JP2011226388A - Spark ignition type internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To properly control mechanical compression ratio, even under fuel cut control.SOLUTION: A spark ignition type internal combustion engine includes a variable compression ratio mechanism A and a variable valve timing mechanism B. A required operating point where an operating point exhibiting the combination of the mechanical compression ratio and an intake valve close timing is to reach in each engine operation condition is set. A target operating point is computed in each fixed time based on the prediction amount of the mechanical compression ratio changeable in a fixed time and the prediction amount of the intake valve close timing changeable in a fixed time so that an operating point when the engine operation condition is changed is changed toward the required operating point set to the engine operation state after the change to change the mechanical compression ratio and the intake valve close timing toward the target operating point. Under the fuel cut control, the prediction amount of the mechanical compression ratio changeable in a fixed time is modified according to at least one of intake air amount and an existing real compression ratio.

Description

  The present invention relates to a spark ignition internal combustion engine.

  A spark having a variable compression ratio mechanism capable of changing the mechanical compression ratio and a variable valve timing mechanism capable of controlling the closing timing of the intake valve so that the actual compression ratio is maintained substantially constant regardless of the engine load. Ignition internal combustion engines are known (see, for example, Patent Document 1). In this internal combustion engine, as the engine load increases, that is, as the required intake air amount increases, the closing timing of the intake valve is advanced so as to approach the intake bottom dead center, and at this time, the actual compression ratio is maintained substantially constant. Therefore, the mechanical compression ratio is lowered as the required intake air amount increases.

  As the variable compression ratio mechanism, for example, a cylinder block and a crankcase are connected so as to be relatively movable, a camshaft is provided at the connecting portion, and the mechanical compression ratio is changed by rotating the camshaft. Is known (see, for example, Patent Document 2).

  In particular, in this variable compression ratio mechanism, the relationship between the rotation angle of the drive motor that drives the camshaft and the mechanical compression ratio is not linear, and the mechanical compression ratio with respect to changes in the rotation angle of the drive motor is particularly low. The rate of change is small. For this reason, in the low mechanical compression region, the amount of the mechanical compression ratio that can be changed in a certain time is small. Therefore, in the internal combustion engine described in Patent Document 2, feedback control is performed so that the mechanical compression ratio becomes the target mechanical compression ratio based on the output of the detection device that detects the mechanical compression ratio, and feedback is performed in the low mechanical compression ratio region. By increasing the control gain in the control, the amount of the mechanical compression ratio that can be changed in a certain time even in the low mechanical compression ratio region is increased.

JP 2009-115060 A JP 2007-303423 A

  By the way, as a method of changing the intake valve closing timing and the mechanical compression ratio in accordance with the change in the required intake air amount in this way, the target intake valve closing timing and the target mechanical compression ratio are calculated at regular time intervals. Further, there is a method of controlling the intake valve closing timing and the mechanical compression ratio so as to be the target intake valve closing timing and the target mechanical compression ratio.

  In such a method, when calculating the target intake valve closing timing and the target mechanical compression ratio at regular time intervals, the amount of intake valve closing timing that can be changed during this fixed time (hereinafter referred to as “the valve timing adjustable amount”). )) And the amount of mechanical compression ratio that can be changed in a certain time (hereinafter referred to as “changeable amount of compression ratio”). In other words, the calculated target intake valve closing timing needs to be an intake valve closing timing that is within the above-mentioned changeable amount of the closing timing from the current intake valve closing timing, and the calculated mechanical compression ratio is the current It is necessary that the mechanical compression ratio is within the compression ratio changeable amount from the mechanical compression ratio.

  Here, the compression ratio changeable amount changes due to various factors. For example, when the variable compression ratio mechanism described in Patent Document 2 is used, the compression ratio changeable amount changes according to the mechanical compression ratio as described above. Further, the compression ratio changeable amount also changes depending on the engine load (or the output torque of the internal combustion engine). That is, the combustion pressure generated by the combustion of the air-fuel mixture in the engine combustion chamber increases as the engine load increases. However, such combustion pressure acts in the direction of lowering the mechanical compression ratio. Therefore, the higher the engine load, the greater the resistance to increasing the mechanical compression ratio, and the lower the resistance to decreasing the mechanical compression ratio. As a result, when the mechanical compression ratio is increased, the compression ratio changeable amount decreases as the engine load increases. On the other hand, when the mechanical compression ratio decreases, the compression ratio changeable amount increases as the engine load increases. .

  On the other hand, during the fuel cut control for stopping fuel injection from the fuel injection valve, the engine load is zero, and the compression ratio changeable amount does not change according to the engine load. On the other hand, during the fuel cut control, the compression ratio changeable amount changes according to the actual compression ratio and the intake air amount. Therefore, if the compression ratio changeable amount is calculated based only on the mechanical compression ratio and the engine load, the compression ratio changeable amount cannot be properly calculated during the fuel cut control. The compression ratio cannot be properly controlled.

  Therefore, in view of the above problems, the present invention can appropriately control the mechanical compression ratio even during fuel cut control in a spark ignition internal combustion engine having a variable compression ratio mechanism and a variable valve timing mechanism. The purpose is to do so.

In order to solve the above-described problem, the first invention includes a variable compression ratio mechanism capable of changing the mechanical compression ratio and a variable valve timing mechanism capable of controlling the closing timing of the intake valve, and each engine operating state is provided. Set the required operating point that should reach the operating point indicating the combination of mechanical compression ratio and intake valve closing timing, and indicate the combination of mechanical compression ratio and intake valve closing timing when the engine operating state changes The predicted amount of mechanical compression ratio that can be changed in a fixed time so that the operating point changes toward the required operating point set for the engine operating state after the change, and the intake valve closing timing that can be changed in a fixed time Fuel supply from a fuel injection valve in a spark ignition type internal combustion engine that calculates a target operating point at regular intervals based on the predicted amount and changes the mechanical compression ratio and the intake valve closing timing toward the target operating point During fuel cut control to stop It was so changed depending on at least one of the predicted amount of current intake air amount and the current actual compression ratio changeable mechanical compression ratio on time.
During fuel cut control, the amount of the mechanical compression ratio that can be changed in a fixed time differs from that during normal operation in accordance with the intake air amount and the actual compression ratio, not the engine load. According to the first invention, the predicted amount of the mechanical compression ratio that can be changed in a certain time is changed according to the current intake air amount or the current actual compression ratio. The mechanical compression ratio that can be changed in a certain time can be calculated appropriately.

  In the second invention, in the first invention, the predicted amount of the mechanical compression ratio that can be changed over a certain period of time is reduced as the current intake air amount increases when the mechanical compression ratio is increased, and the predicted mechanical compression ratio is decreased. When doing so, the larger the current intake air amount, the larger the value.

  In the third invention, in the first or second invention, the predicted amount of the mechanical compression ratio that can be changed over a certain period of time is made smaller as the current actual compression ratio becomes higher when the mechanical compression ratio is made higher. When the ratio is lowered, it is increased as the current actual compression ratio is higher.

  In a fourth invention, in any one of the first to third inventions, an entry prohibition region is set for a combination of the mechanical compression ratio and the intake valve closing timing, and the target operating point is an engine When the operating state changes, the operating point indicating the combination of the mechanical compression ratio and the intake valve closing timing enters the intrusion prohibited area toward the required operating point set for the engine operating state after the change. It is calculated based on the predicted amount of the mechanical compression ratio that can be changed at a certain time so as to change without change, and the predicted amount of the intake valve closing timing that can be changed at a certain time.

  According to a fifth aspect, in the fourth aspect, the target operating point is a predicted amount of mechanical compression ratio that can be changed from a current operating point or a previous target operating point to a predetermined time, and an intake valve that can be changed to a predetermined time. Among the operating points located within the predicted amount of the valve closing timing, the operating point is the farthest from the current operating point in a range that does not enter the intrusion prohibited area.

  According to a sixth aspect of the invention, in any one of the first to fifth aspects, during the normal operation in which the fuel cut control is not performed, the predicted amount of the mechanical compression ratio that can be changed at the predetermined time is set to the engine load. Changed according to.

  In a seventh invention, in any one of the first to sixth inventions, the variable compression ratio mechanism is configured to change a mechanical compression ratio by changing a relative position between a crankcase and a cylinder block by a crank mechanism using a rotating cam. And the predicted amount of the mechanical compression ratio that can be changed in the predetermined time is changed according to the rotation angle of the cam.

  According to the present invention, the mechanical compression ratio can be appropriately controlled even during fuel cut control.

1 is an overall view of a spark ignition 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 figure which shows an intrusion prohibition area | region and a reference | standard operation line. It is a figure which shows the relationship between a throttle opening and intake air amount. It is a figure which shows an intrusion prohibition area | region and a reference | standard operation line. It is a figure which shows an intrusion prohibition area | region. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a figure which shows the change of mechanical compression ratio, intake valve closing timing, and throttle opening. It is a figure which shows the changeable amount of the mechanical compression ratio in a fixed time. It is a figure which shows the changeable amount of the mechanical compression ratio in a fixed time. It is a figure which shows the changeable amount of the mechanical compression ratio in a fixed time. It is a figure which shows the changeable amount of the mechanical compression ratio in a fixed time. It is a figure which shows the change of mechanical compression ratio, intake valve closing timing, and throttle opening. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a time chart which shows changes, such as a mechanical compression ratio, intake valve closing timing, and throttle opening. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a figure which shows a target operating point and an operating point. It is a time chart which shows changes, such as a mechanical compression ratio, intake valve closing timing, and throttle opening. It is a flowchart for calculating a target value. It is a flowchart for performing drive control of a variable compression ratio mechanism or the like. FIG. 36 is an enlarged view of a part of the intake air amount plane shown in FIG. 35. It is a figure which shows the changeable amount of the mechanical compression ratio in a fixed time. It is a figure which shows the changeable amount of the mechanical compression ratio in a fixed time. It is a figure which shows the changeable amount of the mechanical compression ratio in a fixed time. It is a figure which shows the changeable amount of the mechanical compression ratio in a fixed time. It is a flowchart for calculating the predicted value of the compression ratio changeable amount. It is a figure which shows the map used with the flowchart shown in FIG. It is a figure which shows the map used with the flowchart shown in FIG.

  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.

FIG. 1 shows a side sectional view of a spark ignition type internal combustion engine.
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 an 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 a surge tank 12 via an intake branch pipe 11, and a fuel injection valve 13 for injecting fuel into the corresponding intake port 8 is arranged in each intake branch pipe 11. The fuel injection valve 13 may be arranged in each combustion chamber 5 instead of being 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 air flow meter 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 catalytic converter 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. In the following description, portions of the intake branch pipe 11, the surge tank 12, the intake duct 14 and the like from the throttle valve 17 to the intake valve 7 are referred to as an intake pipe portion.

  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 axis 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 operation 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. In the present embodiment, the relative position sensor 22 is used as a mechanical compression ratio detection device for detecting the current mechanical compression ratio, but a detection device other than the relative position sensor 22 is used as the mechanical compression ratio detection device. It is also possible.

