JP2011169226A - Spark-ignited internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To start an internal combustion engine even when a temperature of an engine coolant or the like is not properly detected, in the internal combustion engine including a variable compression ratio mechanism. <P>SOLUTION: A spark-ignited internal combustion engine includes the variable compression ratio mechanism which can vary a mechanical compression ratio, and controls the variable compression ratio mechanism so that the mechanical compression ratio becomes a target mechanical compression ratio which is set according to an engine operating condition. When the internal combustion engine is started, the target mechanical compression ratio is varied between a start of a cranking and fuel supply and a start of the internal combustion engine. <P>COPYRIGHT: (C)2011,JPO&amp;INPIT

Description

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

  A spark ignition type internal combustion engine having a variable compression ratio mechanism capable of changing a mechanical compression ratio is known. In such a spark ignition type internal combustion engine, it has been proposed to control the mechanical compression ratio in a manner different from that during normal operation of the internal combustion engine when the internal combustion engine is started.

  For example, in the internal combustion engine described in Patent Document 1, the higher the mechanical compression ratio at the start of the internal combustion engine, the greater the friction torque required to rotate the crankshaft by the starter motor. The compression ratio is lowered. In particular, in Patent Document 1, considering that the ignition performance of the air-fuel mixture is improved as the actual compression ratio is higher if the engine speed is equal to or higher than a predetermined speed during cranking, the mechanical compression ratio at the start of the internal combustion engine is considered. Is set.

  Further, in the internal combustion engine described in Patent Document 2, if the mechanical compression ratio is changed by the variable compression ratio mechanism when the air-fuel mixture is burned in the combustion chamber, the actuator for driving the variable compression ratio mechanism is large. Since the driving force is required, the mechanical compression ratio is changed when the air-fuel mixture is not combusted in the combustion chamber, that is, when cranking when the engine is started and fuel injection and ignition are not performed. Thereafter, the mechanical compression ratio is kept constant during operation of the internal combustion engine.

JP 2006-52682 A JP 2008-111375 A JP 2002-276446 A JP 2006-342677 A

  By the way, when the internal combustion engine is started, if the mechanical compression ratio is too high, the engine speed does not sufficiently increase during cranking by the starter motor, and as a result, the internal combustion engine cannot be started properly. Conversely, if the mechanical compression ratio is too low, the temperature and pressure of the air-fuel mixture in the combustion chamber when the piston is at the compression top dead center will decrease, and as a result, the internal combustion engine cannot be started properly. Therefore, in order to start the internal combustion engine appropriately, it is necessary to set the mechanical compression ratio to an optimum mechanical compression ratio between these.

  The optimum mechanical compression ratio for starting such an internal combustion engine varies depending on the engine coolant temperature, intake air temperature, atmospheric pressure, battery condition, and the like. For example, the optimum mechanical compression ratio decreases as the temperature of the engine cooling water or the temperature of the intake air increases. Therefore, if the mechanical compression ratio is always set to the same compression ratio when starting the internal combustion engine, the internal combustion engine cannot be started properly depending on the temperature of the engine cooling water or the like.

  On the other hand, it is also conceivable to detect the temperature of the engine cooling water at the time of starting the engine and set the mechanical compression ratio at the time of starting the engine to an optimum compression ratio based on the detected temperature of the engine cooling water. However, in this case, if a failure occurs in a sensor that detects the temperature of engine cooling water or the like, the mechanical compression ratio at the time of starting the engine cannot be set to an optimal compression ratio.

  In view of the above problems, an object of the present invention is to allow an internal combustion engine to be started even in a case where the temperature of engine cooling water or the like cannot be detected properly in an internal combustion engine having a variable compression ratio mechanism. There is to do.

