JP5429136B2 - Spark ignition internal combustion engine - Google Patents

Description

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

  There is known an internal combustion engine in which the intake amount can be reduced by retarding the closing timing of the intake valve in the compression stroke, and thus the intake amount is controlled by making the closing timing of the intake valve variable. (See Patent Document 1). In such an internal combustion engine, the responsiveness of the intake air amount control can be increased as compared with the case where the intake air amount is controlled by the throttle valve in the intake passage, and it is generated by restricting the intake passage by the throttle valve. Pumping loss can be eliminated.

Meanwhile, in an internal combustion engine, by recirculating a portion of exhaust gas into the cylinder, it suppresses the exhaust gas recirculation generation of the NO _X lowers the combustion temperature (EGR) is being performed. In EGR, as the amount of recirculated exhaust gas is large, it is possible to reduce the generation amount of NO _X, while the worsening of combustion. Thus, the EGR rate (recirculation exhaust gas amount / (recirculation exhaust gas amount + new air amount)) is optimally controlled for each engine operating state. In general, the EGR rate is relatively low at high loads where high output is required and at low loads where combustion is likely to be unstable, and the EGR rate is relatively high at medium loads.

  As a result, when EGR is performed in an internal combustion engine that controls the intake air amount by the intake valve, the intake air amount is changed from the first intake air amount at the middle load during a sudden load change from a medium load to a low load. In order to reduce the second intake amount at low load, the intake valve closing timing is retarded from the first closing timing at medium load to the second closing timing at low load, and the EGR rate is set to medium The first EGR rate suitable for the load is reduced to the second EGR rate suitable for the low load. Also, the ignition timing is advanced from the first ignition timing suitable for the medium load to the second ignition timing suitable for the low load.

JP2007-303423 JP 2010-090872 A JP 2004-092639 A JP 2004-197620 A

  Reducing the intake amount by retarding the closing timing of the intake valve in the compression stroke is increasing the intake amount discharged from the cylinder to the intake system in the compression stroke. In this intake air amount control, the recirculated exhaust gas in the cylinder is also discharged to the intake system during the compression stroke, so the EGR rate in the cylinder is reduced during a sudden load change from the middle load to the low load. When the EGR rate of the intake system is changed to a desired value in order to change, the amount of recirculated exhaust gas discharged from the cylinder to the intake system also changes due to a change in the closing timing of the intake valve. The rate is difficult to converge to the desired value. Accordingly, even when the intake valve closing timing is retarded by the actuator to the second valve closing timing and the intake amount is reduced to the second intake amount, the EGR rate of the intake system has not converged to the desired value, A delay time occurs until the EGR rate of the intake system converges to a desired value and the EGR rate in the cylinder becomes the second EGR rate.

  Accordingly, when the ignition timing is set to the closing timing of the intake valve and the intake amount becomes the second intake amount when the intake amount becomes the second intake amount, at this time, the EGR rate in the cylinder is higher than the second EGR rate. Therefore, only an engine output lower than the desired engine output can be generated as it is. In order to generate the desired engine output, the ignition timing must be advanced from the second ignition timing based on the difference between the current EGR rate and the second EGR rate.

  However, at this time, the closing timing of the intake valve is retarded to the second closing timing, and the ignition timing cannot be advanced from the second closing timing of the intake valve. The desired engine output may not be generated by advancing.

  Therefore, an object of the present invention is a spark ignition internal combustion engine that controls the intake air amount by changing the closing timing of the intake valve, and the intake air amount when the required load decreases from the first engine load to the second engine load. In order to reduce the EGR rate in the cylinder from the first EGR rate to the second EGR rate, the intake valve closing timing is started from the first closing timing to the second closing timing at the first time. When the intake valve closing timing is the first closing timing and the EGR rate in the cylinder is the first EGR rate, the ignition timing is the first ignition timing, and the intake valve closing timing is the first closing timing. In a spark ignition internal combustion engine in which the ignition timing is the second ignition timing when the EGR rate in the cylinder is the second EGR rate when the two valve closing timings are the second EGR rate, the closing timing of the intake valve is the second valve closing timing The desired engine output is obtained at the second time.

