JP2012097659A - Internal combustion engine with variable compression ratio mechanism - Google Patents

Abstract

PROBLEM TO BE SOLVED: To prevent the deterioration of convergence to a target idling engine speed caused by the feedback control of ignition timing even if a target mechanical compression ratio is not achieved in an idling operation, in an internal combustion engine with a variable compression ratio mechanism.SOLUTION: The internal combustion engine controls the variable compression ratio mechanism in the idling operation so as to achieve the target mechanical compression ratio, and performs the feedback control of the ignition timing (step 113) so as to achieve the target idling engine speed. When the target mechanical compression ratio is not achieved in the idling operation, the internal combustion engine varies a control amount of the feedback control on the basis of a current mechanical compression ratio (steps 110-112).

Description

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

  An internal combustion engine having a variable compression ratio mechanism is known (see Patent Document 1). In an internal combustion engine having such a variable compression ratio mechanism, each target mechanical compression ratio ((top dead center combustion) is set for each engine load. Chamber volume + stroke volume) / top dead center combustion chamber volume). Thereby, even during the idling operation, the target mechanical compression ratio is usually realized, and the ignition timing feedback control for realizing the target idling speed is performed.

JP2007-303423 JP2009-062886

  In the internal combustion engine having the above-described variable compression ratio mechanism, the target mechanical compression ratio may not be realized during idle operation for some reason. In this case, the convergence to the target idle speed by feedback control of the ignition timing is not achieved. In the worst case, an engine stoppage may occur.

  Therefore, an object of the present invention is to suppress deterioration in convergence to the target idle speed by feedback control of the ignition timing in an internal combustion engine having a variable compression ratio mechanism even if the target mechanical compression ratio is not realized during idle operation. It is to be.

  An internal combustion engine comprising the variable compression ratio mechanism according to the first aspect of the present invention controls the variable compression ratio mechanism so that the target mechanical compression ratio is realized during idle operation, and the target idle speed is realized. In the internal combustion engine that performs feedback control of ignition timing at the same time, when the target mechanical compression ratio is not realized during idle operation, the control amount of the feedback control is changed based on the current mechanical compression ratio. To do.

  An internal combustion engine comprising the variable compression ratio mechanism according to claim 2 according to the present invention is an internal combustion engine comprising the variable compression ratio mechanism according to claim 1, wherein the current mechanical compression ratio is a current mechanical compression for each cylinder. The control amount of the feedback control for each cylinder is changed so that the generated torque of each cylinder is equal.

  According to the internal combustion engine having the variable compression ratio mechanism according to the first aspect of the present invention, the variable compression ratio mechanism is controlled so that the target mechanical compression ratio is realized during the idling operation, and the target idle rotation speed is realized. Thus, in an internal combustion engine that performs feedback control of ignition timing, when the target mechanical compression ratio is not realized during idle operation, the control amount of feedback control is changed based on the current mechanical compression ratio. Yes. As a result, the higher the mechanical compression ratio, the greater the change in the generated torque with respect to the change in the ignition timing.Therefore, when the target mechanical compression ratio is not achieved, the control amount of the ignition timing feedback control is constant. Convergence to the target idle speed may deteriorate due to large or small changes in the generated torque, but the amount of feedback control should be set appropriately based on the current mechanical compression ratio. By changing, the deterioration of the convergence can be suppressed.

  According to an internal combustion engine comprising the variable compression ratio mechanism according to claim 2 of the present invention, in the internal combustion engine comprising the variable compression ratio mechanism according to claim 1, the current mechanical compression ratio is the current machine for each cylinder. This is the compression ratio, and the control amount of feedback control for each cylinder is changed so that the torque generated in each cylinder is equal. Thereby, in the case where the mechanical compression ratio for each cylinder differs due to a difference in the deposit amount or the like, it is possible to suppress deterioration in convergence to the target idle speed while suppressing torque fluctuation.