In order to solve the above-mentioned problem, in the first invention, a variable compression ratio mechanism capable of changing the mechanical compression ratio is provided, and the mechanical compression ratio becomes a target mechanical compression ratio set according to the engine operating state. In the spark ignition internal combustion engine in which the variable compression ratio mechanism is controlled, when starting the internal combustion engine, the target mechanical compression ratio is changed between the start of cranking and fuel supply and the start of the internal combustion engine.
According to the first aspect, the target mechanical compression ratio is changed during cranking, and the actual mechanical compression ratio is also changed accordingly. By changing the actual mechanical compression ratio in this way, the mechanical compression ratio becomes an appropriate mechanical compression ratio for starting the internal combustion engine during the change, and at this time, the start of the internal combustion engine is started.

  In the second invention, in the first invention, the change of the target mechanical compression ratio from the start of the cranking to the start of the start of the internal combustion engine is continued as time elapses from the start of cranking and fuel supply. Done.

  In the third invention, in the first or second invention, the mechanical compression ratio is a predetermined mechanical compression ratio lower than the mechanical compression ratio at the time of engine low load operation before the start of cranking and fuel supply, Immediately after the start of cranking and fuel supply, it is first changed to the high compression ratio side.

  In the fourth invention, in any one of the first to third inventions, the change of the target mechanical compression ratio from the start of the cranking and fuel supply is performed with the passage of time from the start of cranking and fuel supply. This is done by repeatedly increasing and decreasing the target mechanical compression ratio.

  According to a fifth invention, in any one of the first to third inventions, the instantaneous engine speed is a reference with an increase in the mechanical compression ratio between the start of cranking and fuel supply and the start of start of the internal combustion engine. When the engine speed is equal to or lower than the engine speed, the target mechanical compression ratio is controlled to be equal to or lower than the mechanical compression ratio when the instantaneous engine speed is equal to or lower than the reference engine speed.

  According to a sixth invention, in any one of the first to fifth inventions, a variable valve timing mechanism capable of controlling the closing timing of the intake valve is further provided, and when starting the internal combustion engine, cranking and fuel The target valve closing timing of the intake valve is maintained substantially constant between the start of supply and the start of start of the internal combustion engine.

  According to a seventh aspect, in the sixth aspect, the apparatus further comprises elapsed time measuring means for measuring an elapsed time from the start of cranking and fuel supply, and the time measured by the elapsed time measuring means exceeds a predetermined time. However, if the internal combustion engine is not started, the closing timing of the intake valve is moved to the intake bottom dead center side and the target mechanical compression ratio is decreased.

  According to the present invention, in an internal combustion engine having a variable compression ratio mechanism, the internal combustion engine can be started even when the temperature of the engine cooling water or the like cannot be properly detected.

1 is an overall view of a spark ignition 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 time chart, such as a mechanical compression ratio, when starting an internal combustion engine. It is a flowchart which shows the control routine of the drive control of an actuator. It is a figure which shows the map of the integration time for target machine compression ratio calculation, and a target machine compression ratio. It is a time chart, such as a target mechanical compression ratio at the time of performing control in a second embodiment. It is a flowchart which shows the control routine of target mechanical compression ratio calculation control. It is a flowchart which shows the target valve closing timing etc. of an intake valve at the time of performing control in 3rd embodiment.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. In the following description, the same reference numerals are assigned to similar components.

FIG. 1 shows a side sectional view of a spark ignition type internal combustion engine.
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 an 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 air flow meter 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 catalytic converter 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 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 operation 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. In the present embodiment, the relative position sensor 22 is used as a mechanical compression ratio detection device for detecting the current mechanical compression ratio, but a detection device other than the relative position sensor 22 is used as the mechanical compression ratio detection device. It is also possible.

