JP2012197770A - Internal combustion engine with variable compression ratio mechanism - Google Patents

Abstract

PROBLEM TO BE SOLVED: To prevent the occurrence of knocking, even when a closing time of an intake valve is later than a setting time of an ignition timing, in an internal combustion engine with a variable compression ratio mechanism.SOLUTION: When the closing time of the intake valve controlled by a variable valve timing mechanism is earlier than the setting time of the ignition timing, the ignition timing of not generating the knocking is set at the setting time (step 105) with respect to a compression edge pressure (step 104) estimated after the closing time of the intake valve. When the closing time of the intake valve is later than the setting time and the occurrence of knocking is estimated at the ignition time that is set at the setting time with respect to the compression edge pressure (step 108) estimated after the closing time of the intake valve, a mechanical compression ratio is decreased by the variable compression ratio mechanism (step 110).

Description

  The present invention relates to an internal combustion engine including a variable compression ratio mechanism.

  An internal combustion engine in which the mechanical compression ratio is variable by a variable compression ratio mechanism is known. In such an internal combustion engine, when the estimated actual compression ratio becomes higher than the target actual compression ratio, it has been proposed to retard the ignition timing to suppress the occurrence of knocking (see Patent Document 1).

JP2005-069130 JP 2006-046193 A

  Certainly, if the actual compression ratio is higher than the target actual compression ratio, knocking is likely to occur at the ignition timing set for the target actual compression ratio. However, whether or not knocking occurs can be more accurately determined based on the estimated in-cylinder pressure at the compression top dead center when the combustion is not performed, that is, the compression end pressure. For example, if the in-cylinder pressure when the intake valve is closed is higher than normal, the compression end pressure will be high even if the target actual compression ratio is realized, and knocking will be performed with respect to the ignition timing determined for the target actual compression ratio May occur.

  Since the compression end pressure changes not only with the actual compression ratio but also with the in-cylinder pressure when the intake valve is closed, it cannot be accurately estimated unless the actual valve closing timing of the intake valve is reached. As a result, in an internal combustion engine that changes the closing timing of the intake valve, if the closing timing of the intake valve is after the ignition timing setting timing, the accurate compression end pressure cannot be estimated at the ignition timing setting timing. In some cases, knocking may occur at the set ignition timing.

  Therefore, an object of the present invention is to make it possible to suppress the occurrence of knocking in an internal combustion engine having a variable compression ratio mechanism even when the intake valve closing timing is after the ignition timing setting timing.

  An internal combustion engine having a variable compression ratio mechanism according to a first aspect of the present invention is an internal combustion engine having a variable compression ratio mechanism, wherein the closing timing of the intake valve controlled by the variable valve timing mechanism is set to the ignition timing. If it is before the timing, an ignition timing that does not cause knocking to the compression end pressure estimated after the closing timing of the intake valve is set at the set timing, and the intake valve controlled by the variable valve timing mechanism is set. When the closing timing is after the set timing of the ignition timing, when knocking is predicted to occur at the ignition timing set at the set timing with respect to the compression end pressure estimated after the closing timing of the intake valve Further, the mechanical compression ratio is lowered by the variable compression ratio mechanism.

  An internal combustion engine having the variable compression ratio mechanism according to claim 2 according to the present invention is an internal combustion engine having the variable compression ratio mechanism according to claim 1, wherein the internal combustion engine directly injects fuel into a cylinder. When the intake valve closing timing controlled by the variable valve timing mechanism is after the set timing of the ignition timing, the compression end pressure estimated after the intake valve closing timing is provided. In the case where it is predicted that knocking will occur at the ignition timing set at the set timing, additional fuel is further supplied by the fuel injection valve.

