JP5516461B2 - Internal combustion engine having a variable compression ratio mechanism - Google Patents

Description

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

  An internal combustion engine having a variable compression ratio mechanism that makes a mechanical compression ratio variable by changing a combustion chamber volume at a compression top dead center is known. By the way, when the engine is decelerated, in order to reduce fuel consumption and generate a large engine braking force, a fuel cut that interrupts combustion in the cylinder without supplying fuel is generally performed. Due to the fuel cut, the engine speed gradually decreases, and when the engine speed reaches the return speed, fuel is supplied and combustion in the cylinder is resumed. Lowering the fuel cut return speed is more advantageous for reducing fuel consumption, but if the return speed is excessively reduced, the engine may stop even if combustion is resumed.

  In an internal combustion engine equipped with a variable compression ratio mechanism, the higher the actual compression ratio that becomes variable with the mechanical compression ratio, the higher the generated torque and the more difficult it is to stop the engine when returning from fuel cut. It has been proposed that the higher the actual compression ratio at the time of fuel cut, the lower (see Patent Document 1).

JP-A-2005-030223 JP-A-2004-239147 JP2007-107512 JP2010-255585A

  An internal combustion engine that improves thermal efficiency by changing the expansion ratio according to the mechanical compression ratio by changing the combustion chamber volume at the compression top dead center by making the actual compression ratio constant because knocking is likely to occur when the actual compression ratio is increased. In this case, even if the above-described concept is applied, the actual compression ratio is constant, so that the return speed of fuel cut is constant, and fuel consumption due to fuel cut cannot be further reduced.

  Accordingly, an object of the present invention is to further reduce fuel consumption due to fuel cut in an internal combustion engine having a variable compression ratio mechanism in which the mechanical compression ratio is variable by the variable compression ratio mechanism and the actual compression ratio is constant. .

  An internal combustion engine provided with the variable compression ratio mechanism according to claim 1 of the present invention is an internal combustion engine provided with a variable compression ratio mechanism in which the mechanical compression ratio is variable by the variable compression ratio mechanism and the actual compression ratio is constant. The lower the mechanical compression ratio is, the lower the rotational speed of the fuel cut is.

  An internal combustion engine having the variable compression ratio mechanism according to claim 2 according to the present invention is the internal combustion engine having the variable compression ratio mechanism according to claim 1, wherein the return combustion air-fuel ratio at the time of fuel cut return is greater than the stoichiometric air-fuel ratio. The return combustion air-fuel ratio is made smaller as the mechanical compression ratio at the time of fuel cut is made lower.

  According to the internal combustion engine having the variable compression ratio mechanism according to the first aspect of the present invention, in the internal combustion engine having the variable compression ratio mechanism in which the mechanical compression ratio is variable by the variable compression ratio mechanism and the actual compression ratio is constant, The lower the mechanical compression ratio at the time of cutting, the more the actual compression ratio is constant, so the amount of intake air in the cylinder increases, and the generated torque increases when fuel cut is restored. As a result, the lower the mechanical compression ratio at the time of fuel cut, the lower the return speed of the fuel cut, but the engine will not stop, and the fuel consumption due to the fuel cut will be reduced by the amount that the return speed is lowered. Can do.

According to the internal combustion engine provided with the variable compression ratio mechanism according to claim 2 of the present invention, in the internal combustion engine provided with the variable compression ratio mechanism according to claim 1, the return combustion air-fuel ratio at the time of fuel cut return is the theoretical sky. The richer the fuel ratio and the lower the mechanical compression ratio at the time of fuel cut, the more constant the actual compression ratio, so the amount of intake air in the cylinder increases, and the combustion temperature rises at the time of fuel cut recovery, resulting in NO _x generation. Tends to increase. As a result, the lower the mechanical compression ratio at the time of fuel cut, the smaller the return combustion air-fuel ratio, and the combustion temperature is lowered to suppress the increase in the amount of NO _x produced.

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 fuel cut control. It is a map for determining the return rotation speed of a fuel cut. 6 is a map for determining a fuel cut return combustion air-fuel ratio.

  FIG. 1 is a side sectional view of an internal combustion engine having a variable compression ratio mechanism according to the present invention. Referring to FIG. 1, 1 is a crankcase, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is a spark plug disposed at the center of the top surface of the combustion chamber 5, and 7 is intake air. 8 is an intake port, 9 is an exhaust valve, and 10 is an exhaust port. The intake port 8 is connected to a surge tank 12 via an intake branch pipe 11, and a fuel injection valve 13 for injecting fuel into the corresponding intake port 8 is arranged in each intake branch pipe 11. The fuel injection valve 13 may be arranged in each combustion chamber 5 instead of being attached to each intake branch pipe 11.

