily realized by engaging with the tip. Alternatively, a separate rotation restricting member may be mounted between the cylinder head 07 and the body 24.

  The retainer 40 is formed in a cylindrical shape with a lid, a breathing hole 40a is formed in the bottom wall to ensure smooth movement of the regulating pin 41, and a central portion 40b where the breathing hole 40a on the rear end face faces. Are formed flat, but the outer end portions 40c, 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. 5B, the axial length of the retainer 40 is such that the rear end edge of the restricting pin 41 is in contact with the front end edge before the restricting pin 41 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. 5A and 5C, the restriction pin 41 is formed in a covered cylindrical shape and has an outer diameter slightly smaller than the inner diameter of the movement hole 38 to ensure smooth slidability. In addition, the distal end surface 41a of the distal end 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.

  5A, when the restriction pin 41 is moved into the restriction hole 39 by the spring force of the return spring 42, a part of the restriction pin 41 is accommodated in the restriction hole 39. Yes. As a result, the cylinder head 07 is locked, and the vertical movement of the hydraulic lash adjusters 10a and 10b on the F and R sides of the # 1 cylinder, that is, the lost motion is restricted.

  The regulation pin 41 has an outer diameter slightly smaller than the inner diameters of the movement hole 38 and the regulation hole 39 to ensure smooth slidability with respect to them, and the oil passage. When the tip surface 41a serving as a pressure receiving surface receives the hydraulic pressure supplied from the hole 44 to the restriction hole 39, it moves to the left in the figure against the spring force of the return spring 42 as shown in FIG. 5B. When the retainer 40 is brought into contact with the retainer 40 in the axial direction, the entire regulation pin 41 is accommodated in the movement hole 38. As a result, the first and second hydraulic lash adjusters 10a and 10b of the # 1 cylinder are allowed to move in the vertical direction, that is, the lost motion is performed.

  As shown in FIG. 1, the oil pressure fed from the oil pump 64 is supplied to the oil passage hole 44 (regulating hole 39) as a signal oil pressure via an electromagnetic switching valve 65. That is, the electromagnetic switching valve 65 serves as switching energy supply / supply stop conversion means (hydraulic supply / supply stop conversion means) that converts a state of supplying hydraulic pressure as switching energy and a state of stopping supply. Yes.

  The electromagnetic switching valve 65 switches a spool valve slidably provided inside a valve body (not shown) in two stages, on and off, by the electromagnetic force of the solenoid and the spring force of the coil spring. It has become. The solenoid is energized and de-energized (on and off) from the same control unit 63 that controls the drive of the compression ratio control actuator 62 of the variable compression ratio mechanism 02, so that the pump discharge passage and the oil passage hole 44 The hydraulic pressure is supplied to the restriction pin 41 and the pump discharge passage is closed and the oil passage hole 44 and the drain passage 66 are communicated to be controlled.

  Accordingly, when the engine is stopped, the solenoid is not energized from the control unit 63, and the electromagnetic switching valve 65 closes the pump discharge passage and connects the oil passage 44 and the drain passage 66. Therefore, the first and second valve stop mechanisms 11a, The lost motion operation by 11b is impossible. That is, the first and second valve stop mechanisms 11a and 11b are mechanically stable in a valve-operable state (valve operation mode) when supply of hydraulic pressure as switching energy is stopped. It has become.

  The control unit 63 detects the engine operating state based on information signals such as the engine speed, load, and 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 of the electric VTC 05 is driven and controlled by information signals from the crank angle sensor and the cam angle sensor, and the relative rotation angle between the intake camshaft 5 and the crankshaft 50 is controlled. At the same time, the electromagnetic switching valve 65 is controlled to convert and control valve stop and valve operation of the two intake valves 3a and 3a of the # 1 cylinder.

As shown in FIG. 1, the supercharger 04 is disposed on the upstream side of the throttle valve 67 in the intake passage, and is controlled by the control signal output from the control unit 63. The state in which the rotation can be synchronized with the rotation and the state in which the rotation is not possible are controlled by an on / off signal to the clutch mechanism, and the supercharging pressure is controlled by opening / closing control of the intake bypass valve 68.
[Valve device on the exhaust side]
The exhaust side valve gear is basically the same as the intake side, and as shown in FIG. 3, the opening ends of a pair of exhaust ports 70 and 70 per cylinder formed in the cylinder head 07 are respectively opened and closed. Two exhaust valves 71a, 71a, 71b, 71b are provided per cylinder. That is, the first and second exhaust valves 71a and 71a on the F side and the R side are provided in the # 1 cylinder, and the first and second exhaust valves 71b and 71b on the F side and the R side are provided in the # 2 cylinder.

  The exhaust side valve operating device is the same as that on the intake side shown in FIG. 4 and is shown with reference numerals in parentheses. An exhaust camshaft 73 having an egg-shaped rotary cam 73a that opens the exhaust valves 71a to 71b against the spring force of the valve springs 72 is provided. The exhaust valves 71a to 71b and the rotary cams are provided. The exhaust valves 71a to 71b are opened and closed with a constant valve lift amount via rollers 77 and swing arms 74 interposed between the exhaust valve 73a and the swing arm 74a.

  Further, a hydraulic lash adjuster 75a, which is a pivot that is held by the cylinder head 07 and adjusts the gap between each swing arm 74 and each exhaust valve 71a to 71b and the base circle of each rotary cam 73a to zero lash. .About.75d are respectively arranged. That is, there are four hydraulic lash adjusters 75a to 75d on the exhaust side, the first and second hydraulic lash adjusters 75a and 75b are disposed in the # 1 cylinder, and the third and fourth hydraulic lash adjusters 75c are disposed 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 exhaust valves 71a and 71a side of the # 1 cylinder have first and second valve stop mechanisms (lost motion mechanisms) 11a and 11a, respectively. 11b.

  On the other hand, the third and fourth hydraulic lash adjusters 75c and 75d of the exhaust valves 71b and 71b on the F and R sides of the # 2 cylinder do not have a valve stop mechanism.

  The exhaust-side first and second valve stop mechanisms have the same structure as the intake-side first and second valve stop mechanisms 11a and 11b shown in FIG. The detailed explanation is omitted. That is, it is elastically mounted between a cylindrical sliding hole 34 formed continuously on the bottom side of each holding hole 07a of the cylinder head 07, and between the bottom surface of the sliding hole 34 and the 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 regulating mechanisms 76 and 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 exhaust valves 71a and 71a on the # 1 cylinder side provided with the first and second valve stop mechanisms 11a and 11b having the restriction mechanisms 76 and 76 have valve lift amounts due to lost motion. When the valve is stopped, the lift is zero, and when the valve is not operated, the peak lift amount is constant.
[Operating characteristics of this embodiment]
FIG. 7 shows a map of the number of operating cylinders and the mechanical compression ratio ε, where the horizontal axis represents the engine speed and the vertical axis represents the engine torque.