1 is an overall view of an internal combustion engine. It is a disassembled perspective view of a variable compression ratio mechanism. 1 is a schematic side sectional view of an internal combustion engine. It is a figure which shows a variable valve timing mechanism. It is a figure which shows the lift amount of an intake valve and an exhaust valve. It is a figure for demonstrating a mechanical compression ratio, an actual compression ratio, and an expansion ratio. It is a figure which shows the relationship between theoretical thermal efficiency and an expansion ratio. It is a figure for demonstrating a normal cycle and a super-high expansion ratio cycle. It is a figure which shows changes, such as a mechanical compression ratio according to an engine load. It is a flowchart which shows the feedback control of the ignition timing for implement | achieving a target idle speed. It is a graph which shows the relationship between the ignition timing for every mechanical compression ratio, and generated torque. It is a graph which shows the relationship between the ignition timing for every mechanical compression ratio, and generated torque.

  FIG. 1 is a side sectional view of an internal combustion engine having a variable compression ratio mechanism according to the present invention. Referring to FIG. 1, 1 is a crankcase, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is a spark plug disposed at the center of the top surface of the combustion chamber 5, and 7 is intake air. 8 is an intake port, 9 is an exhaust valve, and 10 is an exhaust port. The intake port 8 is connected to 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 intake air amount detector 18 using, for example, heat rays are arranged in the intake duct 14. On the other hand, the exhaust port 10 is connected to a catalyst device 20 containing, for example, a three-way catalyst via an exhaust manifold 19, and an air-fuel ratio sensor 21 is disposed in the exhaust manifold 19.

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

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

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

  2 is an exploded perspective view of the variable compression ratio mechanism A shown in FIG. 1, and FIG. 3 is a side sectional view of the internal combustion engine schematically shown. Referring to FIG. 2, a plurality of protrusions 50 spaced 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 protrusion 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 spaced apart from 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, on both sides of each circular cam 58, as shown in FIG. 3, an eccentric shaft 57 eccentrically arranged with respect to the rotation axis of each camshaft 54, 55 extends. 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 representing 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. 3A, the eccentric shafts 57 are separated from each other. In order to move in the direction, the circular cam 56 rotates in the opposite direction to the circular cam 58 in the cam insertion hole 51, and as shown in FIG. 3B, the position of the eccentric shaft 57 is changed from the high position to the intermediate height position. It becomes. 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 positioned at the compression top dead center. Therefore, by rotating the camshafts 54 and 55, the piston 4 is compressed at the top dead center. The volume of the combustion chamber 5 when it is located at can be changed.

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

  On the other hand, FIG. 4 shows the variable valve timing mechanism B attached to the end of the camshaft 70 for driving the intake valve 7 in FIG. Referring to FIG. 4, the variable valve timing mechanism B includes a timing pulley 71 that is rotated in the direction of an arrow by a crankshaft of an engine via a timing belt, a cylindrical housing 72 that rotates together with the timing pulley 71, an intake valve A rotating shaft 73 that rotates together with the driving camshaft 70 and is rotatable relative to the cylindrical housing 72, and a plurality of partition walls 74 that extend from the inner peripheral surface of the cylindrical housing 72 to the outer peripheral surface of the rotating shaft 73. And a vane 75 extending from the outer peripheral surface of the rotating shaft 73 to the inner peripheral surface of the cylindrical housing 72 between the partition walls 74, and an advance hydraulic chamber 76 on each side of each vane 75. A retarding hydraulic chamber 77 is formed.

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

  When the cam phase of the intake valve driving camshaft 70 should be advanced, the spool valve 85 is moved to the right in FIG. 4 and the hydraulic oil supplied from the supply port 82 is advanced via the hydraulic port 79. The hydraulic oil in the retard hydraulic chamber 77 is discharged from the drain port 84 while being supplied to the hydraulic chamber 76. At this time, the rotary shaft 73 is rotated relative to the cylindrical housing 72 in the direction of the arrow.

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

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

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

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

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

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

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

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

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

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

  The solid line in FIG. 7 shows the change in theoretical thermal efficiency when the actual compression ratio and the expansion ratio are substantially equal, that is, in a normal cycle. In this case, it can be seen that the theoretical thermal efficiency increases as the expansion ratio increases, that is, as the actual compression ratio increases. Therefore, in order to increase the theoretical thermal efficiency in a normal cycle, it is only necessary to increase the actual compression ratio. However, the actual compression ratio can only be increased to a maximum of about 12 due to the restriction of the occurrence of knocking at the time of engine high load operation, and thus the theoretical thermal efficiency cannot be sufficiently increased in a normal cycle.

