JP2017125435A - Internal combustion engine control device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To bring an ignition timing close to an optimal ignition timing while curbing backfire.SOLUTION: An internal combustion engine 100 comprises an engine body 1 mounted with a cylinder 6, an ignition plug 18 which is installed on the engine body 1 and ignites mixed gas in the cylinder 6 and a variable valve timing mechanism B which is installed on the engine body 1 and adapted to variably change a valve closing timing of an intake valve 13 to an arbitrary timing. A control device 200 of the internal combustion engine has an ignition timing control section which controls the ignition plug, when an engine load is in at least a portion of a low load region, in a manner that brings an ignition timing in line with: a timing of when a distance between a valve head 13a and a valve seat of the intake valve 13 becomes not more than a flame-out distance; and the timing prior to a valve closing timing of the intake valve 13.SELECTED DRAWING: Figure 1

Description

  The present invention relates to a control device for an internal combustion engine.

  Patent Document 1 discloses a conventional control device for an internal combustion engine, which is configured to advance the ignition timing when the intake valve closing timing is away from the intake bottom dead center compared to when it is not separated. It is disclosed.

JP 2002-257018 A

  As a result of inventors' earnest research, in some internal combustion engines, the intake valve closing timing is retarded from the intake bottom dead center in the operating region on the low load side, for example, the engine body in a state where the expansion ratio is higher than the actual compression ratio When the engine is operated, the ignition timing at which the output torque of the engine body becomes maximum in a predetermined operating range (hereinafter referred to as “MBT (Minimum advance for the Best Torque)”) is advanced from the intake valve closing timing. I found out there might be.

  However, Patent Document 1 does not mention the relationship between the ignition timing and the intake valve closing timing, and it is unclear how far the ignition timing is advanced. If the ignition timing is set to an advance side with respect to the intake valve closing timing, there is a risk of backfire in which the flame generated in the combustion chamber flows backward to the intake passage side. Generally, it is set on the retard side with respect to the intake valve closing timing. Therefore, there is a problem that the ignition timing cannot be brought close to MBT.

  The present invention has been made paying attention to such problems, and in the operation region where the MBT is on the advance side of the intake valve closing timing, the occurrence of backfire is prevented and the ignition timing is set to the MBT. The purpose is to get closer.

  In order to solve the above problems, according to an aspect of the present invention, an engine body provided with a cylinder, an ignition plug provided in the engine body for igniting an air-fuel mixture in the cylinder, and an intake valve provided in the engine body An internal combustion engine control device for controlling an internal combustion engine comprising a variable valve timing mechanism configured to be capable of changing the valve closing timing to an arbitrary timing, wherein the engine load is in at least a part of the low load region Sometimes, the ignition plug is controlled so that the ignition timing is the timing when the interval between the valve head of the intake valve and the valve seat is less than the extinguishing distance and before the closing timing of the intake valve. An ignition timing control unit is provided.

  According to this aspect of the present invention, it is possible to bring the ignition timing closer to the MBT while preventing the occurrence of backfire even in the operation region where the MBT is on the more advanced side than the intake valve closing timing.

FIG. 1 is a schematic configuration diagram of an internal combustion engine and an electronic control unit that controls the internal combustion engine according to the first embodiment of the present invention. FIG. 2 is an exploded perspective view of the variable compression ratio mechanism. FIG. 3 is a diagram for explaining the operation of the variable compression ratio mechanism. FIG. 4A is a side view of the intake cam fixed to the intake camshaft. FIG. 4B is a diagram illustrating a lift curve of the intake valve. FIG. 5 is a schematic configuration diagram of the variable valve timing mechanism. FIG. 6 is a diagram for explaining the operation of the variable valve timing mechanism. FIG. 7A is a diagram illustrating a mechanical compression ratio, an actual compression ratio, and an expansion ratio. FIG. 7B is a diagram illustrating a mechanical compression ratio, an actual compression ratio, and an expansion ratio. FIG. 7C is a diagram illustrating a mechanical compression ratio, an actual compression ratio, and an expansion ratio. FIG. 8 is a diagram showing the relationship between the theoretical thermal efficiency and the expansion ratio. FIG. 9 is a map showing the operation region of the engine body. FIG. 10 is a diagram illustrating an example of the mechanical compression ratio, the intake valve closing timing, and the optimal ignition timing according to the engine load when the mechanical rotation speed is constant. FIG. 11 is a flowchart illustrating the ignition timing control according to the first embodiment of the present invention. FIG. 12 is a flowchart illustrating the advance limit ignition timing calculation control according to the first embodiment of the present invention. FIG. 13 is a table for calculating the flame extinguishing distance Lq between the two walls based on the initial in-cylinder pressure Pivc. FIG. 14 is a diagram for explaining a method of calculating the crank angle Xg at which the valve lift amount of the intake valve is the flame extinguishing distance Lq between the two walls. FIG. 15 is a table for calculating the correction coefficient K based on the temperature of the cooling water. FIG. 16 is a table for calculating the lift loss Lh based on the engine rotation speed. FIG. 17 is a diagram illustrating the minimum distance d from the center of the spark plug to the intake port opening. FIG. 18 is a diagram showing the relationship between the engine rotation speed and the flame combustion speed v, and is a table for calculating the flame combustion speed v based on the engine rotation speed. FIG. 19 is a flowchart illustrating the advance limit ignition timing calculation control according to the second embodiment of the present invention.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. In the following description, the same reference numerals are assigned to similar components.

(First embodiment)
FIG. 1 is a schematic configuration diagram of an internal combustion engine 100 and an electronic control unit 200 that controls the internal combustion engine 100 according to a first embodiment of the present invention.

  As shown in FIG. 1, the internal combustion engine 100 includes an engine body 1, an intake device 20, and an exhaust device 30.

  The engine body 1 includes a cylinder block 2, a cylinder head 3 attached to the upper part of the cylinder block 2, a crankcase 4 attached to the lower part of the cylinder block 2, and an oil pan 5 attached to the lower part of the crankcase 4. And comprising.

  A plurality of cylinders (cylinders) 6 are formed in the cylinder block 2. Inside the cylinder 6 is housed a piston 7 that reciprocates in the cylinder 6 under combustion pressure. The piston 7 is connected to a crankshaft 9 rotatably supported in the crankcase 4 via a connecting rod 8, and the reciprocating motion of the piston 7 is converted into rotational motion by the crankshaft 9. A space defined by the cylinder head 3, the cylinder 6, and the piston 7 is a combustion chamber 10.

  The cylinder head 3 opens on one side surface (right side in the figure) of the cylinder head 3 and opens to the combustion chamber 10, and opens on the other side surface (left side in the figure) of the cylinder head 3 and burns. An exhaust port 12 that opens to the chamber 10 is formed.

