JP2013238124A - Internal combustion engine including variable compression ratio mechanism - Google Patents

Abstract

PROBLEM TO BE SOLVED: To suppress intake shortage caused by turbo lag of a turbocharger by advancing a valve closing timing of an intake valve without knocking when an engine rotation speed reaches or exceeds a set rotation speed in an internal combustion engine including a turbocharger and a supercharger that is stopped when the engine rotation speed reaches and exceeds the set rotation speed.SOLUTION: When a current engine rotation speed is a set rotation speed or lower, a current target upper limit value of a mechanical compression ratio is set (step 103). The smaller a difference between the set rotation speed and the current rotation speed is, the lower the current target upper limit value is set. When the current mechanical compression ratio exceeds the current target upper limit value, a variable compression ratio mechanism reduces the mechanical compression ratio to the current target upper limit value (step 105). When the engine rotation speed is changed into speed reaching or exceeding the set rotation speed, the variable valve timing mechanism increases an intake amount by advancing a valve closing timing of an intake valve.

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 varies a mechanical compression ratio by moving a cylinder block relative to a crankcase along a cylinder axis is known. In an internal combustion engine having such a variable compression ratio mechanism, it is effective to increase the expansion ratio equal to the mechanical compression ratio in order to increase the thermal efficiency. However, if the actual compression ratio is increased, knocking is likely to occur. End up. Thereby, even when the mechanical compression ratio is increased, the closing timing of the intake valve is retarded as the mechanical compression ratio is increased so that the actual compression ratio does not exceed the set compression ratio.

  Even in an internal combustion engine having a variable compression ratio mechanism, it is desired to increase engine output by supercharging, and a turbocharger that uses exhaust energy may be arranged. However, since the turbocharger does not function sufficiently at low engine speed, it has been proposed to provide an electric supercharger for assisting turbocharger supercharging at low engine speed (see Patent Document 1). .

JP-A-2005-163694 Japanese Patent Laid-Open No. 03-275936 JP 2004-218522 A JP 62-060933 A JP 62-077844 A

  Since the electric supercharger requires a motor mounting space, it is conceivable to replace the electric supercharger with an engine-driven supercharger. However, if the supercharger is driven efficiently at a low engine speed, there is a possibility that the supercharger may be damaged due to an excessive rotation at a high engine speed. As a result, the supercharger is connected to the engine drive shaft only at low engine speeds to assist turbocharger supercharging, and is separated from the engine drive shaft and stopped at high engine speeds. Has been implemented.

  However, even if the turbocharger is switched from a low engine speed to a high engine speed, it is not possible to realize sufficient supercharging immediately by the turbo lag. In this case, in order to eliminate the intake shortage, Although advancement of the valve closing timing is desired, knocking may occur if the actual compression ratio exceeds the set compression ratio.

  Accordingly, an object of the present invention is an internal combustion engine provided with a variable compression ratio mechanism that makes a mechanical compression ratio variable, a variable valve timing mechanism that makes the closing timing of an intake valve variable, a turbocharger, and an engine speed. In an internal combustion engine equipped with a supercharger that stops when the engine speed is greater than or equal to the set speed, when the engine speed is less than the set speed and greater than or equal to the set speed, the intake valve is closed without causing knocking. The shortage of intake caused by the turbo lag of the turbocharger can be suppressed by the advance of the timing.

  An internal combustion engine comprising the variable compression ratio mechanism according to claim 1 of the present invention closes the intake valve so that the actual compression ratio does not exceed the set compression ratio with respect to the mechanical compression ratio controlled by the variable compression ratio mechanism. A variable valve timing mechanism that makes the timing variable, a turbocharger, and a supercharger that stops when the engine speed is equal to or higher than the set speed, and that is used when the current engine speed is lower than the set speed. The current target upper limit value of the compression ratio is set. The smaller the difference between the set engine speed and the current engine speed, the lower the current target upper limit value. The current mechanical compression ratio is higher than the current target upper limit value. Sometimes, the mechanical compression ratio is lowered toward the current target upper limit value by the variable compression ratio mechanism, and when the engine speed is switched from less than the set speed to more than the set speed, Characterized in that to increase the intake air amount by advancing the closing timing of the intake valve by the variable valve timing mechanism.