  Looking at the vertical axis, the area surrounded by the thick solid line indicating the maximum positive torque characteristic (all cylinders) in the all cylinder operation and the maximum negative torque (emblem) characteristic (all cylinders) is the all cylinder operation area (second operation). Area).

  On the horizontal axis, this is a region surrounded by both broken lines of Ni (for example, 700 rpm) at idle speed of the engine speed and Nh (for example, 6000 rpm) of maximum speed. A region surrounded by these four lines is referred to as a second operation region (all cylinder operation region) (however, a region excluding the first operation region described later).

  On the other hand, the area surrounded by the thick solid line indicating the maximum positive torque characteristic (reduced cylinder) in the reduced cylinder operation and the maximum negative torque (emblem) characteristic (reduced cylinder) is the reduced cylinder operation area (first operation area). .

  On the horizontal axis, this is a region surrounded by both solid lines of Nl (for example, 1000 rpm) that is slightly higher than the idling speed of the engine speed and Nm (for example, 3500 rpm) of the medium speed speed. A region surrounded by these four lines is a first operation region (reduced cylinder operation region).

  Here, a two-dot chain line in FIG. 7 shows an equal mechanical compression ratio ε line in the all cylinder operation, and a one dot chain line shows an equal mechanical compression ratio ε line in the reduced cylinder operation.

First, looking at the all-cylinder operation region (second operation region), for example, the equal mechanical compression ratio ε line (two-dot chain line) of ε15 at which the mechanical compression ratio is 15 is the engine torque T15. This T15 is the maximum engine torque (all cylinders) that can be output without causing knocking in all cylinder operation ε15.
As the engine torque increases from T15, the compression ratio (mechanical compression ratio) ε is gradually decreased by the control of VCR02, and until the engine maximum positive torque (all cylinders) T9.5 reaches the compression ratio ε9.5. descend. That is, as the engine torque increases, so-called knocking occurs when the compression ratio ε is large. Therefore, the compression ratio ε is decreased by an amount that can suppress knocking. This is a map (two-dot chain line). In other words, since the compression ratio ε can be increased as much as possible while suppressing knocking, the thermal efficiency can be increased accordingly.
Here, for example, T9.5 is the maximum engine torque (all cylinders) that can be output without generating knocking at ε9.5.

In addition, although only described as compression ratio here, it is the meaning of the mechanical compression ratio mentioned above, and is different from the effective compression ratio mentioned later.
On the other hand, the large compression ratio ε15 is constant in the low engine torque region where the engine torque is less than T15. This is because, in a low engine torque region of less than T15, a large compression ratio ε15 can improve the theoretical thermal efficiency and improve fuel efficiency. However, even if the compression ratio ε15 is further increased, the combustion chamber Since the S / V ratio increases and the cooling loss increases, the fuel consumption does not increase any more. For example, the compression ratio ε15 is kept constant.

Next, looking at the reduced-cylinder operation region (first operation region), the alternate long and short dash line in FIG. 7 shows the equal compression ratio ε line in the reduced-cylinder region.
Here, looking at the ε15 line (one-dot chain line) in the reduced-cylinder operation, the engine torque on the vertical axis is approximately T15, that is, the maximum engine torque that can be output without causing knocking in the all-cylinder operation ε15 (all cylinders). Is almost the same. That is, the engine torque equivalent to that in the all-cylinder operation can be output with the same ε15 and the reduced-cylinder operation. This is because, although the combustion torque that can be generated by the reduced cylinder operation decreases, an engine torque decrease suppression occurs due to the reduction of engine friction and pump loss from the reduced cylinder operation, and the ratio of this decrease suppression is in the low engine torque region. Then it will be bigger.
In the reduced cylinder area on the lower engine torque side where the engine torque is less than T15, the large compression ratio ε15 is constant as in the second operating area (all cylinder operating area), but this is the reduced cylinder operating area. However, even if it is further increased from the compression ratio ε15, the S / V ratio of the combustion chamber is increased and the cooling loss is increased. Therefore, the fuel consumption is not further increased, and for example, the compression ratio ε15 is kept constant. (Same as all cylinder operation)
Next, let us consider a case where the engine torque increases from T15 in the same reduced cylinder operation state.

  As the engine torque increases, the compression ratio ε is suddenly reduced by the control of VCR02, and is reduced to the compression ratio ε9 in the vicinity of the engine torque T12. Here, the engine torque T12 corresponds to the maximum engine torque that can be output without knocking at the compression ratio ε12 in the all-cylinder operation, as described above.

  In the vicinity of T12 in the second operation region (all cylinders), the compression ratio ε12 is still relatively large, whereas in the first operation region (reduction cylinder), the compression ratio ε9 is greatly decreased in the vicinity of the same T12. .

  In the case of reduced cylinder operation, even if the engine torque is the same, the engine torque (load) generated in the operating cylinder is twice as high as that of the operating cylinder, so that knocking occurs especially as the engine torque increases. It becomes easy. Therefore, as the engine torque increases, the compression ratio ε decreases rapidly.

  By doing so, knocking can be avoided in the reduced-cylinder region, and the maximum torque characteristic in the reduced-cylinder can be increased up to T12. Therefore, the frequency of reduced-cylinder operation in actual traveling can be increased, and fuel efficiency performance in actual traveling Can be increased.

Here, consider a case where the accelerator is blown from the point of FIG. 7 (1) (for example, engine torque T15, about 1600 rpm) in the reduced cylinder region to increase the engine torque to T12.
On the reduced-cylinder region map in FIG. 7, the transition from the aforementioned point (1) to the point (2) at T12 is made. During that time, the compression ratio ε greatly decreases from 15 to 9.

  FIG. 8 is a graph in which the horizontal axis represents the mechanical compression ratio ε and the vertical axis represents the engine torque T.

  As shown in FIG. 8, T15 · ε15 (point (1)) is changed to T12 · ε9 (point (2)) while the cylinder is reduced. With respect to the point (2), a large engine torque of T12 can be realized while reducing cylinder operation while suppressing knocking by a low mechanical compression ratio of compression ratio ε9. That is, by this low mechanical compression, it is possible to obtain a large engine torque T12 while avoiding knocking while suppressing the fuel consumption deterioration factors such as the retard of the ignition timing and fuel richness as described above.

