JP5545249B2 - Internal combustion engine having a variable compression ratio mechanism - Google Patents

Description

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

  An internal combustion engine having a variable compression ratio mechanism that moves a cylinder block relative to a crankcase is known (see Patent Document 1). An internal combustion engine having such a variable compression ratio mechanism is also provided with a blow-by gas recirculation device that recirculates the blow-by gas in the crankcase to the intake system.

  The blow-by gas recirculation device communicates the inside of the crankcase and a surge tank on the downstream side of the throttle valve in the intake system via a PCV (Positive Crankcase Ventilation) valve. In the PCV valve, the passage sectional area decreases as the absolute value of the intake pipe negative pressure on the downstream side of the throttle valve increases, and the passage sectional area increases as the absolute value of the intake pipe negative pressure decreases and approaches atmospheric pressure. As a result, the smaller the intake amount and the larger the absolute value of the intake pipe negative pressure downstream of the throttle valve, the smaller the blow-by gas recirculation amount is reduced by the PCV valve. The increase in the amount of blow-by gas that leaks from the cylinder into the crankcase due to pressure is suppressed.

JP2010-261418A JP 2009-293496 A JP2009-114989A JP 2004-076695 A

  Thus, the blow-by gas recirculation amount from the crankcase to the intake system is determined by the magnitude of the absolute value of the intake pipe negative pressure by the PCV valve. By the way, in an internal combustion engine having a variable compression ratio mechanism that moves the cylinder block relative to the crankcase, when the cylinder block is moved closer to the crankcase in order to increase the mechanical compression ratio, the volume in the crankcase decreases. However, at this time, the intake pipe negative pressure does not change, and the amount of blow-by gas recirculation passing through the PCV valve hardly changes, so the crankcase internal pressure becomes abnormally high, and the cylinder block and the crankcase Blow-by gas and lubricating oil in the crankcase are likely to leak from the sliding seal portion.

  Accordingly, an object of the present invention is an internal combustion engine including a variable compression ratio mechanism that moves a cylinder block relative to a crankcase, and a blowby gas recirculation device that communicates the crankcase with a downstream side of a throttle valve in an intake system. In the internal combustion engine provided, it is possible to suppress an abnormal increase in the crankcase internal pressure when the cylinder block is brought close to the crankcase in order to increase the mechanical compression ratio by the variable compression ratio mechanism.

  An internal combustion engine having a variable compression ratio mechanism according to claim 1 of the present invention is an internal combustion engine having a variable compression ratio mechanism for moving a cylinder block relative to a crankcase, wherein the throttle in the crankcase and the intake system. A blow-by gas recirculation device that communicates with the downstream side of the valve is provided, and the blow-by gas recirculation device includes a PCV valve that reduces the cross-sectional area of the passage as the absolute value of the intake pipe negative pressure on the downstream side of the throttle valve increases; A pressure reducing passage that communicates the crankcase and the intake system is provided, and a relief valve that opens only when the inside of the crankcase is equal to or higher than a set pressure is disposed in the pressure reducing passage. Features.

  An internal combustion engine comprising the variable compression ratio mechanism according to claim 2 according to the present invention is the internal combustion engine comprising the variable compression ratio mechanism according to claim 1, wherein the pressure reducing passage is provided with a closing mechanism, and the closing mechanism is The pressure reducing passage is closed when the mechanical compression ratio is lower than a set value by the variable compression ratio mechanism.

  According to the internal combustion engine having the variable compression ratio mechanism according to the first aspect of the present invention, the internal combustion engine includes a variable compression ratio mechanism that moves the cylinder block relative to the crankcase. A blow-by gas recirculation device that communicates with the downstream side of the throttle valve is provided. The blow-by gas recirculation device includes a PCV valve that decreases the cross-sectional area of the passage as the absolute value of the intake pipe negative pressure on the downstream side of the throttle valve increases. A pressure reducing passage that communicates the crankcase and the intake system is provided, and a relief valve that opens only when the inside of the crankcase is equal to or higher than a set pressure is disposed in the pressure reducing passage. As a result, when the cylinder block is moved closer to the crankcase to increase the mechanical compression ratio by the variable compression ratio mechanism, the relief valve opens when the volume of the crankcase decreases and the crankcase internal pressure rises above the set pressure. Then, blow-by gas in the crankcase is released to the intake system through the decompression passage, and the inside of the crankcase can be made less than the set pressure. In this way, it is possible to suppress an abnormal increase in the crankcase internal pressure.