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 air flow meter 18, the air-fuel ratio sensor 21, the relative position sensor 22, the valve timing sensor 23, and the throttle opening degree sensor 24 are input to the input port 35 via the corresponding AD converters 37, respectively. 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, the input port 35 is connected to a crank angle sensor 42 that generates an output pulse every time the crankshaft rotates, for example, 30 °.
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 variable compression ratio mechanism A, and the variable valve timing mechanism B through corresponding drive circuits 38. Further, a starter motor 44 for rotationally driving the crankshaft is attached to a crankshaft (not shown) of the engine body, and the output port 36 is connected to the starter motor 44 via a corresponding drive circuit 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 projecting portions 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 projecting portion 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 with a space between 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, as shown in FIG. 3, eccentric shafts 57 arranged eccentrically with respect to the rotational axes of the cam shafts 54 and 55 extend on both sides of each circular cam 58, and another circular cam is disposed on the eccentric shaft 57. 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 shown by arrows in FIG. 3A from the state shown in FIG. The circular cam 56 rotates in the direction opposite to the circular cam 58 in the cam insertion hole 51 in order to move away from the circular cam 58, and as shown in FIG. Position. 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 located 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 meshing with the worms 61 and 62 are fixed to end portions 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 vanes 75 extending from the outer peripheral surface of the rotary shaft 73 to the inner peripheral surface of the cylindrical housing 72 between the partition walls 74, and advance hydraulic chambers 76 on both sides 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 that performs communication cutoff control 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. 6A, 6B, and 6C 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 action is not performed while the intake valve is open, and the actual compression starts from the time when the intake valve is closed. The action begins. Therefore, the actual compression ratio is expressed as 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 ultra-high 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, the combustion chamber volume is 50 ml and the piston stroke volume is 500 ml as in the examples 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 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, the actual compression ratio should be increased. 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. In this way, the amount of increase in the theoretical thermal efficiency when the expansion ratio is increased while maintaining the actual compression ratio ε at a low value, and the actual compression ratio as shown by the solid line in FIG. 7 are also increased with the expansion ratio. It can be seen that there is no significant difference from the increase in theoretical thermal efficiency.

  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 greatly increase. 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 engine high 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 an 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 catalytic converter 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. Therefore, 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 without depending on 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 reduced in proportion to the reduction 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 L ₁ slightly close to the low load, the mechanical compression ratio reaches the maximum critical mechanical compression ratio that is the structural limit of the combustion chamber 5. . When the mechanical compression ratio reaches the maximum critical mechanical compression ratio, the mechanical compression ratio is maintained at the maximum critical mechanical compression ratio in a region where the load is lower than the engine load L ₁ when the mechanical compression ratio reaches the maximum critical mechanical compression ratio. The Therefore, 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, on 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 L ₁ , the closing timing of the intake valve 7 becomes the limit closing timing at which the amount of intake air supplied into the combustion chamber 5 can be controlled. When the closing timing of the intake valve 7 reaches the limit closing timing, the closing timing of the intake valve 7 in a region where the load is lower than the engine load L ₁ when the closing timing of the intake valve 7 reaches the closing timing. Is held at the limit valve 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 L ₁ when the closing timing of the intake valve 7 reaches the limit closing timing, the throttle valve 17 supplies the fuel into the combustion chamber 5. 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 the direction away from the intake bottom dead center BDC until the limit valve closing timing L ₁ that can control the amount of intake air supplied into the combustion chamber. As described above, 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. Hereinafter, the present invention will be described by taking as an example a case where the closing timing of the intake valve 7 is changed as shown by a solid line in FIG.

  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 the present invention, the variable compression ratio mechanism A is formed so that the expansion ratio is 20 or more.

  By the way, when the engine is started, if the actual compression ratio is too high, the energy required for the compression action by the piston increases, and the engine speed cannot be sufficiently increased by the starter motor 44 during cranking. If the actual compression ratio is too high, knocking or pre-ignition may occur immediately after the engine is started. For this reason, if the actual compression ratio is made too high at the time of starting the engine, the internal combustion engine cannot be started properly.