  An internal combustion engine provided with the variable compression ratio mechanism according to claim 2 of the present invention is an internal combustion engine provided with the variable compression ratio mechanism according to claim 1, and when the engine speed is in transition, the increase speed of the current engine speed is increased. The higher the value, the lower the current target upper limit value.

  An internal combustion engine comprising the variable compression ratio mechanism according to claim 3 according to the present invention is the internal combustion engine comprising the variable compression ratio mechanism according to claim 1 or 2, wherein the turbocharger includes a compressor immediately downstream and a directly upstream side. When the differential pressure is higher than the set pressure, the setting of the current target upper limit value is stopped.

  An internal combustion engine having the variable compression ratio mechanism according to claim 4 according to the present invention is the internal combustion engine having the variable compression ratio mechanism according to any one of claims 1 to 3, wherein the current mechanical compression ratio is When higher than the target upper limit value, the current engine speed is decreased by the transmission.

  According to the internal combustion engine including the variable compression ratio mechanism according to the first aspect of the present invention, the intake valve is controlled so that the actual compression ratio does not exceed the set compression ratio with respect to the mechanical compression ratio controlled by the variable compression ratio mechanism. A variable valve timing mechanism that makes the valve closing timing variable, a turbocharger, and a supercharger that stops when the engine speed is equal to or higher than the set speed, and when the current engine speed is lower than the set speed, The current target upper limit value of the mechanical compression ratio is set, and the smaller the difference between the set engine speed and the current engine speed, that is, the current engine speed is close to the set engine speed and the engine speed changes immediately thereafter. The higher the possibility of exceeding the set speed, the lower the current target upper limit value. When the current mechanical compression ratio is higher than the current target upper limit value, the variable compression ratio mechanism reduces the mechanical compression ratio. Toward the target upper limit of the resident and is adapted to decrease. As a result, when the engine speed is switched from less than the set speed to more than the set speed, the mechanical compression ratio is lowered toward the target upper limit value, and the variable valve timing mechanism is used to suppress the intake shortage. Even if the valve closing timing is advanced, the actual compression ratio does not exceed the set compression ratio, knocking does not occur, the intake amount can be increased, and the intake shortage caused by the turbo lag of the turbocharger is suppressed. be able to.

  According to the internal combustion engine having the variable compression ratio mechanism according to the second aspect of the present invention, in the internal combustion engine having the variable compression ratio mechanism according to the first aspect, when the engine speed is in a transient state, The higher the rising speed, the lower the current target upper limit value. As a result, the current target upper limit value is lowered and the current mechanical compression ratio is reduced as the current engine speed increases at a higher rate and is more likely to exceed the set speed due to a change in the engine speed immediately thereafter. When it is higher than the target upper limit value, the mechanical compression ratio is lowered toward the current target upper limit value by the variable compression ratio mechanism.

  According to the internal combustion engine having the variable compression ratio mechanism according to claim 3 of the present invention, in the internal combustion engine having the variable compression ratio mechanism according to claim 1 or 2, the immediately downstream side and the immediately upstream side of the compressor of the turbocharger. When the differential pressure from the engine is higher than the set pressure, even if the engine speed immediately switches from less than the set speed to more than the set speed, sufficient turbocharging can be achieved only with the turbocharger. Since the intake shortage does not occur, the setting of the current target upper limit value is stopped and the mechanical compression ratio is not changed unnecessarily.

  According to an internal combustion engine including the variable compression ratio mechanism according to claim 4 of the present invention, in the internal combustion engine including the variable compression ratio mechanism according to any one of claims 1 to 3, the current mechanical compression ratio is When the engine speed is higher than the current target upper limit value, if the engine speed immediately switches from less than the set speed to more than the set speed, the mechanical compression ratio is relatively high and the actual compression ratio exceeds the set compression ratio. In addition, the valve closing timing of the intake valve cannot be sufficiently advanced by the variable valve timing mechanism, and intake shortage may occur due to the turbo lag of the turbocharger. Therefore, at this time, the current engine speed is reduced by the transmission without changing the engine output, and it is difficult for the engine speed to switch from less than the set speed to more than the set speed. It is possible to decrease toward the target upper limit value.