  An internal combustion engine having the variable compression ratio mechanism according to claim 2 according to the present invention is the internal combustion engine having the variable compression ratio mechanism according to claim 1, wherein the pressure reducing passage is provided with a closing mechanism, and the closing mechanism is variable. When the mechanical compression ratio is made lower than the set value by the compression ratio mechanism, the pressure reducing passage is closed, so that the mechanical compression ratio is lower than the set value and the crankcase internal pressure is increased above the set pressure. If not, the pressure reducing passage is closed by the closing mechanism, and the possibility that the blow-by gas in the crankcase leaks to the outside air through the pressure reducing passage can be reduced.

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. FIG. 4 is a schematic partial cross-sectional view passing through the axis of each cylinder of the internal combustion engine at the high mechanical compression ratio shown in FIG. FIG. 11 is a part of the schematic partial cross-sectional view of FIG. 10 at the low mechanical compression ratio shown in FIG.

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

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

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

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

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

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

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

  When the concentric portions 58 of the camshafts 54 and 55 are rotated in opposite directions as shown by arrows in FIG. 3A from the state shown in FIG. 3A, the eccentric portions 57 move in directions away from each other. Therefore, the circular cam 56 rotates in the direction opposite to the concentric portion 58 in the cam insertion hole 51, and the position of the eccentric portion 57 is changed from a high position to an intermediate height position as shown in FIG. Next, when the concentric portion 58 is further rotated in the direction indicated by the arrow, the eccentric portion 57 is at the lowest position as shown in FIG.

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

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

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

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

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

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

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

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

  In FIG. 5, the solid line shows the time when the cam phase of the intake valve driving camshaft 70 is advanced 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 reduced to a 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 a structural limit of the combustion chamber Ratio). When the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is held at the limit mechanical compression ratio in a region where the load is lower than the engine load L1 when the mechanical compression ratio reaches the limit mechanical compression ratio. Accordingly, the mechanical compression ratio is maximized and the expansion ratio is maximized at the time of low engine load operation and low engine load operation, that is, at the engine low load operation side. In other words, the mechanical compression ratio is maximized so that the maximum expansion ratio is obtained on the engine low load operation side.

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

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

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

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

  By the way, in an internal combustion engine provided with a variable compression ratio mechanism A, generally, as shown in FIG. 10 which is a schematic partial sectional view passing through the axis of each cylinder of the internal combustion engine, the lower part of the cylinder block 2 is a crankcase 1. The guide wall 1a is covered with a guide wall 1a integral with the crankcase 1 so that the inside does not communicate with the outside and the cylinder block 2 moves without tilting along the cylinder axis when changing the mechanical compression ratio. It is designed to slide in an airtight manner. Therefore, a sliding seal member (not shown) is arranged between the guide wall 1a and the cylinder block 2.

  In FIG. 10, 2a is a cylinder bore formed in the cylinder block 2, and 3 is a cylinder head. Reference numeral 3a denotes a cylinder head cover that covers the upper part of the cylinder head 3 in an airtight manner. Reference numeral 1b denotes a guide wall cover that covers the upper portion of the guide wall 1a. The guide wall cover 1b also slides between the cylinder block 2 and the cylinder head 3 in order to keep the space S in the guide wall 1b airtight. A seal member (not shown) is disposed.

  By the way, the air-fuel mixture in the combustion chamber 5 leaks into the crankcase 1 through the piston ring during the compression stroke, and the burned gas in the combustion chamber 5 enters the crankcase 1 through the piston ring during the expansion stroke. To leak. The gas leaking from the combustion chamber 5 into the crankcase 1 is blow-by gas, and a blow-by gas recirculation device (or blow-by gas reduction device) is provided to recirculate such blow-by gas to the intake system.