  Further, the cylinder head 3 is ignited for igniting an air-fuel mixture of fuel injected from a fuel injection valve 17 attached to each intake branch pipe 23b of the intake manifold 23, which will be described later, and air. The plug 18 is attached so as to face the combustion chamber 10. The fuel injection valve 17 may be attached to the cylinder head 3 so that fuel can be directly injected into the combustion chamber 10.

  The cylinder head 3 is provided with an intake valve 13 for opening and closing the opening between the combustion chamber 10 and the intake port 11 and an intake valve operating device 40 for opening and closing the intake valve 13. The intake valve device 40 includes an intake camshaft 41 extending in the cylinder row direction, an intake cam 42 fixed to the intake camshaft 41, a tappet 43 that contacts the intake cam 42 and pushes down the intake valve, and an intake camshaft 41. And a variable valve timing mechanism B that can change the valve closing timing of the intake valve 13 (hereinafter referred to as “intake valve closing timing”) to an arbitrary timing. Details of the variable valve timing mechanism B will be described later with reference to FIGS.

  Further, the cylinder head is provided with an exhaust valve 14 for opening and closing the opening of the combustion chamber 10 and the exhaust port 12, and an exhaust valve operating device 90 for opening and closing the exhaust valve 14. The exhaust valve operating device 90 includes an exhaust camshaft 91 extending in the cylinder row direction, an exhaust cam 92 fixed to the exhaust camshaft 91, and a tappet 93 that contacts the exhaust cam 92 and pushes down the intake valve.

  Further, the engine body 1 according to the present embodiment includes a variable compression ratio mechanism A at a connecting portion between the cylinder block 2 and the crankcase 4. The variable compression ratio mechanism A according to this embodiment changes the relative position of the cylinder block 2 and the crankcase 4 in the cylinder axial direction, thereby reducing the volume of the combustion chamber 10 when the piston 7 is located at the compression top dead center. To change. A relative position sensor 211 for detecting the relative positional relationship between the cylinder block 2 and the crankcase 4 is attached to a connecting portion between the cylinder block 2 and the crankcase 4. An output signal indicating a change in the interval between the crankcase 2 and the crankcase 4 is output. Details of the variable compression ratio mechanism A will be described later with reference to FIGS.

  The intake device 20 is a device for introducing air into the cylinder 6 through the intake port 11, and includes an air cleaner 21, an intake pipe 22, an intake manifold 23, an electronically controlled throttle valve 24, and a throttle sensor. 212, an air flow meter 213, and an intake pressure sensor 214.

  The air cleaner 21 removes foreign matters such as sand contained in the air.

  One end of the intake pipe 22 is connected to the air cleaner 21, and the other end is connected to a surge tank 23 a of the intake manifold 23.

  The intake manifold 23 includes a surge tank 23a and a plurality of intake branch pipes 23b branched from the surge tank 23a and connected to the openings of the intake ports 11 formed on the side surface of the cylinder head. The air guided to the surge tank 23a is evenly distributed in each cylinder 6 through the intake branch pipe 23b. Thus, the intake pipe 22, the intake manifold 23, and the intake port 11 form an intake passage for guiding air into each cylinder 6.

  The throttle valve 24 is provided in the intake pipe 22. The throttle valve 24 is driven by a throttle actuator (not shown) to change the passage sectional area of the intake pipe 22 continuously or stepwise. By adjusting the opening degree of the throttle valve 24 (hereinafter referred to as “throttle opening degree”) by the throttle actuator, the flow rate of the air sucked into each cylinder 6 can be adjusted. The throttle opening is detected by the throttle sensor 212.

  The air flow meter 213 is provided in the intake pipe 22 upstream of the throttle valve 24. The air flow meter 213 detects the flow rate of air flowing through the intake pipe 22 (hereinafter referred to as “intake intake air amount”).

  The intake pressure sensor 214 is provided in the surge tank 23a. The intake pressure sensor 214 detects the pressure in the surge tank 23a.

  The exhaust device 30 is a device for purifying the combustion gas (exhaust gas) generated in the combustion chamber 10 and discharging it to the outside air. The exhaust device 30 is an exhaust manifold 31, an exhaust aftertreatment device 32, an exhaust pipe 33, An air-fuel ratio sensor 215.

  The exhaust manifold 31 includes a plurality of exhaust branch pipes connected to the openings of the exhaust ports 12 formed on the side surface of the cylinder head, and a collection pipe that collects the exhaust branch pipes into one.

  The exhaust aftertreatment device 32 is connected to the collecting pipe of the exhaust manifold 31. The exhaust aftertreatment device 32 is a device for purifying the exhaust gas and discharging it to the outside air, and carries various catalysts (for example, a three-way catalyst) for purifying harmful substances on a carrier.

  One end of the exhaust pipe 33 is connected to the exhaust aftertreatment device 32, and the other end is an open end. Exhaust gas discharged from each cylinder 6 through the exhaust port 12 to the exhaust manifold 31 flows through the exhaust aftertreatment device 32 and the exhaust pipe 33 and is discharged to the outside air.

  The air-fuel ratio sensor 215 is provided in the collecting pipe of the exhaust manifold 31 and detects the air-fuel ratio of the exhaust.

  The electronic control unit 200 is composed of a digital computer and is connected to each other by a bi-directional bus 201. A ROM (read only memory) 202, a RAM (random access memory) 203, a CPU (microprocessor) 204, an input port 205, and an output port 206.

  The input port 205 detects the temperature of the cooling water that cools the engine body 1 in addition to the output signals of the relative position sensor 211, the throttle sensor 212, the air flow meter 213, the intake pressure sensor 214, the air-fuel ratio sensor 215, and the like. An output signal from the water temperature sensor 216 or the like is input via each corresponding AD converter 207. Further, the output voltage of the load sensor 217 that generates an output voltage proportional to the amount of depression of the accelerator pedal 220 (hereinafter referred to as “accelerator depression amount”) is input to the input port 205 via the corresponding AD converter 207. The In addition, an output signal of a crank angle sensor 218 that generates an output pulse every time the crankshaft 9 of the engine body 1 rotates, for example, 15 ° is input to the input port 205 as a signal for calculating the engine rotation speed and the like. . Further, an output signal of a cam position sensor 219 that generates a signal representing the rotation angle of the intake camshaft is input to the input port 205. As described above, output signals of various sensors necessary for controlling the internal combustion engine 100 are input to the input port 205.

  Control components such as the fuel injection valve 17, the spark plug 18, the variable compression ratio mechanism A, and the variable valve timing mechanism B are electrically connected to the output port 206 through corresponding drive circuits 208.

  The electronic control unit 200 controls the internal combustion engine 100 by outputting a control signal for controlling each control component from the output port 206 based on the output signals of various sensors input to the input port 205.

