JP5754369B2 - 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 capable of increasing a mechanical compression ratio and increasing an expansion ratio in order to improve thermal efficiency at the time of engine low load is known (see Patent Document 1). In this internal combustion engine, if the actual compression ratio is increased at the same time as the expansion ratio, knocking is likely to occur. Therefore, the closing timing of the intake valve is retarded so that the actual compression ratio remains constant even if the mechanical compression ratio is increased. To be controlled.

  Since the thermal efficiency deteriorates as the engine load decreases, the mechanical compression ratio increases as the engine load decreases. As a result, the intake valve closing timing is delayed as the engine load decreases, so that the throttle valve is fully opened. However, the intake air amount can be controlled. Thus, the pumping loss can be reduced by fully opening the throttle valve.

JP2007-303423 JP 2002-276478 A JP 2001-342857 A JP2008-002437

  Incidentally, in general, an internal combustion engine is provided with a gasoline vapor adsorption device in order to suppress release of gasoline vapor generated from a fuel tank to the atmosphere. Such a gasoline vapor adsorbing device cannot adsorb gasoline vapor indefinitely, and it is necessary to discharge gasoline adsorbed before exceeding the limit amount to the intake system.

  In general, in order to allow gasoline to be discharged from the gasoline vapor adsorption device to the intake system without requiring a pump or other device, the intake pipe negative pressure generated on the downstream side of the throttle valve of the intake system is reduced. Used. However, if the throttle valve is fully open, intake pipe negative pressure cannot be generated. As a result, as in the above-described internal combustion engine, if the period during which the throttle valve is fully opened becomes long, the gasoline cannot be discharged from the gasoline vapor adsorption device to the intake system, and the adsorbed gasoline becomes the limit amount. It may become impossible to adsorb gasoline vapor to the gasoline vapor adsorption device.

  Therefore, an object of the present invention is to provide an internal combustion engine having a variable compression ratio mechanism that makes the mechanical compression ratio variable, even when the throttle valve opening is larger than the set opening and the intake pipe negative pressure cannot be generated. This means that gasoline can be discharged from the gasoline vapor adsorption device to the intake system without using an electric pump or the like.

  In the internal combustion engine having the variable compression ratio mechanism according to the first aspect of the present invention, the gasoline vapor adsorbing device is connected to the intake passage downstream of the surge tank of the intake system, and the opening degree of the throttle valve is set to the set opening degree. When the gasoline vapor adsorbing device is to release the gasoline into the intake passage when it is larger, the mechanical compression ratio is changed by the variable compression ratio mechanism so that the piston top surface of the intake top dead center to the surge tank is changed. By changing the length of the intake flow path, a stationary wave of the intake pulsation is generated in the intake flow path, and the connection position of the gasoline vapor adsorption device of the intake system becomes negative pressure due to the stationary wave of the intake pulsation Gasoline is discharged from the gasoline vapor adsorption device to the intake system.

  An internal combustion engine comprising the variable compression ratio mechanism according to claim 2 of the present invention is the internal combustion engine comprising the variable compression ratio mechanism according to claim 1, in order to generate a stationary wave of intake pulsation in the intake flow path. The length of the intake passage is changed by changing the mechanical compression ratio by the variable compression ratio mechanism according to the current engine speed.

  According to the internal combustion engine having the variable compression ratio mechanism according to the first aspect of the present invention, the gasoline vapor adsorption device is connected to the intake passage on the downstream side of the surge tank of the intake system, and the opening degree of the throttle valve is set. When attempting to release gasoline from the gasoline vapor adsorption device to the intake passage when the opening is larger, the mechanical compression ratio is changed by the variable compression ratio mechanism, and the intake flow from the piston top surface of the intake top dead center to the surge tank By changing the length of the path, a stationary wave of intake pulsation is generated in the intake flow path, and when the connection position of the intake system gasoline vapor adsorber becomes negative pressure due to the stationary wave of intake pulsation, the gasoline vapor adsorber When the throttle valve opening is larger than the set opening and intake pipe negative pressure does not occur as it is Can have, without the use of such as an electric pump, gasoline can be released from the gasoline vapor adsorption device to the intake system.