  According to the internal combustion engine including the variable compression ratio mechanism according to the first aspect of the present invention, when the closing timing of the intake valve controlled by the variable valve timing mechanism is earlier than the set timing of the ignition timing, The occurrence of knocking can be suppressed by setting the ignition timing at which the knocking is not generated for the accurate compression end pressure estimated after the valve closing time at the set time. Also, when the intake valve closing timing controlled by the variable valve timing mechanism is after the ignition timing setting timing, it is set at the setting timing with respect to the accurate compression end pressure estimated after the intake valve closing timing. When knocking is predicted to occur at the ignition timing, the mechanical compression ratio is lowered by the variable compression ratio mechanism. Thereby, an actual compression end pressure can be reduced and the occurrence of knocking can be suppressed. Even when the closing timing of the intake valve controlled by the variable valve timing mechanism is earlier than the setting timing of the ignition timing, at the set timing with respect to the accurate compression end pressure estimated after the closing timing of the intake valve When knocking is predicted to occur at the set ignition timing, the mechanical compression ratio can be reduced by the variable compression ratio mechanism to suppress the occurrence of knocking, but the variable compression ratio can be suppressed compared to ignition timing control. When the mechanism is operated, a large amount of electric energy is consumed. At this time, it is preferable to suppress the occurrence of knocking by controlling the ignition timing.

  According to the internal combustion engine including the variable compression ratio mechanism according to claim 2 of the present invention, in the internal combustion engine including the variable compression ratio mechanism according to claim 1, the internal combustion engine directly injects fuel into the cylinder. The fuel injection valve is provided, and when the intake valve closing timing by the variable valve timing mechanism is after the ignition timing setting timing, the compression end pressure estimated after the intake valve closing timing is set at the setting timing. If knocking is predicted to occur at the set ignition timing, additional fuel is supplied by the fuel injection valve, and the in-cylinder temperature at the ignition timing is reduced by the latent heat of vaporization of the additional fuel. And the occurrence of knocking can be suppressed.

1 is an overall view of an internal combustion engine. It is a disassembled perspective view of a variable compression ratio mechanism. 1 is a schematic side sectional view of an internal combustion engine. It is a figure which shows a variable valve timing mechanism. It is a figure which shows the lift amount of an intake valve and an exhaust valve. It is a figure for demonstrating a mechanical compression ratio, an actual compression ratio, and an expansion ratio. It is a figure which shows the relationship between theoretical thermal efficiency and an expansion ratio. It is a figure for demonstrating a normal cycle and a super-high expansion ratio cycle. It is a figure which shows changes, such as a mechanical compression ratio according to an engine load. It is a flowchart which shows the control for suppressing generation | occurrence | production of knocking.

  FIG. 1 is a side sectional view of an internal combustion engine having a variable compression ratio mechanism according to the present invention. Referring to FIG. 1, 1 is a crankcase, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is a spark plug disposed at the center of the top surface of the combustion chamber 5, and 7 is intake air. 8 is an intake port, 9 is an exhaust valve, and 10 is an exhaust port. The intake port 8 is connected to the surge tank 12 via the intake branch pipe 11. Reference numeral 13 denotes a fuel injection valve for directly injecting fuel into each combustion chamber 5. The fuel injection valve 13 may be attached to each intake branch pipe 11.

  The surge tank 12 is connected to an air cleaner 15 via an intake duct 14, and a throttle valve 17 driven by an actuator 16 and an intake air amount detector 18 using, for example, heat rays are arranged in the intake duct 14. On the other hand, the exhaust port 10 is connected to a catalyst device 20 containing, for example, a three-way catalyst via an exhaust manifold 19, and an air-fuel ratio sensor 21 is disposed in the exhaust manifold 19.

  On the other hand, in the embodiment shown in FIG. 1, the piston 4 is positioned at the compression top dead center by changing the relative position of the crankcase 1 and the cylinder block 2 in the cylinder axial direction at the connecting portion between the crankcase 1 and the cylinder block 2. There is provided a variable compression ratio mechanism A capable of changing the volume of the combustion chamber 5 at the time, and further an actual compression action start timing changing mechanism B capable of changing the actual start time of the compression action. In the embodiment shown in FIG. 1, the actual compression action start timing changing mechanism B is composed of a variable valve timing mechanism capable of controlling the closing timing of the intake valve 7.

  As shown in FIG. 1, a relative position sensor 22 for detecting a relative positional relationship between the crankcase 1 and the cylinder block 2 is attached to the crankcase 1 and the cylinder block 2. An output signal indicating a change in the interval between the crankcase 1 and the cylinder block 2 is output. The variable valve timing mechanism B is provided with a valve timing sensor 23 for generating an output signal indicating the closing timing of the intake valve 7, and an output signal indicating the throttle valve opening is provided to the actuator 16 for driving the throttle valve. A throttle opening sensor 24 is attached.