  FIG. 2 is an exploded perspective view of the variable compression ratio mechanism A according to the present embodiment.

  As shown in FIG. 2, a plurality of protrusions 50 spaced from each other are formed below both side walls of the cylinder block 2, and cam insertion holes 51 having a circular cross section are formed in each protrusion 50. ing. On the other hand, a plurality of protrusions 52 are formed on the upper wall surface of the crankcase 4 so as to be fitted between the corresponding protrusions 50 with a space therebetween. A circular cam insertion hole 53 is formed.

  The variable compression ratio mechanism A includes a pair of camshafts 54 and 55, and is circularly inserted into the cam insertion holes 53 on the camshafts 54 and 55 at predetermined intervals. A cam 58 is fixed. These circular cams 58 are coaxial with the rotational axes of the camshafts 54 and 55. On the other hand, an eccentric shaft 57 (see FIG. 3) that is eccentrically arranged with respect to the rotation axis of each camshaft 54, 55 extends on both sides of each circular cam 58, and another circular cam 56 is arranged on this eccentric shaft 57. Is mounted eccentrically and rotatable. As shown in FIG. 2, these circular cams 56 are arranged on both sides of each circular cam 58, and these circular cams 56 are rotatably inserted into the corresponding cam insertion holes 51.

  At one end of each of the camshafts 54 and 55, worm wheels 63 and 64 that engage with a pair of worms 61 and 62 provided on the control shaft 60 are attached. The pair of worms 61 and 62 have their spiral directions reversed so that the camshafts 54 and 55 can be rotated in opposite directions. The control shaft 60 is rotated by a drive motor 65. By rotating the drive motor 65 and rotating the camshafts 54 and 55 in opposite directions, the piston 7 is compressed at top dead center as shown in FIG. The volume of the combustion chamber 10 when it is located at is changed. A cam rotation angle sensor 221 that generates an output signal representing the rotation angle of the cam shaft 55 is attached to the cam shaft 55, and the output signal of the cam rotation angle sensor 221 is passed through the corresponding AD converter 207 to the electronic control unit. 200 is input. Hereinafter, the operation of the variable compression ratio mechanism A will be described with reference to FIG.

  FIG. 3 is a diagram for explaining the operation of the variable compression ratio mechanism A.

  FIG. 3A is a diagram showing a state where the volume of the combustion chamber 10 is maximized when the piston 7 is located at the compression top dead center by the variable compression ratio mechanism A. FIG. FIG. 3B is a view showing a state in which the volume of the combustion chamber 10 is between the maximum and minimum when the piston 7 is located at the compression top dead center by the variable compression ratio mechanism A. FIG. 3C is a view showing a state where the volume of the combustion chamber 10 is minimized by the variable compression ratio mechanism A when the piston 7 is located at the compression top dead center.

  When the circular cams 58 fixed on the camshafts 54 and 55 are rotated in opposite directions as shown by arrows in FIG. 3A from the state shown in FIG. 3 (A), the eccentric shafts 57 move away from each other. In order to move, the circular cam 56 rotates in the opposite direction to the circular cam 58 in the cam insertion hole 51. As a result, as shown in FIG. 3B, the position of the eccentric shaft 57 is changed from a high position to an intermediate height position. Next, when the circular cam 58 is further rotated in the direction indicated by the arrow, the eccentric shaft 57 is at the lowest position as shown in FIG.

  3A to 3C show the positional relationship between the center a of the circular cam 58, the center b of the eccentric shaft 57, and the center c of the circular cam 56 in each state.

  3A to 3C, the relative positions of the crankcase 4 and the cylinder block 2 are determined by the distance between the center a of the circular cam 58 and the center c of the circular cam 56. As the distance between the center a of 58 and the center c of the circular cam 56 increases, the cylinder block 2 moves away from the crankcase 4. That is, the variable compression ratio mechanism A changes the relative position between the crankcase 4 and the cylinder block 2 by a crank mechanism using a rotating cam. When the cylinder block 2 moves away from the crankcase 4, the volume of the combustion chamber 10 when the piston 7 is located at the compression top dead center increases. Thus, the volume of the combustion chamber 10 when the piston 7 is positioned at the compression top dead center can be changed by rotating the camshafts 54 and 55.

  FIG. 4A is a side view of the intake cam 42 fixed to the intake camshaft 41. FIG. 4B is a diagram illustrating a lift curve of the intake valve 13.

  As shown in FIG. 4A, the intake cam 42 includes a cam portion 421, a base circle portion 422, and a buffer portion 423.

  The cam portion 421 is a portion that pushes down the intake valve 13 via the tappet 43 when the intake camshaft 41 rotates.

  The base circle portion 422 is a portion where the valve lift amount of the intake valve 13 becomes zero without pushing down the intake valve 13 when the intake camshaft 41 rotates.

  The buffer portion 423 is formed between the cam portion 421 and the base circle portion 422, and is a portion that pushes down the intake valve 13 via the tappet 43 similarly to the cam portion 421, but as shown in FIG. This is the portion where the valve lift varies in proportion to the crank angle. Since such a buffer portion 423 is configured so that the change in the valve lift amount with respect to the crank angle becomes gradual, the intake valve 13 is opened and closed after passing through the buffer portion 423, thereby It is possible to reduce the noise and impact force when the cam portion 421 of the cam 42 contacts the tappet 43 to reduce noise, and to protect the intake cam 42 and the tappet 43. In the present embodiment, the intake valve closing timing basically indicates the crank angle at the time when the valve lift amount becomes zero. In FIG. 4B, the cam portion 421 and the retard side (the right side in FIG. 4B). An arbitrary crank angle in the boundary with the buffer unit 423 or in the buffer unit 423 on the retard side may be defined as the intake valve closing timing.

  FIG. 5 is a schematic configuration diagram of the variable valve timing mechanism B according to the present embodiment provided at one end portion of the intake camshaft 41.

  As shown in FIG. 5, the variable valve timing mechanism B includes a timing pulley 71 that is rotated by a crankshaft 9 in the direction of an arrow through a timing belt, a cylindrical housing 72 that rotates together with the timing pulley 71, and an intake cam. A rotating shaft 73 that rotates together with the shaft 41 and is rotatable relative to the cylindrical housing 72; a plurality of partition walls 74 that extend from the inner peripheral surface of the cylindrical housing 72 to the outer peripheral surface of the rotating shaft 73; A vane 75 extending from the outer peripheral surface of the rotating shaft 73 to the inner peripheral surface of the cylindrical housing 72 between the partition walls 74. An advance hydraulic chamber 76 and a retard hydraulic chamber 77 are formed on both sides of each vane 75, respectively.