  On the other hand, under such circumstances, it is considered to increase the theoretical thermal efficiency while strictly dividing the mechanical compression ratio and the actual compression ratio. As a result, the theoretical thermal efficiency is governed by the expansion ratio, and the actual compression ratio is compared to the theoretical thermal efficiency. Was found to have little effect. That is, if the actual compression ratio is increased, the explosive force is increased, but a large amount of energy is required for compression. Thus, even if the actual compression ratio is increased, the theoretical thermal efficiency is hardly increased.

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

  Thus, if the actual compression ratio is maintained at a low value, knocking does not occur. Therefore, if the expansion ratio is increased while the actual compression ratio is maintained at a low value, the theoretical thermal efficiency is reduced while preventing the occurrence of knocking. Can be greatly increased. FIG. 8B shows an example where the variable compression ratio mechanism A and variable valve timing mechanism B are used to increase the expansion ratio while maintaining the actual compression ratio at a low value.

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

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

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

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

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

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

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

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

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

  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. Accordingly, when expressing the case shown in FIG. 9 so as to include both the case indicated by the solid line and the case indicated by the broken line, in the embodiment according to the present invention, the valve closing timing of the intake valve 7 becomes smaller as the engine load becomes lower. It is moved in a direction away from the intake bottom dead center BDC until the limit valve closing timing L1 at which the intake air amount supplied into the combustion chamber can be controlled. Thus, 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.

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

  By the way, since the load is low during idle operation, the variable compression ratio mechanism A is controlled with the upper limit mechanical compression ratio as the target mechanical compression ratio. On the other hand, during idle operation, feedback control of ignition timing is performed to make the engine speed coincide with the target idle speed. FIG. 10 is a flowchart showing the feedback control of the ignition timing during the idling operation performed in the internal combustion engine having the variable compression ratio mechanism according to the present invention, and is implemented by the electronic control unit 30.

  First, in step 101, it is determined whether or not it is during idling. For example, if it is detected by the load sensor 41 that the accelerator pedal 40 is released (depression amount L = 0) while the vehicle is stopped, it can be determined that the vehicle is idling. If it is not during idling, the determination in step 101 is negative, and in step 102, the previous mechanical compression ratio E 'during idling, which will be described in detail later, is set to 0, and the process ends.

  On the other hand, when the engine is idling, the determination in step 101 is affirmed, and in step 103, the current mechanical compression ratio E is detected based on the output of the relative position sensor 22, for example. Next, at step 104, it is determined whether or not the current mechanical compression ratio E and the previous mechanical compression ratio E 'are substantially equal. At the beginning of idling, since the previous mechanical compression ratio E ′ is set to 0 in step 102, the determination in step 104 is denied, and in step 105, the reference ignition timing with respect to the current mechanical compression ratio E is determined. ITt is determined. The reference ignition timing ITt is an ignition timing for realizing the target idle speed by generating the reference torque Tr when the desired air amount during the idling operation is realized.

  FIG. 11 is a graph showing the relationship between the ignition timing and the generated torque when the desired air amount is achieved during idle operation, and the solid line indicates the target mechanical compression ratio (for example, the upper limit mechanical compression ratio) during idle operation. The dotted line is a case where the specific mechanical compression ratio is lower than the target mechanical compression ratio. The actual compression ratio in the case of the specific mechanical compression ratio is lower than the actual compression ratio in the case of the target mechanical compression ratio because the intake air amount is the same. In FIG. 11, the reference torque Tt is a torque necessary to realize the target idle speed Nt as described above.

  As shown in FIG. 11, the reference ignition timing corresponding to the reference torque Tt when the upper limit mechanical compression ratio that is the target mechanical compression ratio Et during idle operation is realized is ITA, which is lower than the target mechanical compression ratio Et. The reference ignition timing corresponding to the reference torque Tt when the specific mechanical compression ratio is obtained is ITB. Thus, like the two mechanical compression ratios illustrated in FIG. The graphs showing the relationship with the generated torque are different, the reference ignition timing ITt corresponding to the reference torque Tt differs for each mechanical compression ratio, and the lower the mechanical compression ratio, the reference ignition timing ITt corresponding to the reference torque Tt Become. The reference ignition timing ITt can be mapped to the mechanical compression ratio E.