  The spark ignition internal combustion engine according to claim 1 according to the present invention closes the intake valve by an actuator at a first time in order to reduce the intake amount when the required load decreases from the first engine load to the second engine load. The retard is started from the first closing timing to the second closing timing, the EGR rate in the cylinder is lowered from the first EGR rate to the second EGR rate, and the closing timing of the intake valve is When the valve closing timing is the first EGR rate in the cylinder, the ignition timing is the first ignition timing, and the intake valve closing timing is the second valve closing timing. In a spark ignition internal combustion engine in which the ignition timing is the second ignition timing when the EGR rate in the cylinder is the second EGR rate, the closing timing of the intake valve is set to the second timing due to a delay in response of the EGR rate in the cylinder. At the second time that is the valve closing time, the EGR rate in the cylinder is Even if the EGR rate in the cylinder at the second time is higher than the second EGR rate without decreasing to the second EGR rate, the ignition timing is changed from the second ignition timing to the second valve closing timing. The operation time of the actuator from the first time to the second time is set so that the desired engine output is less than or equal to a specific EGR rate that occurs when the angle is advanced to the maximum.

  A spark ignition internal combustion engine according to a second aspect of the present invention is the spark ignition internal combustion engine according to the first aspect, further comprising a variable compression ratio mechanism capable of changing a mechanical compression ratio, wherein a required load is the first engine load. When the engine load decreases to the second engine load, the mechanical compression ratio is increased by the variable compression ratio mechanism so that the actual compression ratio becomes constant with respect to the decrease in the intake air amount.

  According to the spark ignition internal combustion engine of the first aspect of the present invention, when the required load decreases from the first engine load to the second engine load, the intake valve is operated by the actuator at the first time to reduce the intake air amount. Is started from the first closing timing to the second closing timing, the EGR rate in the cylinder is decreased from the first EGR rate to the second EGR rate, and the closing timing of the intake valve is When the EGR rate in the cylinder is the first EGR rate at the first valve closing timing, the ignition timing is set as the first ignition timing, and the valve closing timing of the intake valve is the second valve closing timing. In the spark ignition internal combustion engine in which the ignition timing is the second ignition timing when the EGR rate of the engine is the second EGR rate, the closing timing of the intake valve becomes the second closing timing due to a delay in response of the EGR rate in the cylinder. At the second time, the EGR rate in the cylinder reaches the second EGR rate. It does not decrease. Here, as the operating time of the actuator from the first time to the second time becomes longer, the EGR rate in the cylinder at the second time can be made closer to the second EGR rate. Even if the EGR rate is higher than the second EGR rate, when the ignition timing is advanced from the second ignition timing to the maximum after the second valve closing timing, the desired engine output is less than or equal to the specific EGR rate. In addition, if the operating time of the actuator is set and the ignition timing at the second time is advanced within a feasible range, the desired engine output can be obtained at this time.

  According to the spark ignition internal combustion engine according to claim 2 of the present invention, the spark ignition internal combustion engine according to claim 1 is provided with a variable compression ratio mechanism capable of changing the mechanical compression ratio, and the required load is the first engine. When the load decreases from the load to the second engine load, the mechanical compression ratio is increased by the variable compression ratio mechanism so that the actual compression ratio becomes constant with respect to the decrease in the intake air amount. The occurrence of misfire due to a decrease in the compression ratio is suppressed.

1 is an overall view of a spark ignition type 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. 3 is a time chart showing load changes, intake valve closing timing changes, mechanical compression ratio changes, EGR rate changes, and ignition timing changes in the spark ignition internal combustion engine according to the present invention.

  FIG. 1 shows a side sectional view of a spark ignition type internal combustion engine. Referring to FIG. 1, 1 is a crankcase, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is a spark plug disposed at the center of the top surface of the combustion chamber 5, and 7 is 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. Reference numeral 100 denotes an exhaust gas recirculation passage that communicates an engine exhaust system such as the exhaust manifold 19 and an engine intake system such as the surge tank 12. The exhaust gas recirculation passage 100 includes an exhaust gas recirculation passage 100. A control valve 101 is provided for controlling the amount of exhaust gas (external EGR amount) recirculated into the cylinder.

  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, the variable valve timing mechanism B, and the exhaust gas recirculation passage 100 via the corresponding drive circuit 38. The control valve 101 is connected.

  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 attached 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 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. When the engine load is reduced to the low load L2, the mechanical compression ratio reaches a limit mechanical compression ratio (upper limit mechanical compression ratio) that is a structural limit of the combustion chamber 5. 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 L2 when the mechanical compression ratio reaches the limit mechanical compression ratio. Therefore, at the time of engine low load operation, that is, at the engine low load operation side, the mechanical compression ratio is maximized and the expansion ratio is also maximized. 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 L2, 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 in a region where the load is lower than the engine load L2 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 L2 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.

  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.

