JP5640753B2 - Spark ignition internal combustion engine - Google Patents

Description

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

  A spark internal combustion engine having a variable compression ratio mechanism is known. In such an internal combustion engine, the variable compression ratio mechanism lowers the mechanical compression ratio at high engine loads with high thermal efficiency and increases the mechanical compression ratio at low and medium engine loads with low thermal efficiency (see Patent Document 1).

  By the way, when the catalyst device arranged in the engine exhaust system has a temperature equal to or higher than the set temperature, when the unburned fuel in the exhaust gas burns in the catalyst device, the catalyst device may be melted by the combustion heat. As a result, when the catalyst device is at a set temperature or higher, the amount of fuel supplied into the cylinder is increased so that most of the oxygen in the intake air is consumed in the cylinder, and the unburned fuel in the exhaust gas is the catalyst. Almost no combustion occurs in the device.

  In particular, when the catalyst device reaches a set temperature or higher during an engine load, the mechanical compression ratio is increased and the combustion chamber volume at the compression top dead center is reduced during the engine load. As a result, the amount of fuel that flows into the clevis immediately after ignition and does not burn increases, so that the amount of oxygen consumed in the cylinder decreases, and a large amount of unburned fuel and unconsumed oxygen remain in the cylinder. Therefore, it is necessary to suppress the decrease in the amount of oxygen consumed in the cylinder even if the fuel flows into the clevis immediately after ignition and the fuel is increased by a relatively large amount.

JP2007-303423 JP2009-185669 JP2009-221855A

  When the catalyst device is at a set temperature or higher, the mechanical compression ratio remains high or the mechanical compression ratio is further increased even when the load changes from the middle engine load to the low engine load. The ratio of the clevis volume at the dead center remains high, and it is necessary to suppress the decrease in the amount of oxygen consumed in the cylinder by continuing to increase the fuel in a relatively large amount so as not to increase the amount of oxygen in the exhaust gas. Rather, fuel consumption is greatly exacerbated.

  Accordingly, an object of the present invention is to provide a spark ignition internal combustion engine having a variable compression ratio mechanism and capable of increasing the mechanical compression ratio when the engine is under load and when the engine is low, compared to when the engine is high. When the load in the cylinder is changed to a low load when the fuel in the cylinder is increased in order to suppress a decrease in the amount of oxygen consumed in the cylinder because the exhaust system catalytic device is above the set temperature, It is to improve the deterioration of fuel consumption by suppressing the increase of fuel.

  The spark ignition internal combustion engine according to claim 1 of the present invention includes a variable compression ratio mechanism, and when the engine load is an engine load that is equal to or higher than the first set load and less than the second set load, and less than the first set load. In a spark-ignition internal combustion engine in which the mechanical compression ratio is increased when the engine is low and the engine compression ratio is higher than that when the engine is high or higher than the second set load, the engine exhaust system catalyst device is set at the medium load. When the load in the cylinder is increased to reduce the amount of oxygen in the exhaust gas because the temperature is higher than the temperature, and when the load changes from the medium engine load to the low engine load, the low engine load In this case, the ignition timing is retarded from the maximum torque ignition timing to reduce the increased amount of fuel in the cylinder.

  The spark ignition internal combustion engine according to claim 1 of the present invention is a fuel in a cylinder for reducing the amount of oxygen in exhaust gas because the catalyst device of the engine exhaust system is at a set temperature or higher when the engine is loaded. When the load changes from medium load to low engine load, the mechanical compression ratio is high even at low engine load, so the ratio of the clevis volume at the compression top dead center is high. In order not to increase the amount of oxygen in the exhaust gas, it is necessary to suppress the decrease in the amount of oxygen consumed in the cylinder by continuing to increase the fuel in a relatively large amount, but the ignition timing is delayed from the maximum torque ignition timing. By reducing the pressure in the cylinder immediately after ignition, the amount of fuel that flows into the clevis and does not burn immediately after ignition can be reduced. It is possible to improve.