  The electronic control unit 30 is composed of a digital computer, and is connected to each other by a bidirectional bus 31. A ROM (read only memory) 32, a RAM (random access memory) 33, a CPU (microprocessor) 34, an input port 35 and an output port 36. It comprises. Output signals of the intake air amount detector 18, the air-fuel ratio sensor 21, the relative position sensor 22, the valve timing sensor 23, and the throttle opening sensor 24 are input to the input port 35 via the corresponding AD converters 37. A load sensor 41 that generates an output voltage proportional to the depression amount L of the accelerator pedal 40 is connected to the accelerator pedal 40, and the output voltage of the load sensor 41 is input to the input port 35 via the corresponding AD converter 37. Is done. Further, a crank angle sensor 42 that generates an output pulse every time the crankshaft rotates, for example, 30 ° is connected to the input port 35. On the other hand, the output port 36 is connected to the spark plug 6, the fuel injection valve 13, the throttle valve driving actuator 16, the variable compression ratio mechanism A, and the variable valve timing mechanism B through corresponding drive circuits 38.

  2 is an exploded perspective view of the variable compression ratio mechanism A shown in FIG. 1, and FIG. 3 is a side sectional view of the internal combustion engine schematically shown. Referring to FIG. 2, a plurality of protrusions 50 spaced apart from each other, that is, cylinder block side supports, are formed below both side walls of the cylinder block 2, and each protrusion 50 has a circular cross section. The cam insertion hole 51 is formed. On the other hand, on the upper wall surface of the crankcase 1, there are formed a plurality of protrusions 52 that are fitted between the corresponding protrusions 50 spaced apart from each other, that is, the crankcase side support. A cam insertion hole 53 having a circular cross section is also formed in each protrusion 52.

  As shown in FIG. 2, a pair of camshafts 54, 55 are provided, and on each camshaft 54, 55, a concentric portion 58 is rotatably inserted into each cam insertion hole 53. positioned. Each concentric portion 58 is coaxial with the rotational axis of each camshaft 54, 55. On the other hand, as shown in FIG. 3, eccentric portions 57 that are eccentrically arranged with respect to the rotation axes of the camshafts 54 and 55 are positioned on both sides of each concentric portion 58. A cam 56 is eccentrically mounted for rotation. That is, the eccentric portion 57 is fitted into an eccentric hole formed in the circular cam 56, and the circular cam 56 rotates around the eccentric portion 57 around the eccentric hole. As shown in FIG. 2, the circular cams 56 are disposed on both sides of each concentric portion 58, and the circular cams 56 are rotatably inserted into the corresponding cam insertion holes 51. As shown in FIG. 2, a cam rotation angle sensor 25 that generates an output signal representing the rotation angle of the camshaft 55 is attached to the camshaft 55.

  When the concentric portions 58 of the camshafts 54 and 55 are rotated in opposite directions as shown by arrows in FIG. 3A from the state shown in FIG. 3A, the eccentric portions 57 move in directions away from each other. Therefore, the circular cam 56 rotates in the direction opposite to the concentric portion 58 in the cam insertion hole 51, and the position of the eccentric portion 57 is changed from a high position to an intermediate height position as shown in FIG. Next, when the concentric portion 58 is further rotated in the direction indicated by the arrow, the eccentric portion 57 is at the lowest position as shown in FIG.

  3A, FIG. 3B, and FIG. 3C show the center line a of the concentric portion 58 (that is, the center line of the camshaft) a and the center line b of the eccentric portion 57 in each state. The positional relationship with the center line c of the circular cam 56 is shown.

  3A to 3C, the relative positions of the crankcase 1 and the cylinder block 2 are determined by the distance between the center line a of the concentric part 58 and the center line c of the circular cam 56. The cylinder block 2 moves away from the crankcase 1 as the distance between the center line a of the concentric part 58 and the center line c of the circular cam 56 increases. That is, the variable compression ratio mechanism A changes the relative position between the crankcase 1 and the cylinder block 2 by a crank mechanism using a rotating cam. When the cylinder block 2 moves away from the crankcase 1, the volume of the combustion chamber 5 increases when the piston 4 is positioned at the compression top dead center. Therefore, by rotating the camshafts 54 and 55, the piston 4 is compressed at the top dead center. The volume of the combustion chamber 5 when it is located at can be changed.