  The hydraulic oil supply control to the hydraulic chambers 76 and 77 is performed by a hydraulic oil supply control valve 78 driven by the electronic control unit 200. The hydraulic oil supply control valve 78 includes hydraulic ports 79 and 80 connected to the hydraulic chambers 76 and 77, a hydraulic oil supply port 82 discharged from the hydraulic pump 81, a pair of drain ports 83 and 84, And a spool valve 85 that performs communication cutoff control between the ports 79, 80, 82, 83, and 84.

  When the phase of the intake cam 42 of the intake camshaft 41 should be advanced, the spool valve 85 is moved to the right in FIG. 5 and the hydraulic oil supplied from the supply port 82 is advanced via the hydraulic port 79. The hydraulic oil in the retard hydraulic chamber 77 is discharged from the drain port 84 while being supplied to the hydraulic chamber 76. At this time, the rotary shaft 73 is rotated relative to the cylindrical housing 72 in the direction of the arrow.

  On the other hand, when the phase of the intake cam 42 of the intake camshaft 41 should be retarded, the spool valve 85 is moved to the left in FIG. 5 and the hydraulic oil supplied from the supply port 82 causes the hydraulic port 80 to move. And the hydraulic oil in the advance hydraulic chamber 76 is discharged from the drain port 83. At this time, the rotating shaft 73 is rotated relative to the cylindrical housing 72 in the direction opposite to the arrow.

  When the spool valve 85 is returned to the neutral position shown in FIG. 5 while the rotation shaft 73 is rotated relative to the cylindrical housing 72, the relative rotation operation of the rotation shaft 73 is stopped. The relative rotation position at that time is held. In this way, the variable valve timing mechanism B can advance or retard the phase of the intake cam 42 of the intake camshaft 41 by a desired amount.

  FIG. 6 is a diagram for explaining the operation of the variable valve timing mechanism B. FIG.

  The solid line in FIG. 6 shows the lift curve when the phase of the intake cam 42 of the intake camshaft 41 is advanced most by the variable valve timing mechanism B, and the broken line in FIG. The lift curve when the phase of the cam 42 is most retarded is shown. Therefore, the valve opening period of the intake valve 13 can be arbitrarily set between the range indicated by the solid line and the range indicated by the broken line in FIG. 5, and the intake valve closing timing is also set to an arbitrary value within the range indicated by the arrow C in FIG. The crank angle can be set.

  That is, by the variable valve timing mechanism B, the intake valve closing timing is changed from the valve closing timing when the phase of the intake cam 42 of the intake camshaft 41 is most advanced (hereinafter referred to as “advanced side limit closing timing”). The timing can be changed to any timing up to the valve closing timing when the phase of the intake cam 42 of the intake camshaft 41 is most retarded (hereinafter referred to as “retard-side limit closing timing”).

  The variable valve timing mechanism B shown in FIG. 1 and FIG. 5 is an example. For example, a variable valve timing mechanism that can change only the intake valve closing timing while keeping the intake valve opening timing constant, etc. Various types of variable valve timing mechanisms can be used.

  Next, the meanings of the terms “mechanical compression ratio”, “actual compression ratio”, and “expansion ratio” used in this specification will be described with reference to FIGS. 7A to 7C. 7A to 7C show the engine body 1 in which the combustion chamber volume is 50 ml and the stroke volume of the piston 7 is 500 ml for explanation of each term. In these FIGS. 7A to 7C, the combustion chamber is shown. The volume represents the volume of the combustion chamber 10 when the piston 7 is located at the compression top dead center.

  FIG. 7A is a diagram for explaining the mechanical compression ratio.

  The mechanical compression ratio is a value mechanically determined only from the stroke volume of the piston 7 and the combustion chamber volume during the compression stroke, and is represented by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 7A, the mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11.

  FIG. 7B is a diagram for explaining the actual compression ratio.

  The actual compression ratio is a value determined from the actual piston stroke volume and the combustion chamber volume from when the compression action is actually started until the piston 7 reaches the top dead center, and (combustion chamber volume + actual stroke volume). ) / Combustion chamber volume. That is, as shown in FIG. 7B, even if the piston 7 starts to rise during the compression stroke, the compression action is not performed while the intake valve 13 is open, and the actual compression starts from the time when the intake valve 13 is closed. The action begins. Therefore, the actual compression ratio is expressed as above using the actual stroke volume. In the example shown in FIG. 7B, the actual compression ratio is (50 ml + 450 ml) / 50 ml = 10.

  FIG. 7C is a diagram illustrating the expansion ratio.

  The expansion ratio is a value determined from the stroke volume of the piston 7 and the combustion chamber volume during the expansion stroke, and is represented by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 7C, the expansion ratio is (50 ml + 500 ml) / 50 ml = 11.

  FIG. 8 is a diagram showing the relationship between the theoretical thermal efficiency and the expansion ratio.

  The solid line in FIG. 8 shows the change in the theoretical thermal efficiency when the actual compression ratio and the expansion ratio are substantially equal, that is, in a normal cycle. In this case, it can be seen that the theoretical thermal efficiency increases as the expansion ratio increases, that is, as the actual compression ratio increases. Therefore, in order to increase the theoretical thermal efficiency in a normal cycle, the actual compression ratio should be increased. However, the actual compression ratio can only be increased to some extent due to the restriction of the occurrence of knocking during engine high-load operation, and thus the theoretical thermal efficiency cannot be sufficiently increased in a normal cycle.

  On the other hand, under such circumstances, it was studied to increase the theoretical thermal efficiency while strictly dividing the mechanical compression ratio and the actual compression ratio. As a result, the theoretical thermal efficiency is governed by the expansion ratio, and the actual thermal efficiency is compared to the theoretical thermal efficiency. The ratio was found to have little effect. That is, when the actual compression ratio is increased, the explosive force is increased, but the energy required for the compression is increased, and as a result, it is found that the theoretical thermal efficiency is hardly increased even if the actual compression ratio is increased.

  On the other hand, when the expansion ratio is increased, the period during which the pressing force acts on the piston 7 during the expansion stroke becomes longer, and the period during which the piston 7 applies the rotational force to the crankshaft 9 becomes longer. Therefore, the theoretical thermal efficiency increases as the expansion ratio increases. The broken line ε = 10 in FIG. 8 indicates the theoretical thermal efficiency when the expansion ratio is increased with the actual compression ratio fixed at 10. Thus, the theoretical thermal efficiency when the expansion ratio is increased while maintaining the actual compression ratio ε at a low value, and the theoretical thermal efficiency when the actual compression ratio shown by the solid line in FIG. 8 increases with the expansion ratio. It can be seen that there is no significant difference from the amount of increase.

  Thus, if the actual compression ratio is maintained at a low value, knocking does not occur. Therefore, if the expansion ratio is increased while the actual compression ratio is maintained at a low value, the theoretical thermal efficiency is reduced while preventing the occurrence of knocking. Can greatly increase. In general, internal combustion engines tend to have lower thermal efficiency when the engine load is lower. Therefore, in order to improve the thermal efficiency during engine operation and improve fuel efficiency, improve the thermal efficiency when the engine load is low. It is effective.