  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 air flow meter 18, the air-fuel ratio sensor 21, the relative position sensor 22, the valve timing sensor 23, and the throttle opening degree sensor 24 are input to the input port 35 via the corresponding AD converters 37, respectively. 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 projecting portions 50 spaced from each other are formed below both side walls of the cylinder block 2, and cam insertion holes 51 each having a circular cross section are formed in each projecting portion 50. Has been. On the other hand, a plurality of protrusions 52 are formed on the upper wall surface of the crankcase 1 so as to be fitted between the corresponding protrusions 50 with a space between each other. Cam insertion holes 53 each having a circular cross section are formed.

  As shown in FIG. 2, a pair of camshafts 54 and 55 are provided, and on each camshaft 54 and 55, a circular cam 58 is rotatably inserted into each cam insertion hole 53. It is fixed. These circular cams 58 are coaxial with the rotational axes of the camshafts 54 and 55. On the other hand, as shown in FIG. 3, eccentric shafts 57 arranged eccentrically with respect to the rotational axes of the cam shafts 54 and 55 extend on both sides of each circular cam 58, and another circular cam is disposed on the eccentric shaft 57. 56 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. As shown in FIG. 2, a cam rotation angle sensor 25 that generates an output signal indicating the rotation angle of the camshaft 55 is attached to the camshaft 55.

  When the circular cams 58 fixed on the camshafts 54 and 55 are rotated in opposite directions as indicated by arrows in FIG. 3A from the state shown in FIG. The circular cam 56 rotates in the direction opposite to the circular cam 58 in the cam insertion hole 51 in order to move in the away direction, and as shown in FIG. 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, 3B, and 3C show the positional relationship among 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. It 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 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 1. 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 located 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 meshing with the worms 61 and 62 are fixed to end portions 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 can rotate 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 vanes 75 extending from the outer peripheral surface of the rotary shaft 73 to the inner peripheral surface of the cylindrical housing 72 between the partition walls 74, and advance hydraulic chambers 76 on both sides 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 that performs communication cutoff control 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 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. 6A, 6B, and 6C 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 action is not performed while the intake valve is open, and the actual compression is started from the time when the intake valve 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. 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, the combustion chamber volume is 50 ml and the piston stroke volume is 500 ml as in the examples 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 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 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 to 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 engine high 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 catalytic converter 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. 8A is executed during engine high load operation. Therefore, as shown in FIG. 9, the expansion ratio is low because the mechanical compression ratio is lowered at this time. As shown by the solid line in FIG. 9, the 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 without depending on 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 reduced in proportion to the reduction 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, since the air-fuel ratio in the combustion chamber 5 is the stoichiometric air-fuel ratio, 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 L ₁ slightly close to the low load, the mechanical compression ratio reaches the maximum limit mechanical compression ratio that is the structural limit of the combustion chamber 5. . When the mechanical compression ratio reaches the maximum critical mechanical compression ratio, the mechanical compression ratio is maintained at the maximum critical mechanical compression ratio in a region where the load is lower than the engine load L ₁ when the mechanical compression ratio reaches the maximum critical mechanical compression ratio. The Therefore, 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, on 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 L ₁ , 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 in a region where the load is lower than the engine load L ₁ when the closing timing of the intake valve 7 reaches the closing timing. Is held at the limit valve 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 L ₁ when the closing timing of the intake valve 7 reaches the limit closing timing, the throttle valve 17 supplies the fuel into the combustion chamber 5. 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.

On the other hand, as shown by the broken line in FIG. 9, the intake air amount can be controlled without depending on the throttle valve 17 by advancing the closing timing of the intake valve 7 as the engine load becomes lower. Therefore, if expressed in such a manner that both the case indicated by the solid line and the case indicated by the broken line in FIG. 9 can be included, in the embodiment according to the present invention, the valve closing timing of the intake valve 7 increases as the engine load decreases. It is moved in the direction away from the intake bottom dead center BDC until the limit valve closing timing L ₁ that can control the amount of intake air supplied into the combustion chamber. As described above, the intake air amount can be controlled by changing the closing timing of the intake valve 7 as shown by the solid line in FIG. 9 or by changing it as shown by the broken line. Hereinafter, the present invention will be described by taking as an example a case where the closing timing of the intake valve 7 is changed as shown by a solid line in FIG.

  Incidentally, 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 the present invention, the variable compression ratio mechanism A is formed so that the expansion ratio is 20 or more.

  Next, the reference prohibition line for the intrusion prohibited area, the mechanical compression ratio, and the intake valve closing timing will be described with reference to FIGS.

  FIG. 10 shows the intake air amount necessary to obtain the required engine load, that is, the required intake air amount, the mechanical compression ratio, and the intake valve closing timing. In FIG. 10, the required intake air amount increases as the distance from the origin 0 increases, and the mechanical compression ratio increases as the distance from the origin 0 increases. Further, in FIG. 10, the intake valve closing timing is represented by the crank angle after the intake bottom dead center (ABDC). Therefore, the intake valve closing timing is retarded as the distance from the origin 0 is increased.

On the other hand, in FIG. 10, Q ₁ , Q ₂ , Q ₃ , Q ₄ , and Q ₅ each represent the same intake air amount surface. Further, θ _max represents the throttle fully open surface where the throttle valve 17 is fully open, and as can be seen from FIG. 10, this throttle fully open surface θ _max is formed of a curved surface convex upward. In the region below the throttle fully open surface θ _max, the throttle opening decreases as it goes downward.

This is shown in FIG. Curved surfaces θ ₁ and θ ₂ in FIG. 11 are the same throttle opening surfaces when the throttle openings are θ ₁ and θ ₂ , respectively. As can be seen from FIG. 11, the throttle opening surfaces θ ₁ and θ ₂ consists of an upwardly convex curved surface. The relationship between the throttle openings θ _max , θ ₂ , θ ₁ is θ _max > θ ₂ > θ _1, and the smaller the throttle opening, the intake air at the same mechanical compression ratio and the same intake valve closing timing. The amount is reduced.

In FIG. 10, the hatched areas indicate intrusion prohibited areas in the same intake air quantity surfaces Q ₁ , Q ₂ , Q ₃ , Q ₄ , and Q ₅ . On the other hand, FIG. 12 shows a view from the top of FIG. 10, FIG. 13A shows a view of the left side surface S ₁ in FIG. 10 from the direction of the arrow, and FIG. the right side surface S ₂ shows the place viewed from the arrow direction, these 12 and FIG. 13 (a), the shows areas forbidden entry area shown by hatching also in (B).

10, 12, and 13 (A) and (B), the intrusion prohibition area is three-dimensionally expanded. Further, the intrusion prohibition area includes a high load side area X ₁ and a low load side area X ₂ . It can be seen that it consists of two regions. Incidentally, FIGS. 10, 12, FIG. 13 (A), the mechanical compression ratio at high load forbidden entry area X ₁ As can be seen from (B) is a number required intake air amount, the intake valve closing timing is advanced side is formed on the high side, low-load forbidden entry area X ₂ is less required intake air amount, the mechanical compression ratio is formed on a lower side in the intake valve closing timing is retarded side.

FIG. 9 shows the relationship between the intake valve closing timing, the mechanical compression ratio, the actual compression ratio, and the throttle opening, at which the minimum fuel consumption can be obtained with respect to the required intake air amount. This is indicated by a solid line W in FIGS. This line W As can be seen from FIG. 10 on the side is larger intake air amount than the same intake air amount plane Q ₃ extends over the full throttle surface theta _max, the amount of intake air than the same intake air amount plane Q ₃ less the side extends over the right side S _2. The same intake air amount surface Q ₃ corresponds to the load L ₁ in FIG.

That is, in the region where the engine load is higher than L ₁ in FIG. 9, the intake valve closing timing is advanced with the throttle valve 17 being held fully open as the engine load increases, that is, as the required intake air amount increases. At this time, the mechanical compression ratio is lowered as the required intake air amount increases so that the actual compression ratio becomes constant. The relationship between the mechanical compression ratio and the intake valve closing timing at this time is represented by a line W on the throttle fully open surface θ _max of FIG. That is, as shown in FIG. 10, on the side where the intake air amount is larger than the same intake air amount surface Q ₃ , the intake valve closing timing is set while the throttle valve 17 is held fully open as the required intake air amount increases. At this time, the mechanical compression ratio is lowered as the required intake air amount increases so that the actual compression ratio becomes constant.

On the other hand, in the region where the engine load is lower than L ₁ in FIG. 9, the mechanical compression ratio and the intake valve closing timing are kept constant, and the throttle valve 17 opens as the engine load decreases, that is, the required intake air amount decreases. The degree is reduced. Relationship, the mechanical compression ratio and the intake valve closing timing at this time is represented by a line W on the right side S ₂ of Figure 10. That is, as shown in FIG. 10, on the side where the intake air amount is smaller than the same intake air amount surface Q ₃ , the mechanical compression ratio and the intake valve closing timing are kept constant, and the lower the engine load, that is, the required intake air. As the amount decreases, the opening degree of the throttle valve 17 is decreased.

  In the present specification, a line that the mechanical compression ratio and the intake valve closing timing follow when the required intake air amount changes is referred to as an operation line, and in particular, the line W illustrated in FIG. 10 is referred to as a reference operation line. ing. Note that, as described above, this reference operation line indicates the minimum fuel consumption operation line with which the minimum fuel consumption can be obtained.

As described above, the actual compression ratio is constant on the reference operation line W. Since the actual compression ratio is independent of the opening degree of the throttle valve 17 and is determined only by the mechanical compression ratio and the intake valve closing timing, the actual compression ratio is the same on the curved surface extending in the vertical direction through the reference operation line W in FIG. Become. In this case, the actual compression ratio becomes higher on the side where the mechanical compression ratio is higher than the curved surface, and the actual compression ratio becomes lower on the side where the mechanical compression ratio is lower than the curved surface. That is, roughly speaking, the high load side intrusion prohibited region X ₁ is located in a region where the actual compression ratio is higher than the actual compression ratio on the reference operation line W, and the low load side intrusion prohibited region X ₂ is the reference operation. It is located in a region where the actual compression ratio is lower than the actual compression ratio on the line W.

If the actual compression ratio is increased to improve fuel consumption, knocking occurs. If the ignition timing is retarded to prevent knocking, combustion becomes unstable and torque fluctuation occurs. High load forbidden entry area X ₁ is operating regions produce such torque variations, hence at the time of engine operation the operating state of the engine is required to prevent from entering the operating area of produce such torque fluctuations. On the other hand, when the amount of intake air is small and the actual compression ratio is low, combustion is difficult, and when the opening of the throttle valve 17 is small and the compression end pressure is low, combustion is deteriorated and torque fluctuation occurs. The low load side forbidden entry area X ₂ are operating region resulting such torque variations at the time of engine operation therefore it is necessary that the operating state of the engine in this operating region from entering.