  According to the internal combustion engine having the variable compression ratio mechanism according to claim 2 of the present invention, in the internal combustion engine having the variable compression ratio mechanism according to claim 1, in order to generate a stationary wave of the intake pulsation in the intake passage. Depending on the current engine speed, the variable compression ratio mechanism changes the mechanical compression ratio to change the length of the intake flow path. In addition, a stationary wave of intake pulsation can be generated in the intake flow path for each engine speed.

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. 7 is a flowchart showing control of a variable compression ratio mechanism and a control valve for a gasoline vapor discharge passage for enabling the release of gasoline vapor from the gasoline vapor adsorption device even when the opening of the throttle valve is relatively large. It is a figure for demonstrating the length of an intake flow path. It is a map which shows the relationship between an engine speed and the mechanical compression ratio which generates the stationary wave of intake pulsation. It is a time chart which shows the pressure change of the connection position of the gasoline vapor | steam discharge passage of an intake branch pipe.

  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. A gasoline vapor adsorption device 90 is connected to the intake branch pipe 11 on the downstream side of the surge tank 12 via a gasoline vapor discharge passage 91. A control valve 92 is disposed in the gasoline vapor discharge passage 91.

  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 drive actuator 16, the control valve 92 of the gasoline vapor discharge passage 91, the variable compression ratio mechanism A, and the variable valve timing mechanism via the corresponding drive circuit 38. Connected to B.

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

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

  When the circular cams 58 fixed on the camshafts 54 and 55 are rotated in opposite directions as indicated by arrows in FIG. 3A from the state shown in FIG. 3A, the eccentric shafts 57 are separated from each other. In order to move in the direction, the circular cam 56 rotates in the opposite direction to the circular cam 58 in the cam insertion hole 51, and as shown in FIG. It becomes. Next, when the circular cam 58 is further rotated in the direction indicated by the arrow, the eccentric shaft 57 is at the lowest position as shown in FIG.

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

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

  When the engine load is further reduced, the mechanical compression ratio is further increased, 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.

  As described with reference to FIG. 9, the mechanical compression ratio is increased in order to increase the expansion ratio by the variable compression ratio mechanism A as the engine load decreases, but the actual compression ratio becomes too high and knocking or pre-ignition is caused. In order not to generate this, the valve closing timing of the intake valve is retarded by the variable valve timing mechanism B so that the actual compression ratio is constant. Until the closing timing of the intake valve is most retarded (more than the engine load L1), the throttle valve is fully opened, and the intake air amount is controlled by the retardation of the closing timing of the intake valve.

  By the way, the gasoline vapor adsorption device 90 described above is also communicated with a fuel tank (not shown), and adsorbs gasoline vapor generated in the fuel tank with activated carbon or the like so as not to be released to the atmosphere. However, it is not possible to adsorb gasoline vapor indefinitely, and it is necessary to release gasoline vapor adsorbed before exceeding the limit amount to the intake system.

  In this embodiment, the gasoline vapor adsorption device 90 is connected to the downstream side of the surge tank 12 of the intake branch pipe 11 by the gasoline vapor discharge passage 91, and when the opening of the throttle valve 17 is relatively small, the intake branch pipe The gasoline vapor can be discharged to the gasoline vapor adsorption device 90 intake branch pipe 11 by using the intake negative pressure generated in the gas generator 11. However, if the throttle valve 17 is fully open, if no action is taken, no intake negative pressure is generated in the intake branch pipe 11, so that gasoline vapor cannot be released from the gasoline vapor adsorption device 90.

  In the internal combustion engine of the present embodiment, as explained in FIG. 9, when the engine load is L1 or more, the throttle valve 17 is fully opened, and thus the period during which the throttle valve is fully opened is long. As a result, the gasoline vapor adsorption device 90 may limit the amount of adsorbed gasoline, and the gasoline vapor adsorption device 90 may not be able to adsorb the gasoline vapor.