  As a result, as described above, the reduced-cylinder region having a large fuel efficiency effect can be expanded to the high engine torque side, the frequency of reduced-cylinder operation during actual traveling can be increased, and the fuel efficiency performance during actual traveling can be improved.

  Here, a conventional cylinder deactivation system which is not the present invention will be considered.

  In this case, the mechanical compression ratio ε is fixed, and an intermediate value such as a mechanical compression ratio ε11 is taken. In this case, the point at which the engine torque T15 is the same as point (1) in the reduced cylinder region in FIG. 8 is T15 · ε11 indicated by point (a).

When the accelerator is blown from this point (A), the engine torque reaches the point (A) of T13. However, if the engine torque is increased further, knocking will occur due to the compression ratio ε11. If it exceeds, it will be forced to switch to all-cylinder operation.
Here, the point (a) in FIG. 8 is a correlation line between the mechanical compression ratio ε and the maximum engine torque that can be generated without knocking from the point (1) to the point (2) in the reduced cylinder operation, Corresponds to the intersection with the ε11 constant line.

  On the other hand, in this embodiment, after passing through point (b), the compression ratio ε decreases toward point (2), so that the engine torque can be increased while suppressing knocking.

  At this time, as shown from the points (1) to (2) in FIG. 9, the intake valve closing timing (IVC) is brought closer to the bottom dead center, so that the charging efficiency is further improved and the engine torque is knocked. It can be further increased toward the maximum engine torque that can be generated without being generated.

  Furthermore, if the clutch of the supercharger 04 is turned on to close the bypass valve 68 and increase the supercharging pressure, the charging efficiency is further improved, and the engine torque in the reduced cylinder region can be increased to the engine torque T12 shown in (2).

  As a result, the maximum engine torque in the reduced cylinder can be increased by Tg indicated by the large arrow (outlined) in FIGS.

  Here, as a method of suppressing knocking other than the reduction of the mechanical compression ratio, there is a delay in the ignition timing. However, this causes problems such as a decrease in engine torque and a significant decrease in thermal efficiency. In addition to insufficient improvement in torque during operation, it goes against the improvement in fuel consumption, which is the purpose of reducing the number of cylinders.

  Even if the air-fuel ratio is made rich, the engine torque can be slightly increased while suppressing knocking. However, it is needless to say that this is also an inappropriate technique because it causes a deterioration in fuel consumption.

  In the present embodiment, knock suppression can be suppressed by reducing the mechanical compression ratio without particularly using these inappropriate knock suppression techniques.

  Here, in order to suppress knocking, the effective compression ratio can be reduced by greatly separating the intake valve closing timing (IVC) from the bottom dead center instead of reducing the mechanical compression ratio described above. However, this reduction in the effective compression ratio is accompanied by a reduction in the intake charging efficiency regardless of whether the IVC is separated from the bottom dead center toward the advance side or the retard side. The cylinder operation region cannot be expanded to the high engine torque side.

  Next, returning to FIG. 8 and FIG. 9, when reaching the point (2) (torque T12; reduced cylinder) in the reduced cylinder area of the present embodiment, the supercharging capacity (supercharger 04 depends on the number of revolutions) or more. Since air cannot be pushed in, even if the compression ratio is further lowered to further increase the knocking resistance, not only the engine torque cannot be improved, but also the thermal efficiency is lowered and the engine torque is lowered.

  Therefore, all-cylinder operation (point (3)) is switched so that higher engine torque can be produced.

  Specifically, as shown in the control flowchart (acceleration side) of the control unit 63 shown in FIG. 10, a control signal (cylinder deactivation release signal) is sent to the cylinder deactivation mechanism 03, the control hydraulic pressure is turned off, and the valve operation mode is set. Transition.

  That is, in step 1 of FIG. 10, the current engine operating state is read, and in step 2, it is determined whether or not the reduced cylinder operating state is in effect.

  If it is determined that the reduced-cylinder operation is not performed, the process returns as it is. If it is determined that the reduced-cylinder operation is performed, the process proceeds to Step 3.

  In this step 3, based on the map for reduced cylinder operation (region 1), the compression ratio ε control is performed by the VCR, the IVC control is performed by the electric VTC 05, and the clutch mechanism of the supercharger 04 (S / C) and the intake bypass valve 68 are controlled. Take control.

  In step 4, it is determined whether or not the boundary line between region 1 (first region) and region 2 (second region) shown in FIG. 7 has been reached, and if not, the process returns. If yes, go to Step 5.

  In step 5, a cylinder deactivation cancellation signal is output to the electromagnetic switching valve 65 (control current off), and a control signal for converting IVC to the retard side is output to the electric VTC05. Further, a signal for increasing the opening degree of the S / C intake bypass valve 68 is output.

  In Step 6, after outputting the control signal in Step 5, it is determined whether or not the predetermined time t has elapsed. If NO, the process returns to Step 6, and if it is determined that it has elapsed, the process proceeds to Step 7. .

  In step 7, a signal for increasing the compression ratio ε is output by VCR02, and the process proceeds to step 8.

  In step 8, based on the all-cylinder operation map, the compression ratio ε is controlled by the VCR02, the IVC is controlled by the electric VTC05, and the opening degree of the S / C clutch mechanism and the intake bypass valve 68 is controlled. To return.

  In this case, since all cylinders are operated, the engine torque is greatly increased and a torque shock is generated. Therefore, as shown in FIG. 9 (3), the intake valve closing timing (IVC) is converted to the retard side. Controlling, lowering the charging efficiency and increasing the opening degree of the intake bypass valve 68 of the supercharger 04 to lower the actual supercharging pressure, further lowering the charging efficiency, resulting from the transition to all-cylinder operation Increase in engine torque is suppressed.

  As a result, it is possible to suppress the throttle valve 67 from being throttled to reduce engine torque, thereby suppressing increase in engine torque while suppressing pump loss.

  After these processes are performed, the compression ratio ε is increased to 12 by VCR02 as shown in FIG.

  In other words, this compression ratio ε12 is the maximum mechanical compression ratio at which knocking does not occur in the engine torque T12 in all cylinder operation, and even in all cylinder operation that is disadvantageous to fuel consumption, the high compression ratio ε12 allows fuel consumption (thermal efficiency) as much as possible. ) Can be improved.

  By the way, assuming that the compression ratio ε is first increased and then the above-mentioned all-cylinder transition sequence is performed as this conversion order, there is an instant of the high compression ratio ε in the reduced cylinder state. Thus, large knocking (transient knocking) occurs.