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 the melting damage prevention of a catalyst apparatus.

  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 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, when the unburned fuel in the exhaust gas burns in the catalyst device 20 when the catalyst device 20 is equal to or higher than the set temperature, the catalyst device 20 may be melted by the combustion heat. Thus, when the catalyst device 20 is at a set temperature or higher, the amount of fuel supplied into the cylinder is increased so that most of the oxygen in the intake air is consumed in the cylinder, and unburned fuel in the exhaust gas is consumed. It is desirable to have little combustion in the catalytic device.

  The correction coefficient for increasing the amount of fuel in the cylinder when the catalyst device 20 is equal to or higher than the set temperature is changed according to the mechanical compression ratio. The higher the mechanical compression ratio, the smaller the combustion chamber volume at the compression top dead center. Therefore, the ratio of the clevis volume at which the flame such as a gap between the cylinder block and the cylinder head cannot enter at the compression top dead center is high. Therefore, the proportion of fuel that flows into the clevis immediately after ignition and does not burn increases.

  Accordingly, in order to prevent the amount of oxygen consumed in the cylinder from being increased by increasing the amount of fuel that flows into the clevis and does not burn, the higher the mechanical compression ratio, the higher the correction coefficient k ( > = 1) is increased, and the fuel injection amount Q ′ when the catalyst device 20 is equal to or higher than the set temperature is increased by multiplying the required fuel amount Q in the current engine operating state by the correction coefficient k ( Q ′ = Q * k).

  When the catalyst device 20 is equal to or higher than the set temperature, the mechanical compression ratio remains high (or the mechanical compression ratio is further increased) even when the load is changed from the medium engine load to the low engine load. The ratio of the clevis volume at the compression top dead center remains high, and the correction coefficient k for increasing the fuel is kept large. Thereby, the fuel consumption at the time of engine low load is greatly deteriorated at the time of load change when the catalyst device 20 is higher than the set temperature.

  In the spark ignition internal combustion engine of the present embodiment, the electronic control unit 30 controls the ignition timing of the spark plug 6 according to the flowchart shown in FIG. 10, thereby reducing the engine load from the engine load when the catalyst device 20 is equal to or higher than the set temperature. When the load changes to the load, the deterioration of the fuel consumption at the time of engine low load is suppressed.

First, in step 101, it is determined based on the output of the load sensor 41 whether or not the current engine load _Li is equal to or less than the first set load La. The first set load La is an upper limit value of a low load smaller than the medium load L1 on the low load side in FIG. If the determination in step 101 is negative, this time the process is terminated as it is, not the engine low load operation.

When the determination in step 101 is affirmative, it is determined in step 102 whether or not the previous engine load L _i-1 is greater than the first set load La and equal to or less than the second set load Lb based on the output of the load sensor 41. Is done. The second set load Lb is an upper limit value of a medium load larger than the medium load L1 on the low load side in FIG. When the determination in step 102 is negative, the previous operation is not an engine load operation but an engine high load operation, and the process ends.

  If the determination in step 102 is affirmative, the previous operation is an engine load operation, the present operation is an engine low load operation, and the load changes from an engine load to an engine low load. It is determined whether or not the temperature T of the device 20 is equal to or higher than the set temperature T1. The temperature T of the catalyst device 20 can be directly measured or estimated based on the temperature of the exhaust gas flowing into the catalyst device 20, and the temperature of the exhaust gas flowing out from the catalyst device 20 is determined as the catalyst device. It can be used as a temperature of 20.

  When the determination in step 103 is negative, even if the unburned fuel discharged from the cylinder in the catalyst device 20 burns using unconsumed oxygen discharged from the cylinder, the catalyst device 20 does not melt. In particular, it is not necessary to increase the amount of fuel to reduce the amount of unconsumed oxygen in the cylinder. Thereby, in step 104, the target ignition timing ITt is set to the ignition timing IT0 (for example, MBT) that generates the maximum torque in the engine operating state based on the current engine speed and the current engine load (low load). The correction coefficient k for increasing the fuel is set to 1, and the required fuel amount Q in the current engine operating state is not increased.