On the other hand, the higher the actual compression ratio, the better the fuel consumption. Therefore, the minimum fuel consumption operation line for obtaining the minimum fuel consumption without causing knocking or torque fluctuation is prohibited from entering the high load side as indicated by W in FIGS. extends while along the outer edge of the high-load side forbidden entry area X ₁ outside the region X _1. As described above, in the embodiment according to the present invention, this minimum fuel consumption operation line is the reference operation line W, and basically shows a combination of the mechanical compression ratio and the intake valve closing timing according to the required intake air amount. The mechanical compression ratio, the intake valve closing timing, and the opening degree of the throttle valve 17 are controlled so that the operating point moves on the reference operating line W. The current operating point is always detected by the relative position sensor 22, the valve timing sensor 23, and the throttle opening sensor 24.

  Next, how to control the mechanical compression ratio, the intake valve closing timing and the opening degree of the throttle valve 17 according to the present invention will be described from the basic control method. This basic control method is shown in FIGS.

That is, FIG. 14 shows a case where the required intake air amount is increased when the mechanical compression ratio and the intake valve closing timing are maintained at the values at the point m on the reference operation line W. Incidentally, in the embodiment according to the present invention, for example, the required intake air amount is calculated every predetermined time, and the operating point on the reference operation line W satisfying the required intake air amount calculated every predetermined time. Are sequentially calculated. Examples of operating points that satisfy this required intake air amount, that is, required operating points, are indicated by a ₁ , a ₂ , a ₃ , a ₄ , a ₅ , and a ₆ in FIG. That is, in this example, the requested operating point that satisfies the requested intake air amount that is detected first after the requested intake air amount is increased is a ₁ , and the requested operating point that satisfies the detected requested intake air amount is the next. is a _2, then fill the detected required amount of intake air required operating point is a _3.

When the required operating point changes, the operating point indicating the mechanical compression ratio and the intake valve closing timing changes toward a new required operating point. That is, the operating point showing a mechanical compression ratio and the intake valve closing timing in the example shown in FIG. 14 is changed toward the _point a from the required operating point when is the a ₁ m point, the required operating point is a _When it is ₂ , the operating point indicating the mechanical compression ratio and the intake valve closing timing changes toward a ₂ . In this case, if the mechanical compression ratio and intake valve closing timing reach the required operating point before the required operating point changes, the mechanical compression ratio and intake valve closing timing follow the change in the required operating point without any problem. And change. However, problems may arise if the mechanical compression ratio and the intake valve closing timing do not reach the required operating point before the required operating point changes.

That is, in FIG. 14, when the mechanical compression ratio and the intake valve closing timing are at the point m and the required operating point a ₁ is reached, the mechanical compression ratio and the intake valve timing do not change, and at this time, the required intake air amount is satisfied. Accordingly, the opening degree of the throttle valve 17 is increased. The response of the opening change of the throttle valve 17 by the actuator 16 is very fast. Therefore, when the required operating point becomes a ₁ , the operating point indicating the mechanical compression ratio and the intake valve closing timing immediately shifts from the m point to the a ₁ point.

Then the required operating point while it is just slightly advanced mechanical compression ratio is made to decrease only slightly and the intake valve closing timing becomes the a ₂ is the throttle valve 17 is fully opened. At this time, the mechanical compression ratio and the intake valve closing timing reach close to the required operating point a ₂ when the next required operating point a ₃ is calculated. Indicated by the operating point b ₂ in FIG. 15 showing the place mechanical compression ratio and the intake valve closing timing is viewed from above in FIG. 14 to reach this time.

When the required operating point a ₃ is calculated, the mechanical compression ratio and the intake valve closing timing start moving from the operating point b ₂ toward the required operating point a ₃ . That is, the mechanical compression ratio is lowered while the throttle valve 17 is fully open, and the intake valve closing timing is advanced. However, the response of the mechanical compression ratio change by the variable compression ratio mechanism A and the response of the valve closing timing change of the intake valve 7 by the variable valve timing mechanism B are not so fast. Sex is quite slow. Therefore, when the increase rate of the required intake air amount is high, the required operating point and the operating point indicating the actual values of the mechanical compression ratio and the intake valve closing timing are gradually separated. For example, in FIG. 15, when the required operating point moves to a _{6, the} operating point indicating the actual values of the mechanical compression ratio and the intake valve closing timing still remains in the vicinity of b ₂ .

However, when such a mechanical compression ratio and the intake valve closing timing when the move by the feedback control toward the required operating point without entering the closing timing mechanical compression ratio and the intake valve forbidden entry region X ₁ It takes time to reach the required operating point. That is, in this case, the advance action of the intake valve closing timing when the operating point is about to penetrate the forbidden entry area X ₁ is stopped by advancing the intake valve closing timing, then the mechanical compression ratio is It can be reduced by a certain amount. The mechanical compression ratio is again advanced the intake valve closing timing and used to lower by a certain amount, the operating point is advancing action of the intake valve closing timing is stopped to become likely to invade forbidden entry area X _1. This is repeated below.

That is, the operating point showing a mechanical compression ratio and the intake valve closing timing when the move by the feedback control toward the mechanical compression ratio and the intake valve closing timing to the required operating point along the outer edge of the forbidden entry area X ₁ Therefore, it takes time for the mechanical compression ratio and the intake valve closing timing to reach the required operating point. As a result, good engine responsiveness cannot be obtained with respect to changes in the required intake air amount.

Therefore, in the present invention, when the required intake air amount changes, the operating point is constant so as to change without entering the intrusion prohibited areas X ₁ and X ₂ toward the required operating point that satisfies the required intake air amount. Based on the predicted amount of mechanical compression ratio that can be changed in time and the predicted amount of intake valve closing timing that can be changed in a certain period of time, the target operating point is calculated at regular intervals, and the mechanical compression ratio and intake valve closing timing are calculated. It is made to change toward this target operating point.

Next, an embodiment embodying the present invention will be described with reference to FIG. 15 showing the throttle fully open surface θ _max . As described above, FIG. 15 shows a case where the operating point indicating the mechanical compression ratio and the intake valve closing timing is b ₂ when the required operating point becomes a ₃ . In this case, the arrow R ₂ represents the predicted amount of the mechanical compression ratio that can be changed at a predetermined time toward the required operating point a ₃ , and the arrow T ₂ is predetermined toward the required operating point a _3. This represents the predicted amount of the intake valve closing timing that can be changed at a predetermined time. In FIG. 15, c ₂ is a target operating point that can be reached after a certain time without entering the intrusion prohibited area X ₁ from the current operating point b ₂ toward the required operating point a ₃ that satisfies the required intake air amount. Show.

As shown in FIG. 15, when the required intake air amount is increased and the operating point b ₂ and the required operating point a ₃ are on the throttle fully open surface θ _max , the target operating point c ₂ is on the reference operating line W, In the example shown in FIG. 15, it is located on the minimum fuel consumption operation line W. That is, FIG. In the example shown in 15, the minimum fuel consumption operating target operating point which extends along the outer edge of the outside and a by forbidden entry area X ₁ of the forbidden entry area X ₁ when the throttle valve 17 is maintained fully open It can be moved on the line W.

Further, in FIG. 15, if the operating point indicating the mechanical compression ratio and the intake valve closing timing is b _i when the required operating point is a ₆ , the target operating point is also a point on the reference operating line W in this case. c _i . In FIG. 15, the arrow R _i similarly represents the predicted amount of the mechanical compression ratio that can be changed at a fixed time, and the arrow T _i represents the predicted amount of the intake valve closing timing that can be changed at a fixed time. .

In this way, in the example shown in FIG. 15, when the target operating point c ₂ is calculated when the operating point is b ₂ , the operating point indicating the mechanical compression ratio and the intake valve closing timing after a certain time is the target operating point c. Reach ₂ At this time, the next new target operating point that can be reached after a certain time without entering the intrusion prohibited area X ₁ from the current operating point c ₂ toward the required operating point that satisfies the required intake air amount is calculated, The point reaches this new target operating point after a certain time. In this case, in the embodiment according to the present invention, the mechanical compression ratio, the intake valve closing timing, and the opening degree of the throttle valve 17 are made to reach the target operating point by PID (proportional integral derivative) control.

  Thus, in the example shown in FIG. 15, the operating point indicating the mechanical compression ratio and the intake valve closing timing moves smoothly along the reference operation line W without stagnation. That is, when the required intake air amount is increased when the mechanical compression ratio and the intake valve closing timing are maintained at the point m in FIG. 14, the mechanical compression ratio and the intake valve closing timing are indicated by arrows in FIG. Thus, it can be smoothly changed along the reference operation line W without stagnating. As a result, it is possible to ensure a good engine response to changes in the required intake air amount.

In this case, in order to further improve the responsiveness of the engine with respect to the required intake air amount, it is preferable that the target operating points c ₂ and c _i be separated from the corresponding current operating points b ₂ and b _i as much as possible. Therefore, in the embodiment according to the present invention, the target operating points c ₂ and c _i are the predicted amount of the mechanical compression ratio that can be changed from the corresponding current operating points b ₂ and b _i to a fixed time and the intake valve that can be changed to a fixed time. Of the operating points located within the predicted amount of the closing timing, the furthest away from the current operating points b ₂ and b _i toward the required operating point that satisfies the required intake air amount in a range that does not enter the intrusion prohibited area X ₁ . It is considered as an operating point.

That is, when the current operating point is b ₂ , the reach limit of the mechanical compression ratio from the operating point b ₂ is set as the target operating point c _2, and this target operating point c ₂ is the operating point for the intake valve closing timing. the front of the arrival limit of the intake valve closing timing from b _2. Therefore, at this time, the mechanical compression ratio is lowered at the maximum possible speed, and the intake valve closing timing is advanced at a speed slower than the maximum possible speed. In contrast, the arrival limit of the intake valve closing timing from the operating point b _i in the case of the b _i the current operating point is a target operating point c _i, the target operating point c _i for the mechanical compression ratio the front of the arrival limit of the intake valve closing timing from the operating point b _i. Therefore, at this time, the intake valve closing timing is advanced at the maximum possible speed, and the mechanical compression ratio is decreased at a speed slower than the maximum possible speed.

  The maximum possible change speed of the intake valve closing timing, that is, the amount of the intake valve closing timing that can be changed at a certain time is hardly affected by the operating state of the engine, and therefore the intake valve closing timing that can be changed at a certain time. The amount is constant regardless of the engine operating condition. On the other hand, the maximum change speed at which the mechanical compression ratio can be changed, that is, the amount of the mechanical compression ratio that can be changed in a fixed time is strongly influenced by the operating state of the engine. Next, this will be described with reference to FIGS.