  On the other hand, if the actual compression ratio is too low when the engine is started, the temperature (compression end temperature) and pressure (compression end pressure) of the air-fuel mixture in the combustion chamber 5 when the piston 4 is at the compression top dead center during cranking. And the air-fuel mixture cannot be combusted properly in the combustion chamber 5. For this reason, the internal combustion engine cannot be started properly even if the actual compression ratio is made too low when the engine is started. Therefore, in order to start the internal combustion engine appropriately, it is necessary to set the actual compression ratio to an optimum actual compression ratio between these.

  It should be noted that the amount of intake air during cranking at the time of engine start is always substantially constant even though there is some variation depending on the temperature of the engine cooling water or the like. That is, in order to keep the increase in the engine speed immediately after starting the engine within a certain range, it is necessary to set the fuel supply amount at the time of starting the engine within a certain range. Accordingly, the intake air amount must be within a certain range when the engine is started.

  As described above, since the intake air amount is always substantially constant when the engine is started, the closing timing of the intake valve 7 and the opening degree of the throttle valve 17 are also substantially constant when the engine is started. Therefore, in order to start the internal combustion engine appropriately, it can be said that the mechanical compression ratio needs to be set to an optimum mechanical compression ratio.

  The optimal mechanical compression ratio (actual compression ratio) for starting such an internal combustion engine varies depending on the engine cooling water temperature, intake air temperature, atmospheric pressure, battery condition, and the like. For example, the higher the engine cooling water temperature and the intake air temperature, the higher the compression end temperature and the more likely knocking occurs. Therefore, when the engine cooling water temperature and the intake air temperature increase, The optimum mechanical compression ratio for starting is low. For this reason, when the internal combustion engine is started, if the mechanical compression ratio is always set to the same compression ratio, the internal combustion engine cannot be started properly depending on the temperature of the engine cooling water or the like.

  On the other hand, immediately before starting the internal combustion engine, the temperature of the engine cooling water, the temperature of the intake air, etc. are detected using a water temperature sensor, an atmospheric temperature sensor, etc., and the detected temperature of the engine cooling water, the temperature of the intake air, etc. Based on this, it is conceivable to set the mechanical compression ratio at the time of starting the engine to an optimum mechanical compression ratio. However, in this case, if a failure occurs in the water temperature sensor, the atmospheric temperature sensor, or the like, the mechanical compression ratio at the time of starting the engine cannot be set to an optimum mechanical compression ratio. In addition, if the relative position sensor 22 fails and the current mechanical compression ratio cannot be accurately detected, the actual mechanical compression ratio is set to the target machine even if the optimum target mechanical compression ratio is set when the engine is started. The compression ratio cannot be controlled. As a result, if the water temperature sensor, the atmospheric temperature sensor, the relative position sensor 22 or the like fails, the internal combustion engine cannot be optimally started.

  Therefore, in the embodiment of the present invention, when starting the internal combustion engine, the target mechanical compression ratio is changed between the start of cranking and fuel supply and the start of the internal combustion engine.

  FIG. 10 is a time chart of the presence or absence of cranking, the presence or absence of fuel supply, the target mechanical compression ratio, the engine speed, the target closing timing of the intake valve 7 and the target opening of the throttle valve 17 when starting the internal combustion engine. It is. When the motoring in the figure is ON, it indicates that the crankshaft is driven by the starter motor 44, that is, cranking is being performed, and when the fuel supply in the figure is ON, the fuel from the fuel injection valve 13 is shown. This shows that fuel is being supplied.

In the example shown in FIG. 10, the ignition is turned on at time t ₀ , and simultaneously, cranking by the starter motor 44 and fuel supply from the fuel supply valve 13 are started. Note that the start of cranking and the start of fuel supply are not necessarily started simultaneously.

Before the time t ₀ , that is, before the start of cranking and fuel supply, the mechanical compression ratio is set to a predetermined target compression ratio at starting that is relatively low. Specifically, for example, the mechanical compression ratio is set to the starting target compression ratio immediately after the end of the previous operation of the internal combustion engine. The target compression ratio at start-up is, for example, a value lower than the mechanical compression ratio set when the engine load is in the low load operation region or the medium load operation region during engine operation, or the structural limit of the variable compression ratio mechanism A It is set as the value near the minimum limit mechanical compression ratio which becomes. In the example shown in FIG. 10, the starting target compression ratio is the minimum critical mechanical compression ratio εmmin.