  For this reason, in this embodiment, as shown in Step 5 to Step 7 in FIG. 10, the whole cylinder shift, the IVC change, and the control of the bypass valve 68 are performed first, and immediately after that (after a slight predetermined time t has elapsed). A control signal for increasing the compression ratio ε is output.

  Next, when the accelerator is further depressed, the engine torque in all-cylinder operation further increases and increases to T9.5 at (4) point in FIG.

  At this time, in order to increase the filling efficiency, the retarded IVC again approaches the bottom dead center. Further, the supercharger 04 keeps the intake bypass valve 68 fully closed while the clutch is ON.

  In the process from point (3) to point (4), the compression ratio ε traces the maximum value that can suppress knocking according to each engine torque, and at point (4) of the maximum engine torque T9.5, The compression ratio is ε9.5.

  On the other hand, as described above, when the conventional cylinder is deactivated, it is necessary to switch to all-cylinder operation at the point (a) in FIG. 8, but when the accelerator is further depressed, the engine torque moves toward the point (c). Will increase.

  However, since the compression ratio ε11 is fixed, when the point (c) of T11 is reached, it becomes the knocking limit in the all cylinder operation, and no further engine torque can be obtained.

  On the other hand, when returning to the present embodiment, as described above, the reduction control is performed to the compression ratio ε9.5 by the VCR02, so that the engine torque can be increased to T9.5 of (4) point while suppressing knocking.

  A supplementary explanation of the maximum engine torque T9.5 for all cylinders will be described. Since the mechanical supercharger (supercharger 04) rotates in synchronization with the engine rotation (in the embodiment, about 1600 rpm), the supercharging capability is limited by the rotational speed. The There is a charging efficiency determined under this restriction, that is, engine torque (all cylinders). At this time, the limit compression ratio ε at which knocking can be suppressed is 9.5 in this embodiment, and the engine torque at that time is T9. It is five. Therefore, although the engine torque can be improved up to T9.5, in other words, it cannot reach T9. The reason for this is that in all-cylinder operation, the air pushed by the supercharger 04 is distributed to all cylinders, so the effect of increasing the charging efficiency per cylinder is less than in the reduced-cylinder operation distributed to a small number of cylinders. It is.

  Therefore, the mechanical compression ratio at the maximum engine torque (T9.5) in all the cylinders does not have to be lowered as much as the compression ratio ε9 at the time of the reduced cylinder, and is slightly larger as the compression ratio ε9.5.

  In other words, as a feature when using the supercharger 04, when the cylinder is reduced, the air pushed by the supercharger 04 is distributed to a small number of operating cylinders. In addition, since it is a rotation-dependent type and a sufficient supercharging effect can be obtained from the low engine torque range, the maximum engine torque at the time of cylinder reduction is T12 (maximum torque that can be output without causing knocking in all cylinder operation ε12). It was big enough.

  By reducing the compression ratio ε to the minimum compression ratio ε9 in the control range with the VCR02, it is possible to generate the maximum engine torque that can be output without causing knocking in the large T12, that is, all the cylinders ε12, while reducing the cylinder. Moreover, the occurrence of knocking can be suppressed. Accordingly, the reduced-cylinder engine torque is sufficiently increased (T12), and the reduced-cylinder operation region is sufficiently expanded to the high engine torque side, so that the frequency of reduced-cylinder operation can be sufficiently increased and the fuel efficiency performance in actual driving can be sufficiently increased. .

  Next, the deceleration side operation scene that changes from the second operation region to the first operation region will be described.

  Consider the case where the engine torque decreases from the point (4) of the maximum engine torque of all cylinders, reaches the point (3), and further decreases as shown by the downward arrow (white) in FIG. .

  (3) When reaching the point, if the accelerator opening is still high to some extent, it is determined that the cruise is continuing, and as shown in the control flowchart (deceleration side) in FIG. Switch to the (reduced cylinder) state.

  That is, in step 11, the engine operating state is read, and in step 12, it is determined whether or not all cylinders are currently in operation (region 2).

  If all-cylinder operation is not performed, the process returns to step 11. However, if it is determined that all-cylinder operation is performed, the process proceeds to step 13 where the compression is performed to VCR02 based on the all-cylinder operation map. While outputting the signal of ratio control, the signal which performs IVC control is output to electric VTC05. A control signal is output to the clutch mechanism and the intake bypass valve 68.

  In step 14, it is determined whether or not the boundary line of region 1 to region 2 has been reached. If it is determined that the boundary line has not been reached, the process returns. If it is determined that the boundary line has been reached, the process proceeds to step 15.

  In step 15, it is determined whether or not the accelerator opening is equal to or greater than a predetermined value. If the accelerator opening is not equal to or greater than the predetermined value (less than the predetermined value), the process proceeds to step 21 and is determined to be equal to or greater than the predetermined value. If yes, go to Step 16. In step 16, a target mechanical compression ratio ε (ε0) is calculated for VCR02 from a target engine torque (accelerator opening) and the like.

  Thus, the target mechanical compression ratio ε0 is calculated (in FIGS. 7 and 8, for example, the target engine torque does not change and the target mechanical compression ratio ε0 is ε9), and the VCR is converted into the target ε0 in advance. Is output.

  Thereafter, in step 17, a decrease signal is output to the VCR02 toward the target mechanical compression ratio ε0, and in step 18, it is determined whether or not the actual mechanical compression ratio ε has reached the target ε0 by the VCR02. Here, the actual machine compression ratio ε may be detected from the phase of the control shaft 61 of the VCR 02 and the rotation angle of the motor 62.

If the target ε0 has been reached, the routine proceeds to step 19 where a cylinder deactivation signal, a signal for advancing the IVC, and an opening reduction signal for the intake bypass valve 68 are output, and the routine proceeds to step 20. (If the target ε0 has not been reached, return to step 17 again and repeat.)
In step 20, the mechanical compression ratio ε is controlled by the VCR 02 based on the map for the reduced cylinder operation (area 1), the IVC is controlled by the electric VTC 05, and the opening degree control is further performed by the intake bypass valve 68.

  That is, after confirming that the actual mechanical compression ratio ε of the VCR02 has actually decreased to the target compression ratio ε0 (for example, ε9 described above), the cylinder deactivation signal output, the IVC advance signal output, and the intake bypass of the supercharger 04 An opening reduction signal output of the valve 68 is performed.