  As shown in FIG. 2, in order to rotate the camshafts 54 and 55 in opposite directions, a pair of worms 61 and 62 having opposite spiral directions are attached to the rotation shaft of the drive motor 59, respectively. Worm wheels 63 and 64 that mesh with the worms 61 and 62 are fixed to the ends of the camshafts 54 and 55, respectively. In this embodiment, by driving the drive motor 59, the volume of the combustion chamber 5 when the piston 4 is located at the compression top dead center can be changed over a wide range.

  On the other hand, FIG. 4 shows the variable valve timing mechanism B attached to the end of the camshaft 70 for driving the intake valve 7 in FIG. Referring to FIG. 4, the variable valve timing mechanism B includes a timing pulley 71 that is rotated in the direction of an arrow by a crankshaft of an engine via a timing belt, a cylindrical housing 72 that rotates together with the timing pulley 71, an intake valve A rotating shaft 73 that rotates together with the driving camshaft 70 and is rotatable relative to the cylindrical housing 72, and 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. And 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, and an advance hydraulic chamber 76 on each side of each vane 75. A retarding hydraulic chamber 77 is formed.

  The hydraulic oil supply control to the hydraulic chambers 76 and 77 is performed by a hydraulic oil supply control valve 78. 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 for controlling communication between the ports 79, 80, 82, 83, and 84.

  When the cam phase of the intake valve driving camshaft 70 should be advanced, the spool valve 85 is moved to the right in FIG. 4 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 cam phase of the intake valve driving camshaft 70 should be retarded, the spool valve 85 is moved to the left in FIG. 4, and the hydraulic oil supplied from the supply port 82 causes the hydraulic port 80 to move. The hydraulic oil in the advance hydraulic chamber 76 is discharged from the drain port 83 while being supplied to the retard hydraulic chamber 77. At this time, the rotating shaft 73 is rotated relative to the cylindrical housing 72 in the direction opposite to the arrow.

  If the spool valve 85 is returned to the neutral position shown in FIG. 4 while the rotation shaft 73 is rotated relative to the cylindrical housing 72, the relative rotation operation of the rotation shaft 73 is stopped, and the rotation shaft 73 is The relative rotation position at that time is held. Therefore, the variable valve timing mechanism B can advance and retard the cam phase of the intake valve driving camshaft 70 by a desired amount.

  In FIG. 5, the solid line shows the time when the cam phase of the intake valve driving camshaft 70 is advanced the most by the variable valve timing mechanism B, and the broken line shows the cam phase of the intake valve driving camshaft 70 being the most advanced. It shows when it is retarded. Therefore, the valve opening period of the intake valve 7 can be arbitrarily set between the range indicated by the solid line and the range indicated by the broken line in FIG. 5, and therefore the closing timing of the intake valve 7 is also the range indicated by the arrow C in FIG. Any crank angle can be set.

  The variable valve timing mechanism B shown in FIG. 1 and FIG. 4 shows an example. For example, the variable valve timing that can change only the closing timing of the intake valve while keeping the opening timing of the intake valve constant. Various types of variable valve timing mechanisms, such as mechanisms, can be used.

  Next, the meanings of terms used in the present application will be described with reference to FIG. 6 (A), (B), and (C) show an engine having a combustion chamber volume of 50 ml and a piston stroke volume of 500 ml for the sake of explanation. , (B), (C), the combustion chamber volume represents the volume of the combustion chamber when the piston is located at the compression top dead center.

  FIG. 6A explains the mechanical compression ratio. The mechanical compression ratio is a value mechanically determined only from the stroke volume of the piston and the combustion chamber volume during the compression stroke, and this mechanical compression ratio is expressed by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 6A, this mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11.