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 setting the current target upper limit value of the mechanical compression ratio. It is a map which shows the relationship between the difference between the setting engine speed which stops a supercharger, the present engine speed, and the current target upper limit of an engine compression ratio. It is a map which shows the relationship between the change speed of an engine speed, and the coefficient for correct | amending a target upper limit.

  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.

  An upstream supercharger compressor 28 and a downstream turbocharger compressor 29 are arranged in series between the throttle valve 17 of the intake duct 14 and the intake air amount detector 18. An air bypass passage 28a that bypasses the compressor 28 of the supercharger is provided. The air bypass passage 28a controls the amount of air that passes through the air bypass passage 28a without passing through the compressor 28 of the supercharger. 28b is arranged. The turbocharger compressor 29 is driven by using exhaust pressure, and is connected to a turbine (not shown) disposed in the engine exhaust system. The compressor 28 of the supercharger is an engine drive type, and is connected to the engine drive shaft via an electromagnetic clutch (not shown), and can be stopped by being separated from the engine drive shaft by the electromagnetic clutch. . The exhaust port 10 is connected via an exhaust manifold 19 to, for example, a catalyst device 20 containing a three-way catalyst, 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, air-fuel ratio sensor 21, relative position sensor 22, valve timing sensor 23, and throttle opening sensor 24 are input to the input port 35 via 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 drive actuator 16, an electromagnetic clutch disposed between the compressor 28 of the supercharger and the engine drive shaft via a corresponding drive circuit 38, variable. The compression ratio mechanism A and the variable valve timing mechanism B are 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 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. 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, 3B, and 3C show the positional relationship between the center a of the concentric portion 58, the center b of the eccentric portion 57, and the center c of the circular cam 56 in each state. It is shown.

  3A to 3C, the relative position of the crankcase 1 and the cylinder block 2 is determined by the distance between the center a of the concentric part 58 and the center c of the circular cam 56, and the concentric part. As the distance between the center a of 58 and the center c of the circular cam 56 increases, the cylinder block 2 moves away from the crankcase 1. That is, the variable compression ratio mechanism A changes the relative position between the crankcase 1 and the cylinder block 2 by a crank mechanism using a rotating cam. When the cylinder block 2 moves away from the crankcase 1, the volume of the combustion chamber 5 increases when the piston 4 is positioned at the compression top dead center. Therefore, by rotating the camshafts 54 and 55, the piston 4 is compressed at the top dead center. The volume of the combustion chamber 5 when it is located at can be changed.

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

  In the internal combustion engine of the present embodiment, as described above, the turbocharger compressor 29 and the supercharger compressor 28 are arranged in the intake duct 14, and the engine output is intended to be increased by supercharging. The turbocharger compressor 29 cannot sufficiently increase the supercharging pressure when the exhaust pressure is low as in the case of low engine speed, and the supercharger compressor 29 assists the supercharging of the turbocharger at low engine speed. 28 is provided.

  If the compressor 28 of the supercharger is driven efficiently by the engine drive shaft at the time of low engine speed, it may be damaged due to excessive rotation at the time of high engine speed. When the number exceeds a certain value, the compressor 28 of the supercharger is separated from the engine drive shaft by an electromagnetic clutch.

  In this way, supercharging is performed only by the turbocharger above the set rotational speed. However, even if the engine speed is less than the set speed or more than the set speed, the turbocharger does not immediately increase the turbine speed to the desired speed, and this turbo lag realizes sufficient supercharging immediately. I can't. At this time, in order to eliminate the shortage of intake, it is desired to advance the valve closing timing of the intake valve 7 by the variable valve timing mechanism B. However, the mechanical compression ratio cannot be reduced immediately by the variable compression ratio mechanism A. Depending on the current mechanical compression ratio, the actual compression ratio becomes higher than the set compression ratio due to the advance timing of the closing timing of the intake valve 7, Knocking may occur.

