JP5472136B2 - Spark ignition internal combustion engine - Google Patents

Description

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

  Even if the amount of fuel supplied to each cylinder is uniformly controlled to the desired amount of fuel and the total intake amount of all cylinders is controlled to the desired all-cylinder intake amount by using an air flow meter or the like, it is supplied to each cylinder. If the intake air amount cannot be uniformly controlled to the desired intake air amount, the desired air-fuel ratio can be achieved as the average air-fuel ratio of all the cylinders, but the intended combustion with the desired air-fuel ratio is achieved in at least two cylinders. Cannot be realized.

  In a spark internal combustion engine equipped with a variable compression ratio mechanism, it is possible to make the combustion chamber volume of each cylinder equal by forming a built-up portion on the piston top surface of each cylinder and deleting a part of the built-up portion of each cylinder. It has been proposed (see Patent Document 1).

JP2009-008016 JP 09-072809 JP 2009-281236 A JP2007-231798 JP 2004-52620 A

  Even if the combustion chamber volumes of the respective cylinders are made equal as described above, the intake air amount of each cylinder will not be uniform if a slight deviation occurs in the intake valve closing timing of each cylinder. Even if it is attempted to estimate the intake air amount of each cylinder from the generated torque of each cylinder, it is difficult to detect a slight difference in the intake air amount for each cylinder.

  Accordingly, an object of the present invention is to make it possible to detect a variation in the intake air amount of each cylinder even in a slight difference in a spark internal combustion engine having a variable compression ratio mechanism.

  The spark ignition internal combustion engine according to claim 1 of the present invention is a spark ignition internal combustion engine having a variable compression ratio mechanism, and when detecting a variation in the intake air amount of each cylinder, the intake air amount of each cylinder is changed. The mechanical compression ratio is increased by the variable compression ratio mechanism, the ignition timing is retarded, and a value based on the torque generated in each cylinder is measured.

  A spark ignition internal combustion engine according to a second aspect of the present invention is characterized in that, in the spark ignition internal combustion engine according to the first aspect, variation in intake air amount of each cylinder is detected during idling operation.

  According to the spark ignition internal combustion engine of the first aspect of the present invention, in the spark ignition internal combustion engine having the variable compression ratio mechanism, the ignition timing is retarded when detecting the variation in the intake amount of each cylinder. As a result, the difference in the torque generated by each cylinder is enlarged, and even if there is a slight difference, it is possible to detect the variation in the intake air amount of each cylinder. However, if the ignition timing is simply retarded, the air-fuel mixture cannot be ignited and burned well. Therefore, the actual compression ratio can be increased by increasing the mechanical compression ratio by the variable compression ratio mechanism without changing the intake air amount of each cylinder. Therefore, even if the ignition timing is retarded, the air-fuel mixture is ignited and burned well.

  According to the spark ignition internal combustion engine according to claim 2 of the present invention, in the spark ignition internal combustion engine according to claim 1, the time between the expansion strokes becomes longer as the engine speed is lower, and the generated torque of each cylinder is easily detected. Therefore, the variation in the intake air amount of each cylinder is detected during the idling operation.

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 for detecting the dispersion | variation in the intake air amount of each cylinder. It is a graph which shows the relationship between ignition timing and the generated torque of one cylinder. 6 is a map for calculating an intake air amount deviation rate.

  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 apart from each other, that is, cylinder block side supports, are formed below both side walls of the cylinder block 2, and each protrusion 50 has a circular cross section. The cam insertion hole 51 is formed. On the other hand, on the upper wall surface of the crankcase 1, there are formed a plurality of protrusions 52 that are fitted between the corresponding protrusions 50 spaced apart from each other, that is, the crankcase side support. A cam insertion hole 53 having a circular cross section is also formed in each protrusion 52.

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

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

  3A, 3B, and 3C show the positional relationship between the center a of the concentric portion 58, the center b of the eccentric portion 57, and the center c of the circular cam 56 in each state. It is shown.

  3A to 3C, the relative position of the crankcase 1 and the cylinder block 2 is determined by the distance between the center a of the concentric part 58 and the center c of the circular cam 56, and the concentric part. 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 the theoretical thermal efficiency when the actual compression ratio and the expansion ratio are substantially equal, that is, in a normal cycle. In this case, it can be seen that the theoretical thermal efficiency increases as the expansion ratio increases, that is, as the actual compression ratio increases. Therefore, in order to increase the theoretical thermal efficiency in a normal cycle, it is only necessary to increase the actual compression ratio. However, the actual compression ratio can only be increased to a maximum of about 12 due to the restriction of the occurrence of knocking at the time of engine high load operation, and thus the theoretical thermal efficiency cannot be sufficiently increased in a normal cycle.