  Here, as described above, the reason why the compression ratio ε is reduced first is as follows. That is, if the cylinder is deactivated when the compression ratio ε is large, the IVC is brought close to bottom dead center, or the opening degree of the intake bypass valve 68 is reduced, the charging efficiency of the operating cylinder is high, so that knocking is not performed. It will occur. Therefore, after confirming that the compression ratio ε has actually decreased to the target compression ratio ε0, these controls are performed.

  On the other hand, returning to step 15, if the accelerator opening is less than the predetermined value, the process goes to step 21 to output a signal for maintaining all cylinder operation (continuing the cylinder deactivation OFF signal), and at step 22 A compression ratio ε control signal is output to the VCR 02 based on a map for all cylinder operation (region 2) (two-dot chain line in FIG. 7). In addition, an IVC control signal is output to the electric VTC 05 and a control signal is output to the clutch mechanism and the intake bypass valve 68. In other words, at the point (3) of the all-cylinder operation, when the accelerator opening becomes less than the predetermined amount, it is determined that the deceleration is relatively abrupt and the all-cylinder operation state is maintained.

With all cylinders operating, the point decreases from the (3) point (T12, all cylinders) to the (1 ') point (T15, all cylinders) in FIG. 8, and the maximum negative torque Ta (emblem) characteristic is obtained. As a result, sufficient engine braking is produced and braking ability is increased.
The points (1 ′) in FIGS. 7 and 8 are points where ε15 and T15 are obtained in the all-cylinder operation. As described above, the points (1 ′) in FIG. 7 and FIG. Match.

  Here, if the cylinder shifts to the reduced cylinder state in this relatively abrupt deceleration process, the pump loss and valve friction of the idle cylinder are sufficiently small, so that the absolute value of the maximum negative engine torque also reaches Tb. As a result, the engine brake (emblem) is less effective and the braking ability is reduced (Tb <Ta).

  On the other hand, as described above, since the all-cylinder operation (maximum negative engine torque Ta) in which the engine brake is easily effective is achieved, the braking ability can be sufficiently increased (Ta> Tb).

  On the other hand, consider the scene where the accelerator is blown again and reaccelerated before the torque falls to the negative engine torque range.

  For example, let us assume a case in which re-acceleration is performed at a time point when the entire cylinder operation is continued to the point (1 ') (T15, all cylinders).

  Here, if it is assumed that the reduced-cylinder state has occurred, it is point (1), and since it switches to all-cylinder operation during reacceleration from there ((1) → (2) → (3)), the switching response Acceleration response will deteriorate by the amount of time.

  On the other hand, as in this embodiment, if all-cylinder operation is maintained during sudden deceleration, all-cylinder operation ((1 ' ) Point), and it is not necessary to switch the number of cylinders, so that a good acceleration response can be obtained. This embodiment also has such incidental effects.

  As described above, in this embodiment, in the reduced-cylinder operation region, the knocking resistance can be improved by a low compression ratio. Therefore, the reduced-cylinder operation region that is advantageous in fuel efficiency while suppressing knocking is expanded to the engine high torque side. it can. That is, it is possible to increase the driving frequency in the reduced-cylinder driving region where the fuel efficiency is good, and to improve the fuel efficiency when the vehicle is actually running.

  In addition, the frequency of operation in the all-cylinder operation region, which is disadvantageous in terms of fuel consumption, decreases, while in this all-cylinder operation region, thermal efficiency can be improved as much as possible by high compression ratio control. Therefore, it is possible to improve the fuel efficiency performance in the total during actual traveling of the vehicle.

Further, the accompanying effects as described above can be obtained including the acceleration scene and the deceleration scene.
[Second Embodiment]
12 to 17 show a second embodiment. As a supercharger, instead of the mechanical supercharger 04, for example, a so-called turbocharger using exhaust gas described in Japanese Patent Application Laid-Open No. 2009-209880 is used. 69 is used.

  An exhaust bypass valve 80 for bypassing the turbine inlet and outlet of the turbocharger is provided. This can control the opening degree by electronic control, and corresponds to what is called a so-called electric control waste gate valve.

  The exhaust bypass valve 80 reduces the turbine work caused by the exhaust gas, thereby reducing the supercharging work of the compressor.

  VCR02 is the same as that of the first embodiment, but the control range of the compression ratio ε is expanded, and the minimum control compression ratio ε is set as low as 8.5.

  Further, on the intake valve side of this embodiment, as shown in FIG. 13, an intake VEL 81 that variably controls the valve lift amount and the operating angle of each of the intake valves 3a to 3b is used. The intake VEL 81 is described in, for example, Japanese Patent No. 4998523 filed earlier by the applicant of the present application. Therefore, a specific description thereof is omitted, but as a basic configuration, a drive shaft 82 that is rotationally driven by the crankshaft 50 and The rotary cam 82a press-fitted and fixed to the outer periphery of the drive shaft 82, and the roller 14 of each swing arm 6 supported on the outer periphery of the drive shaft 82 so as to be swingable and provided at the upper ends of the intake valves 3a to 3b. And a swing cam 83 that opens the intake valves 3a to 3b in sliding contact with each other, and is interposed between the rotary cam 82a and the swing cam 83 to convert the rotational force of the rotary cam 82a into a swing motion. A transmission mechanism 84 that transmits a swinging force to the swing cam 83 and a control mechanism 85 that controls the attitude of the transmission mechanism 84 according to the engine operating state are provided.

  The control mechanism 85 is configured such that the rotational angle position is controlled by an electric motor via a speed reduction mechanism (not shown), and the electric motor is controlled to be rotated by the control unit 63.

  FIG. 14 shows a map of the number of operating cylinders and the mechanical compression ratio (ε). Looking at the maximum engine torque characteristics (all cylinders), this engine torque is low in the low rotation side region. This is because the exhaust gas absolute amount to be discharged is small in the low rotation range, the exhaust pressure cannot be sufficiently increased, and the turbine cannot be sufficiently rotated.

  On the other hand, in the high rotation range, the engine torque is higher than in the first embodiment using the supercharger 04, and the maximum output and the maximum torque are also increased. This is because the exhaust gas amount increases on the high rotation side of the high load region, the turbine is rotated at high speed, and the supercharging pressure by the compressor is sufficiently increased. Further, since the increase in operating friction accompanying the increase in rotation is smaller than that in the first embodiment using the supercharger 04, the torque of the engine is further improved.

  Here, when looking at the mechanical compression ratio (ε) map, the maximum positive torque in the reduced cylinder, that is, the engine torque at the boundary between the region 1 and the region 2 is T12.5, which is slightly compared with T12 of the supercharger 04. It is low. This is because the number of cylinders exhausting exhaust gas is halved, so the amount of exhaust gas as an engine is small even when the load is high, and it is difficult to increase the boost pressure, and the charging efficiency is sufficient. It is difficult to raise.