  The surge tank 12 is connected to an air cleaner 15 via an intake duct 14, and a throttle valve 17 driven by an actuator 16 and an intake air amount detector 18 using, for example, heat rays are arranged in the intake duct 14. On the other hand, the exhaust port 10 is connected to a catalyst device 20 containing, for example, a three-way catalyst via an exhaust manifold 19, and an air-fuel ratio sensor 21 is disposed in the exhaust manifold 19.

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

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

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

  2 is an exploded perspective view of the variable compression ratio mechanism A shown in FIG. 1, and FIG. 3 is a side sectional view of the internal combustion engine schematically shown. Referring to FIG. 2, a plurality of protrusions 50 spaced from each other are formed below both side walls of the cylinder block 2, and cam insertion holes 51 each having a circular cross section are formed in each protrusion 50. Has been. On the other hand, a plurality of protrusions 52 are formed on the upper wall surface of the crankcase 1 so as to be fitted between the corresponding protrusions 50 spaced apart from each other. Cam insertion holes 53 each having a circular cross section are formed.

  As shown in FIG. 2, a pair of camshafts 54, 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. That is, the eccentric portion 57 is fitted into an eccentric hole formed in the circular cam 56, and the circular cam 56 rotates around the eccentric portion 57 around the eccentric hole. 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, FIG. 3B, and FIG. 3C show the center line a of the concentric portion 58 (that is, the center line of the camshaft) a and the center line b of the eccentric portion 57 in each state. The positional relationship with the center line c of the circular cam 56 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 line a of the concentric part 58 and the center line c of the circular cam 56. The cylinder block 2 moves away from the crankcase 1 as the distance between the center line a of the concentric part 58 and the center line c of the circular cam 56 increases. That is, the variable compression ratio mechanism A changes the relative position between the crankcase 1 and the cylinder block 2 by a crank mechanism using a rotating cam. When the cylinder block 2 moves away from the crankcase 1, the volume of the combustion chamber 5 increases when the piston 4 is positioned at the compression top dead center. Therefore, by rotating the camshafts 54 and 55, the piston 4 is compressed at the top dead center. The volume of the combustion chamber 5 when it is located at can be changed.

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

  On the other hand, as shown by the broken line in FIG. 9, the intake air amount can be controlled without depending on the throttle valve 17 by advancing the closing timing of the intake valve 7 as the engine load becomes lower. Accordingly, when expressing the case shown in FIG. 9 so as to include both the case indicated by the solid line and the case indicated by the broken line, in the embodiment according to the present invention, the valve closing timing of the intake valve 7 becomes smaller as the engine load becomes lower. It is moved in a direction away from the intake bottom dead center BDC until the limit valve closing timing L1 at which the intake air amount supplied into the combustion chamber can be controlled. Thus, the intake air amount can be controlled by changing the closing timing of the intake valve 7 as shown by the solid line in FIG. 9 or by changing it as shown by the broken line.

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

  By the way, when the engine is decelerated, in order to reduce fuel consumption and generate a large engine braking force, a fuel cut that interrupts combustion in the cylinder without supplying fuel is generally performed. The engine speed gradually decreases due to the fuel cut, and when the engine speed reaches the return speed at which the fuel cut is stopped, fuel is supplied and combustion in the cylinder is resumed. Lowering the fuel cut return speed is more advantageous in reducing fuel consumption. However, if the return speed is excessively low, the combustion interval of each cylinder will become very long. However, the decrease in the engine speed cannot be stopped, and the engine may be stopped as it is.

  The higher the actual compression ratio, the higher the generated torque of each cylinder. Therefore, the higher the actual compression ratio, the lower the engine speed at the time of fuel cut recovery, even if the return speed of fuel cut is lowered and the combustion interval of each cylinder is lengthened. Never stop. However, in the internal combustion engine of this embodiment, the mechanical compression ratio is made variable by the variable compression ratio mechanism A, but the actual compression ratio is made constant by the variable valve timing mechanism B.

  FIG. 10 is a flowchart showing fuel cut control in the internal combustion engine of the present embodiment, which is performed by the electronic control unit 30. First, in step 101, it is determined whether or not there is a fuel cut request at the time of engine deceleration or the like. When this judgment is denied, the process is terminated as it is. However, when there is a request for fuel cut, in step 102, the return speed N1 of fuel cut is determined based on the current mechanical compression ratio E estimated from the output of the relative position sensor 22.

  Next, in step 103, fuel supply is stopped, combustion of each cylinder is stopped, and fuel cut is performed. The engine speed gradually decreases due to the fuel cut. In step 104, it is determined whether or not the current engine speed N detected by the crank angle sensor 42 has decreased to the return speed N1 determined in step 102. The The fuel cut is continued until this determination is affirmed. When the engine speed N decreases to the return speed N1, the determination in step 104 is affirmed. In step 105, fuel supply is started and combustion of each cylinder is started. Resume.