Meanwhile, in a spark ignition internal combustion engine of the present embodiment, to suppress by recirculating a portion of exhaust gas into the cylinders through the exhaust gas recirculation passage 100, the generation of the NO _X lowers the combustion temperature exhaust Gas recirculation (EGR) is performed. The amount of exhaust gas recirculated through the exhaust gas recirculation passage 100 is controlled by the opening of the control valve 101 as an external EGR amount.

  In addition, if there is a valve overlap period in which both the intake valve and the exhaust valve are open from the end of the exhaust stroke to the beginning of the intake stroke, this valve overlap period is compared to the exhaust system. Since the pressure in the intake system is low, no gas is supplied from the intake system into the cylinder, and the exhaust gas in the cylinder and in the exhaust system flows back to the intake system, and this backflow exhaust gas is used as internal EGR. In the intake stroke, it is supplied into the cylinder after the exhaust valve is closed. The longer the valve overlap period, the greater the amount of internal EGR. Thereby, for example, if a variable valve timing mechanism similar to that described above is provided so that the valve closing timing of the exhaust valve is variable, the internal EGR amount can be controlled by changing the valve overlap period. In this case, even if the exhaust gas recirculation passage 100 is omitted and the external EGR is not performed, the EGR can be performed only by the internal EGR.

  Both the external EGR and the internal EGR or the exhaust gas recirculation passage are omitted, and the EGR rate (recirculation exhaust gas amount / (recirculation exhaust gas amount + new air amount)) in the cylinder is controlled only by the internal EGR. In this internal combustion engine, the intake amount is controlled by making the closing timing of the intake valve 7 variable, that is, by retarding the closing timing of the intake valve 7 in the compression stroke, The exhaust gas is discharged to the intake system to control the intake amount. At that time, part of the recirculated exhaust gas supplied into the cylinder in the intake stroke is also discharged to the intake system. As a result, the recirculation exhaust gas from the external EGR and the internal EGR or the recirculation exhaust gas from the internal EGR alone and the exhaust gas discharged from the cylinder in the compression stroke are mixed in the intake system, and these are the next intake air. In the stroke, the EGR rate in the cylinder is determined together with the exhaust gas that is sucked into the cylinder and remains in the cylinder in the intake stroke.

  Since EGR is accompanied by some deterioration of combustion, the EGR rate is relatively low at high loads that require high output and at low loads where combustion is likely to be unstable, and the EGR rate is relatively high at medium loads. . As a result, in the case of a sudden load change from the medium load L1 to the low load L2 shown in FIG. 9, in order to reduce the intake air amount, the closing timing of the intake valve 7 is changed from the first closing timing to the second closing timing. The EGR rate in the cylinder is reduced from the first EGR rate to the second EGR rate by delaying the timing and controlling the valve overlap period. The ignition timing is advanced from the first ignition timing at the middle load L1 to the second ignition timing at the low load L2.

  When the closing timing of the intake valve 7 is changed in this way, the amount of recirculated exhaust gas discharged from the cylinder to the intake system during the compression stroke changes, so that the EGR rate of the intake system is difficult to converge to a desired value. Therefore, even when the closing timing of the intake valve 7 is retarded to the second closing timing and the intake amount is reduced to the second intake amount, the EGR rate of the intake system does not converge to the desired value, and the intake system A delay time is generated until the EGR rate of the engine converges to a desired value and the EGR rate in the cylinder reaches the second EGR rate R2.

  Accordingly, when the ignition timing is set to the closing timing of the intake valve and the intake amount becomes the second intake amount when the intake amount becomes the second intake amount, at this time, the EGR rate in the cylinder decreases to the second EGR rate. However, as it is, only an engine output lower than the desired engine output can be generated. In order to generate the desired engine output, the ignition timing must be advanced from the second ignition timing based on the difference between the current EGR rate and the second EGR rate.

  FIG. 10 shows a spark ignition internal combustion engine according to the present invention, in which the intake valve closing timing IVC, mechanical compression ratio E, EGR rate R, and ignition when the engine load L suddenly changes from the medium load L1 to the low load L2. It is a time chart which shows each control of time IT. When the engine load L changes from the medium load L1 to the low load L2, at the same time (or immediately after), at the first time t1, the valve closing timing of the intake valve is decreased by the variable valve timing mechanism B in order to decrease the intake amount. Is started from the first valve closing timing IVC1 of the medium load L1 to the second valve closing timing IVC2 of the low load L2. Further, at the first time t1 or just after that, the EGR rate in the cylinder is decreased from the first EGR rate R1 of the medium load L1 toward the second EGR rate R2 of the low load L2.