Similarly, before the time ₀ is reached, the closing timing of the intake valve 7 and the opening of the throttle valve 17 are also set to the predetermined target closing timing at starting and the target opening at starting, respectively. With respect to the closing timing of the intake valve 7 and the opening degree of the throttle valve 17 as well as the mechanical compression ratio, the starting target closing timing and the starting target opening degree are set immediately after the end of the previous operation of the internal combustion engine, respectively. . The target closing timing at the start of the intake valve 7 and the target opening at the start of the throttle valve 17 are, for example, a closing timing and an opening at which the engine speed is appropriately increased immediately after the engine is started.

Thereafter, when cranking is started and fuel supply is started at time t ₀ , the engine speed is increased. Thereafter, when the engine speed reaches a constant speed by driving the starter motor 44 (time t ₁ ), the change of the target mechanical compression ratio is started. In the embodiment shown in FIG. 10, the target mechanical compression ratio is gradually increased from time t ₁ .

Thereafter, the target mechanical compression ratio reaches the maximum limit mechanical compression ratio εmmax at time t _2, the target mechanical compression ratio after being kept constant the maximum limit mechanical compression ratio εmmax period, caused to gradually decrease from time t _3. When the target mechanical compression ratio is decreased to the minimum limit mechanical compression ratio εmmin, the target mechanical compression ratio is maintained at the minimum limit mechanical compression ratio εmmin for a certain period, and then gradually increased again.

In the example shown in FIG. 10, as a result of increasing / decreasing the target mechanical compression ratio in this way, at time t ₄ , the engine speed is changed to the engine start determination speed (for example, 400 rpm) by combustion of the air-fuel mixture in the combustion chamber 5. ) And the internal combustion engine is started. With the start of the internal combustion engine, the drive of the crankshaft by the starter motor 44 is stopped.

  In the example shown in FIG. 10, the valve closing timing of the intake valve 7 and the opening of the throttle valve 17 are substantially constant from the start of cranking and fuel supply to the start of the internal combustion engine. Maintained. Thereby, when the start of the internal combustion engine is started, the engine speed can be appropriately increased. Note that how the engine speed increases at the start of the internal combustion engine changes depending on the temperature of the engine cooling water, the atmospheric temperature, etc. at the start of the engine, so the internal combustion engine starts from the start of cranking and fuel supply. The closing timing of the intake valve 7 and the opening degree of the throttle valve 17 may be changed based on the engine cooling water temperature, the atmospheric temperature, and the like.

  The above example shows a case where the internal combustion engine is started when the target mechanical compression ratio is increased and decreased once and then increased again. If the start of the internal combustion engine is not started at this time, the target mechanical compression ratio is repeatedly increased and decreased thereafter as shown by the broken line in FIG.

  Thus, in the embodiment of the present invention, when starting the internal combustion engine, the target mechanical compression ratio is continuously changed from the start of cranking and fuel supply until the start of the internal combustion engine is started, In particular, in the above embodiment, it is repeatedly increased and decreased.

  Here, when the relative position sensor 22 has not failed or the like, the actual mechanical compression ratio is also changed following the change of the target mechanical compression ratio. Therefore, the actual mechanical compression ratio is also the target shown in FIG. It can be changed in the same way as the compression ratio. When the actual mechanical compression ratio is changed in this way, there is always a time when the mechanical compression ratio becomes an appropriate mechanical compression ratio for starting the internal combustion engine during the change. At this time, the internal combustion engine starts to start. Will be. Therefore, according to the embodiment of the present invention, it is possible to start the internal combustion engine without using the output values of the water temperature sensor, the atmospheric temperature sensor, etc., and therefore, even if the water temperature sensor or the atmospheric temperature sensor fails, it is ensured. The internal combustion engine can be started.