  FIG. 6B describes the actual compression ratio. This 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 reaches top dead center, and this actual compression ratio is (combustion chamber volume + actual (Stroke volume) / combustion chamber volume. That is, as shown in FIG. 6B, even if the piston starts to rise in the compression stroke, the compression operation is not performed while the intake valve is open, and the actual compression is performed from the time when the intake valve is closed. The action begins. Therefore, the actual compression ratio is expressed as described above using the actual stroke volume. In the example shown in FIG. 6B, the actual compression ratio is (50 ml + 450 ml) / 50 ml = 10.

  FIG. 6C illustrates the expansion ratio. The expansion ratio is a value determined from the stroke volume of the piston and the combustion chamber volume during the expansion stroke, and this expansion ratio is expressed by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 6C, this expansion ratio is (50 ml + 500 ml) / 50 ml = 11.

  Next, the super expansion ratio cycle used in the present invention will be described with reference to FIGS. FIG. 7 shows the relationship between the theoretical thermal efficiency and the expansion ratio, and FIG. 8 shows a comparison between a normal cycle and an ultrahigh expansion ratio cycle that are selectively used according to the load in the present invention.

  FIG. 8A shows a normal cycle when the intake valve closes near the bottom dead center and the compression action by the piston is started from the vicinity of the intake bottom dead center. In the example shown in FIG. 8A as well, the combustion chamber volume is set to 50 ml, and the stroke volume of the piston is set to 500 ml, similarly to the example shown in FIGS. 6A, 6B, and 6C. As can be seen from FIG. 8A, in a normal cycle, the mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11, the actual compression ratio is almost 11, and the expansion ratio is also (50 ml + 500 ml) / 50 ml = 11. That is, in a normal internal combustion engine, the mechanical compression ratio, the actual compression ratio, and the expansion ratio are substantially equal.

  The solid line in FIG. 7 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, it is only necessary to increase the actual compression ratio. However, the actual compression ratio can only be increased to a maximum of about 12 due to the restriction of the occurrence of knocking at the time of 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 is considered 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 compression ratio is compared to the theoretical thermal efficiency. Was found to have little effect. That is, if the actual compression ratio is increased, the explosive force is increased, but a large amount of energy is required for compression. Thus, even if the actual compression ratio is increased, the theoretical thermal efficiency is hardly increased.

  On the other hand, when the expansion ratio is increased, the period during which the pressing force acts on the piston during the expansion stroke becomes longer, and thus the period during which the piston applies the rotational force to the crankshaft becomes longer. Therefore, the larger the expansion ratio, the higher the theoretical thermal efficiency. The broken line ε = 10 in FIG. 7 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 the actual compression ratio ε is maintained at a low value, and the theoretical thermal efficiency when the actual compression ratio is increased with the expansion ratio as shown by the solid line in FIG. 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 be greatly increased. FIG. 8B shows an example where the variable compression ratio mechanism A and variable valve timing mechanism B are used to increase the expansion ratio while maintaining the actual compression ratio at a low value.

  Referring to FIG. 8B, in this example, the variable compression ratio mechanism A reduces the combustion chamber volume from 50 ml to 20 ml. On the other hand, the variable valve timing mechanism B delays the closing timing of the intake valve until the actual piston stroke volume is reduced from 500 ml to 200 ml. As a result, in this example, the actual compression ratio is (20 ml + 200 ml) / 20 ml = 11, and the expansion ratio is (20 ml + 500 ml) / 20 ml = 26. In the normal cycle shown in FIG. 8A, the actual compression ratio is almost 11 and the expansion ratio is 11, as described above. Compared with this case, only the expansion ratio is shown in FIG. 8B. It can be seen that it has been increased to 26. This is why it is called an ultra-high expansion ratio cycle.

  Generally speaking, in an internal combustion engine, the lower the engine load, the worse the thermal efficiency. Therefore, in order to improve the thermal efficiency during engine operation, that is, to improve fuel efficiency, it is necessary to improve the thermal efficiency when the engine load is low. Necessary. On the other hand, in the ultra-high expansion ratio cycle shown in FIG. 8B, since the actual piston stroke volume during the compression stroke is reduced, the amount of intake air that can be sucked into the combustion chamber 5 is reduced. The expansion ratio cycle can only be adopted when the engine load is relatively low. Therefore, in the present invention, when the engine load is relatively low, the super high expansion ratio cycle shown in FIG. 8B is used, and during the high engine load operation, the normal cycle shown in FIG. 8A is used.