FIG. 17 shows the relationship between the engine load and the amount of the mechanical compression ratio that can be changed at a certain time, that is, the compression ratio difference between the current mechanical compression ratio and the mechanical compression ratio that can be reached after a certain time. FIG. 17 shows the amount of the mechanical compression ratio that can be changed in a certain time when the mechanical compression ratio is a certain mechanical compression ratio (hereinafter referred to as “the compression ratio changeable amount”). A chain line F ₀ indicates the amount by which the compression ratio can be changed when the engine is stopped. In FIG. 17, the torque applied to the variable compression ratio mechanism A by the combustion pressure is indicated by a broken line. This torque acts in the direction of pulling the cylinder block 2 away from the crankcase 1, that is, in the direction of reducing the compression ratio. As shown by the broken line, this torque increases as the combustion pressure increases, that is, as the engine load increases.

Thus, this torque acts on the variable compression ratio mechanism A in the direction of lowering the compression ratio. Therefore, when the mechanical compression ratio is lowered, the mechanical compression ratio is easily lowered, and in this case, the compression ratio is changed. The possible amount increases. In FIG. 17, the solid line F ₁ indicates the compression ratio changeable amount in this case, and the compression ratio changeable amount in this case increases as the engine load increases. On the other hand, since this torque resists an increase in the mechanical compression ratio, the amount by which the compression ratio can be changed is smaller when the mechanical compression ratio is increased than when the mechanical compression ratio is decreased. In FIG. 17, the solid line F ₂ indicates the compression ratio changeable amount when the mechanical compression ratio is increased. In this case, the compression ratio changeable amount decreases as the engine load increases.

In this embodiment, the reference compression ratio changeable amount indicated by F ₀ in FIG. 17 is stored in advance, and this reference compression ratio changeable amount is corrected by the relationship indicated by F ₁ and F ₂ in FIG. Thus, an expected value of the compression ratio changeable amount corresponding to the engine load is calculated. That is, in the present embodiment, the expected value of the compression ratio changeable amount is changed according to the engine load.

  FIG. 18 shows the relationship between the amount of mechanical compression ratio that can be changed in a certain time and the rotation angle of the camshafts 54 and 55, that is, the rotation angle of the circular cam 58. In FIG. 18, the left side of the horizontal axis shows the state where the mechanical compression ratio is the lowest shown in FIG. 3A, and in FIG. 18, the right end of the horizontal axis shows the mechanical compression shown in FIG. It shows the state with the highest ratio. FIG. 18 shows the compression ratio changeable amount when the engine load is a certain load, and the broken line in FIG. 18 shows the torque applied to the variable compression ratio mechanism A by the combustion pressure.

In the embodiment shown in FIG. 2, the worms 61 and 62 are of a type in which the worms 61 and 62 are not rotated by the worm wheels 63 and 64 as the worm gears. A worm gear of a type is used, and a one-dot chain line G _{0 in} FIG. 18 indicates the amount by which the compression ratio can be changed when the operation of the engine is stopped when such a worm gear is used. As can be seen from FIGS. 3A, 3B, and 3C, the change in compression ratio per unit rotation angle of the camshafts 54 and 55 when the mechanical compression ratio is intermediate, that is, as shown in FIG. Therefore, when the mechanical compression ratio is intermediate as shown by the one-dot chain line G ₀ in FIG. 18, the amount by which the compression ratio can be changed becomes the largest.

Further, as shown by the broken line in FIG. 18, the torque applied to the variable compression ratio mechanism A by the combustion pressure is highest when it is shown in FIG. 3B, that is, when the mechanical compression ratio is intermediate. On the other hand, in FIG. 18, a solid line G ₁ indicates a case where the mechanical compression ratio is decreased, and a solid line G ₂ indicates a case where the mechanical compression ratio is increased. As shown in FIG. 18, the compression ratio changeable amount G ₁ when the mechanical compression ratio is decreased is larger than the compression ratio changeable amount G ₂ when the mechanical compression ratio is increased. At this time the compression ratio changeable amount G ₁ is higher because the mechanical compression ratio is torque based on the compression ratio is highest at the middle compression ratio changeable quantity G ₂ is lowered.

In the present embodiment, the compression ratio changeable amount serving as a reference as shown in G ₀ 18 is stored in advance, corrected by the relationship shown the basic compression ratio changeable amount G ₁ and G ₂ in FIG. 18 Thus, the compression ratio changeable amount corresponding to the rotation angle of the camshafts 54 and 55 is calculated. Further, by correcting this compression ratio changeable amount by the relationship indicated by F ₁ and F ₂ in FIG. 17, a predicted value of the compression ratio changeable amount corresponding to the rotation angle of the camshafts 54 and 55 and the engine load is calculated. Is done. That is, in this embodiment, the predicted value of the compression ratio changeable amount is changed according to the rotation angle of the rotating cam 58 and the engine load.

  On the other hand, FIG. 19 shows the influence of the lubrication state by the slide bearing on the compression ratio changeable amount when all the bearings of the variable compression ratio mechanism A are constructed from the slide bearings. That is, the higher the engine load, the easier it becomes a boundary lubrication region where oil film breakage starts, and the lower the operation speed on the bearing surface, the easier it becomes the boundary lubrication region. Accordingly, as shown in FIG. 19, when (engine load / operating speed) exceeds a certain limit value, the lubrication state becomes the boundary lubrication region, and as a result, the frictional force in the sliding bearing increases, so that the compression ratio changeable amount is small. Become.

In another embodiment according to the present invention, the predicted value of the compression ratio changeable amount is calculated in consideration of the lubrication state of the slide bearing. For example, the reference compression ratio changeable amount indicated by F ₀ in FIG. 17 is corrected by the relationship indicated by F ₁ and F ₂ in FIG. 17, and the corrected compression ratio changeable amount is indicated by G ₁ and G ₂ in FIG. Correction is made according to the relationship shown, and the compression according to the engine load, the rotation angle of the camshafts 54 and 55, and (engine load / operation speed) is corrected by correcting the corrected compression ratio changeable amount according to the relationship shown in FIG. A predicted value of the ratio changeable amount is calculated.

  FIG. 20 shows an embodiment in which a change in the engine load is detected and a compression ratio changeable amount is calculated based on the detected change in the engine load. When the combustion pressure varies between cycles or cylinders, the eccentric shaft 57 bends, and the relative position between the cylinder block 2 and the crankcase 1 changes. The relative position sensor 22 detects the change in the relative position between the cylinder block 2 and the crankcase 1, that is, the change in the distance between the cylinder block 2 and the crankcase 1. The distance between the cylinder block 2 and the crankcase 1 increases as the combustion pressure increases.

As described above, torque is applied to the variable compression ratio mechanism A by the combustion pressure, and this torque acts on the variable compression ratio mechanism A in the direction of lowering the compression ratio. Therefore, when the combustion pressure increases, the mechanical compression ratio can be easily lowered by the variable compression ratio mechanism A. H _{1 in} FIG. 20 indicates a compression ratio changeable amount when the mechanical compression ratio is lowered, and H ₂ in FIG. 20 indicates a compression ratio changeable amount when the mechanical compression ratio is increased.

  In this embodiment, the predicted value of the compression ratio changeable amount can be appropriately calculated according to the fluctuation of the combustion pressure. In particular, the predicted value of the compression ratio changeable amount is calculated according to the engine load as shown in FIG. 17, and at this time, the predicted value of the compression ratio changeable amount is based on the fluctuation of the combustion pressure as shown in FIG. Further calculation allows the predicted value of the compression ratio changeable amount to be precisely controlled to an optimum value. If a worm gear of the type capable of rotating the worms 61, 62 by the worm wheels 63, 64 is used as the worm gear, the deflection of the eccentric shaft 57 when the combustion pressure fluctuates further increases. As a result, the variable compression ratio mechanism Variations in torque acting on A can be detected with higher accuracy by the relative position sensor 22.

  Next, a case where the required intake air amount is reduced will be described with reference to FIGS. Of FIGS. 21 to 36, FIGS. 21 and 22 show cases where the required intake air amount is slowly decreased, and FIGS. 23 to 30 show that the required intake air amount is decreased relatively quickly. FIG. 31 to FIG. 36 show a case where the required intake air amount is sharply reduced. FIGS. 21 to 36 show a case where the action of reducing the required intake air amount is started when the operating point indicating the combination of the mechanical compression ratio and the intake valve closing timing is n point on the reference operating line W. ing.

First, a case where the required intake air amount is slowly reduced will be described with reference to FIGS. 21 and 22. FIG. 22 shows the throttle fully open surface θ _max similar to FIG.

FIG. 22 shows the relationship between the current operating point and the requested operating point in this case. That is, in FIG. 22 requirements operating point when the current operating point is e _i is indicated by d _i, the predicted amount of time can change the mechanical compression ratio in a predetermined time indicated by R _i At this time, a predicted amount of the intake valve closing timing that can be changed at a certain time is indicated by T _i . Further in FIG. 22 is shown requesting operating point when the current operating point is e _j is in d _j, the predicted amount of changeable mechanical compression ratio constant time this time is indicated by R _j At this time, a predicted amount of the intake valve closing timing that can be changed at a predetermined time is indicated by T _j .

In this case, the required operating point d _i is before the reaching limit of the mechanical compression ratio and is before the reaching limit of the intake valve closing timing, so that the required operating point d _i becomes the target operating point. Similarly, the required operating point _dj is before the reaching limit of the mechanical compression ratio and is before the reaching limit of the intake valve closing timing, so that the required operating point _dj is the target operating point. Therefore, in this case, the operating point moves along the reference operating line W. That is, when the required intake air amount decreases slowly, the intake valve closing timing is gradually retarded while the throttle valve 17 is kept fully open, and the mechanical compression ratio gradually increases so that the actual compression ratio becomes constant. Will be increased.

Next, a case where the required intake air amount is decreased relatively quickly will be described with reference to FIGS. As described above, in the embodiment according to the present invention, for example, the required intake air amount is calculated every predetermined time, and the required operation point on the reference operation line W that satisfies the sequentially calculated required intake air amount is shown in FIG. D ₁ , d ₂ , d ₃ , d ₄ , and d ₅ .

In order to easily understand the control according to the present invention, FIG. 23 shows that the required intake air amount at the required operation point d ₁ is Q ₅ , and the required intake air amounts at the required operation point d ₂ are Q ₅ and Q ₄ . It is an intermediate value, the required intake air amount at the required operating point d ₃ is Q ₄ , the required intake air amount at the required operating point d ₄ is an intermediate value between Q ₄ and Q ₃ , and at the required operating point d ₅ The case where the required intake air amount is Q ₃ is shown. That is, from the sequentially calculated required intake air amount Q ₆ (n point on the throttle fully open surface θ _max ), Q ₅ , Q ₅ and Q ₄ , Q ₄ , Q ₄ and Q ₃ , shows a case where the change in Q _3.

The Figure 24 shows a wide open throttle surface theta _max, Figure 25 is the intake air amount represents the same intake air amount plane of Q _5, FIG. 26 is the intake air amount of the intermediate value Q ₅ and Q ₄ 27 shows the same intake air amount surface, FIG. 27 shows the same intake air amount surface where the intake air amount is Q ₄ , and FIG. 28 shows the same intake air amount where the intake air amount is an intermediate value between Q ₄ and Q _3. It shows a plane, Figure 29 is the intake air amount indicates the same intake air amount plane of Q _3.