  In the internal combustion engine of the present embodiment, the variable compression ratio mechanism A and the control valve 92 of the gasoline vapor discharge passage 91 are controlled according to the flowchart shown in FIG. Gasoline vapor can be released. This flowchart is executed by the electronic control unit 30 simultaneously with the engine start.

  First, in step 101, it is determined whether or not the current gasoline vapor adsorption amount A of the gasoline vapor adsorption device 90 has become a set amount A 'slightly smaller than the limit amount. If this determination is negative, the gasoline vapor adsorbing device 90 still has a margin for adsorbing gasoline vapor, and the process is terminated. In the determination of step 101, it can be utilized that the longer the elapsed time from the completion of the previous release of gasoline vapor, the greater the current gasoline vapor adsorption amount A of the gasoline vapor adsorption device 90, and the like.

  On the other hand, when the determination in step 101 is affirmative, the gasoline vapor adsorbing device 90 does not have much room to adsorb gasoline vapor, and it is desirable to release the gasoline vapor to the intake system. At this time, the process proceeds to step 102. In step 102, based on the output of the throttle opening sensor 24, it is determined whether or not the opening TA of the throttle valve 17 is equal to or greater than a set opening A 'near the fully open position. When this determination is negative, since a relatively large intake pipe negative pressure is generated in the intake branch pipe 11, the control valve 92 is opened and the gasoline vapor discharge passage 91 is opened in step 103. The gasoline vapor can be released from the gasoline vapor adsorption device 90 to the intake system.

  Next, in step 104, it is determined whether or not the elapsed time T from the start of the release of gasoline vapor from the gasoline vapor adsorption device 90 has reached the first set time T1, and when this determination is negative, Assuming that the gasoline vapor is not sufficiently released from the gasoline vapor adsorption device 90, the process returns to step 103 and the control valve 92 is kept open. On the other hand, if the determination in step 104 is affirmative, it is determined that all gasoline vapor adsorbed by the gasoline vapor adsorption device 90 has been released, and the control valve 92 is closed in step 108 and the process is terminated.

  Further, when the opening degree TA of the throttle valve 17 is equal to or larger than the set opening degree TA ′ and the determination in step 102 is affirmed, the intake pipe negative pressure is not generated in the intake branch pipe 11 as it is. In this flowchart, in step 105, the mechanical compression ratio E is changed by the variable mechanical compression ratio mechanism A, and a stationary wave of the intake pulsation is generated by the air column natural vibration in the intake passage.

  Here, as shown in FIG. 11, the intake flow path is from the top surface of the piston 4 at the intake top dead center including the intake branch pipe 11 and the intake port 8 to the surge tank 12, and the mechanical compression ratio E is By changing the distance in the cylinder from the top surface of the top dead center piston to the intake port 8 to change the length LE of the intake flow path, the intake top dead center is generated in the cylinder at each intake stroke. The intake pulsation reflected as a positive pressure wave in the surge tank 12 where the negative pressure wave is at the open end can be a stationary wave.

  As shown in FIG. 12, the lower the current engine speed NE, the lower the mechanical compression ratio E and the longer the length LE of the intake passage, thereby generating a stationary wave of intake pulsation. ing. In this embodiment, at the time of high rotation (for example, 5000 rpm), the intake branch pipe 11 is set so that a stationary wave of intake pulsation is generated in the intake flow path at the upper limit mechanical compression ratio at which the length LE of the intake flow path is the shortest. And the length of the intake port 8 is designed.