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

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

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

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

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

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

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

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

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

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

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

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

  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, the total amount of intake air supplied to all cylinders can be detected by the intake air amount detector 18, and the closing timing of the intake valve 7 is controlled for each engine operating state (engine speed and engine load). Alternatively, if a desired amount of fuel is supplied by the fuel injection valve 13 of each cylinder by realizing a desired total intake amount by controlling the opening degree of the throttle valve 17, on the downstream side of the exhaust merging portion of the exhaust manifold 19 The air-fuel ratio of the exhaust gas detected by the air-fuel ratio sensor 21 can be controlled to a desired air-fuel ratio (for example, the theoretical air-fuel ratio).

  However, the desired total intake air amount for each engine operating state is not always uniformly distributed to each cylinder, and the intake air amount may vary among the cylinders. As a result, the desired intake air amount for at least two cylinders may vary. The intended combustion of the air / fuel ratio may not be realized. In particular, when the valve closing timing of the intake valve 7 of each cylinder is controlled by the variable valve timing mechanism B as in the present embodiment, the intake amount of each cylinder is reduced by a slight deviation in the valve closing timing of the intake valve. It will vary. Further, as in the present embodiment, when the mechanical compression ratio is increased by the variable compression ratio mechanism A and the combustion chamber volume at the top dead center is decreased, it is possible to cope with a slight shift in the closing timing of the intake valve 7. The variation in the intake air amount of each cylinder becomes remarkable.

  In each cylinder, the amount of fuel supplied from the fuel injection valve 13 into the cylinder can be easily controlled. Therefore, it can be assumed that a desired amount of fuel is supplied to each cylinder. The generated torque increases in a large number of cylinders, and the generated torque decreases in a cylinder whose intake air amount is smaller than a desired intake air amount. As a result, it is conceivable to detect variations in the intake air amount of each cylinder by detecting the generated torque of each cylinder, but if the intake air amount of each cylinder is only slightly different, the generated torque of each cylinder However, there is a case where the variation in the intake air amount of each cylinder cannot be detected simply by detecting the torque generated by each cylinder.

  In the spark ignition internal combustion engine of this embodiment, the electronic control unit 30 controls the variable compression ratio mechanism A and the spark plug 6 according to the flowchart shown in FIG. .

  First, in step 101, it is determined whether or not the current engine operation state is idle operation. For example, when the vehicle is stopped, the depression amount of the accelerator pedal 40 detected by the load sensor 41 is zero (0), that is, the accelerator pedal is not depressed, and the engine speed is detected by the crank angle sensor 42. Thus, it can be determined that the vehicle is idling. When this determination is denied, the process is terminated as it is, but when the idling operation is being performed, the determination at step 101 is affirmed and the process proceeds to step 102.

  In step 102, it is determined whether or not the flag F is 1. The flag F is set to 1 when detecting the variation in the intake air amount of each cylinder. For example, the flag F is set to 1 only once every engine operation or each time a plurality of engine operations are performed. In idling immediately after engine startup, combustion is unstable, so it is difficult to accurately detect variations in the intake air amount of each cylinder, and it is preferable not to set the flag F to 1. If the determination in step 102 is negative, the processing is terminated as it is. However, when the variation in the intake air amount of each cylinder is detected, the determination in step 102 is affirmed, and in step 103, the target mechanical compression ratio Et is set to E1.

  During idle operation where combustion is stable, not immediately after engine startup, the closing timing of the intake valve is retarded to the maximum and the throttle valve is set to the minimum opening (see FIG. 9). Further, the actual compression ratio during the idling operation is set lower than the constant value shown in FIG. 9 by setting the mechanical compression ratio to an idle mechanical compression ratio E0 smaller than the upper limit value in order to suppress vibration.

  The mechanical compression ratio E1 determined as the target mechanical compression ratio Et in step 103 is a mechanical compression ratio higher than the idle mechanical compression ratio E0, and is preferably an upper limit value of the mechanical compression ratio. Next, in step 104, the variable compression ratio mechanism A is controlled to increase the mechanical compression ratio E. At this time, since the intake air amount is not changed in each cylinder and the actual compression ratio is increased by increasing the mechanical compression ratio E, the ignition timing is delayed so that the air-fuel mixture can be ignited and burned well in each cylinder. The angle limit expands to the retard side. In step 105, the ignition timing IT of each cylinder is controlled so as to be a retard limit with respect to the current actual compression ratio. Next, at step 106, it is determined whether or not the mechanical compression ratio E has reached the target mechanical compression ratio Et (E1), and the processing of steps 104 and 105 is repeated until this determination is affirmed.