Now, when the mechanical compression ratio and the intake valve closing timing are held at the operating point n shown in FIG. 23, the required intake air amount changes from Q ₆ to Q _5, and as a result, the required operating point becomes d ₁ . If this is the case, first, as shown in FIG. 24, the target operating point e ₁ is calculated on the throttle fully open surface θ _max . The calculation method of the target operating point e ₁ is the same as the calculation method described so far, and the prediction amount of the mechanical compression ratio that can be changed at a certain time and the prediction amount of the intake valve closing timing that can be changed at a certain time. target operating point e ₁ closest to the required operating point d ₁ without penetration is calculated forbidden entry region X ₁ from. In the example shown in FIG. 24, the target operation point e ₁ is located on the reference operation line W.

Incidentally, the intake air amount at the target operating point e ₁ is an intermediate value between Q ₆ and Q ₅ described above and is larger than the required intake air amount Q ₅ . However, it is preferable that the intake air amount matches the required intake air amount as much as possible. However, when the required intake air amount is decreased, the intake air amount can be adjusted by changing the opening of the throttle valve 17. Therefore, when the intake air amount at the target operating point e ₁ is larger than the required intake air amount Q _{5, the} required intake air amount is set without changing the target values for the mechanical compression ratio and the intake valve closing timing. until the target opening degree required for the amount Q ₅ so that to close the throttle valve 17.

That is, in FIG. 23, the final target operating point e ₁ is a point on the same intake air amount surface Q ₅ located immediately below the target operating point e ₁ on the throttle fully open surface θ _max shown in FIG. . The final target operating point e ₁ on the same intake air amount surface Q ₅ is shown in FIGS. 23 and 25. The mechanical compression ratio, the intake valve closing timing, and the opening degree of the throttle valve 17 are the final values. The target operating point e ₁ is changed. That is, at this time, the mechanical compression ratio is increased, the intake valve closing timing is retarded, and the opening degree of the throttle valve 17 is decreased from the fully opened state.

Next, when the required intake air amount becomes an intermediate value between Q ₅ and Q ₄ and the required operating point becomes d ₂ , this time, on the same intake air amount surface at the current intake air amount Q ₅ as shown in FIG. A target operating point e ₂ is calculated. The calculation method of the target operating point e ₂ is the same as the calculation method described so far, and the prediction amount of the mechanical compression ratio that can be changed at a certain time and the prediction amount of the intake valve closing timing that can be changed at a certain time. the target operating point e ₂ closest to the required operating point d ₂ without entering calculated in forbidden entry region X ₁ from. In the example shown in FIG. 25, this target operating point e ₂ is located on the reference operating line W in the same intake air amount surface Q ₅ (note that the reference operating line W at this time is the reference operating line W shown in FIG. It is different from the operation wire W, and the minimum fuel consumption operation line at the same intake air amount plane Q _5).

In this case as well, the intake air amount at the target operating point e ₂ is larger than the required intake air amount. Therefore, also in this case, in FIG. 23, on the same intake air amount surface (intermediate value between Q ₅ and Q ₄ ) located directly below the target operating point e ₂ on the same intake air amount surface Q ₅ shown in FIG. Is the final target operating point e ₂ . The final target operating point e ₂ on the same intake air amount surface (intermediate value between Q ₅ and Q ₄ ) is shown in FIGS. 23 and 26, and the mechanical compression ratio, intake valve closing timing and throttle valve are shown. The opening degree of 17 is changed toward this final target operating point e ₂ . Also at this time, the mechanical compression ratio is increased, the intake valve closing timing is retarded, and the opening degree of the throttle valve 17 is reduced from the fully opened state.

Next, when the required intake air amount becomes Q ₄ , then becomes an intermediate value between Q ₄ and Q ₃ , and then becomes Q ₃ , the same thing is sequentially repeated. That is, when the required intake air amount becomes Q ₄ , the final target operating point e ₃ on the same intake air amount surface Q ₄ is calculated as shown in FIG. 27, and the required intake air amounts are equal to Q ₄ and Q ₃ . When the intermediate value is reached, the final target operating point e ₄ on the same intake air amount surface (intermediate value between Q ₄ and Q ₃ ) is calculated as shown in FIG. 28, and then the required intake air amount becomes Q ₃ . Then, as shown in FIG. 29, the final target operating point e ₅ on the same intake air amount surface Q ₃ is calculated.

During this time, that is, while the mechanical compression ratio, the intake valve closing timing, and the opening degree of the throttle valve 17 are sequentially changed toward the final target operating points e ₃ , e ₄ , e ₅ , the mechanical compression ratio increases. The intake valve closing timing is retarded, and the opening degree of the throttle valve 17 is reduced.

When the required intake air amount reaches Q ₃ , final target operating points e ₆ , e ₇ , e ₈ , e ₉ , e ₁₀ are sequentially calculated on the same intake air amount surface Q ₃ as shown in FIG. The mechanical compression ratio, the intake valve closing timing, and the opening degree of the throttle valve 17 are sequentially changed to the required operating point d ₅ through these final target operating points e ₆ , e ₇ , e ₈ , e ₉ , e _10. It will be. During this time, the mechanical compression ratio is increased, the intake valve closing timing is retarded until reaching e ₈ , the opening of the throttle valve 17 is gradually increased, and when it reaches e _8, it is fully opened.

FIG. 30 shows the intake valve closing timing and the mechanical compression ratio when the target intake air amount is decreased relatively quickly from Q ₆ (n point) to Q ₃ (target operating point d ₅ ) as shown in FIG. , Changes in actual compression ratio and throttle opening. From FIG. 30, in this case, after the required intake air amount reaches the final target value Q ₃ (operating point e ₄ ), the retarding operation of the intake valve closing timing is completed (operating point e ₈ ). It can be seen that the increasing action of the mechanical compression ratio is completed (target operating point d ₅ ). On the other hand, the actual compression ratio gradually decreases until the retarding action of the intake valve closing timing is completed (operating point d ₈ ), and then gradually increases. Further, the throttle opening is gradually decreased from the fully opened state until the operating point becomes the operating point e ₅ on the same intake air amount surface Q ₃ , and then until the retarding operation of the intake valve closing timing is completed (operation The valve is gradually opened until it is fully opened up to point e ₈ ).

As shown in FIGS. 23 to 30, when the required intake air amount is decreased relatively quickly, the throttle opening is also controlled in addition to the control of the mechanical compression ratio and the intake valve closing timing. In the present invention, at this time, the three-dimensional intrusion prohibited areas X ₁ and X ₂ are set for the combination of the mechanical compression ratio, the intake valve closing timing, and the throttle opening, and the mechanical compression ratio, the intake valve closing timing, An operating point indicating a combination with the throttle opening is prohibited from entering the three-dimensional intrusion prohibited areas X ₁ and X ₂ .

Also in this case, when the required intake air amount changes, the mechanical compression ratio and the intake valve closing timing are set such that the operating point reaches the required operating point where the required intake air amount satisfies the required intake air amount X. ₁ , target at every fixed time based on the predicted amount of mechanical compression ratio that can be changed at a fixed time and the predicted amount of intake valve closing timing that can be changed at a fixed time so as to change without entering X ₂ The operating point is calculated, and the mechanical compression ratio and the intake valve closing timing are changed toward the calculated target operating point. Further, in this case, when the required intake air amount changes, the throttle opening is changed according to the required intake air amount so as not to enter the three-dimensional intrusion prohibited areas X ₁ and X ₂ .

Even in this case, the target operating point is changed from the current operating point to a fixed time so that the mechanical operating ratio, intake valve closing timing, and throttle opening reach the required operating point that satisfies the required intake air amount as soon as possible. Of operating points located within the predicted amount of possible mechanical compression ratio and the estimated amount of intake valve closing timing that can be changed at a certain time, required within a range that does not enter the three-dimensional intrusion prohibited areas X ₁ and X ₂ The operating point is the furthest away from the operating point toward the required operating point that satisfies the intake air amount.

Further, in this case, in the embodiment according to the present invention, when the required intake air amount decreases, the mechanical compression ratio and the intake valve closing timing are changed from the current operating point toward the operating point that satisfies the required intake air amount. A target operating point that can be reached after a certain time without entering the intrusion prohibited areas X ₁ and X ₂ in the intake air amount is calculated, and toward the target operating point where the mechanical compression ratio and the intake valve closing timing are calculated Can be changed. On the other hand, in this case, as for the throttle opening, the target opening satisfying the required intake air amount is calculated at the calculated target operating point, and the throttle is not limited unless the target opening is in the three-dimensional intrusion prohibited areas X ₁ and X _2. The opening is changed to the target opening.

Next, a case where the required intake air amount is suddenly reduced to the minimum intake air amount Q ₁ will be described with reference to FIGS. As described above, in the embodiment according to the present invention, for example, the required intake air amount is calculated every predetermined time, and the required operation point on the reference operation line W that satisfies the sequentially calculated required intake air amount is shown in FIG. D ₁ , d ₂ , and d ₃ .

Also in this case, in order to easily understand the control according to the present invention, FIG. 31 shows that the required intake air amount at the required operation point d ₁ is Q ₄ and the required intake air amount at the required operation point d ₂ is Q _2. And Q ₃ , and the required intake air amount at the required operating point d ₃ is Q ₁ . That is, a case is shown where the sequentially calculated required intake air amount changes from Q ₆ (n point on the throttle fully open surface θ _max ) to Q ₄ , an intermediate value between Q ₃ and Q ₂ , Q ₁ .

FIG. 32 shows the throttle fully open surface θ _max , FIG. 33 shows the same intake air amount surface with the intake air amount Q ₄ , and FIG. 34 shows the intake air amount with an intermediate value between Q ₃ and Q ₂ . It shows the same intake air amount plane, Figure 35 is the intake air amount indicates the same intake air amount plane of Q _1.

Now, when the mechanical compression ratio and the intake valve closing timing are held at the operating point n shown in FIG. 31, the required intake air amount changes from Q ₆ to Q _4, and as a result, the required operating point becomes d ₁ . If this is the case, first, as shown in FIG. 32, the target operating point e ₁ is calculated on the throttle fully open surface θ _max . The calculation method of the target operating point e ₁ is the same as the calculation method shown in FIG. 24, and the predicted amount of the mechanical compression ratio that can be changed at a fixed time and the predicted amount of the intake valve closing timing that can be changed at a fixed time. target operating point e ₁ closest to the required operating point d ₁ without penetration is calculated forbidden entry region X ₁ from. In the example shown in FIG. 32, the target operation point e ₁ is located on the reference operation line W.

On the other hand, as in the case shown in FIG. 23, the target values for the mechanical compression ratio and the intake valve closing timing are not changed, and the target opening degree required to set the intake air amount to the required intake air amount Q ₄ is not changed. The throttle valve 17 is closed.