  On the other hand, when the output value of the relative position sensor 22 is shifted or malfunctioned, the actual mechanical compression ratio does not accurately follow the target mechanical compression ratio. However, for example, even if there is a deviation in the output value of the relative value sensor 22, the actual mechanical compression ratio also varies as the target mechanical compression ratio varies. Even if the relative value sensor 22 is out of order and the output value of the relative position sensor 22 is a constant value, if the target mechanical compression ratio increases and is higher than this constant value, the actual value The mechanical compression ratio is increased, while the actual mechanical compression ratio is decreased when the target mechanical compression ratio decreases and is lower than this constant value. For this reason, even if the relative value sensor 22 is out of order, the actual mechanical compression ratio can be varied. As described above, as long as the actual mechanical compression ratio varies as described above, there is a time when the mechanical compression ratio becomes the optimum mechanical compression ratio for starting the internal combustion engine. For this reason, the internal combustion engine can be reliably started even if the output value of the relative position sensor 22 is deviated or malfunctions.

  Further, when the internal combustion engine is restarted at a high temperature, if the mechanical compression ratio, that is, the actual compression ratio is high, the compression end temperature rises, which may cause knocking or pre-ignition. In the above embodiment, when starting the crankshaft by the starter motor 44 and starting the fuel supply from the fuel injection valve 13, the mechanical compression ratio is set to a relatively low compression ratio, so that the internal combustion engine is restarted at a high temperature. Even so, the occurrence of knocking and pre-ignition can be suppressed.

  In the above embodiment, when the mechanical compression ratio reaches the maximum limit mechanical compression ratio and the minimum mechanical compression ratio while the mechanical compression ratio is varied up and down, the maximum limit mechanical compression ratio and the minimum limit machine are respectively set. After maintaining the compression ratio for a certain period, the mechanical compression ratio is decreased and increased. However, it is not always necessary to maintain for a certain period of time, and the mechanical compression ratio may be decreased and increased as soon as the mechanical compression ratio reaches the maximum critical mechanical compression ratio and the minimum mechanical compression ratio.

  Further, in the above-described embodiment, the mechanical compression ratio at the start of cranking and fuel supply is always set to a predetermined starting target compression ratio. However, the target mechanical compression ratio at the start of cranking and fuel supply may be changed according to the engine cooling water temperature, atmospheric temperature, atmospheric pressure, battery condition, etc. immediately before the start of cranking and fuel supply. Good. In this case, cranking and fuel supply are started after a lapse of a short time after the ignition is turned on, and the target mechanical compression ratio is changed after the ignition is turned on until cranking and fuel supply are started. Done.

  Thus, by changing the target mechanical compression ratio at the start of cranking and fuel supply based on the engine cooling water temperature or the like, if there is no failure in the sensor for detecting the engine cooling water temperature or the like, cranking and At the start of fuel supply, the mechanical compression ratio can be set to an optimum mechanical compression ratio for starting the internal combustion engine, and the internal combustion engine can be started appropriately. On the other hand, even when a sensor that detects the temperature or the like of the engine coolant has failed, the internal combustion engine can be appropriately started by changing the mechanical compression ratio as described above.

  FIG. 11 is a flowchart showing a drive control routine of actuators such as the drive motor 59 of the variable compression ratio mechanism A, the hydraulic oil supply control valve 78 of the variable valve timing mechanism B, and the actuator 16 for driving the throttle valve 17.

  As shown in FIG. 11, first, in step S11, it is determined whether or not driving of the crankshaft by the starter motor 44 is started. If the drive by the starter motor 44 has not been started yet, the control routine is ended.