  FIG. 11 is a graph showing the relationship between the ignition timing IT during idle operation and the torque generated by one cylinder. The solid line indicates the case where the desired intake air amount during idle operation is supplied into the cylinder, and the dotted line indicates the cylinder. This is a case where intake air slightly larger than the desired intake air amount during idle operation is supplied into the cylinder, and the alternate long and short dash line is when intake air slightly smaller than the desired intake air amount during idle operation is supplied into the cylinder. In any of the solid line, the dotted line, and the alternate long and short dash line, a desired fuel amount during idling is supplied into the cylinder. When the desired fuel amount is supplied, the generated torque increases as the intake air amount supplied to the cylinder increases.

  IT0 is the optimum ignition timing common to all cylinders during idle operation, and is not only based on the cylinder supplied with the intake air of the desired intake amount during idle operation indicated by the solid line, but also from the desired intake amount during idle operation indicated by the dotted line. Even in a cylinder to which a slightly larger amount of intake air is supplied, even in a cylinder in which an intake air slightly smaller than the desired intake air amount during idle operation is supplied into the cylinder indicated by the alternate long and short dash line, the generated torque at the optimal ignition timing IT0 Nearly maximum. As shown in FIG. 11, when the ignition timing is controlled to the optimum ignition timing IT0, if the difference in the intake air amount of each cylinder is slight, the generated torque of each cylinder is almost the same, and such each cylinder has no difference. It is difficult to accurately detect a slight generated torque difference.

  However, as shown in FIG. 11, as the ignition timing is retarded from the optimal ignition timing IT0, even if the difference in the intake air amount between the cylinders is small, the difference in the generated torque of each cylinder becomes remarkable. It is easy to accurately detect the difference in torque generated. Therefore, in the flowchart shown in FIG. 10, the ignition timing is retarded in step 105, and the mechanical compression ratio is increased in step 104 in order to reliably ignite and burn the air-fuel mixture even if the ignition timing is retarded. This increases the actual compression ratio.

  When the determination in step 106 is affirmative, the mechanical compression ratio E is increased to the target mechanical compression ratio Et, the ignition timing IT is retarded to the retard limit IT1 at this time, and the generated torque difference of each cylinder is It has been enlarged. Thus, in step 107, the TDC time Ti of each cylinder is detected as a value based on the torque generated by each cylinder. For example, when the ignition order is # 1, # 3, # 4, and # 2 cylinders, the TDC time T1 of the # 1 cylinder is from the compression top dead center of the # 1 cylinder to the compression top dead center of the # 3 cylinder. # 2 cylinder TDC time T2 is the time from compression top dead center of # 2 cylinder to # 1 cylinder compression top dead center, and # 3 cylinder TDC time T3 is # 3 cylinder From the compression top dead center of the cylinder # 4 to the compression top dead center of the cylinder # 4. The TDC time T4 of the cylinder # 4 is the time from the compression top dead center of the cylinder # 4 to the compression top dead center of the cylinder # 2. is there.

  The shorter the TDC time Ti, the greater the generated torque. Based on the TDC time Ti of each cylinder, the TDC time ratio ΔTi of each cylinder is calculated. The # 1 cylinder TDC time ratio ΔT1 is calculated by (average # 1 cylinder TDC time T1-TDC time average) / TDC time average, and # 2 cylinder TDC time ratio ΔT2 is (# 2 cylinder TDC time T2- The TDC time ratio ΔT3 of the # 3 cylinder is calculated by (TDC time T3−TDC time average of the # 3 cylinder) / TDC time average, and the TDC time ratio of the # 4 cylinder. ΔT4 is calculated by (TDC time T4-TDC time average of # 4 cylinder) / TDC time average. Here, the TDC time average is an average value of the TDC time Ti, is calculated by (T1 + T2 + T3 + T4) / 4, and is the TDC time of the cylinder to which a desired amount of intake air is supplied.

  A cylinder having a positive TDC time ratio ΔTi is a cylinder to which intake air smaller than a desired air amount is supplied. As the absolute value of the TDC time ratio ΔTi increases, only a small amount of intake air is supplied. A cylinder having a TDC time ratio ΔTi of zero is a cylinder to which intake air of a desired air amount is supplied. A cylinder having a negative TDC time ratio ΔTi is a cylinder to which intake air larger than the desired air amount is supplied, and a larger amount of intake air is supplied as the absolute value of the TDC time ratio ΔTi is larger. If the calculated TDC time ratio ΔTi of each cylinder is not all zero, it can be determined that the intake air amount of each cylinder varies, and the cylinder is supplied to each cylinder based on the TDC time ratio ΔTi of each cylinder. The intake air amount can be estimated.