Next, the overall operation control will be schematically described with reference to FIG. FIG. 9 shows changes in the intake air amount, the intake valve closing timing, the mechanical compression ratio, the expansion ratio, the actual compression ratio, and the opening degree of the throttle valve 17 according to the engine load at a certain engine speed. . 9 shows that the average air-fuel ratio in the combustion chamber 5 is the output signal of the air-fuel ratio sensor 21 so that unburned HC, CO and NO _x in the exhaust gas can be simultaneously reduced by the three-way catalyst in the catalyst device 20. This shows a case where feedback control is performed to the stoichiometric air-fuel ratio based on the above.

  As described above, the normal cycle shown in FIG. 8 (A) is executed during engine high load operation. Accordingly, as shown in FIG. 9, the expansion ratio is low because the mechanical compression ratio is lowered at this time, and the valve closing timing of the intake valve 7 is advanced as shown by the solid line in FIG. ing. At this time, the amount of intake air is large, and at this time, the opening degree of the throttle valve 17 is kept fully open, so that the pumping loss is zero.

  On the other hand, as shown by the solid line in FIG. 9, when the engine load becomes low, the closing timing of the intake valve 7 is delayed in order to reduce the intake air amount. Further, at this time, as shown in FIG. 9, the mechanical compression ratio is increased as the engine load is lowered so that the actual compression ratio is kept substantially constant. Therefore, the expansion ratio is also increased as the engine load is lowered. At this time, the throttle valve 17 is kept fully open, and therefore the amount of intake air supplied into the combustion chamber 5 is controlled by changing the closing timing of the intake valve 7 regardless of the throttle valve 17. Has been.

  As described above, when the engine load is reduced from the engine high load operation state, the mechanical compression ratio is increased as the intake air amount is decreased while the actual compression ratio is substantially constant. That is, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center is decreased in proportion to the decrease in the intake air amount. Therefore, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center changes in proportion to the intake air amount. At this time, in the example shown in FIG. 9, the air-fuel ratio in the combustion chamber 5 is the stoichiometric air-fuel ratio, so the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center is proportional to the fuel amount. Will change.

  When the engine load is further reduced, the mechanical compression ratio is further increased, and when the engine load is lowered to the medium load L1 slightly close to the low load, the mechanical compression ratio becomes a limit mechanical compression ratio (upper limit mechanical compression) that becomes the structural limit of the combustion chamber 5 Ratio). When the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is held at the limit mechanical compression ratio in a region where the load is lower than the engine load L1 when the mechanical compression ratio reaches the limit mechanical compression ratio. Accordingly, the mechanical compression ratio is maximized and the expansion ratio is maximized at the time of low engine load operation and low engine load operation, that is, at the engine low load operation side. In other words, the mechanical compression ratio is maximized so that the maximum expansion ratio is obtained on the engine low load operation side.

  On the other hand, in the embodiment shown in FIG. 9, when the engine load decreases to L1, the closing timing of the intake valve 7 becomes the limit closing timing that can control the amount of intake air supplied into the combustion chamber 5. When the closing timing of the intake valve 7 reaches the limit closing timing, the closing timing of the intake valve 7 is reduced in a region where the load is lower than the engine load L1 when the closing timing of the intake valve 7 reaches the closing timing. It is held at the limit closing timing.

  When the closing timing of the intake valve 7 is held at the limit closing timing, the amount of intake air can no longer be controlled by changing the closing timing of the intake valve 7. In the embodiment shown in FIG. 9, at this time, that is, in a region where the load is lower than the engine load L1 when the valve closing timing of the intake valve 7 reaches the limit valve closing timing, the intake valve 7 is supplied into the combustion chamber 5 by the throttle valve 17. The amount of intake air to be controlled is controlled, and the opening degree of the throttle valve 17 is made smaller as the engine load becomes lower. In this embodiment, the intake air amount is controlled even when the closing timing of the intake valve 7 is changed as shown by the solid line in FIG.