  The blow-by gas recirculation device communicates the crankcase 1 with a surge tank 12 on the downstream side of the throttle valve 17 in the intake system via a PCV (Positive Crankcase Ventilation) valve V1, and also in the crankcase 1 and the intake system. The upstream side of the throttle valve 17 communicates with the check valve V2. In this embodiment, the PCV valve V1 is disposed on the guide wall 1a, and the check valve V2 is disposed on the cylinder head cover 3a. The inside of the cylinder head cover 3a and the inside of the crankcase 1 are communicated via communication passages 3b and 2b formed in the cylinder head 3 and the cylinder block 2. The check valve V2 allows only gas flow from the intake system into the crankcase 1.

  The blow-by gas recirculation device recirculates the blow-by gas in the crankcase 1 to the intake system via the PCV valve V1 using the intake pipe negative pressure on the downstream side of the throttle valve, and the intake system via the check valve V2. From the air into the crankcase.

  In the PCV valve V1, the passage cross-sectional area decreases as the absolute value of the intake pipe negative pressure on the downstream side of the throttle valve increases, and the passage cross-sectional area increases as the absolute value of the intake pipe negative pressure decreases and approaches atmospheric pressure. . Thereby, the smaller the intake amount and the larger the absolute value of the intake pipe negative pressure on the downstream side of the throttle valve, the lower the amount of blow-by gas recirculation by the PCV valve V1, and the greater the absolute value of the intake pipe negative pressure, This suppresses an increase in the amount of blow-by gas that leaks from the combustion chamber 5 into the crankcase 1 due to negative pressure.

  In this way, the blow-by gas recirculation amount from the crankcase 1 to the intake system is determined by the absolute value of the intake pipe negative pressure by the PCV valve V1. When the internal combustion engine includes the variable compression ratio mechanism A that moves the cylinder block 2 relative to the crankcase 1 as in this embodiment, the cylinder block 2 is moved relative to the crankcase 1 in order to increase the mechanical compression ratio. When approaching, the volume in the crankcase 1 decreases. However, the intake pipe negative pressure does not change at this time, and the amount of blow-by gas recirculation passing through the PCV valve V1 hardly changes. Therefore, if nothing is done, the internal pressure of the crankcase 1 becomes abnormally high, and the cylinder block Blowby gas and lubricating oil in the crankcase 1 are likely to leak from the sliding seal portion between the crankcase 1 and the crankcase 1.

  In the internal combustion engine of the present embodiment, as shown in FIG. 10, a decompression passage PA that communicates the inside of the crankcase 1 and the surge tank 12 is provided, and the inside of the crankcase 1 is equal to or higher than the set pressure in the decompression passage PA. A relief valve V3 is provided that opens only when the valve is open.

  As a result, when the cylinder block 2 is brought close to the crankcase 1 in order to increase the mechanical compression ratio by the variable compression ratio mechanism A, as shown in FIG. 10 showing the case of the high mechanical compression ratio in FIG. When the volume of the case 1 decreases and the internal pressure of the crankcase 1 rises above the set pressure, the relief valve V3 opens and the blow-by gas in the crankcase 1 is released to the surge tank 12 through the pressure reducing passage PA. Since the inside of 1 becomes less than the set pressure, it is possible to suppress an abnormal increase in the internal pressure of the crankcase 1. In the present embodiment, the pressure reducing passage PA provided with the relief valve V3 is connected to the surge tank 12 while sharing a part of the blow-by gas recirculation device. However, this does not limit the present invention, and the air cleaner 15 may be connected to any position of the intake system as long as it is on the downstream side.

  The inside of the crankcase 1 and the intake system may always be communicated with each other by the pressure reducing passage PA having the relief valve V3. For example, as in the case of the low mechanical compression ratio in FIG. When the mechanical compression ratio is made lower than the set value by the variable compression ratio mechanism A, the internal pressure of the crankcase 1 does not increase beyond the set pressure. In this embodiment, the pressure reducing passage PA is closed by the closing mechanism. It is like that. Thereby, the possibility that the blow-by gas in the crankcase 1 leaks to the outside air via the decompression passage PA can be reduced.