  Nonetheless, by controlling the compression ratio ε to be slightly low 10, knocking can be suppressed while improving the charging efficiency to some extent, and the engine torque T12.5 larger than that of the conventional technology and a considerable T13 can be achieved (FIGS. 14 and 15). Large white arrow). Therefore, similarly to the first embodiment, the reduced-cylinder operation region is expanded, the frequency of use of the reduced-cylinder region is increased, and the fuel efficiency performance during actual traveling of the vehicle can be improved.

  Next, consider a case where the accelerator is depressed to accelerate from the point (1) in the reduced-cylinder operation state (similar to the point (1) in the first embodiment). In the case of the first embodiment using the supercharger 04, as shown in FIG. 7, the change in the engine speed is slight, but in the second embodiment using the turbocharger 69, as shown in FIG. The engine torque increases while increasing the engine speed, that is, with a relatively large inclination. This is because the exhaust pressure is low and the supercharging pressure does not increase on the low rotation and low torque side, so it is better to increase the engine rotation while controlling it relatively to the low gear side with a transmission such as CVT. This is because acceleration can be improved. For such a purpose, such an inclination is provided.

  Then, as in the first embodiment, when the point (2) which is the boundary between the first operation region and the second operation region is reached, as shown in FIG. 15, the operation is switched to the all-cylinder operation and the compression is performed from the compression ratio ε10. The ratio changes to ε12.5.

  Specifically, as shown in the control flowchart (acceleration side) of FIG. 17, a cylinder deactivation control signal (cylinder deactivation release signal) is sent to the electromagnetic valve 65, the control hydraulic pressure of the cylinder deactivation mechanism 03 is turned OFF, and the intake / exhaust valve is operated. It moves to the aspect.

  The flowchart of FIG. 17 is almost the same as that of FIG. 10 of the first embodiment, but there are some differences. First, as a method of supercharging pressure control, the first embodiment is the clutch ON-OFF control of the supercharger 04 and the control of the intake bypass valve 68, whereas in this embodiment, the exhaust bypass valve of the turbocharger 69 is used. Although the control is 80, the basic concept of control is the same.

  Further, as a concept of intake valve closing timing (IVC) control, in the first embodiment, the IVC is moved away from the bottom dead center as a measure for reducing the charging efficiency and the effective compression ratio. In the form, the IVC is moved away from the bottom dead center, but the idea is the same.

  The configuration for changing the IVC is different from the first embodiment in which the electric motor VTC05 having a variable phase is provided, but the present embodiment is different in that an intake air VEL81 capable of changing an operating angle and a lift is also provided. The degree of freedom in controlling the opening timing (IVO) and valve overlap can be improved.

  That is, first, at step 31, the engine operating state is read at present, and at step 32, it is determined whether or not it is reduced cylinder operation (region 1). Here, when it is determined that the reduced-cylinder operation is not performed, the process returns, but when it is determined that the reduced-cylinder operation is performed, the process proceeds to step 33.

  In step 33, control signals are output to the VCR02, the electric VTC05, the intake VEL 81, and the exhaust bypass valve 80 based on the reduced-cylinder operation (region 1) map.

  In step 34, it is determined whether or not the engine torque shown in FIG. 14 has reached the boundary line of region 1 (first operation region) -region 2 (second operation region). However, if the boundary line is reached, the process proceeds to step 35.

  In this step 35, a cylinder deactivation cancellation signal and an IVC advance signal are output to the electromagnetic switching valve 65 and the electric VTC 05, and a signal for increasing the valve opening is output to the exhaust bypass valve 80.

  In step 36, after outputting the control signal in step 35, it is determined whether or not the predetermined time t has elapsed. If NO, the process returns to step 36, and if it is determined that it has elapsed, the process proceeds to step 37. .

  In step 37, a signal for increasing the compression ratio ε is output by VCR02, and the routine proceeds to step 38.

  In step 38, based on the all-cylinder operation map, the compression ratio ε is controlled by the VCR02, the IVC and the like are controlled by the electric VTC05 and VEL81, the opening degree of the exhaust bypass valve 80 is controlled, and then the process returns.

  Supplementally, since all cylinder operation is performed in the process of steps 35 to 37, the engine torque is greatly increased and a torque shock is generated. Therefore, as shown in FIG. Advancing, that is, unlike the first embodiment, the charging efficiency is lowered by separating it toward the bottom dead center side, and the actual boost pressure is lowered by increasing the opening of the exhaust bypass valve 80, Further, the charging efficiency is lowered, and an increase in engine torque is suppressed.

  Accordingly, it is possible to suppress the throttle valve 67 from being throttled, thereby suppressing the increase in engine torque while suppressing the pump loss.

  After these processes are performed, in step 37, the compression ratio ε is increased to 12.5 by the VCR as indicated by point (3) in FIG.

  This compression ratio ε12.5 is the maximum mechanical compression ratio at which knocking does not occur at engine torque T12.5 in all cylinder operation, and even in all cylinder operation that is disadvantageous in fuel efficiency, this high compression ratio ε12.5 is as much as possible. It improves fuel efficiency (thermal efficiency).

  By the way, if it is assumed that the compression ratio ε increases first and then the above-mentioned all-cylinder transition sequence is performed as this conversion order, there is an instant of the high compression ratio ε in the reduced cylinder state. Therefore, large knocking (transient knocking) occurs.

  Therefore, in the present embodiment, as shown in FIG. 17, in step 35, all cylinder transitions, IVC changes, and control of the exhaust bypass valve 80 are performed first, and immediately after that, after a short predetermined time t has elapsed in step 36, compression is performed. A control signal for increasing the ratio ε is output, which is the same concept as in the first embodiment.

  Next, when the accelerator is further depressed, the engine torque (charging efficiency) in all-cylinder operation further increases, and when the compression ratio ε is decreased, it increases to T9 at (4) point.

  In the case of the turbocharger 69, as compared with the mechanical supercharger, since the exhaust pressure becomes high in a high load region, the supercharging pressure can be easily increased, and the increase in operating friction due to the increase in engine rotation is small. And the engine torque can be increased compared to the supercharger. At this time, in order to further increase the charging efficiency, the exhaust bypass valve 80 is throttled and the advanced IVC approaches the bottom dead center again. And it controls when the bottom dead center which becomes the maximum filling efficiency passes a little.