  Accordingly, in step 105, a reference ignition timing ITt for the current mechanical compression ratio E is determined by using a map or the like, and in step 106, the ignition timing is set to the reference ignition timing ITt determined in step 105. Next, at step 107, the current engine speed N is detected using the crank angle sensor 42. Next, at step 108, a rotational speed deviation ΔN between the target idle rotational speed Nt and the current engine rotational speed N is calculated. Next, at step 109, it is determined whether or not the rotational speed deviation ΔN calculated at step 108 is substantially zero. If the target idle speed Nt is realized as a result of setting the ignition timing IT to the reference ignition timing ITt with respect to the current mechanical compression ratio E, the determination in step 109 is affirmed and it is not necessary to change the ignition timing. Accordingly, in step 114, the current mechanical compression ratio E detected in step 102 is ended as the previous mechanical compression ratio E '.

  On the other hand, when the determination in step 109 is negative, the target idle speed Nt is not realized, and in step 110, the control amount ΔIT that changes the ignition timing IT is set. When the rotational speed deviation ΔN is positive, the control amount ΔIT must advance the ignition timing IT to increase the current engine rotational speed N, and the control amount ΔIT becomes negative. Further, when the rotational speed deviation ΔN is negative, the control amount ΔIT must retard the ignition timing IT to lower the current engine rotational speed N, and the control amount ΔIT becomes positive. The absolute value of the control amount ΔIT may be a constant value regardless of the absolute value of the rotation speed deviation ΔN, but it is preferable to increase the absolute value of the rotation speed deviation ΔN as the absolute value of the rotation speed deviation ΔN increases. As a result, the larger the engine speed deviation ΔN, the greater the ignition timing IT can be changed, and the engine engine speed N can be converged to the target idle engine speed Nt earlier.

  As shown in FIG. 11, when the upper limit mechanical compression ratio that is the target mechanical compression ratio Et during idle operation is realized, compared to when the specific mechanical compression ratio is lower than the target mechanical compression ratio Et, The change in the generated torque with respect to the change in the ignition timing (advance angle or retard angle) is large. The lower the mechanical compression ratio, the smaller the change in the generated torque with respect to the change in the ignition timing. As a result, when the mechanical compression ratio is lower than the target mechanical compression ratio Et, the control amount ΔIT set on the assumption that the target mechanical compression ratio Et is realized is used as it is, and the ignition timing feedback control is performed. If it is implemented, the generated torque can only be changed small, and the convergence to the target idle speed Nt will deteriorate. Accordingly, it is necessary to appropriately change the feedback control amount ΔIT based on the current mechanical compression ratio.

  In this flowchart, in step 111, a compression ratio deviation ΔE between the target mechanical compression ratio Et during idle operation and the current mechanical compression ratio E detected in step 103 is calculated. In the case of the present embodiment, since the target mechanical compression ratio Et during idle operation is the upper limit mechanical compression ratio, the calculated compression ratio deviation ΔE does not become negative. If the target mechanical compression ratio Et is a mechanical compression ratio lower than the upper limit mechanical compression ratio, the compression ratio deviation ΔE calculated in step 111 may be negative.

  For example, when the operating state changes from engine high load operation to idle operation and the mechanical compression ratio is being changed, or when the variable compression ratio mechanism A is broken, the target mechanical compression ratio Et during idle operation is realized. In the case where the compression ratio deviation ΔE cannot be performed, the compression ratio deviation ΔE is calculated as a positive value in Step 111. In Step 112, the correction amount ΔIT determined in Step 110 based on the compression ratio deviation ΔE is changed by the coefficient k. To do. Next, at step 113, the ignition timing IT is corrected by the changed correction amount ΔIT.