  Hereinafter, basic control of the variable compression ratio mechanism A and the variable valve timing mechanism B according to the present embodiment will be described with reference to FIG.

  FIG. 9 is a map showing an operation region of the engine body 1. Hereinafter, for the sake of convenience, the region below the first load line when the operating region of the engine body 1 is divided into three equal parts by the first load line and the second load line is referred to as a low load region. The area below the second load line and excluding the low load area is referred to as a medium load area. A region where the engine load is higher than that of the second load line is referred to as a high load region.

  In the electronic control unit 200, when the engine operating state determined based on the engine speed and the engine load (accelerator depression amount) is in a region below the load line L1 existing slightly on the first load line side in the middle load region. The intake valve closing timing is fixed at the retard side limit closing timing that is most retarded from the intake bottom dead center, the intake air amount is controlled by the throttle valve 24, and the mechanical compression ratio is fixed at the upper limit mechanical compression ratio. . The upper limit mechanical compression ratio is a mechanical compression ratio when the combustion chamber volume is minimized (the state shown in FIG. 3C).

  Thus, when the engine operating state is in the region below the load line L1, the electronic control unit 200 maintains the expansion ratio at the maximum expansion ratio by fixing the mechanical compression ratio to the upper limit mechanical compression ratio, and closes the intake valve. By fixing the timing to the retard side limit closing timing, the actual compression ratio is maintained at a predetermined value (11 in this embodiment) that does not cause knocking or pre-ignition.

  When applied to the engine body 1 shown in FIGS. 7A to 7C, by fixing the intake valve closing timing to the retard side limit closing timing, for example, the actual piston stroke volume is changed from 500 ml to 200 ml. By fixing the ratio to the upper limit mechanical compression ratio, for example, the combustion chamber volume is changed from 50 ml to 20 ml. Therefore, in the engine body 1 shown in FIGS. 7A to 7C, when the engine operating state is in the low load region, the actual compression ratio is (20 ml + 200 ml) / 20 ml = 11 and the expansion ratio is (20 ml + 500 ml) / 20 ml = 26.

  As a result, in the region below the load line L1, the expansion ratio can be maintained at the maximum expansion ratio while maintaining the actual compression ratio at a value at which knocking does not occur. Thermal efficiency can be greatly increased.

  The electronic control unit 200 controls the throttle valve 24 so that the intake air amount becomes the target intake air amount corresponding to the engine load when the engine operating state is in the region below the load line L1. Specifically, if the engine rotational speed is constant, the engine load is high so that the throttle valve 24 is fully opened when the engine load exists at point A on the load line L1, as shown in FIG. Indeed, the throttle opening is increased. Therefore, when the engine load becomes higher than the load line L1, the throttle valve 24 can no longer control the intake air amount. Therefore, when the engine load becomes higher than the load line L1, the intake air amount is increased by advancing the intake valve closing timing from the retard side limit closing timing to the intake bottom dead center side.

  That is, when the engine operating state is in a region higher than the load line L1, the electronic control unit 200 fixes the throttle valve 24 fully open, controls the intake air amount by the variable valve timing mechanism B, and has an actual compression ratio. The mechanical compression ratio is lowered from the upper limit mechanical compression ratio so as to be maintained at a predetermined value.

  Specifically, if the engine speed is constant, the electronic control unit 200 determines that the intake valve closing timing is advanced when the engine load exists at point B on the entire load line as shown in FIG. The intake air amount is increased by advancing the intake valve closing timing from the retard side limit closing timing to the intake bottom dead center side as the engine load increases so as to reach the limit closing timing. Then, the electronic control unit 200 decreases the mechanical compression ratio from the upper limit mechanical compression ratio as the engine load increases so that the actual compression ratio is maintained at a predetermined value.

  In the engine main body 1 shown in FIGS. 7A to 7C, when the intake valve closing timing is advanced from the retarded limit closing timing to the intake bottom dead center, for example, the actual piston stroke volume is changed from 500 ml to 400 ml. Then, in order to maintain the actual compression ratio at a constant 11, the electronic control unit 200 reduces the mechanical compression ratio so that the combustion chamber volume becomes 40 ml.

  Thus, in the region higher than the load line L1, the intake valve closing timing is controlled toward the advance side limit closing timing, and the controlled valve closing timing is used to maintain the actual compression ratio at a predetermined value. Accordingly, the mechanical compression ratio is made smaller than the upper limit mechanical compression ratio. Therefore, although the expansion ratio is smaller than the maximum expansion ratio even in a region higher than the load line L1, the engine body 1 can be operated with the expansion ratio maintained at a value higher than the actual compression ratio. . Therefore, even in a region higher than the load line L1, it is possible to increase the theoretical thermal efficiency while suppressing the occurrence of knocking. Further, since the throttle valve 24 is fixed fully open in the region higher than the load line L1, the pumping loss can be made substantially zero.

  As described above, in the present embodiment, the variable compression ratio mechanism A and the variable valve timing mechanism B are cooperatively controlled based on the engine operation state, so that the actual compression ratio does not cause knocking in the entire operation region. The engine body 1 is operated with the expansion ratio higher than the actual compression ratio.

  In addition to the variable compression ratio mechanism A and the variable valve timing mechanism B, the electronic control unit 200 controls the ignition timing of the spark plug 18. In general, the ignition timing of the spark plug 18 is controlled to an ignition timing that is closest to the MBT within a range in which knocking does not occur (MBT when knocking does not occur in the MBT, hereinafter referred to as “optimum ignition timing”).

  Here, if the engine load is constant, the MBT basically tends to change toward the advance side as the engine speed increases. This is because the higher the engine speed, the higher the rising speed of the piston and the shorter the combustion period.

  If the engine rotation speed is constant, the MBT basically tends to change toward the advance side as the engine load decreases. This is because as the engine load decreases, the amount of intake air decreases, the compression end temperature of the air-fuel mixture decreases, and the combustion speed decreases.

  If the engine speed and the engine load are constant, the MBT basically tends to change toward the advance side as the mechanical compression ratio increases. This is because, as the mechanical compression ratio is increased, the cross-sectional shape of the combustion chamber 10 along the cylinder axial direction in the vicinity of the compression top dead center becomes flat, the attenuation of gas flow of the in-cylinder gas increases, and the combustion speed decreases. It is.

  For this reason, the variable compression ratio mechanism A is provided as in the present embodiment and has a high compression ratio when the engine load is low. In an internal combustion engine in which the mechanical compression ratio is fixed at a high compression ratio (for example, 15) without A, the combustion speed becomes very slow on the low load side, and the MBT tends to change further to the advance side.