  Thus, when the mechanical compression ratio E is changed so as to generate a stationary wave of the intake pulsation in the intake path based on the current engine speed, in step 106, the opening / closing control of the control valve 92 of the gasoline vapor discharge passage 91 is performed. The FIG. 13 is a time chart showing a change in pressure generated at the connection position of the gasoline vapor discharge passage 91 of the intake branch pipe 11, that is, the downstream side of the surge tank 12 in the intake passage when a stationary wave of intake pulsation is generated. It is. The connection position of the gasoline vapor discharge passage 91 is preferably in the vicinity of a node of a standing wave to be generated, whereby the pressure changes greatly at the connection position, and as shown in FIG. 13, between time t1 and time t3, time t3 From t5 to t4 and from t5 to t6. As described above, since the connecting position of the gasoline vapor discharge passage 91 periodically becomes a negative pressure, the control valve 92 is controlled to be opened and closed so that the gasoline vapor discharge passage 91 is opened only during the negative pressure. .

  Thereby, even when the opening degree TA of the throttle valve 17 is equal to or larger than the set opening degree TA ′ and the intake pipe negative pressure cannot be generated as it is, the gasoline vapor is released from the gasoline vapor adsorption device 90 to the intake system. be able to.

  Next, in step 107, it is determined whether or not the elapsed time T from the start of the release of gasoline vapor from the gasoline vapor adsorption device 90 has reached the second set time T2, and when this determination is negative, Assuming that the gasoline vapor is not sufficiently released from the gasoline vapor adsorption device 90, the process returns to step 106, and the control valve 92 continues to be opened and closed. On the other hand, when the determination in step 107 is affirmed, it is determined that all gasoline vapor adsorbed by the gasoline vapor adsorption device 90 has been released, and the control valve 92 is closed in step 108 and the process is terminated. As described above, when the gasoline vapor is released from the gasoline vapor adsorption device 90 by generating a stationary wave of the intake pulsation, the intake pipe negative pressure generated when the throttle valve opening is relatively small is used as compared with the case where the intake pipe negative pressure is used. Since it is difficult to release gasoline vapor, the second set time T2 is set longer than the first set time T1.

  When gasoline vapor is released to the intake system, the amount of fuel injected from the fuel injection valve 13 is reduced by air-fuel ratio control using the air-fuel ratio sensor 21. In the present embodiment, since the gasoline vapor discharge passage 91 is connected to the intake branch pipe 11 of the specific cylinder, the gasoline amount of the fuel injection valve 13 of the specific cylinder is controlled when the gasoline vapor is discharged. Is required. When the air-fuel ratio sensor 21 is arranged downstream of the exhaust collecting portion of each cylinder, the air-fuel ratio of the exhaust gas immediately after the exhaust stroke of the specific cylinder is detected in order to detect the air-fuel ratio of the exhaust gas of the specific cylinder. What is necessary is just to detect. Of course, the air-fuel ratio sensor for a specific cylinder may be arranged in the exhaust passage of the specific cylinder upstream of the exhaust collecting portion of each cylinder.

DESCRIPTION OF SYMBOLS 1 Crankcase 2 Cylinder block 4 Piston 12 Surge tank 17 Throttle valve 90 Gasoline vapor adsorption apparatus 91 Gasoline vapor discharge passage 92 Control valve A Variable compression ratio mechanism

Claims (2)

  When the gasoline vapor adsorbing device is connected to the intake passage downstream of the intake system surge tank and the throttle valve opening is larger than the set opening, let the gasoline vapor adsorbing device release the gasoline to the intake passage. In this case, by changing the mechanical compression ratio by the variable compression ratio mechanism and changing the length of the intake passage from the top surface of the piston at the intake top dead center to the surge tank, the intake pulsation is generated in the intake passage. Is generated, and when the connection position of the gasoline vapor adsorption device of the intake system becomes negative pressure due to the stationary wave of the intake pulsation, gasoline is discharged from the gasoline vapor adsorption device to the intake system. An internal combustion engine provided with a variable compression ratio mechanism.   In order to generate a stationary wave of intake pulsation in the intake passage, the length of the intake passage is changed by changing the mechanical compression ratio by the variable compression ratio mechanism according to the current engine speed. An internal combustion engine comprising the variable compression ratio mechanism according to claim 1.

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