  The internal combustion engine of the present embodiment sets the current target upper limit value of the mechanical compression ratio according to the flowchart shown in FIG. 10, and when the current mechanical compression ratio is higher than the current target upper limit value, the variable compression ratio mechanism A performs mechanical compression. The ratio is reduced toward the current target upper limit. Thus, when the engine speed is less than the set speed and greater than or equal to the set speed, the intake valve is within the range where the actual compression ratio does not exceed the set compression ratio even if the target upper limit value of the mechanical compression ratio is not realized. It is intended to suppress the intake shortage caused by the turbo lag of the turbocharger by the advance angle of the closing timing of the intake valve without causing knocking.

  This flowchart is repeatedly performed by the electronic control unit 30 every set time. First, in step 101, it is determined whether or not the current engine speed N detected based on the output of the crank angle sensor 42 is lower than the set speed N '. The set rotational speed N 'is a threshold value for separating the supercharger compressor 28 from the engine drive shaft by the electromagnetic clutch, and when the determination in step 101 is negative, the process is terminated.

  On the other hand, when the determination in step 101 is affirmative, the supercharger compressor 28 is currently driven by the engine drive shaft, and the supercharger compressor 28 is disconnected from the engine drive shaft depending on the fluctuation of the engine speed in the future. there is a possibility. In step 102, it is determined whether or not the differential pressure P between the downstream side and the downstream side of the compressor 29 of the turbocharger is equal to or higher than a set pressure P '. The differential pressure P is the supercharging pressure by only the turbocharger compressor 29 ignoring the supercharging amount of the supercharger compressor 28, and pressure sensors are provided on the upstream side and the downstream side of the compressor 29 in the exhaust duct 14. It can be measured by arranging.

  The set pressure P ′ is the maximum supercharging pressure intended to be realized only by the compressor 29 of the turbocharger when the engine speed N is equal to or higher than the set speed N ′. If the determination in step 102 is affirmed Immediately, even if the supercharger compressor 28 is disconnected from the engine drive shaft, a desired supercharging pressure can be realized only by the turbocharger compressor 29, and intake shortage does not occur. As a result, it is not necessary to prepare for when the compressor 28 of the supercharger is disconnected from the engine drive shaft, and the process is terminated without unnecessarily lowering the mechanical compression ratio and deteriorating the thermal efficiency.

  On the other hand, when the determination in step 102 is negative, in step 103, based on the difference ΔN between the set rotational speed N ′ and the current engine speed N, the current target upper limit Et of the mechanical compression ratio is determined from the map shown in FIG. Set '.

  In the map shown in FIG. 11, the target upper limit Et ′ is lowered as the difference ΔN is smaller. The target upper limit value Et ′ is preferably changed from a value slightly smaller than the upper limit mechanical compression ratio shown in FIG. 3C to a value slightly larger than the lower limit mechanical compression ratio shown in FIG. When the difference ΔN is sufficiently large (larger than the set difference), the upper limit mechanical compression ratio shown in FIG. 3C may be set as the target upper limit value Et ′, or the target upper limit value Et ′ may not be set. good.

  In this way, the current target upper limit Et ′ of the mechanical compression ratio becomes higher as the current engine speed N is closer to the set speed N ′ and is more likely to become equal to or higher than the set speed N ′ due to the change in the engine speed immediately after that. Lowered. Next, at step 104, it is determined whether or not the current mechanical compression ratio E is higher than the current target upper limit value Et '. If this determination is negative, since the current mechanical compression ratio E is equal to or less than the current target upper limit value Et ′, the process ends as it is.

  However, when the determination in step 104 is affirmative, in step 105, the current mechanical compression ratio E is lowered toward the target upper limit value Et 'by the variable compression ratio mechanism A. At this time, if the mechanical compression ratio is being changed toward another target value Et due to a change in engine load or the like, the mechanical compression ratio decreases to the lower of the current target upper limit value Et ′ and the target value Et. Reduce towards.