  In such an internal combustion engine, the variable valve timing mechanism B is provided as in the present embodiment, and the intake valve closing timing is retarded from the intake bottom dead center on the low load side, and the expansion ratio is made higher than the actual compression ratio. If the engine body 1 is to be operated in this state, the MBT may be on the more advanced side than the intake valve closing timing. In particular, when the variable compression ratio mechanism A and the variable valve timing mechanism B are provided and controlled cooperatively as in this embodiment, the mechanical compression ratio can be increased on the low load side than on the high load side. Therefore, the retard amount from the intake bottom dead center at the intake valve closing timing on the low load side can be increased. For this reason, the reduction in the combustion speed on the low load side becomes remarkable, and the MBT tends to come to the advance side with respect to the intake valve closing timing.

  FIG. 10 shows an intake valve and a mechanical compression ratio according to the engine load when the mechanical rotation speed is constant when the above-described basic control is performed on the variable compression ratio mechanism A and the variable valve timing mechanism B. It is the figure which showed an example of the closing timing and the optimal ignition timing.

  As shown in FIG. 10, it can be seen that the optimum ignition timing is closer to the advance side than the intake valve closing timing in a part of the operation region on the low load side. Thus, when the optimal ignition timing is advanced from the intake valve closing timing, the problem is how far the ignition timing is advanced. This is because if the optimal ignition timing is on the more advanced side than the intake valve closing timing, if ignition is performed at the optimal ignition timing, the flame propagates from the combustion chamber 10 into the intake port 11. This is because there is a possibility that backfire occurs in which the flame flows backward to the intake passage side.

  Therefore, in this embodiment, the valve lift amount of the intake valve 13, which is the distance between the valve head (valve head) 13a of the intake valve 13 and the valve seat on which the valve head 13a is seated when the intake valve 13 is closed, If the flame extinguishing distance Lq between the two walls defined by the following equation (1) is equal to or less than that, there is no risk of backfire, and the crank angle when the valve lift amount of the intake valve 13 becomes the extinguishing distance Lq between the two walls Is the advance limit ignition timing Xg. If the optimum ignition timing SA is on the more advanced side than the advance limit ignition timing Xg, the advance limit ignition timing Xg is set as the target ignition timing. On the other hand, if the optimum ignition timing SA is the same as the advance limit ignition timing Xg or on the retard side with respect to the advance limit ignition timing Xg, the optimum ignition timing SA is set as the target ignition timing. There are two crank angles at which the valve lift amount of the intake valve 13 becomes the flame extinguishing distance Lq between the two walls, the intake valve opening timing side and the closing timing side. Is the advance limit ignition timing Xg.

Lq = 14.8 × (Pivc) ^−0.9 × (Tw) ^−0.5 (1)
Lq [mm]: Flame extinguishing distance between two walls Pivc [MPa]: In-cylinder pressure at the time of intake valve closing (initial in-cylinder pressure)
Tw [K]: Flame extinguishing wall surface temperature

  Hereinafter, the ignition timing control of the spark plug 18 according to this embodiment will be described with reference to FIGS.

  FIG. 11 is a flowchart illustrating the ignition timing control according to the present embodiment. The electronic control unit 200 repeatedly executes this routine at a predetermined calculation cycle (for example, 10 [ms]) during engine operation.

  In step S1, the electronic control unit 200 reads the engine rotational speed and the engine load, and detects the current engine operating state.

  In step S2, the electronic control unit 200 detects the mechanical compression ratio based on the detection value of the relative position sensor 211 of the variable compression ratio mechanism A. Further, the electronic control unit 200 reads the detection values of the crank angle sensor 218 and the cam position sensor 219 of the intake camshaft 41, and detects the intake valve closing timing based on these detection values.

  In step S <b> 3, the electronic control unit 200 refers to a map previously created by experiment or the like and stored in the ROM 202, and calculates the optimum ignition timing SA based on the engine operating state, the mechanical compression ratio, and the intake valve closing timing.

  In step S4, the electronic control unit 200 reads the advance limit ignition timing Xg calculated as needed during engine operation separately from this routine. The advance limit ignition timing calculation control for calculating the advance limit ignition timing Xg will be described later with reference to FIG.

  In step S5, the electronic control unit 200 determines whether or not the optimum ignition timing SA is on the advance side with respect to the advance limit ignition timing Xg. The electronic control unit 200 proceeds to the process of step S6 if the optimum ignition timing SA is on the advance side with respect to the advance limit ignition timing Xg. On the other hand, if the optimum ignition timing SA is the same as the advance limit ignition timing Xg or is on the retard side with respect to the advance limit ignition timing Xg, the electronic control unit 200 proceeds to step S7.

  In step S6, the electronic control unit 200 sets the target ignition timing to the advance limit ignition timing Xg.

  In step S7, the electronic control unit 200 sets the target ignition timing to the optimum ignition timing SA.

  In step S8, the electronic control unit 200 controls the spark plug 18 so that the ignition timing becomes the target ignition timing.

  FIG. 12 is a flowchart illustrating the advance limit ignition timing calculation control according to this embodiment. The electronic control unit 200 repeatedly executes this routine at a predetermined calculation cycle (for example, 10 [ms]) during engine operation.

  In step S11, the electronic control unit 200 reads the engine rotational speed and the engine load, and detects the current engine operating state.

  In step S12, the electronic control unit 200 detects the mechanical compression ratio based on the detection value of the relative position sensor 211 of the variable compression ratio mechanism A. Further, the electronic control unit 200 reads the detection values of the crank angle sensor 218 and the cam position sensor 219 of the intake camshaft 41, and detects the intake valve closing timing based on these detection values.

  In step S13, the electronic control unit 200 estimates the in-cylinder pressure (hereinafter referred to as “initial in-cylinder pressure”) Pivc at the intake valve closing timing. The initial in-cylinder pressure Pivc correlates with the pressure in the surge tank 23a at the intake valve closing timing, and the pressure in the surge tank 23a changes according to the engine operating state (strictly, the throttle opening and the engine speed). . Therefore, in the present embodiment, a map in which the engine operating state and the initial in-cylinder pressure Pivc are associated with each other is created in advance through experiments or the like, and the map is stored in the ROM 202 in advance. The electronic control unit 200 estimates the initial in-cylinder pressure Pivc based on the engine operating state by referring to the map.

  In step S14, the electronic control unit 200 refers to the table of FIG. 13 previously created by experiment or the like and stored in the ROM 202, and calculates the flame extinguishing distance Lq between the two walls based on the estimated initial in-cylinder pressure Pivc. . As indicated by the above equation (1), the flame extinguishing distance Lq between the two walls correlates with the initial in-cylinder pressure Pivc and the extinguishing wall temperature Tw, and as the initial in-cylinder pressure Pivc increases, the extinguishing wall temperature Tw. The higher the is, the shorter the flame extinguishing distance Lq between the two walls.