That is, in FIG. 31, the final target operating point e ₁ is a point on the same intake air amount surface Q ₄ located directly below the target operating point e ₁ on the throttle fully open surface θ _max shown in FIG. . The final target operating point e ₁ on the same intake air amount surface Q ₄ is shown in FIGS. 31 and 33. The mechanical compression ratio, the intake valve closing timing, and the opening degree of the throttle valve 17 are the final values. The target operating point e ₁ is changed. At this time, the mechanical compression ratio is increased, the intake valve closing timing is retarded, and the opening degree of the throttle valve 17 is reduced from the fully opened state.

Next, when the required intake air amount becomes an intermediate value between Q ₃ and Q ₂ and the required operation point becomes d ₂ , this time, on the same intake air amount surface in the current intake air amount Q ₄ as shown in FIG. A target operating point e ₂ is calculated. The calculation method of the target operating point e ₂ is the same as the calculation method described so far, and the prediction amount of the mechanical compression ratio that can be changed at a certain time and the prediction amount of the intake valve closing timing that can be changed at a certain time. the target operating point e ₂ closest to the required operating point d ₂ without entering calculated in forbidden entry region X ₁ from. Also in this case, in FIG. 31, on the same intake air amount surface (intermediate value of Q ₃ and Q ₂ ) located directly below the target operating point e ₂ on the same intake air amount surface Q ₄ shown in FIG. The point is set as the final target operation point e ₂ . FIG. 31 and FIG. 34 show the final target operating point e ₂ on the same intake air amount surface (an intermediate value between Q ₃ and Q ₂ ).

Next, when the required intake air amount becomes Q ₁ and the required operation point becomes d ₃ , the target operation point e ₃ is calculated on the same intake air amount surface (an intermediate value between Q ₃ and Q ₂ ) as shown in FIG. Then, as shown in FIG. 35, the final target operating point e ₃ on the same intake air amount surface Q ₁ is calculated. When the final target operating point e ₃ is calculated, the mechanical compression ratio, the intake valve closing timing, and the opening degree of the throttle valve 17 are changed toward the final target operating point e ₃ . Also at this time, the mechanical compression ratio is increased, the intake valve closing timing is retarded, and the opening degree of the throttle valve 17 is reduced from the fully opened state.

Meanwhile the low load side forbidden entry area X ₂ comes appear in the same intake air amount plane in this way required intake air amount is reduced. Low load forbidden entry area X ₂ appearing on the same intake air amount plane increases as the amount of intake air is small, the low load side forbidden entry area X ₂ appearing on the same intake air amount plane as shown in FIG. 35 becomes maximum when the required intake air amount is minimized Q ₁ in. In the embodiment according to the present invention, the operating point into the low load intrusion prohibition region X ₂ is slightly spaced from the low load side intrusion prohibition region X ₂ around the low load side intrusion prohibition region X ₂ . An intrusion prevention surface for preventing intrusion is set in advance, and an intrusion prevention line that is an intersection line between the intrusion prevention surface and the same intake air amount surface is indicated by WX in FIG.

Now, an embodiment according to the present invention and the predicted amount of changeable mechanical compression ratio constant time on the same intake air amount plane Q ₁ as shown in FIG. 35 when the amount of intake air is required intake air quantity Q ₁ constant Each target operating point e ₄ , e ₅ , e ₆ , e ₇ , e ₈ , e ₉ , e ₁₀ , e ₁₁ , closest to the required operating point d ₃ from the predicted amount of intake valve closing timing that can be changed in time. e ₁₂ is sequentially calculated. In this case, the calculated target operating point when the calculated target operating point closest to the required operating point d ₃ , such as the target operating point e ₄ , is located on the opposite side of the intrusion prevention area X ₂ with respect to the intrusion prevention line WX. Is the target operating point e ₄ . On the other hand, when the calculated target operating point closest to the required operating point d ₃ is closer to the intrusion prohibition region X ₂ than the intrusion prevention line WX, either the mechanical compression ratio or the intake valve closing timing is reached. The points on the intrusion prevention line WX that are the limits are set as the target operating points e ₅ , e ₆ , e ₇ , e ₈ , and e ₉ .

That is, when the required intake air amount becomes Q ₁ , the mechanical compression ratio, the intake valve closing timing, and the opening degree of the throttle valve 17 are sequentially set to the final target operating points e ₄ , e ₅ , on the same intake air amount surface Q ₁ , The required operating point d _{3 is} changed through e ₆ , e ₇ , e ₈ , e ₉ , e ₁₀ , e ₁₁ , e ₁₂ . During this time, the mechanical compression ratio is increased, the intake valve closing timing is retarded until reaching e _10, and the opening of the throttle valve 17 is gradually increased, and when it reaches e _10, it is fully opened.

FIG. 36 shows the intake valve closing timing, the mechanical compression ratio, when the target intake air amount is sharply decreased from Q ₆ (point n) to Q ₁ (target operating point d ₃ ), as shown in FIG. Changes in actual compression ratio and throttle opening are shown. From FIG. 36, after the required intake air amount reaches the final target value Q ₁ (operating point e ₂ ), the retarding action of the intake valve closing timing is completed (operating point e ₁₀ ). It can be seen that the increasing action of the mechanical compression ratio is completed (target operating point d ₃ ). On the other hand, the actual compression ratio gradually decreases until the retarding action of the intake valve closing timing is completed (operating point e ₁₀ ), and then gradually increases. Further, the throttle opening is lowered from the fully opened state until the operating point reaches the operating point e ₃ on the same intake air amount surface Q ₁ , and then until the retarding action of the intake valve closing timing is completed (operating point e _The valve is gradually opened until ₁₀ ).

It should be noted that the demanded intake meet air amount throttle valve 17 opening of the three-dimensional forbidden entry area, that is, if a low load side forbidden entry area X ₂ when required intake air amount changes. In this case, the opening degree of the throttle valve 17 is changed to the above-described intrusion prevention surface, that is, just before entering the three-dimensional intrusion prohibition region, and then the mechanical compression ratio, intake valve closing timing, and throttle opening degree are changed. The operating point indicating the combination is changed without entering the three-dimensional intrusion prohibited area toward the operating point satisfying the required intake air amount.

  In the above embodiment, the required operating point is the operating point on the reference operating line W that satisfies the required intake air amount (required engine load). Therefore, it can be said that the required operation line is set based on the required intake air amount, that is, based on the engine operating state. Further, the reference operation line W described above indicates an operation line through which an operating point should pass in an engine operation state in which it is necessary to minimize fuel consumption. For example, at the time of engine cold start, an internal combustion engine or a three-way catalyst In an engine operating state where a temperature increase is required, this operation line is different from the reference operation line W. Therefore, the required operating point differs between an operating state that minimizes fuel consumption and an operating state that raises the temperature of the internal combustion engine. When these are expressed together, it can be said that the required operating point is the optimal operating point in each engine operating state and is set for each engine operating state.

  FIG. 37 shows a routine for calculating a target operating point that can be reached after a predetermined time from the current operating point, that is, a target value for the mechanical compression ratio, intake valve closing timing, and throttle opening. Is shown.

  In this routine, a target operating point that can be reached after a predetermined time is calculated every predetermined time. Therefore, the routine shown in FIG. 37 is executed by interruption every predetermined time. The predetermined time can be arbitrarily determined, but in the embodiment according to the present invention, the predetermined time is 8 msec. Therefore, in the embodiment according to the present invention, the target value calculation routine shown in FIG. 37 is executed every 8 msec, and the target operating point that can be reached after 8 msec from the current operating point is calculated every 8 msec.

  Referring to FIG. 37, first, at step 100, the required intake air amount GX is calculated. The required intake air amount GX is stored in advance in the ROM 32 as a function of, for example, the depression amount of the accelerator pedal 40 and the engine speed. Next, at step 101, a required operating point on the reference operating line W corresponding to the required intake air amount GX is calculated. Next, at step 102, it is determined whether or not the current operating point is the requested operating point. When the current operating point is the requested operating point, the processing cycle is completed. On the other hand, when the current operating point is not the required operating point, the routine proceeds to step 103, where it is determined whether or not the required intake air amount GX is larger than the intake air amount GA at the current operating point.

When GX> GA, that is, when the intake air amount is to be increased, the routine proceeds to step 104, where the target operating point is determined as described with reference to FIGS. That is, in step 104, the target throttle opening degree corresponding to the required intake air amount GX is calculated. When the target throttle opening degree required operating point is located on the throttle full open plane theta _max normally fully opened. Next, at step 105, a predicted amount of the intake valve closing timing that can be changed at a fixed time is calculated, and then at step 106, a predicted amount of a mechanical compression ratio that can be changed at a fixed time is calculated.

  Next, at step 107, the target operating point is determined by the method described with reference to FIG. Next, at step 108, the target value of the mechanical compression ratio and the target value of the intake valve closing timing are calculated from the determined target operating point. The target value of the throttle opening is already calculated as the target throttle opening in step 104.

  On the other hand, when it is determined in step 103 that GX ≦ GA, that is, when the intake air amount should be reduced or when the intake air amount is the required intake air amount, the routine proceeds to step 109, based on FIG. 21 to FIG. The target operating point is determined as described above. That is, in step 109, a predicted amount of the intake valve closing timing that can be changed at a fixed time is calculated, and then in step 110, a predicted amount of a mechanical compression ratio that can be changed at a fixed time is calculated. Next, at step 111, the target operating point is determined by the method described with reference to FIGS. 22, 24 to 29, and FIGS.

  Next, at step 112, a target throttle opening satisfying the required intake air amount is calculated, and this target throttle opening is set as a target value of the throttle opening. However, when the throttle opening satisfying the required intake air amount GX falls within the intrusion prohibition region, the target throttle opening is set to the value on the intrusion prevention surface described above, and the mechanical compression ratio and the intake valve closing timing are set as the required operating points. The target throttle opening is changed along the intrusion prevention surface as it approaches.

Although not described so far, the same can occur when the required intake air amount increases. For example, if the required intake air amount increases when the operating point is located below the high load side intrusion prohibition region X ₁ in FIG. 14, the target throttle opening may be within the high load side intrusion prohibition region X ₁ . . At this time, the target throttle opening is set to a value on the reference operation surface including each reference operation line W set in advance for each same intake air amount surface, and the mechanical compression ratio and the intake valve closing timing are set as the required operation points. As the distance approaches, the target throttle opening is changed along this reference operation surface.

  FIG. 38 shows variable compression ratio mechanism A, variable valve timing mechanism B, and the like so that the mechanical compression ratio, intake valve closing timing, and throttle opening using PID control become the target values calculated in the routine shown in FIG. A drive routine for driving the throttle valve 17 is shown. This routine is repeatedly executed when the engine is started.