  Thus, when the differential pressure P between the downstream side and the downstream side of the compressor 29 of the turbocharger is less than the set pressure P ′, the mechanical compression is increased as the current engine speed N is closer to the set speed N ′. Since the ratio E is lowered, when the engine speed N is switched from less than the set speed N ′ to more than the set speed N ′, the valve closing timing of the intake valve 7 is advanced by the variable valve timing mechanism B. However, the actual compression ratio exceeds the set compression ratio, knocking does not occur, the intake air amount can be increased, and intake shortage caused by the turbo lag of the turbocharger can be suppressed.

  Of course, when the engine speed N is switched from less than the set speed N ′ to more than the set speed N ′, the target upper limit value Et ′ of the mechanical compression ratio at this time may not be realized. The intake valve closing timing can be advanced by a small amount so that the actual compression ratio does not exceed the set compression ratio, thereby completely eliminating the intake shortage caused by the turbo lag of the turbocharger. It is possible to suppress even without it.

  In the variable compression ratio mechanism A of this embodiment, the mechanical compression ratio cannot be changed instantaneously, and a certain amount of time is required for changing the mechanical compression ratio. Thereby, depending on the decrease in the mechanical compression ratio in step 104, the mechanical compression ratio E cannot be sufficiently reduced when the engine speed N is switched from less than the set speed N ′ to more than the set speed N ′. There is.

  In the flowchart of FIG. 10, in step 106, the current engine speed N is reduced by the transmission without changing the engine output, and the engine speed N is less than the set speed N ′ to the set speed N ′ or more. Therefore, it is possible to sufficiently reduce the mechanical compression ratio E toward the current target upper limit value Et ′.

  In Step 103, the target upper limit Et ′ is set lower as the difference between the set engine speed N ′ and the current engine speed N is smaller. However, when the engine speed is in transition, the current engine speed N ′ is set lower. The current target upper limit value Et ′ may be further decreased as the speed of increase in the rotational speed N is higher. As a result, the current target upper limit Et ′ is lowered as the current increase speed of the engine speed N is higher and the possibility that the engine speed N immediately exceeds the set speed N ′ is higher.

  Here, the current increase speed V of the engine speed N is, for example, the difference between the current engine speed N measured this time and the engine speed N1 measured last time in the flowchart of FIG. It can be calculated by dividing by time (execution interval). Based on the rising speed V thus calculated, a correction coefficient k (a positive number less than 1) is set from the map shown in FIG. 12, and the target upper limit value Et ′ set from FIG. The current target upper limit value Et ′ (= Et ′ * k) may be used. Of course, the new target upper limit value Et ′ must not be set lower than the lower limit mechanical compression ratio shown in FIG.

28 Supercharger compressor 29 Turbocharger compressor 30 Electronic control unit A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (4)

  The variable valve timing mechanism that makes the intake valve closing timing variable so that the actual compression ratio does not exceed the set compression ratio with respect to the mechanical compression ratio controlled by the variable compression ratio mechanism, the turbocharger, and the engine speed A supercharger that stops when the engine speed is greater than or equal to the set speed, and the current target upper limit value of the mechanical compression ratio is set when the current engine speed is lower than the set speed, the set speed and the current The smaller the difference from the engine speed, the lower the current target upper limit value. When the current mechanical compression ratio is higher than the current target upper limit value, the variable compression ratio mechanism directs the mechanical compression ratio to the current target upper limit value. When the engine speed is switched from less than the set speed to above the set speed, the variable valve timing mechanism advances the intake valve closing timing to Internal combustion engine having a variable compression ratio mechanism, characterized in that to increase.   2. The internal combustion engine having a variable compression ratio mechanism according to claim 1, wherein when the engine speed is in transition, the current target upper limit value is lowered as the current engine speed increases at a higher speed.   3. The setting of the current target upper limit value is stopped when the differential pressure between the immediately downstream side and the immediately upstream side of the compressor of the turbocharger is higher than a set pressure. 4. An internal combustion engine provided with a variable compression ratio mechanism.   The variable compression ratio mechanism according to any one of claims 1 to 3, wherein when the current mechanical compression ratio is higher than a current target upper limit value, the current engine speed is reduced by the transmission. Internal combustion engine.

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