  When the compression ratio deviation ΔE is positive, the coefficient k is such that the change in the generated torque with respect to the change in the ignition timing is too small. Therefore, the correction amount ΔIT is increased as a positive value larger than 1, and the ignition timing control in step 113 is performed. To improve the convergence to the target idle speed. In addition, when the compression ratio deviation ΔE is negative, the coefficient k is a positive value smaller than 1 and the correction amount ΔIT is reduced because the change in the generated torque with respect to the change in the ignition timing is too large. In the ignition timing control, the engine speed is not changed excessively so that the engine is not stopped or the engine speed is not diverged. Here, the coefficient k is preferably increased as the compression ratio deviation ΔE is larger (the compression ratio deviation ΔE is negative and smaller (the larger the absolute value is)), the coefficient k is decreased in a positive number smaller than 1. ).

  Thus, after the ignition timing is controlled in step 113, in step 114, the current mechanical compression ratio E detected in step 103 is ended as the previous mechanical compression ratio E '. If the mechanical compression ratio is being changed during idle operation, the determination in step 104 is again denied, and in step 105, the reference ignition timing ITt for the current mechanical compression ratio E is newly determined. The ignition timing IT is set as a new reference ignition timing ITt. Thereby, when the mechanical compression ratio changes greatly, the convergence to the target idle speed Nt can be improved. Of course, if the mechanical compression ratio does not change during idling, the determination in step 104 is affirmed, and the ignition timing is fed back after step 107 without returning the ignition timing to the reference ignition timing for the current mechanical compression ratio E. Control is implemented.

  By the way, in the flowchart shown in FIG. 10, the current mechanical compression ratio E detected based on the output of the relative position sensor 22 in step 103 is a mechanical compression ratio common to each cylinder in design. However, in reality, deposits may adhere to each cylinder, and the design combustion chamber volume when the piston is located at the compression top dead center may be changed. The actual mechanical compression ratio may vary from cylinder to cylinder because the deposit amount varies from cylinder to cylinder. Hereinafter, feedback control of the ignition timing in consideration of the difference in mechanical compression ratio for each cylinder will be described.

An in-cylinder pressure sensor is disposed in each cylinder. For example, the first in-cylinder pressure P1 and the second in-cylinder are set for each cylinder at the first crank angle and the second crank angle from the intake valve closing to the ignition timing. If the pressure P2 is measured, the following equation (1) is established as adiabatic compression of the air-fuel mixture in the cylinder.
P1 * (V1 + VC) ^k = P2 * (V2 + VC) ^k (1)

Here, VC is an actual combustion chamber volume narrowed by a deposit or the like, V1 is a piston stroke volume from the first crank angle to the compression top dead center, and V2 is from the second crank angle to the compression top dead center. Piston stroke volume, and k is the specific heat ratio of the air-fuel mixture in the cylinder.
Since VC = (P2 ^k * V1−P1 ^k * V2) / (P1 ^k −P2 ^k ) is derived from the above equation (1), and P1, P2, V1, V2, and k are all known, the cylinder The actual combustion chamber volume VC for each can be calculated.

  Thereby, instead of detecting the mechanical compression ratio E based on the output of the relative position sensor 22 in step 103 of the flowchart described above, for example, in the case of a four-cylinder internal combustion engine, # 1 cylinder, # 2 cylinder, # The actual mechanical compression ratios E1, E2, E3, and E4 of the three cylinders and the # 4 cylinder can be calculated. As described with reference to FIG. 11, the reference ignition timing ITt corresponding to the reference torque Tt for realizing the target idle speed Nt is different for each mechanical compression ratio. Reference ignition timings IT1, IT2, IT3, IT4 corresponding to the reference torque Tt as shown in FIG. 12 are determined for the compression ratios E1, E2, E3, E4.

  Next, similarly to step 106 in the flowchart described above, ignition is performed at the reference ignition timing IT1, IT2, IT3, IT4 for each cylinder. Thereby, the generated torque of each cylinder becomes substantially equal, and the torque fluctuation can be sufficiently suppressed. Even when the ignition timing of each cylinder is set to the reference ignition timing IT1, IT2, IT3, IT4, the target idle speed Nt cannot be realized, and the speed deviation ΔN with respect to the current engine speed N is not substantially zero. Actually, the reference torque Tt is not generated in each cylinder, and the correction amount ΔT of the generated torque T is determined instead of determining the control amount ΔIT of the feedback control in step 110 of the above-described flowchart.