Referring to FIG. 38, in step 200, a difference ΔIT (= IT ₀ −IT) between the target value IT ₀ of the intake valve closing timing and the current intake valve closing timing IT is calculated, and the target value CR _{0 of the} mechanical compression ratio is calculated. And the current mechanical compression ratio CR is calculated ΔCR (= CR ₀ −CR), and the difference Δθ (θ ₀ −θ) between the throttle opening target value θ ₀ and the current throttle opening θ is calculated. The

Then the proportional term E _p1 of the drive voltage for the variable valve timing mechanism B is calculated by multiplying the proportionality constant K _p1 to ΔIT In step 201, the drive for the variable compression ratio mechanism A is multiplied by the proportional constant K _p2 to ΔCR A proportional term E _{p2 of} voltage is calculated, and a proportional term E _p3 of driving voltage for the throttle valve 17 is calculated by multiplying Δθ by a proportional constant K _p3 .

Next, at step 202, ΔIT is multiplied by an integral constant K _i1 and this multiplication result (K _i1 · ΔIT) is multiplied to calculate the integral term E _i1 of the drive voltage for the variable valve timing mechanism B, and ΔCR is integrated with the integral constant K. The integral term E _i2 of the drive voltage for the variable compression ratio mechanism A is calculated by multiplying _i2 and multiplying the multiplication result (K _i2 · ΔCR), and Δθ is multiplied by an integral constant K _i3 to obtain the multiplication result ( By integrating (K _i3 · Δθ), the integral term E _i3 of the drive voltage for the throttle valve 17 is calculated.

Then differential term E _d1 of the drive voltage for the variable valve timing mechanism B is calculated by multiplying the derivative constant K _d1 to the difference between DerutaIT ₁ that is currently DerutaIT and previously calculated in step 203 (ΔIT-ΔIT _1), A differential term E _d2 of the driving voltage for the variable compression ratio mechanism A is calculated by multiplying the difference between the current ΔCR and the previously calculated ΔCR ₁ (ΔCR−ΔCR ₁ ) by a differential constant K _d2 , and the current Δθ and differential term E _d3 of the drive voltage for the throttle valve 17 is calculated by multiplying the derivative constant K _d3 to the difference between [Delta] [theta] ₁ calculated the last time (Δθ-Δθ _1).

Then the drive voltage E ₁ for the variable valve timing mechanism B is calculated by adding the differential term and the proportional term E _p1 in step 204 and the integral term E _i1 E _d1, derivative and proportional term E _p2 and the integral term E _i2 term The drive voltage E ₂ for the variable compression ratio mechanism A is calculated by adding E _d2, and the drive voltage E ₃ for the throttle valve 17 is added by adding the proportional term E _p3 , the integral term E _i3, and the differential term E _d3. Is calculated.

When the variable valve timing mechanism B, the variable compression ratio mechanism A, and the throttle valve 17 are driven according to the drive voltages E ₁ , E ₂ , and E ₃ , the intake valve closing timing, the mechanical compression ratio, and the throttle opening are sequentially set. It will change towards the changing target value.

  By the way, in many internal combustion engines, fuel cut control for stopping the fuel supply to the combustion chamber 5 is performed in accordance with the engine operating state for the purpose of improving the fuel consumption. When performing fuel cut control, for example, when the engine load decreases and deceleration operation is performed, the engine speed and the vehicle speed of the vehicle equipped with the internal combustion engine increase excessively, and the engine speed and the vehicle speed need to be reduced. In such a case, there may be mentioned a case where the engine speed is decreased when a shift up is performed in an automatic transmission of a vehicle equipped with an internal combustion engine.

While the fuel cut control is being performed, the air-fuel mixture is not combusted in the combustion chamber 5. Therefore, even if the operating point enters the intrusion prohibited areas X ₁ and X ₂ , there is no substantial problem. Does not occur. However, if the operating point is located in the intrusion prohibited areas X ₁ and X ₂ when the fuel cut control is finished and the normal operation is resumed, abnormal combustion or torque fluctuation occurs when the normal operation is resumed. Therefore, it is necessary to prevent the operating point from entering the intrusion prohibited areas X ₁ and X ₂ during fuel cut control as in the normal operation.

  Incidentally, in the above-described embodiment, as described with reference to FIG. 17, the predicted amount of the mechanical compression ratio that can be changed in a certain time is calculated based on the engine load in consideration of the combustion pressure, and the target is calculated based on this. The operating point is calculated. Therefore, as shown in FIG. 17, when calculating the predicted amount of the mechanical compression ratio that can be changed in a certain time based on the engine load, the prediction of the mechanical compression ratio that can be changed in a certain time during the fuel cut control. The amount will be calculated as constant.

In the above-described embodiment, as described with reference to FIG. 18, the rotation angle of the camshafts 54 and 55 and the combustion pressure are taken into consideration, and the time can be changed based on the rotation angle of the camshafts 54 and 55. The predicted amount of mechanical compression ratio is calculated. However, no combustion pressure is generated during fuel cut control. Therefore, the predicted amount of possible mechanical compression ratio change in a certain time during the fuel cut control is to be calculated as indicated by a chain line G ₀ in FIG. 18 based on the rotation angle of the cam shafts 54, 55 Become.

However, during the fuel cut control, the amount of the mechanical compression ratio that can be changed in a certain time does not necessarily change according to only the rotation angle of the camshafts 54 and 55, and the intake air supplied into the combustion chamber 5 It varies depending on the amount and actual compression ratio. Thus, during the combustion cutoff control, when calculated as the predicted amount of changeable mechanical compression ratio constant time indicated by a chain line G ₀ in FIG. 18, based only on the rotation angle of the cam shafts 54, 55, a constant The predicted amount of the mechanical compression ratio that can be changed over time cannot be accurately calculated.

When the predicted amount of the mechanical compression ratio that can be changed in a certain time cannot be accurately calculated in this way, the operating point indicating the combination of the mechanical compression ratio and the intake valve closing timing becomes the intrusion prohibited areas X ₁ and X _2. There is a case to invade. This will be briefly described below.

Considering the fuel cut control when the engine load is reduced and the deceleration operation is performed, in this case, the engine load is zero, and therefore the required intake air amount is also the minimum, that is, Q ₁ shown in FIG. ing. For this reason, during the fuel cut control, the operating point indicating the combination of the mechanical compression ratio and the intake valve closing timing is moved, for example, as shown in FIGS.

At this time, after the operating point becomes the operating point on the same intake air amount plane Q ₁ , the target operating point is moved on the same intake air amount plane Q ₁ as shown in FIGS. 31 and 35 (e ₃ To e ₁₂ ), in particular, the operating points e _{5 to} e ₉ are located on the intrusion prevention line WX.

FIG. 39 is an enlarged view of a part of the intake air amount plane Q ₁ shown in FIG. When the next target operating point is calculated in the state where the current operating point is at e ₅ in the same manner as shown in FIG. 35, the predicted amount R of the mechanical compression ratio that can be changed from the current operating point to a predetermined time and the constant Of the operating points located within the predicted amount T of the intake valve closing timing that can be changed in time, the operating point e _{6 that} is not on the intrusion prohibition region X ₂ side from the intrusion prevention line WX is calculated as the next target operating point. It will be.

However, if there is an error in the predicted amount R of the mechanical compression ratio that can be changed at a certain time, and only the mechanical compression ratio R ′ can be actually changed at a certain time, the target operating point is the point e ₆ in FIG. On the other hand, the operating point actually reached is the point e ₆ ′ in FIG. As can be seen from FIG. 39, the operating point e ₆ ′ is located in the intrusion prohibition region X ₂ , so if the fuel cut control is terminated and the normal operation is resumed when the operating point is at e ₆ ′, Abnormal combustion and torque fluctuations will occur when returning to operation.

  Therefore, in the present embodiment, during the fuel cut control, the predicted amount of the mechanical compression ratio that can be changed in a certain time is calculated based on the intake air amount and the actual compression ratio instead of the engine load. This will be described below with reference to FIGS. 40 to 43.

  FIG. 40 shows the relationship between the amount of mechanical compression ratio that can be changed over a certain period of time (the amount by which the compression ratio can be changed) and the amount of intake air. FIG. 40 shows the compression ratio changeable amount when the mechanical compression ratio and the intake valve closing timing are a certain mechanical compression ratio and a certain intake valve closing timing.

  Here, during fuel cut control, that is, even when the engine load is zero, compression of the air-fuel mixture in the combustion chamber 5 after the intake valve 7 is closed and before the piston 4 reaches the compression top dead center. Action occurs. In other words, while the piston 4 is moving up, a force acts on the piston 4 in the direction of lowering the piston 4 by the air-fuel mixture in the combustion chamber 5, and this force pulls the cylinder block 2 away from the crankcase 1. Acting in the direction, i.e. in the direction of reducing the compression ratio. As indicated by a broken line in FIG. 40, this force increases as the amount of intake gas supplied into the combustion chamber 5 when the intake valve 7 is closed, that is, the amount of intake air increases.

In this way, this force acts on the variable compression ratio mechanism A in the direction of lowering the compression ratio. Therefore, when the mechanical compression ratio is lowered, the mechanical compression ratio is easily lowered, and in this case, the compression ratio is changed. The possible amount increases. In FIG. 40, the solid line I ₁ indicates the compression ratio changeable amount in this case, and the compression ratio changeable amount in this case increases as the intake air amount increases. On the other hand, since this force resists an increase in the mechanical compression ratio, the amount by which the compression ratio can be changed is smaller when the mechanical compression ratio is increased than when the mechanical compression ratio is decreased. In FIG. 40, the solid line I ₂ indicates the amount by which the compression ratio can be changed when the mechanical compression ratio is increased. In this case, the amount by which the compression ratio can be changed decreases as the intake air amount increases.

  On the other hand, FIG. 41 shows the relationship between the amount of mechanical compression ratio that can be changed in a certain period of time (amount that can be changed in compression ratio) and the actual compression ratio. FIG. 41 shows the compression ratio changeable amount when the intake air amount is a certain intake air amount.

As described above, the force acts in the direction of lowering the compression ratio due to the compression action of the piston 4 even during the fuel cut control, but this force also changes depending on the actual compression ratio as shown by the broken line in FIG. That is, as the actual compression ratio increases, the pressure in the combustion chamber 5 near the compression top dead center increases, and accordingly, the force acting in the direction of decreasing the compression ratio increases. Therefore, as in the case of the above-described intake air amount, if reducing the mechanical compression ratio, the compression ratio can be changed amount as the actual compression ratio increases as shown by the solid line J ₁ in FIG. 41 increases. On the other hand, when the mechanical compression ratio is increased, as shown by the solid line J ₂ in FIG. 41, the compression ratio changeable amount decreases as the actual compression ratio increases.

  In summary, during the fuel cut control, the relationship between the compression ratio changeable amount, the intake air amount, and the actual compression ratio is as shown in FIG. FIG. 42A shows a case where the mechanical compression ratio is lowered, and FIG. 42B shows a case where the mechanical compression ratio is increased. As shown in FIG. 42A, when the mechanical compression ratio is lowered, the amount of change in the compression ratio increases as the intake air amount increases and the actual compression ratio increases. On the other hand, as shown in FIG. 42B, when the mechanical compression ratio can be increased, the greater the intake air amount and the higher the actual compression ratio, the smaller the compression ratio changeable amount.