  Next, at step 108, based on the TDC time ratio ΔTi of each cylinder, for example, the intake air amount deviation rate Ci (%) of each cylinder is determined from the map shown in FIG. In order to eliminate an error in the measurement of the TDC time Ti and the like, the cylinder in which the TDC time ratio ΔTi is close to zero has an intake air amount deviation rate Ci of 0%. Except when the TDC time ratio ΔTi is near zero, the intake air amount deviation rate Ci is negative when the TDC time ratio ΔTi is positive, and is negative indicating that the intake air amount is short, and when the TDC time ratio ΔTi is negative, the intake air amount is negative. The absolute value of the intake air amount deviation rate is increased as the absolute value of the TDC time ratio ΔTi is increased. If the intake amount deviation amount Ci of each cylinder thus determined is not all zero, it can be determined that the intake amount of each cylinder varies, and the cylinder is supplied to each cylinder based on the intake amount deviation rate Ci of each cylinder. The intake air amount can be estimated. The map shown in FIG. 12 is a unique map for determining the intake air amount deviation rate Ci with respect to the TDC time ratio ΔTi when the actual compression ratio is increased and the ignition timing is retarded to the retard limit IT1 during idle operation. It is.

  Next, at step 109, in order to set the air-fuel ratio of each cylinder to the desired air-fuel ratio, the fuel injection correction coefficient Ki for each cylinder is calculated. The fuel injection correction coefficient Ki of each cylinder is calculated by 1 + Ci-d, where d is an average value of the intake air amount deviation rate Ci of each cylinder (d = (C1 + C2 + C3 + C4) / 4). Even if there is a variation in the intake air amount of each cylinder, since it is assumed that the average intake air amount of each cylinder is a desired intake air amount, correction by d is required. For example, when the intake amount deviation rate C1 of the # 1 cylinder is 4% (0.04) and the intake amount deviation rates C2, C3, and C4 of the # 2 cylinder to the # 4 cylinder are 0% (0). D is 1%, the fuel injection correction coefficient K1 of the # 1 cylinder is 1.03, and the fuel injection correction coefficients K2, K3, and K4 of the # 2 cylinder to the # 4 cylinder are 0.09.

  As a result, the supplied fuel is increased by 3% in the # 1 cylinder, and the supplied fuel is decreased by 1% in the # 2 to # 4 cylinders. As a result, the difference in generated torque among the cylinders increases not only in the amount of intake air but also in the amount of fuel that increases as the intake air amount increases. However, combustion at a desired air-fuel ratio can be realized in each cylinder. If it is preferable not to increase the torque difference between the cylinders, the process of step 109 may be omitted. Next, at step 110, since the variation in the intake air amount of each cylinder has been detected, the flag F is set to 0 and the processing ends.

  During idle operation, since the engine speed is low, the TDC time Ti, which is a value based on the torque generated in each cylinder, becomes long and it becomes easy to measure accurately. In addition, since the intake air amount is small, the variation in the intake air amount becomes large with respect to a slight shift in the closing timing of the intake valve, so that it is easy to detect the variation in the intake air amount. Thus, in this embodiment, the variation in the intake air amount of each cylinder is detected during the idling operation.

  However, this does not limit the present invention, and the variation in the intake air amount of each cylinder may be detected in the same manner during steady operation other than during idle operation. In this case, even if the generated torque difference between the cylinders is small at the current optimal ignition timing, the ignition timing retardation that can be achieved by increasing the mechanical compression ratio and increasing the actual compression ratio as in the idling state, The generated torque difference between the cylinders can be increased, and the variation in the intake air amount of each cylinder can be reliably detected from the value based on the generated torque of each cylinder.

  At this time, if it is necessary to calculate the intake air amount deviation rate Ci of each cylinder, the actual compression ratio is increased in the engine operation state when detecting the intake air amount variation, and the ignition timing is set to the specific retardation limit IT ′. A map as shown in FIG. 12 for determining the intake air amount deviation rate Ci with respect to the TDC time ratio ΔTi when the retardation is made may be set in advance.

1 Crankcase 2 Cylinder block 6 Spark plug A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (2)

  In a spark ignition internal combustion engine equipped with a variable compression ratio mechanism, when detecting variations in intake air amount of each cylinder, the mechanical compression ratio is increased and ignition is performed by the variable compression ratio mechanism without changing the intake air amount of each cylinder. A spark ignition internal combustion engine characterized in that the timing is retarded and a value based on torque generated in each cylinder is measured.   The spark ignition internal combustion engine according to claim 1, wherein a variation in intake air amount of each cylinder is detected during idling.

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