  Here, the flame extinguishing wall surface temperature Tw can be generally regarded as the temperature of cooling water for cooling the engine body 1 (temperature of cooling water for cooling the cylinder head 3 around the intake valve 13). The temperature of the cooling water is basically kept at a constant reference water temperature (for example, 90 ° C.). Therefore, in the present embodiment, the initial in-cylinder pressure shown in FIG. 13 is calculated by calculating the flame extinguishing distance Lq between the two walls when the extinguishing wall temperature Tw is the reference water temperature according to the initial in-cylinder pressure Pivc through experiments or the like. A table is created in which Pivc and the extinguishing distance Lq between two walls are associated with each other.

  In step S15, the electronic control unit 200 calculates a crank angle Xg at which the valve lift amount of the intake valve 13 becomes the flame extinguishing distance Lq between the two walls. In this embodiment, since the lift curve of the intake valve 13 is constant, if the intake valve closing timing is known, the valve lift amount of the intake valve 13 corresponding to the crank angle is known as shown in FIG. Can calculate the crank angle Xg at which the flame extinguishing distance Lq between the two walls becomes.

  In step S16, the electronic control unit 200 sets the crank angle Xg at which the valve lift amount of the intake valve 13 is the two-wall extinguishing distance Lq as the advance limit ignition timing.

  According to the present embodiment described above, the engine main body 1 including the cylinder 6 (cylinder), the spark plug 18 provided in the engine main body 1 for igniting the air-fuel mixture in the cylinder 6, and the engine main body 1 are provided. An electronic control unit 200 (control device) that controls the internal combustion engine 100 that includes a variable valve timing mechanism B configured to be able to change the valve closing timing of the intake valve 13 to an arbitrary timing. The ignition timing is a time when the interval between the valve head 13a (valve head) of the intake valve 13 and the valve seat is equal to or less than the two-wall extinction distance Lq (extinguishing distance). And an ignition timing control unit that controls the spark plug 18 so that the timing is earlier than the intake valve closing timing.

  Therefore, even in the operation region where the MBT is on the more advanced side than the intake valve closing timing, it is possible to improve fuel efficiency by preventing the occurrence of backfire and bringing the ignition timing closer to the optimal ignition timing, and thus MBT.

(Second Embodiment)
Next, a second embodiment of the present invention will be described. The ignition timing control according to this embodiment is the same as that of the first embodiment, but the advance angle limit ignition timing calculation control is different from that of the first embodiment. Hereinafter, the difference will be mainly described.

  In the first embodiment described above, assuming that the temperature of the cooling water is basically kept at a constant reference water temperature, the extinguishing distance Lq between the two walls when the extinguishing wall temperature Tw is the reference water temperature is shown in FIG. The calculation was made in accordance with the initial in-cylinder pressure Pivc with reference to the 13 tables. However, for example, during the warm-up operation, the temperature of the cooling water is lower than the reference water temperature. Further, for example, when the low load operation is continued after the high load operation is continued, the temperature of the cooling water may be transiently higher than the reference water temperature. When the temperature of the cooling water is lower than the reference water temperature, that is, when the flame-extinguishing wall temperature Tw is lower than the reference water temperature, the flame extinguishing distance Lq between the two walls is increased in view of the above equation (1). On the other hand, when the temperature of the cooling water is higher than the reference water temperature, that is, when the flame-extinguishing wall temperature Tw is higher than the reference water temperature, the quenching distance Lq between the two walls is reduced in view of the above equation (1).

  Therefore, in the present embodiment, the extinguishing wall surface temperature Tw calculated according to the initial in-cylinder pressure Pivc with reference to the table of FIG. 13 is set to the two-wall extinguishing distance Lq according to the cooling water temperature. I decided to correct it. Specifically, the correction coefficient K is calculated based on the cooling water temperature with reference to the table shown in FIG. 15, and the correction coefficient K is calculated as the flame extinguishing distance Lq between the two walls calculated with reference to the table of FIG. Multiply.

  Thereby, the flame extinguishing distance Lq between the two walls can be accurately calculated according to the temperature of the cooling water. Specifically, for example, when the temperature of the cooling water is lower than the reference water temperature, the flame extinguishing distance Lq between the two walls becomes large, so that the advance limit ignition timing can be set to the advance side. As a result, the ignition timing can be made closer to the optimal ignition timing, and fuel efficiency can be improved. On the other hand, when the temperature of the cooling water is lower than the reference water temperature, the flame extinguishing distance Lq between the two walls becomes small. Therefore, it is possible to improve fuel efficiency by bringing the ignition timing closer to the optimal ignition timing while suppressing the occurrence of backfire by suppressing the ignition timing from being advanced excessively.

  In the first embodiment described above, the intake valve 13 is pushed down by the intake cam 42 via the tappet 43. At this time, the thermal expansion of the components of the cylinder head 3, the intake valve 13, the intake valve device 40, etc. In consideration of the above, a predetermined gap (hereinafter referred to as “valve clearance”) is formed between the base circle portion 422 of the intake cam 42 and the tappet 43. However, if the valve clearance changes due to deterioration over time or the like, the valve lift amount and the valve closing timing of the intake valve 13 change, so that periodic adjustment is necessary.

  Therefore, in the present embodiment, as the tappet 43, a hydraulic tappet having an automatic valve clearance adjustment mechanism for keeping the valve clearance at zero is used. Since the structure of the hydraulic tappet is well known, it will be briefly described here, not shown. The hydraulic tappet communicates with a plunger whose tip is in contact with the tappet 43, a body that houses the plunger, a first hydraulic chamber formed inside the plunger, and a second hydraulic chamber formed inside the body. A check ball that seals the passage, and a spring that is arranged in the second hydraulic chamber and constantly presses the plunger toward the intake cam 42 (the tappet 43).

  When the tappet 43 is not pressed by the intake cam 42, the hydraulic tappet pushes the plunger by the spring force of the spring to bring the tappet 43 into contact with the base circle portion 422 of the intake cam 42, thereby reducing the valve clearance. It is kept at zero. On the other hand, when the pressing force by the intake cam 42 is applied to the tappet 43, the plunger is pushed down, and the second hydraulic chamber is sealed by the check ball and becomes high pressure. As a result, the position of the plunger (= the position of the tappet 43) is fixed at a predetermined position by the hydraulic pressure in the second hydraulic chamber. When this state is reached, the intake valve 13 is pushed down by the intake cam 42 via the tappet 43, and the valve Lifting begins.