  In the present embodiment, the pressure reducing passage relief valve V3 is disposed in the guide wall 1a, and the pressure reducing passage PA extends from the space S in the guide wall 1a to the sliding portion between the guide wall 1a and the cylinder block 2. And leads into the crankcase 1. As shown in FIG. 10, when the cylinder block 2 is at the highest mechanical compression ratio, a portion PA1 of the pressure reducing passage in the guide wall 1a and a portion PA2 of the pressure reducing passage in the cylinder block 2 slide. It is designed to communicate with each other.

  Thus, when the cylinder block 2 is lifted to reduce the mechanical compression ratio, a part of the decompression passage in the guide wall 1a is shown in FIG. 11 showing the case of the low mechanical compression ratio in FIG. PA1 and part of the decompression passage PA2 in the cylinder block 2 are not communicated with each other, and the decompression passage PA is closed. Thus, the sliding portion between the guide wall 1a and the cylinder head 2 functions as a closing mechanism.

  A boss 1c around the opening of the relief valve V3 protrudes from the guide wall 1a into the space S in the guide wall 1a. Further, a boss 2c protrudes from the cylinder block 2 into the space S in the guide wall 1a. When the cylinder block 2 is lifted to reduce the mechanical compression ratio, the low mechanical compression shown in FIG. As shown in FIG. 11 showing the ratio, the boss 1c of the guide wall 1a and the boss 2c of the cylinder block 2 come into contact with each other, and the opening of the relief valve V3 is closed from the space S in the guide wall 1a. The space S in the guide wall 1a is also part of the decompression passage PA, and thus the boss 1c of the guide wall 1a and the boss 2c of the cylinder block 2 function as a closing mechanism.

  Thus, in the present embodiment, when the mechanical compression ratio is set lower than the set value by the variable compression ratio mechanism A, the pressure reducing passage PA is reliably closed by the two closing mechanisms.

  Further, in the pressure reducing passage PA, a control valve is provided in an external passage closer to the intake system than the relief valve V3, and this control valve is closed when the variable compression ratio mechanism A makes the mechanical compression ratio lower than the set value. If controlled in this way, it can function as a closing mechanism for the pressure reducing passage PA.

  In the present embodiment, as described above, when the cylinder block 2 is at the highest mechanical compression ratio, a part PA1 of the pressure reducing passage in the guide wall 1a and a part PA2 of the pressure reducing passage in the cylinder block 2 Are communicated at the sliding portion, and blowby gas in the crankcase 1 is discharged to the surge tank 12 when the inside of the crankcase 1 is equal to or higher than the set pressure. At this time, since the engine oil is contained in the blow-by gas that passes through parts PA1 and PA2 of the decompression passage, the engine oil is supplied to the sliding portion between the guide wall 1a and the cylinder block 2, and the cylinder block 2 The relative movement with respect to the crankcase 1 can be improved. As a result, the cylinder block 2 is not tilted with respect to the crankcase 1 during relative movement, and the amount of blow-by gas from the combustion chamber 5 into the crankcase 1 is suppressed.

1 Crankcase 1a Guide wall 2 Cylinder block A Variable compression ratio mechanism V1 PCV valve V2 Check valve V3 Relief valve PA Pressure reducing passage

Claims (1)

An internal combustion engine having a variable compression ratio mechanism that moves a cylinder block relative to a crankcase, wherein a blowby gas recirculation device that communicates the crankcase and a downstream side of a throttle valve in an intake system is provided, and the blowby gas recirculation The apparatus includes a PCV valve that decreases the cross-sectional area of the passage as the absolute value of the intake pipe negative pressure on the downstream side of the throttle valve increases, and further includes a pressure reducing passage that communicates the crankcase with the intake system. A relief valve that opens only when the inside of the crankcase is equal to or higher than a set pressure is disposed in the passage, and a closing mechanism is provided in the decompression passage, and the closing mechanism includes the variable compression ratio. Bei variable compression ratio mechanism, characterized by closing the pressure decreasing passage when the mechanical compression ratio is lower than the set value by a mechanism Internal combustion engine that.

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