  On the other hand, when the determination in step 103 is affirmative, if the temperature T of the catalyst device 20 is equal to or higher than the set temperature T1 and a certain amount of unburned fuel is burned in the catalyst device 20, the catalyst device 20 is dissolved by the combustion heat. Therefore, the amount of unconsumed oxygen in the cylinder must be reduced by increasing the amount of fuel in the cylinder. The temperature T of the catalyst device 20 does not become equal to or higher than the set temperature T1 only by the engine low load operation. When the determination in step 103 is affirmative, the temperature T of the catalyst device 20 in the previous engine medium load operation is The fuel temperature is increased at the set temperature T1 or higher.

  Since the mechanical compression ratio is high during engine load, the fuel increase correction coefficient k is increased (eg, 1.2). In order to keep the mechanical compression ratio high or even higher in the engine low load operation this time, the correction factor k for increasing the fuel is kept large or even larger as it is. It will make fuel consumption worse.

  In this flowchart, in step 106, the target ignition timing ITt is set to IT1 that is retarded from the ignition timing IT0 that generates the maximum torque in the current engine operating state. By retarding the ignition timing from the maximum torque ignition timing in this way, the pressure in the cylinder immediately after ignition can be reduced, and the amount of fuel that flows into the clevis and does not burn immediately after ignition can be reduced. Even if the amount is reduced, the amount of unconsumed oxygen can be sufficiently reduced. As a result, in step 107, the correction coefficient k for increasing the fuel amount is made smaller by Δk from the value k based on the current mechanical compression ratio, and is made closer to 1 (preferably 1). Thereby, only a slight fuel increase is performed (or the fuel increase is stopped).

  The processing of Step 106 and Step 107 is continued until the temperature T of the catalyst device 20 becomes lower than the set temperature T1. By retarding the ignition timing from the maximum torque ignition timing IT0 to IT1, the exhaust gas temperature at the time of low load rises, but it is lower than that at the time of medium load when the temperature T of the catalyst device 20 is set to the set temperature T1 or higher. Since the exhaust gas temperature at the time of load is low and the amount of exhaust gas at the time of low load is small, the temperature T of the catalyst device 20 becomes lower than the set temperature T1 relatively early. Accordingly, if the determination in step 103 is negative, the target ignition timing ITt is set to the ignition timing IT0 that generates the maximum torque in step 104, and the correction coefficient k for increasing the fuel is set to 1 in step 105. If the increase has been carried out, the fuel increase is stopped.

  Thus, when the fuel in the cylinder is increased by the above-described control to suppress a decrease in the amount of oxygen consumed in the cylinder because the catalyst device 20 is at or above the set temperature T1 at the time of engine load. When the load changes to a low engine load, the increase in fuel in the cylinder is suppressed without reducing the efficiency by lowering the engine compression ratio so as to reduce the ratio of the clevis volume at the compression top dead center. Deterioration of consumption can be improved.

1 Crankcase 2 Cylinder block 6 Spark plug 30 Electronic control unit A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (1)

An engine having a variable compression ratio mechanism, wherein the engine load is equal to or greater than the first set load and less than the second set load, and when the engine load is less than the first set load and the engine is equal to or greater than the second set load. In a spark ignition internal combustion engine in which the mechanical compression ratio is increased as compared with a high load, since the temperature of the catalyst device of the engine exhaust system is equal to or higher than the set temperature at the medium load, the amount of oxygen in the exhaust gas is reduced. when to reduce has increased the fuel in the cylinder, the increasing fuel in the cylinder together with the time from the engine during load and load change to the low engine load retards the maximum torque ignition timing point fire timing A spark ignition internal combustion engine characterized in that , when the temperature of the catalyst device is lower than a set temperature, the ignition timing is set to the maximum torque ignition timing and the fuel increase is stopped .

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