  In such a hydraulic tappet, due to its structure, a small amount of oil in the second hydraulic chamber leaks into the first hydraulic chamber from the gap between the inner diameter of the body and the outer diameter of the plunger during the valve lift, and the second hydraulic chamber in the valve lift. The hydraulic pressure in the hydraulic chamber decreases. Therefore, the plunger is slightly lowered toward the intake valve 13 from the predetermined position at the start of the valve lift during the valve lift, so that the intake valve 13 is closed earlier by the amount corresponding to the decrease. That is, the intake valve closing timing changes to the advance side by the amount of the plunger decrease. When the amount of decrease in the plunger position during the valve lift is referred to as lift loss Lh [mm], the lift loss Lh is because the time during which the intake cam 42 presses the tappet 43 becomes longer as the engine speed is lower. growing. Therefore, the lift loss Lh increases as the engine speed decreases, and the intake valve closing timing changes to the advance side.

  Therefore, in the present embodiment, the lift loss Lh is calculated based on the engine rotational speed with reference to the table shown in FIG. Thus, the advance limit ignition timing can be changed to the advance side by the amount that the intake valve closing timing is changed to the advance side by the lift loss Lh of the hydraulic tappet. Therefore, the fuel consumption can be improved by bringing the ignition timing closer to the optimum ignition timing.

  In the first embodiment described above, the time until the flame generated in the combustion chamber 10 reaches the intake port opening in the combustion chamber 10 (hereinafter referred to as “flame arrival time”) T [s] is considered. It wasn't. Considering such flame arrival time T, when the flame arrival time T has elapsed from the ignition timing of the air-fuel mixture by the spark plug 18, the valve lift amount of the intake valve 13 is less than or equal to the flame extinguishing distance Lq between the two walls. There is no risk of backfire. Therefore, when the flame arrival time T is taken into consideration, the advance limit ignition timing is further increased by the crank angle amount Xf corresponding to the flame arrival time T than the crank angle Xg at which the valve lift amount of the intake valve 13 becomes the flame extinguishing distance Lq between the two walls. Can be set to the advance side.

  Here, the flame arrival time T can be calculated by dividing the minimum distance d [m] from the center of the spark plug 18 shown in FIG. 17 to the intake port opening by the flame combustion speed v [m / s]. . The flame combustion speed v increases as the engine speed increases, as shown in FIG. This is because the gas flow of the in-cylinder gas increases as the engine speed increases.

  Therefore, in the present embodiment, correction is made to advance the advance angle limit ignition timing Xg by the crank angle amount Xf corresponding to the flame arrival time T. Thus, the advance limit ignition timing can be further set to the advance side only by the flame arrival time T, that is, the time until the flame generated in the combustion chamber 10 reaches the intake port opening in the combustion chamber 10. As a result, the ignition timing can be made closer to the optimal ignition timing, and fuel efficiency can be improved.

  Hereinafter, the advance limit ignition timing calculation control according to this embodiment will be described.

  FIG. 19 is a flowchart illustrating the advance limit ignition timing calculation control according to this embodiment. The electronic control unit 200 repeatedly executes this routine at a predetermined calculation cycle (for example, 10 [ms]) during engine operation.

  Since the process from step S11 to step S14 is the same as that of the first embodiment, the description thereof is omitted here.

  In step S21, the electronic control unit 200 detects the temperature of the cooling water based on the detection value of the water temperature sensor 216.

  In step S <b> 22, the electronic control unit 200 calculates a correction coefficient K based on the temperature of the cooling water with reference to the table of FIG. 15 created in advance by experiments or the like and stored in the ROM 202.

  In step S23, the electronic control unit 200 calculates the lift loss Lh of the hydraulic tappet based on the engine rotational speed with reference to the table of FIG.

  In step S24, the electronic control unit 200 calculates a correction value Lq ′ for the two-wall extinction distance Lq from the following equation (2) based on the two-wall extinction distance Lq, the correction coefficient K, and the lift loss Lh.

    Lq ′ = Lq × K + Lh (2)

  In step S25, the electronic control unit 200 calculates a crank angle Xg at which the valve lift amount of the intake valve 13 becomes the correction value Lq ′ of the flame extinguishing distance Lq between the two walls.

  In step S <b> 26, the electronic control unit 200 refers to the table of FIG. 18 created in advance by experiments or the like and stored in the ROM 202, and calculates the flame combustion speed v based on the engine speed.

  In step S27, the electronic control unit 200 calculates the flame arrival time T by dividing the minimum distance d from the center of the spark plug 18 to the intake port opening by the flame combustion speed v.

  In step S28, the electronic control unit 200 calculates a crank angle amount Xf corresponding to the flame arrival time T from the following equation (3) based on the flame arrival time T and the engine rotational speed Ne.

    Xf = 360 × Ne / 60 × T (3)

  In step S29, the electronic control unit 200 calculates a crank angle Xg 'obtained by advancing the crank angle Xg from the crank angle Xg by the crank angle amount Xf according to the following equation (4) based on the crank angle Xg and the crank angle amount Xf.

    Xg '= Xg-Xf (4)

  In step S30, the electronic control unit 200 sets the crank angle Xg 'as the advance limit ignition timing.

  The embodiment of the present invention has been described above. However, the above embodiment only shows a part of application examples of the present invention, and the technical scope of the present invention is limited to the specific configuration of the above embodiment. Absent.

  For example, in each of the above embodiments, the so-called direct-acting intake valve operating device 40 that directly opens and closes the intake valve 13 by the intake cam 42 has been described as an example. However, the so-called intake valve 13 is indirectly opened and closed by a rocker arm. It may be a rocker arm type intake valve operating device. In addition, as a valve clearance automatic adjustment mechanism of the rocker arm type intake valve operating device, there is a hydraulic adjuster having a configuration similar to that of a hydraulic tappet, and a lift loss Lh occurs in the hydraulic adjuster as in the hydraulic tappet. Therefore, even when the rocker arm type intake valve operating device is used, the extinguishing distance Lq between the two walls may be corrected according to the lift loss Lh as in the second embodiment.

  In the second embodiment, various corrections are performed according to the temperature of the cooling water, the lift loss Lh, and the flame arrival time T. However, at least one of these corrections may be performed.

B Variable valve timing mechanism 1 Engine body 6 Cylinder
13 Intake valve 13a Valve head (valve head)
18 Spark plug 100 Internal combustion engine 200 Electronic control unit (control device)

Claims (1)

An engine body with cylinders;
A spark plug provided in the engine body for igniting an air-fuel mixture in the cylinder;
A variable valve timing mechanism provided in the engine body and configured to change the closing timing of the intake valve to an arbitrary timing;
An internal combustion engine control device for controlling an internal combustion engine comprising:
When the engine load is in at least a part of the low load region, the ignition timing is a timing when the interval between the valve head of the intake valve and the valve seat is equal to or less than the extinguishing distance, and the intake valve An ignition timing control unit for controlling the spark plug so as to be a timing before the valve closing timing;
Control device for internal combustion engine.

Images