  Therefore, in the present embodiment, the relationship as shown in FIG. 42 is obtained in advance, and during the fuel cut control, without calculating the predicted value of the compression ratio changeable amount based on the engine load shown in FIG. Based on the relationship as shown in FIG. 42, the predicted value of the compression ratio changeable amount is calculated according to the intake air amount and the actual compression ratio. That is, in this embodiment, during the fuel cut control, the predicted value of the compression ratio changeable amount is changed according to the intake air amount and the actual compression ratio. Note that the predicted value of the compression ratio changeable amount does not necessarily have to be changed according to both the intake air amount and the actual compression ratio, and may be changed according to one of them.

  Further, during the normal operation in which the fuel cut control is not performed, the compression ratio changeable amount is affected by the combustion pressure and changes according to the rotation angle of the camshafts 54 and 55 as shown in FIG. On the other hand, when the fuel cut control is performed, similarly, the compression ratio changeable amount is affected by the compression action of the piston 4 and changes according to the rotation angle of the camshafts 54 and 55. This is shown in FIG.

FIG. 43 is a view similar to FIG. 18, and the alternate long and short dash line K _{0 in} FIG. 18 indicates when the compression action by the piston 4 does not occur (for example, when the intake valve 7 and the exhaust valve 9 remain open). ), And the broken line in the figure indicates the force acting in the direction of lowering the compression ratio by the compression action of the piston 4. Although the magnitude of this force is smaller than the combustion pressure in the case shown in FIG. 18, it acts similarly to the combustion pressure. Therefore, when the mechanical compression ratio is lowered, the compression ratio changeable amount increases as shown by the solid line K ₁ in FIG. 43, and when the mechanical compression ratio is increased, the actual amount as shown by the solid line K ₂ is increased. The compression ratio changeable amount becomes smaller. In addition, when the mechanical compression ratio is intermediate, the extent to which the compression ratio changeable amount when the mechanical compression ratio is reduced becomes large and the extent to which the compression ratio changeable amount becomes small when the mechanical compression ratio is increased becomes the largest. .

  Therefore, in the present embodiment, the relationship as shown in FIG. 43 is obtained in advance, and during the fuel cut control, the predicted value of the compression ratio changeable amount based on the rotation angle and engine load shown in FIG. 18 is calculated. Instead, the predicted value of the compression ratio changeable amount is calculated according to the rotation angle based on the relationship as shown in FIG. In particular, in the present embodiment, the prediction value of the compression ratio changeable amount calculated based on the relationship shown in FIG. 43 is corrected by the relationship shown in FIG. A value is calculated.

  In the above embodiment, even during the fuel cut control, the operating point is controlled in the same manner as in the examples shown in FIGS. 21 to 36, so that the required intake air amount can be promptly drawn even during the fuel cut control. It is possible to reach the required amount of air and to quickly reach the required operating point operating point that satisfies the required intake air amount. However, since the air-fuel mixture is not burned in the combustion chamber 5 during fuel cut control, the intake air amount does not necessarily reach the required intake air amount quickly. Therefore, during the fuel cut control, for example, the mechanical compression ratio and the intake valve closing timing may be controlled so that the operating point moves on the reference operation line W. As a result, the operating point is located on the reference operating line W when returning from the fuel cut control, so that a fuel-efficient driving can be performed when returning from the fuel cut control.

  FIG. 44 is a flowchart showing a routine for calculating a predicted value of the compression ratio changeable amount. The predicted value of the compression ratio changeable amount calculated by the routine shown in FIG. 44 is used in step 106 and step 110 of the routine shown in FIG. The routine shown in FIG. 44 is executed by interruption every predetermined time (for example, 8 msec).

  First, in step 201, it is determined whether or not the mechanical compression ratio is currently changed. When the mechanical compression ratio is not changed, it is not necessary to calculate a predicted value of the compression ratio changeable amount. Complete the processing cycle. In contrast, when the mechanical compression ratio is currently being changed, the routine proceeds to step 202. In step 202, it is determined whether or not the fuel cut control is currently in progress. When the fuel cut control is not being executed, the routine proceeds to step 203, where the compression ratio changeable amount is explained as described with reference to FIGS. The predicted value of is calculated.

That is, in step 203, it is determined whether or not the mechanical compression ratio is currently decreased. When the mechanical compression ratio is decreased, the basic predicted value and the correction coefficient are calculated in steps 204 to 206. Specifically, in step 204, the rotation angles of the camshafts 54 and 55 are generated using a map such as that indicated by G ₁ 'in FIG. 45A created based on the relationship indicated by G ₁ in FIG. Based on the basic prediction value R _base is calculated. In step 205, a correction coefficient R ₁ is calculated based on the engine load using a map such as that shown by F ₁ ′ in FIG. 45B created based on the relationship shown by F ₁ in FIG. . In step 206, the correction coefficient R ₂ on the basis of engine load / operation speed by using a map shown in FIG. 45 (C) that is created on the basis of the relationship shown in FIG. 19 is calculated.

On the other hand, when it is determined in step 203 that the mechanical compression ratio is increased, the basic predicted value and the correction coefficient are calculated in steps 207 to 209. Specifically, in step 207, the rotation angles of the camshafts 54 and 55 are generated using a map such as that shown by G ₂ 'in FIG. 45A created based on the relationship shown by G ₂ in FIG. Based on the basic prediction value R _base is calculated. In step 208, the correction coefficient R ₁ is calculated based on the engine load by using a map such as shown by F ₂ ′ in FIG. 45B created based on the relationship shown by F ₂ in FIG. . In step 209, the correction coefficient R ₂ on the basis of engine load / operation speed by using a map shown in FIG. 45 (C) that is created on the basis of the relationship shown in FIG. 19 is calculated.

  On the other hand, when it is determined in step 202 that the fuel cut control is being performed, the routine proceeds to step 210, where the predicted value of the compression ratio changeable amount is calculated as described with reference to FIGS.

That is, in step 210, it is determined whether or not the mechanical compression ratio is currently decreased. When the mechanical compression ratio is decreased, the basic predicted value and the correction coefficient are calculated from step 211 to step 213. Specifically, in step 211, the rotation angles of the camshafts 54 and 55 are generated using a map such as shown by K ₁ 'in FIG. 46A created based on the relationship shown by K ₁ in FIG. Based on the basic prediction value R _base is calculated. In step 212, the correction coefficient R ₁ is calculated based on the engine load using the map shown in FIG. 46B created based on the relationship shown in FIG. In step 213, when the combustion of the air-fuel mixture in the combustion chamber 5 is not performed, since no load on the bearing of the more variable compression ratio mechanism A becomes boundary lubrication, the correction coefficient R ₂ is a 1.

On the other hand, when it is determined in step 210 that the mechanical compression ratio is increased, the basic predicted value and the correction coefficient are calculated in steps 214 to 216. Specifically, in step 214, the rotation angles of the camshafts 54 and 55 are created using a map such as shown by K ₂ 'in FIG. 46A created based on the relationship shown by K ₂ in FIG. Based on the basic prediction value R _base is calculated. In step 215, the correction coefficient R ₁ is calculated based on the engine load using the map as shown in FIG. 46C created based on the relationship shown in FIG. 42B. In step 216, the correction coefficient R ₂ is set to 1 as in step 213.

Thus, in step 217 after the basic predictive value and the correction coefficient is calculated, it is calculated by the correction factor R ₁ and steps 206 etc. calculated in step 205 or the like to the basic predictive value R _base calculated in step 204 or the like By multiplying the correction coefficient R ₂ , the predicted value R of the compression ratio changeable amount is calculated (R = R _base · R ₁ · R ₂ ).

  In the routine shown in FIG. 44, correction based on the relationship shown in FIG. 20 is not performed, but the predicted value of the compression ratio changeable amount calculated by the routine shown in FIG. 44 is the relationship shown in FIG. Further correction may be made based on this.

DESCRIPTION OF SYMBOLS 1 Crankcase 2 Cylinder block 3 Cylinder head 4 Piston 5 Combustion chamber 7 Intake valve 17 Throttle valve A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (7)

A variable compression ratio mechanism capable of changing the mechanical compression ratio, and a variable valve timing mechanism capable of controlling the closing timing of the intake valve,
For each engine operating state, the operating point indicating the combination of the mechanical compression ratio and the intake valve closing timing is set as the required operating point to be reached, and when the engine operating state changes, the mechanical compression ratio and the intake valve closing timing Of the mechanical compression ratio that can be changed in a certain time and an intake valve that can be changed in a certain time so that the operating point indicating the combination of the engine changes toward the required operating point set for the engine operating state after the change In a spark ignition internal combustion engine that calculates a target operating point at regular intervals based on a predicted amount of valve closing timing, and changes the mechanical compression ratio and the intake valve closing timing toward the target operating point.
During the fuel cut control for stopping the fuel supply from the fuel injection valve, the predicted amount of the mechanical compression ratio that can be changed at the predetermined time depends on at least one of the current intake air amount and the current actual compression ratio. A spark-ignition internal combustion engine designed to change.   The predicted amount of the mechanical compression ratio that can be changed over a certain period of time is decreased as the current intake air amount is increased when the mechanical compression ratio is increased, and is increased as the current intake air amount is increased when the mechanical compression ratio is decreased. The spark ignition internal combustion engine according to claim 1.   The predicted amount of the mechanical compression ratio that can be changed over a certain period of time is decreased as the current actual compression ratio is increased when the mechanical compression ratio is increased, and is increased as the current actual compression ratio is increased when the mechanical compression ratio is decreased. The spark ignition internal combustion engine according to claim 1 or 2.   An intrusion prohibition area is set for the combination of the mechanical compression ratio and the intake valve closing timing, and the target operating point is the combination of the mechanical compression ratio and the intake valve closing timing when the engine operating state changes. The predicted amount of the mechanical compression ratio that can be changed over a certain period of time so that the operating point changes without entering the intrusion prohibited area toward the required operating point set for the engine operating state after the change. The spark ignition internal combustion engine according to any one of claims 1 to 3, wherein the spark ignition internal combustion engine is calculated based on a predicted amount of an intake valve closing timing that can be changed in time.   The target operating point is an operating point located within the predicted amount of the mechanical compression ratio that can be changed in a certain time from the current operating point or the previous target operating point and the predicted amount of the intake valve closing timing that can be changed in a certain time. The spark ignition internal combustion engine according to claim 4, wherein the operating point is the farthest from the current operating point within a range that does not enter the intrusion prohibited area.   The normal amount of the mechanical compression ratio that can be changed during the predetermined time is changed according to the engine load during the normal operation in which the fuel cut control is not performed. The spark ignition internal combustion engine described.   The variable compression ratio mechanism changes the mechanical compression ratio by changing the relative position between the crankcase and the cylinder block by a crank mechanism using a rotating cam, and sets the predicted amount of the mechanical compression ratio that can be changed at the predetermined time. The spark ignition internal combustion engine according to any one of claims 1 to 6, wherein the internal combustion engine is changed in accordance with a rotation angle of the cam.

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