  As described above, since the intake valve closing timing changes when the load changes from the medium load L1 to the low load L2, the amount of recirculated exhaust gas discharged from the cylinder to the intake system changes during the compression stroke. The EGR rate of the intake system is difficult to converge to a desired value. Thus, when the variable valve timing mechanism B, that is, the actuator, changes the valve closing timing IVC of the intake valve from the first valve closing timing IVC1 to the second valve closing timing IVC2 at a constant speed indicated by a dotted line, The EGR rate R in the cylinder that is changed from the EGR rate R1 to the second EGR rate R2 is a second time t2 when the valve closing timing IVC of the intake valve becomes the second valve closing timing IVC2 of the low load L2, as indicated by a dotted line. A delay time DA is generated from 'and becomes the second EGR rate R2.

  Since the intake air amount in the cylinder follows the change in the intake valve closing timing almost without delay, the ignition timing IT is matched with the intake valve closing timing from the first ignition timing IT1 of the medium load L1. When the ignition timing is advanced to the second ignition timing IT2, the EGR rate R in the cylinder is the third EGR rate R3 higher than the second EGR rate R2 of the low load L2 at the second time t2 ′. Output cannot be generated. Accordingly, it is necessary to advance the ignition timing from the second ignition timing IT2 to increase the engine output.

  However, the third EGR rate R3 in the cylinder at the second time t2 ′ is significantly higher than the second EGR rate R2, and the ignition timing advance amount ΔITA corresponding to the difference (R3−R2) becomes large. As indicated by the dotted line, when the ignition timing is advanced so as to generate the desired engine output from the first time t1 to the second time t2 ′, at the second time t2 ′, the closing timing IVC2 of the intake valve at that time This is not feasible because it becomes more advanced, that is, ignition must be performed before the intake valve is closed.

  Thereby, in this embodiment, the closing timing IVC of the intake valve is changed from the first closing timing IVC1 to the second closing timing IVC2 at a constant speed indicated by a solid line, and thus the operating speed of the actuator is changed. As shown by the solid line, the EGR rate in the cylinder that changes from the first EGR rate R1 to the second EGR rate R2 by delaying is the same as the second valve closing timing IVC2 when the intake valve closing timing is low load L2. A delay time DB is generated from the second time t2 (which is later than the second time t2 ′ in the case of the dotted line because the actuator operating speed is slow), and becomes the second EGR rate R2.

  The delay time DB is actually shorter because the EGR rate of the intake system tends to converge to a desired value as the operating speed of the actuator that retards the closing timing of the intake valve is delayed. The delay time (the time from the second time when the intake valve closing timing becomes the second valve closing timing IVC2 until the EGR rate in the cylinder becomes the second EGR rate) is constant regardless of the operating speed of the actuator. However, since the second time itself with the closing timing of the intake valve as the second closing timing is delayed, the EGR rate at the second time is close to the second EGR rate R2. Further, in practice, the delay time becomes shorter as the second time when the operation speed of the actuator is slowed down and the closing timing of the intake valve is set as the second valve closing timing becomes shorter, so the EGR rate R at the second time becomes shorter. Is remarkably close to the second EGR rate R2.

  Accordingly, when the operating speed of the actuator is slowed down, the EGR rate at the second time t2 is the fourth EGR rate R4 smaller than the third EGR rate R3 as shown by the solid line, and is shown by the dotted line In comparison, the difference (R4−R2) from the second EGR rate becomes smaller, and the advance amount ΔITB of the ignition timing corresponding to this difference also becomes smaller. Therefore, it is desired from the first time t1 to the second time t2. When the ignition timing is advanced so as to generate engine output, at the second time t2, the intake valve closing timing IVC2 is retarded, that is, the air-fuel mixture is compressed after the intake valve is closed. After that, ignition is performed, which is feasible. The ignition timing IT is gradually retarded from the second time t2 as the EGR rate in the cylinder decreases, and is set to the second ignition timing IT2 at the end of the delay time DB, and the desired engine output is also generated during this period. .

  As described above, the delay time from the second time when the intake valve closing timing IVC becomes the second closing timing IVC2 until the EGR rate in the cylinder becomes the second EGR rate R2 is as described above. From the first valve closing timing IVC1 to the second valve closing timing IVC2, the shorter the actuator operating time from the first time to the second time at which the retardation is started, the shorter is the first EGR rate of the medium load The smaller the difference between R1 and the low-load second EGR rate R2, the shorter.