  In this fuel cut control, in step 102, the return rotational speed N1 is set using the map shown in FIG. 11 based on the mechanical compression ratio E at the start of fuel cut, and is lowered as the mechanical compression ratio E is lower. The mechanical compression ratio E at the start of the fuel cut is maintained during the fuel cut and becomes the mechanical compression ratio at the time of return. Here, since the actual compression ratio is constant, the lower the mechanical compression ratio E, the more the intake valve closing timing is advanced, and the intake amount in the cylinder increases. In particular, in the first combustion of each cylinder when returning from a fuel cut, the lower the mechanical compression ratio E, the larger the combustion chamber volume at the intake top dead center as well as the compression top dead center. In order to be filled with energy, the amount of fresh air in the cylinder becomes very large.

  Accordingly, the lower the mechanical compression ratio E, the higher the generated torque of each cylinder at the time of return from the fuel cut, and even if the return rotation speed N1 is lowered and the combustion interval of each cylinder becomes longer, the combustion of each cylinder If the engine is restarted, the engine speed can be reliably increased and the engine will not be stopped. In this way, the fuel cut time is lengthened by the amount that the return rotational speed is lowered, and fuel consumption can be reduced.

During the fuel cut, since the intake air is discharged from each cylinder as it is to the exhaust system, the three-way catalyst device 20 stores a large amount of oxygen due to the O ₂ storage capability. As a result, the oxidation action of the three-way catalyst device 20 becomes active, but the reduction action becomes inactive. When combustion is resumed at the time of return from the fuel cut, the three-way catalyst device 20 causes the HC in the exhaust gas. and CO can be favorably oxidized, nO _X in the exhaust gas can not be satisfactorily reduced.

Thus, at the time of return from the fuel cut, it is preferable to reduce the NO _x production amount for several cycles even if the combustion air fuel ratio is made richer than the stoichiometric air fuel ratio and the production amounts of HC and CO increase. Further, NO _X is the amount produced and the combustion temperature is high increases. Accordingly, the lower the mechanical compression ratio E at the time of return from the fuel cut, the larger the intake air amount, and the combustion temperature becomes higher as it is. Therefore, the target combustion air-fuel ratio AFt at the time of return from the fuel cut is As shown in FIG. 12, it is preferable that the lower the mechanical compression ratio E in the richer range than the theoretical air-fuel ratio S, the smaller. The combustion temperature can be lowered as the combustion air-fuel ratio is reduced in a richer range than the theoretical air-fuel ratio. As a result, at the time of return from fuel cut, the higher the mechanical compression ratio E and the larger the intake air amount, the smaller the return combustion air-fuel ratio becomes, and it is possible to suppress the increase in the amount of NO _x produced due to the higher combustion temperature. it can.

In returning from the fuel cut, the first combustion of each cylinder is particularly likely to increase the combustion temperature because the entire combustion chamber at the intake top dead center is filled with fresh air as described above. Accordingly, the target combustion air-fuel ratio AFt in the first combustion of each cylinder at the time of return is determined by MAP1 in FIG. 12, and is made considerably smaller than the theoretical air-fuel ratio S at each mechanical compression ratio E, and the combustion temperature becomes high. This suppresses an increase in the amount of NO _x produced.

Further, the second combustion of each cylinder at the time of return is performed by the first combustion in each cylinder so that the entire combustion chamber at the intake top dead center is almost filled with exhaust gas, and thus the exhaust gas remaining in the cylinder causes the first combustion. Since the combustion temperature is lower than the combustion, the target combustion air-fuel ratio AFt in the second combustion of each cylinder is determined by MAP2 in FIG. 12, and at each mechanical compression ratio E, the target combustion air-fuel ratio AFt in the first combustion is determined. Made bigger. Thereby, the amount of the NO _X in the combustion temperature becomes higher without consuming unnecessarily fuel is prevented from increasing.

The target combustion air-fuel ratio AFt determined by MAP2 in FIG. 12 may be applied to the combustion of each cylinder after the third time from the return of the fuel cut. Thus, when returning from the fuel cut, combustion is performed at a combustion air-fuel ratio richer than the stoichiometric air-fuel ratio in each cylinder for several cycles, and the oxygen stored by the O ₂ storage capability in the three-way catalyst device 20 is limited. If it is reduced to about half of the amount, the target combustion air-fuel ratio AFt of each cylinder is gradually made the stoichiometric air-fuel ratio S. During these several cycles, the mechanical compression ratio is preferably maintained at the mechanical compression ratio at the start of fuel cut.

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

Claims (2)

  In an internal combustion engine having a variable compression ratio mechanism in which the mechanical compression ratio is variable by the variable compression ratio mechanism and the actual compression ratio is constant, the lower the mechanical compression ratio at the time of the fuel cut, the lower the return rotation speed of the fuel cut. An internal combustion engine comprising a variable compression ratio mechanism characterized by the above.   2. The return combustion air-fuel ratio at the time of fuel cut return is made richer than the stoichiometric air-fuel ratio, and as the mechanical compression ratio at the time of fuel cut is lower, the return combustion air-fuel ratio is made smaller. An internal combustion engine provided with a variable compression ratio mechanism.

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