  On the other hand, the intake angle VEL81 increases the operating angle and the valve lift, and further improves the charging efficiency. As a result, the valve overlap section can be increased incidentally, and the intake air is sucked into the cylinder by synchronizing with this section and the negative pressure wave of the exhaust pulsation, thereby further increasing the charging efficiency.

  The maximum compression ratio ε at which knocking does not occur with this increased charging efficiency is 9, and the engine torque at that time is T9 described above, which is T9.5 of the first embodiment using the supercharger 04. It is getting bigger.

  Further, as the rotation increases, the engine speed changes along with the maximum engine torque characteristic and operates to the maximum output point (maximum torque point at the maximum rotation; T8.5 · ε8.5) indicated by point (5) in FIG. Points are transferred.

  By the way, the compression ratio ε9 at the point (4) in FIG. 15 and the compression ratio ε8.5 at the point (5) are smaller than the minimum compression ratio ε10 at the reduced cylinder at the time of the reduced cylinder maximum engine torque, and a large compression ratio ε. Near the minimum value in the control range. Therefore, it has sufficient knocking resistance even at a high charging efficiency, can increase the maximum engine torque and the maximum output in the second operating region (all cylinder regions), and can improve the acceleration performance. That is, the maximum specific output / maximum specific torque (maximum output / maximum torque per unit displacement) can be increased.

  Alternatively, if the effect of improving the maximum specific output / maximum specific torque is set so as to reduce the total engine displacement, fuel efficiency can be further improved.

  In the present embodiment, as described above, the electric VTC 05 and the intake VEL 81 are provided side by side, thereby improving the degree of freedom in controlling the intake valve opening timing (IVO) and the valve overlap. This will be specifically described.

  In FIG. 16, when changing from the point (1) to the point (2) in the reduced cylinder state, the IVC approaches the bottom dead center in order to increase the charging efficiency. At that time, the operating angle is reduced by the intake air VEL81. By doing so, the change of valve overlap can be suppressed and the transient change of the residual gas amount can also be suppressed, so that the transient performance is stabilized.

  Also, when changing from the point (2) in the reduced cylinder operation to the point (3) in the all cylinder operation, the IVC is immediately moved away from the bottom dead center toward the advance side in order to reduce the charging efficiency of the operating cylinder. In this embodiment, since the intake VEL 81 and the electric VTC 05 are provided side by side, as shown in FIG. 16, it is possible to control the change in IVO so as to suppress the change in valve overlap.

  In other words, the change in valve overlap between (2) and (3) when switching between reduced cylinder and all cylinder operation is suppressed, and the residual gas amount between (2) and (3), where transient performance instability is likely to occur. The transient performance can be suppressed and the transient performance can be stabilized.

  Note that the broken-line intake lift curves shown in (2) and (3) of FIG. 16 show an example in which the electric VTC 05 is provided but the VEL 81 is not provided as in the first embodiment. On the other hand, the change of valve overlap can be suppressed.

Further, at the points (4) and (5) of the all-cylinder operation maximum engine torque, as described above, a large valve overlap can be actively used to increase the engine torque.
Further, at the time of deceleration, for example, the sequence from (4) point → (3) point → (2) point in FIG. 15 or (4) point → (3) point → (1 ′) point sequence. Is the same as that of the first embodiment, and the same effect can be obtained.

  The present invention is not limited to the configuration of each of the above-described embodiments. For example, in each embodiment, a two-cylinder internal combustion engine is shown, but a three-cylinder, four-cylinder, six-cylinder, eight-cylinder, ten-cylinder, and twelve-cylinder are used. You may apply to. In this case, the reduced cylinder operation means that some cylinders are in combustion operation and the remaining cylinders are cylinder deactivation (intake and exhaust valve deactivation).

  In each of the above-described embodiments, the mechanism for losing the rassia adjuster is shown as the cylinder deactivation mechanism, but any cylinder deactivation mechanism may be used. For example, as shown in Japanese Patent Application Laid-Open No. 10-82334, the operating cam may be switched by a hydraulic pin, or the operating cam may be selected by moving the camshaft in the axial direction as shown in Special Table 2010-20395. But it ’s okay.

  Further, although the variable phase VTC is an electric type, it may be a hydraulic type using a vane or the like. Although VEL is shown as the variable lift mechanism, other lift variable mechanisms may be used. The VCR of the variable mechanical compression ratio mechanism is not particularly limited, and for example, a method of changing the height of the piston itself as shown in JP-A-2003-65090 may be used. That is, the present invention can be applied to various systems, configurations, and structures 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.
[Claim a] The control apparatus for a multi-cylinder internal combustion engine according to claim 1,
While providing an intake valve timing mechanism that can change the closing timing of the intake valve as an engine valve,
In the first operating region on the lower torque side than the boundary line between the two regions, the intake valve timing mechanism is controlled so that the closing timing of the intake valves of the remaining cylinders excluding the some cylinders approaches the bottom dead center. ,
A control apparatus for a multi-cylinder internal combustion engine, characterized in that, in a second operating region on a higher torque side than a boundary line between the two regions, the closing timing of the intake valves of all the cylinders is controlled so as to be away from the bottom dead center.

  According to the present invention, since the intake charge efficiency can be increased in the first operation region (reduction cylinder region), the engine torque can be increased correspondingly, and the operation frequency in the reduction cylinder region with good fuel consumption can be further increased. This can further improve the fuel efficiency when the vehicle is actually running.

  On the other hand, when shifting to the second operation region (all cylinder region), the number of operating cylinders increases and the engine torque increases. Therefore, there has been a problem that a torque increase shock occurs or the pump valve is increased because the throttle valve is greatly throttled to suppress it.

On the other hand, since the intake valve closing timing of all cylinders is away from the bottom dead center, the intake charging efficiency is lowered, so that it is possible to suppress the torque shock described above and throttle the throttle valve, thereby suppressing pump loss. Fuel consumption can be improved.
[Claim b]
The control apparatus for a multi-cylinder internal combustion engine according to claim 1,
When the operating state of the engine changes from the first operating region on the low torque side to the second operating region on the high torque side, the operation of the engine valve in the partial cylinder is started by the cylinder deactivation mechanism in advance. Then, immediately after that, the variable compression ratio mechanism changes the relatively low first mechanical compression ratio to the relatively high second mechanical compression ratio.

According to the present invention, it is possible to avoid the moment when the high mechanical compression ratio is achieved in the reduced cylinder operation state, and it is possible to avoid the occurrence of transient knocking.
[Claim c]
The control apparatus for a multi-cylinder internal combustion engine according to claim 1,
A control device for a multi-cylinder internal combustion engine, comprising a supercharger that performs supercharging in a first operating region on a lower torque side than a boundary line between the two regions.