  If the delay time is estimated in this way, for example, the EGR rate in the cylinder at the second time can be estimated assuming that the EGR rate in the cylinder changes at a constant speed. As a result, when the EGR rate in the cylinder at the second time advances the ignition timing IT to the maximum after the second closing timing of the intake valve from the second ignition timing IT2 (the intake valve is in the second closing timing). After the valve is closed, the actuator operating time (from the first time to the second time) at which the desired engine output is generated when the crank angle range for compressing the air-fuel mixture in the cylinder to some extent is required. The time to the time) can be estimated in reverse.

  As a result, when the required load decreases from the medium load L1 to the low load L2, the actuator closes the intake valve closing timing from the first closing timing IVC1 to the second closing timing by the actuator at the first time t1 in order to reduce the intake amount. The retard is started to the valve timing IVC2, the EGR rate in the cylinder is decreased from the first EGR rate R1 to the second EGR rate R2, and the closing timing of the intake valve is the first closing timing IVC1 and When the EGR rate of the engine is the first EGR rate R1, the ignition timing is the first ignition timing IT1, the closing timing of the intake valve is the second closing timing IVC2, and the EGR rate in the cylinder is the second EGR rate. When the ignition timing is set to the second ignition timing IT2 at the rate R2, the second time t2 when the closing timing of the intake valve is set to the second closing timing IVC2 due to a delay in response of the EGR rate in the cylinder. In the EGR in the cylinder Is not decreased to the second EGR rate R2, but the desired engine output is generated when the EGR rate in the cylinder is advanced to the maximum after the second ignition timing IT2 and the second valve closing timing IVC2. The actuator operating time can be set based on the first EGR rate R1, the second EGR rate R2, the first valve closing timing IVC1, and the second valve closing timing IVC2 so as to be equal to or lower than the EGR rate.

  Thus, when the required load decreases from the medium load L1 to the low load L2, the valve closing timing of the intake valve is changed from the first valve closing timing IVC1 to the second valve closing timing IVC2 during the set operation time of the actuator. If the ignition timing at the second time t2 is advanced within a feasible range, a desired engine output can be obtained at this time. The actuator operation time is preferably shorter in order to realize the intake amount of the low load L2 at an early stage, and the desired engine output is generated when the EGR rate in the cylinder at the second time advances the ignition timing to the maximum. It is preferable that the specific EGR rate be set.

  As described above, when the load changes from the medium load to the low load, the desired engine is set at the second time based on the first EGR rate R1, the second EGR rate R2, the first valve closing timing IVC1, and the second valve closing timing IVC2. The operation time of an actuator (actuator for delaying the closing timing of the intake valve) that can realize the advance of the ignition timing that generates the output can be set and mapped in advance. However, when the load changes from the medium load to the low load, the desired engine output is output at the second time based on the first EGR rate R1, the second EGR rate R2, the first valve closing timing IVC1, and the second valve closing timing IVC2. It is also possible to calculate and set the operating time of the actuator that can realize the advance of the ignition timing to be generated.

  Further, as shown in FIG. 10, the mechanical compression ratio E is variable so that the actual compression ratio becomes constant with respect to the decrease in the intake air amount when the required load decreases from the medium load L1 to the low load L2. The compression ratio mechanism increases the first mechanical compression ratio E1 to the second mechanical compression ratio E2, thereby suppressing the occurrence of misfire due to a decrease in the actual compression ratio.

A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (2)

  When the required load drops from the first engine load to the second engine load, the actuator closes the intake valve closing timing from the first closing timing to the second closing timing at the first time to reduce the intake air amount. And the EGR rate in the cylinder is decreased from the first EGR rate to the second EGR rate, and the closing timing of the intake valve is the first closing timing, and the EGR rate in the cylinder is the first EGR rate. When the EGR rate is one, the ignition timing is the first ignition timing, and when the intake valve closing timing is the second closing timing and the EGR rate in the cylinder is the second EGR rate. In a spark ignition internal combustion engine in which the ignition timing is the second ignition timing, due to a delay in response of the EGR rate in the cylinder, at the second time when the closing timing of the intake valve is the second closing timing, The EGR rate does not decrease to the second EGR rate, and the cylinder at the second time Even if the EGR rate of the engine is higher than the second EGR rate, a specific EGR rate at which a desired engine output is generated when the ignition timing is advanced to the maximum after the second valve closing timing from the second ignition timing An operation time of the actuator from the first time to the second time is set so as to be as follows.   A variable compression ratio mechanism capable of changing the mechanical compression ratio is provided, and when the required load decreases from the first engine load to the second engine load, the actual compression ratio becomes constant with respect to the decrease in the intake air amount. The spark ignition internal combustion engine according to claim 1, wherein the mechanical compression ratio is increased by the variable compression ratio mechanism.

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