  As described above, the expansion ratio is 26 in the ultra-high expansion ratio cycle shown in FIG. The higher the expansion ratio, the better. However, as can be seen from FIG. 7, a considerably high theoretical thermal efficiency can be obtained if it is 20 or more with respect to the practically usable lower limit actual compression ratio ε = 5. Therefore, in this embodiment, the variable compression ratio mechanism A is formed so that the expansion ratio is 20 or more.

  By the way, in the internal combustion engine, it is necessary to suppress the occurrence of knocking. For this reason, in the internal combustion engine of the present embodiment, the electronic control unit 30 performs the control of the flowchart shown in FIG. First, in step 101, as described above, the intake valve closing timing IVC controlled by the variable valve timing mechanism B is set based on the current engine load. In step 102, it is determined whether or not the set intake valve closing timing IVC is before the ignition timing setting timing ITT. Here, the ignition timing setting timing ITT is set to a predetermined crank angle in the middle of the compression stroke because a predetermined energization time to the coil of the spark plug 6 is necessary before the actual ignition timing. The

If the determination in step 102 is affirmative, the intake valve is closed before the ignition timing setting timing ITT. In step 103, the in-cylinder pressure disposed in each cylinder at the closing timing of the intake valve. An intake pressure Pm in a cylinder (corresponding cylinder) in which the intake valve is closed is measured by a sensor (not shown) (or an intake valve closed of a pressure sensor (not shown) arranged in the surge tank 12). The measured value at that time may be the intake pressure Pm when the intake valve of the corresponding cylinder is closed). Next, at step 104, the compression end pressure Pe of the corresponding cylinder is estimated by the following equation.
Pe = Pm (Vm / Ve) ^k
Here, Vm is the combustion chamber volume when the intake valve is closed (combustion chamber volume at compression top dead center + actual stroke volume), Ve is the combustion chamber volume at compression top dead center, and k is the specific heat ratio. is there. (Vm / Ve) is the actual compression ratio described above. The combustion chamber volume at the compression top dead center can be accurately estimated based on the output of the relative position sensor 22, and the actual stroke volume can be accurately estimated based on the actual intake valve closing timing. The compression end pressure Pe of the corresponding cylinder can be accurately estimated at the intake valve closing timing or thereafter.

  Next, at step 105, when the ignition timing setting timing ITT is reached, the ignition timing IT that generates the highest torque without setting knocking to the compression end pressure Pe of the corresponding cylinder estimated at step 104 is set. If ignition is performed at the ignition timing IT thus set, the occurrence of knocking in the corresponding cylinder can be sufficiently suppressed.

  On the other hand, if the intake valve closing timing IVC is after the ignition timing setting timing ITT, the determination in step 102 is denied, and in step 106, the ignition timing setting timing ITT is used as the ignition timing IT. A predetermined optimal ignition timing (MBT) is set for the engine operating state (engine load, engine speed, target mechanical compression ratio, intake valve closing timing IVC, etc.). Here, in step 105 described above, when the compression end pressure Pe estimated in step 104 is higher than the compression end pressure assumed in the current engine operation state, the optimal ignition timing (MBT) in the current engine operation state is determined. ) May be corrected to the retard side.

  In step 107, when the intake valve closing timing IVC is reached, the intake pressure Pm of the corresponding cylinder when the intake valve is closed is measured in the same manner as in step 103, and then the compression end pressure Pe of the corresponding cylinder is determined in the same manner as in step 104. Presumed. Next, at step 109, it is predicted whether or not knocking will occur at the optimal ignition timing set at the setting timing ITT with respect to the accurate compression end pressure Pe estimated at the closing timing of the intake valve or thereafter. . For example, when the intake pressure Pm when the intake valve is closed is higher than the pressure assumed in the current engine operation state, the combustion chamber volume at the compression top dead center is smaller than the volume assumed in the current engine operation state Or, when the valve closing timing of the intake valve is on the advance side of the timing assumed in the current engine operating state, the estimated compression end pressure Pe is higher than the pressure assumed in the current engine operating state. Become. When the estimated compression end pressure Pe is higher than the set value by the compression end pressure assumed in the current engine operating state, it is predicted that knocking will occur at the optimal ignition timing set for the current engine operating state. can do.