  When the rotational speed deviation ΔN is positive, the correction amount ΔT of the generated torque T must be increased to increase the current engine speed N, and the correction amount ΔT becomes positive. Further, when the rotational speed deviation ΔN is negative, the correction amount ΔT must be reduced by reducing the generated torque T and the current engine rotational speed N, and the correction amount ΔT becomes negative. The absolute value of the correction amount ΔT may be a constant value regardless of the absolute value of the rotational speed deviation ΔN, but it is preferable to increase the absolute value of the rotational speed deviation ΔN as the absolute value of the rotational speed deviation ΔN increases. As a result, the larger the rotational speed deviation ΔN is, the larger the generated torque T is changed, and the engine rotational speed N can be quickly converged to the target idle rotational speed Nt.

  In order to generate the torque corrected by the correction amount ΔT, in step 113, the ignition timing determined for the actual mechanical compression ratio is determined for each cylinder in the same manner as in step 113 of the flowchart described above. Ignition is performed. As shown in FIG. 12, for example, when the correction amount is positive ΔT ′, the ignition timing of each cylinder is IT1 ′, IT2 ′, IT3 ′, IT4 ′, for example, ΔT with a negative correction amount. ", The ignition timing of each cylinder is IT1", IT2 ", IT3", IT4 ". Thus, the feedback control amount of the ignition timing of each cylinder is not the same, Based on the current mechanical compression ratio, the torque generated in each cylinder has been changed to be equal, and when the mechanical compression ratio for each cylinder differs due to differences in the amount of deposit, etc., the target idle In order to set the ignition timing in this way, not only the four mechanical compression ratios shown in FIG. It is necessary to set in advance a graph showing the relationship with the generated torque.

  When the correction amount of the generated torque is ΔT ′ (positive), the feedback control amount of the ignition timing is (IT1-IT1 ′) for the # 1 cylinder, (IT2-IT2 ′) for the # 2 cylinder, (IT3-IT3 ′) and # 4 cylinder becomes (IT4-IT4 ′), which are different from each other, but both are negative values, and the ignition timing is advanced in each cylinder. Further, when the correction amount of the generated torque is ΔT ″ (negative), the feedback control amount of the ignition timing is (IT1−IT1 ″) for the # 1 cylinder, (IT2−IT2 ″) for the # 2 cylinder, and # 3 The cylinder is (IT3-IT3 ") and the # 4 cylinder is (IT4-IT4"), which are different from each other, but both are positive values, and the ignition timing is retarded in each cylinder.

  In this way, the ignition timing feedback control is performed so that the generated torque T is corrected by the same correction amount ΔT in each cylinder until the rotational speed deviation ΔN becomes substantially zero.

  If the variable compression ratio mechanism A functions normally and the designed target mechanical compression ratio is realized during idle operation, the actual mechanical compression ratio of each cylinder is reduced by the deposit adhesion. When E1, E2, E3, and E4 are higher than the target mechanical compression ratio, the variable compression ratio mechanism A reduces the mechanical compression ratio, and as a result, the average value of the actual mechanical compression ratio of each cylinder that decreases ( E1 ′ + E2 ′ + E3 ′ + E4 ′) / 4 is preferably the target mechanical compression ratio.

  Thus, when the actual mechanical compression ratio of each cylinder is changed during the idling operation, initially, the ignition timing of each cylinder is generated and the reference torque Tt is generated with respect to the actual mechanical compression ratio of each cylinder. It is preferable that the control to be the reference ignition timing ITt is performed.

1 Crankcase 2 Cylinder block 3 Cylinder head 30 Electronic control unit A Variable compression ratio mechanism

Claims (2)

  In an internal combustion engine that controls a variable compression ratio mechanism so that a target machine compression ratio is realized during idle operation and performs feedback control of ignition timing so as to achieve a target idle speed, the target machine is controlled during idle operation. An internal combustion engine comprising a variable compression ratio mechanism, wherein a control amount of the feedback control is changed based on a current mechanical compression ratio when a compression ratio is not realized.   The current mechanical compression ratio is a current mechanical compression ratio for each cylinder, and the control amount of the feedback control for each cylinder is changed so as to make the generated torque of each cylinder equal. An internal combustion engine comprising the variable compression ratio mechanism according to 1.

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