According to the present invention, the first operating region (reduced cylinder region) advantageous for fuel consumption can be further expanded to the higher engine torque side by supercharging, and the fuel efficiency performance during actual vehicle travel can be further improved.
[Claim d]
The control apparatus for a multi-cylinder internal combustion engine according to claim c,
The control device for a multi-cylinder internal combustion engine, wherein the supercharger is a mechanical supercharger.

According to this, even when the engine torque is low (low exhaust pressure), the boost pressure can be sufficiently increased, and the first operating region (reduced cylinder region) advantageous for fuel efficiency can be further increased. The fuel efficiency can be further improved when the vehicle is actually running.
[Claim e]
The control apparatus for a multi-cylinder internal combustion engine according to claim c,
A control device for a multi-cylinder internal combustion engine, wherein the supercharger is a turbocharger.

  According to this, in addition to the fuel efficiency improvement shown in claim c, the turbocharger can easily increase the supercharging pressure in a high load range (exhaust pressure high) as compared with the mechanical supercharger. Since the increase in operating friction with the increase in rotation is small, the maximum specific output / maximum specific torque (maximum output / maximum torque per unit displacement) can be increased, and acceleration performance is improved. Alternatively, the total engine displacement can be set to be small as the acceleration performance is improved, and in this case, the fuel efficiency can be further improved.

01 ... Internal combustion engine 02 ... Variable mechanical compression ratio mechanism (VCR)
03 ... Cylinder deactivation mechanism 04 ... Supercharger (mechanical supercharger)
05 ... Intake electric VTC
06 ... Cylinder block 07 ... Cylinder head 07a ... Holding hole 3a, 3a ... # 1 cylinder side first and second intake valves 3b, 3b ... # 2 cylinder side first and second intake valves 5 ... Intake camshaft 5a ... Rotary cam 6 ... Swing arm 6a ... One end 6b ... Other end 10a, 10b ... Intake side first and second hydraulic lash adjusters 10c, 10d of # 1 cylinder # 3 Intake side third and fourth oil pressure of # 2 cylinder Rashia adjuster 11a ... 1st valve stop mechanism 11b ... 2nd valve stop mechanism 12 ... Intake side valve spring 13 ... Bearing part 14 ... Roller 24 ... Body 27 ... Plunger 27b ... Tip head 34 ... Sliding hole 35 ... Lost motion spring 36 ... Restriction mechanism 38 ... Movement hole 39 ... Restriction hole 40 ... Retainer 41 ... Restriction pin 42 ... Return spring 44 ... Oil passage hole 66 ... Drain passage 6 DESCRIPTION OF SYMBOLS 4 ... Oil pump 65 ... Electromagnetic switching valve 71a, 71a ... 1st, 2nd exhaust valve 71b, 71b on the # 1 cylinder side ... 1st, 2nd exhaust valve on the # 2 cylinder side 72 ... Exhaust side valve spring 73 ... Exhaust Side camshaft 73a ... Rotating cam 74 ... Exhaust side swing arm 75a, 75b ... # 1 cylinder exhaust side first and second hydraulic lash adjusters 75c, 75d ... # 2 cylinder exhaust side third and fourth hydraulic lash adjusters

Claims (6)

Among a plurality of cylinders, a cylinder deactivation mechanism capable of stopping the operation of intake valves and exhaust valves of some cylinders, a variable compression ratio mechanism capable of changing a mechanical compression ratio of all cylinders, the cylinder deactivation mechanism, A controller for controlling a variable compression ratio mechanism, and a control device for a multi-cylinder internal combustion engine,
A first operation region in which the operation of the intake valves and exhaust valves of the partial cylinders is stopped by the cylinder deactivation mechanism, and a second operation region in which the intake valves and exhaust valves of all cylinders are operated,
When the engine operating state changes from the first operating region on the engine low torque side to the second operating region on the engine high torque side, the variable compression ratio mechanism causes the relatively low first mechanical compression ratio. Change to a relatively high second mechanical compression ratio,
When the engine operating state changes from the second operating region to the first operating region, the mechanical compression ratio is reduced by the variable compression mechanism in advance, and then the part is stopped by the cylinder deactivation mechanism. A control apparatus for a multi-cylinder internal combustion engine, wherein operation of an intake valve and an exhaust valve of a cylinder is stopped . The control apparatus for a multi-cylinder internal combustion engine according to claim 1,
The control apparatus for a multi-cylinder internal combustion engine, wherein the first mechanical compression ratio is set in the vicinity of a minimum mechanical compression ratio in a variable control range of the variable compression ratio mechanism. The control apparatus for a multi-cylinder internal combustion engine according to claim 1,
A control device for a multi-cylinder internal combustion engine, wherein a mechanical compression ratio in the vicinity of the maximum engine torque in the second operating region is set lower than the first mechanical compression ratio. The control apparatus for a multi-cylinder internal combustion engine according to claim 1,
When the operating state of the engine changes from the first operating region on the low torque side to the second operating region on the high torque side, the operation of the engine valve in the partial cylinder is started by the cylinder deactivation mechanism in advance. Then, immediately after that, the variable compression ratio mechanism changes the relatively low first mechanical compression ratio to the relatively high second mechanical compression ratio. The control apparatus for a multi-cylinder internal combustion engine according to claim 1,
A control device for a multi-cylinder internal combustion engine, comprising a supercharger that performs supercharging in a first operating region on a lower torque side than a boundary line between the two regions. A multi-cylinder internal combustion engine having a cylinder deactivation mechanism capable of stopping the operation of intake valves and exhaust valves of some cylinders, and a variable compression ratio mechanism capable of changing the mechanical compression ratio of all cylinders. A controller,
A first operation region in which the operation of the intake valves and exhaust valves of the partial cylinders is stopped by the cylinder deactivation mechanism, and a second operation region in which the intake valves and exhaust valves of all cylinders are operated,
When the engine operating state changes from the first operating region on the engine low torque side to the second operating region on the engine high torque side, the variable compression ratio mechanism causes the relatively low first mechanical compression ratio. Change to a relatively high second mechanical compression ratio,
When the engine operating state changes from the second operating region to the first operating region, the mechanical compression ratio is reduced by the variable compression mechanism in advance, and then the part is stopped by the cylinder deactivation mechanism. A controller for a multi-cylinder internal combustion engine, wherein operation of an intake valve and an exhaust valve of a cylinder is stopped .

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