  If the determination in step 109 is negative, the process ends as it is. However, if it is predicted that knocking will occur, in step 110, the variable compression ratio mechanism A is immediately activated to lower the mechanical compression ratio. Thereby, in the corresponding cylinder, the estimated compression end pressure can be reduced, and the occurrence of knocking can be suppressed. Here, it is preferable to lower the mechanical compression ratio until the estimated compression end pressure becomes the compression end pressure assumed in the current engine operating state by the ignition timing IT set in step 106, but the ignition timing IT If the mechanical compression ratio can be reduced even slightly, the estimated compression end pressure can be reduced, and the occurrence of knocking can be suppressed.

  Further, when the mechanical compression ratio cannot be reduced until the estimated compression end pressure becomes the compression end pressure assumed in the current engine operating state by the ignition timing IT, the mechanical compression is also performed after the ignition timing IT. It is preferable to keep the ratio down. Thereby, the in-cylinder pressure can be reduced even after the start of combustion, and the occurrence of knocking can be suppressed. Further, if it is during combustion, it is preferable to reduce the in-cylinder pressure by reducing the mechanical compression ratio even after compression top dead center, so that the occurrence of knocking can be suppressed and the occurrence of knocking Even so, the duration of knocking can be suppressed.

  Thus, when it is predicted that knocking will occur, it is preferable to start the reduction of the mechanical compression ratio as early as possible, but even after the ignition timing, if it is in combustion, it will be after the compression top dead center. Even if it starts, knocking can be suppressed if the reduction of the mechanical compression ratio is started.

  Next, in the cylinder where the intake valve is closed, if the mechanical compression ratio is sufficiently lowered, the combustion chamber volume at the compression top dead center is increased, and the compression end pressure estimated when the intake valve is closed is In step 109, the occurrence of knocking is not predicted because the compression end pressure is assumed to be lowered in the current engine operating state.

  Further, if the fuel injection valve 13 directly injects fuel into the cylinder, when the determination in step 109 is affirmative, additional fuel is immediately handled by the fuel injection valve 13 as in step 111. You may make it supply in a cylinder. Thereby, the in-cylinder temperature at the ignition timing can be lowered by the vaporization latent heat of the additional fuel, and the occurrence of knocking can be suppressed.

  The supply of additional fuel when knocking is predicted to occur is preferably started as early as possible, but even after the ignition timing, or even after compression top dead center during combustion, The in-cylinder temperature can be reduced to suppress knocking, and even if knocking occurs, the duration of knocking can be suppressed.

  In this embodiment, even when the intake valve closing timing IVC is earlier than the ignition timing setting timing ITT, the current engine is used for the accurate compression end pressure Pe estimated after the intake valve closing timing. When knocking is predicted to occur at the optimum ignition timing in the operating state, it is possible to reduce the mechanical compression ratio by the variable compression ratio mechanism A and suppress the occurrence of knocking as in step 110. When the variable compression ratio mechanism A is operated as compared with the above, a large amount of electric energy is consumed. At this time, it is preferable to suppress the occurrence of knocking by the ignition timing control in step 105.

DESCRIPTION OF SYMBOLS 1 Crankcase 2 Cylinder block 6 Spark plug 13 Fuel injection valve A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (2)

  An internal combustion engine having a variable compression ratio mechanism, and when the closing timing of the intake valve controlled by the variable valve timing mechanism is before the ignition timing setting timing, the compression estimated after the closing timing of the intake valve An ignition timing that does not cause knocking with respect to the end pressure is set at the set timing, and when the intake valve closing timing controlled by the variable valve timing mechanism is after the ignition timing set timing, the intake valve closing is not performed. The variable compression ratio mechanism reduces the mechanical compression ratio when knocking is predicted to occur at the ignition timing set at the set timing with respect to the compression end pressure estimated after the valve timing. An internal combustion engine provided with a variable compression ratio mechanism.   The internal combustion engine includes a fuel injection valve that directly injects fuel into the cylinder, and the intake valve controlled by the variable valve timing mechanism is inhaled when the closing timing of the intake valve is after the set timing of the ignition timing. When knocking is predicted to occur at the ignition timing set at the set timing with respect to the compression end pressure estimated after the valve closing timing, additional fuel is supplied by the fuel injection valve. An internal combustion engine comprising the variable compression ratio mechanism according to claim 1.

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