  On the other hand, when the drive by the starter motor 44 is started, the process proceeds from step S11 to step S12. In step S12, it is determined whether or not the internal combustion engine has been started. The determination of the start of the internal combustion engine is made, for example, based on whether or not the engine speed has reached a predetermined speed (for example, 400 rpm). If it is determined that the engine speed has not reached the predetermined speed and the internal combustion engine has not yet been started, the routine proceeds from step S12 to step S13. In step S13, the remainder obtained by dividing the accumulated time t after the start of the drive by the starter motor 44 by the fluctuation period T of the mechanical compression ratio is set as the accumulated time t 'for calculating the target mechanical compression ratio. The fluctuation cycle T of the mechanical compression ratio means a cycle in which the mechanical compression ratio is increased and decreased once, that is, T in FIG.

  Next, in step S14, the target mechanical compression ratio is calculated using the map as shown in FIG. 12 based on the target mechanical compression ratio calculation integration time t 'calculated in step S13. Next, in step S15, the target valve closing timing of the intake valve 7 and the target opening of the throttle valve 17 are calculated. In the embodiment described above, the target valve closing timing of the intake valve 7 and the target opening of the throttle valve 17 are set to a predetermined starting target valve closing timing and starting target opening, respectively. Next, in step S16, the drive motor 59 of the variable compression ratio mechanism A, the target mechanical compression ratio calculated in steps S14 and S15, the target valve closing timing of the intake valve 7 and the target opening of the throttle valve 17 are obtained. The hydraulic oil supply control valve 78 of the variable valve timing mechanism B and the actuator 16 for driving the throttle valve 17 are controlled.

  Thereafter, when the internal combustion engine is started and the engine rotational speed becomes equal to or higher than the predetermined rotational speed, in the next routine, it is determined in step S12 that the internal combustion engine has been started, and the process proceeds to step S17. In step S17, normal control as shown in FIG. 9 is performed.

  Next, a second embodiment of the present invention will be described with reference to FIG. The configuration of the spark ignition internal combustion engine of the second embodiment is basically the same as the configuration of the spark ignition internal combustion engine of the first embodiment. However, the spark ignition internal combustion engine of the second embodiment is different from the spark ignition internal combustion engine of the first embodiment in that the mechanical compression ratio is not necessarily changed to the maximum limit mechanical compression ratio when starting the internal combustion engine. ing.

  As described above, if the mechanical compression ratio is excessively increased during cranking, the energy required for the compression action of the piston 4 increases, and the engine speed can be sufficiently increased by the starter motor 44 during cranking. become unable. On the other hand, in order to start the internal combustion engine, it is necessary to set the engine speed to a certain fixed speed or higher by cranking. For this reason, if the mechanical compression ratio is excessively increased during cranking and the engine speed is reduced, it becomes difficult to start the internal combustion engine properly.

  Therefore, in the second embodiment of the present invention, when the mechanical compression ratio is changed from the start of cranking to the start of the internal combustion engine, the engine speed is lower than a predetermined limit engine speed (reference speed) NEmin. In such a case, the mechanical compression ratio is maintained below the mechanical compression ratio at that time.

FIG. 13 is a time chart of the presence / absence of cranking, the presence / absence of fuel supply, the target mechanical compression ratio, and the engine speed when the internal combustion engine is started in the second embodiment. In the example shown in FIG. 13, the ignition is turned on at time t ₅ , and simultaneously, cranking by the starter motor 44 and fuel supply from the fuel injection valve 13 are started. Prior to time t ₅ , the mechanical compression ratio is set to the target mechanical compression ratio at the time of start, as in the first embodiment.

At time t _5, when the fuel supply from the cranking and the fuel supply valve 13 by the starter motor 44 is started, the engine speed is raised. Thereafter, when a predetermined time Ta has elapsed from the start of cranking or the like (time t ₆ ), the change of the target mechanical compression ratio is started. This fixed time Ta is, for example, a time normally required for the engine speed to reach a fixed speed (for example, a limit engine speed NEmin described later) from the start of cranking. The change of the target mechanical compression ratio may be started when the engine speed reaches a certain speed by driving the starter motor 44.

When the target mechanical compression ratio is gradually increased from the time t _6, the actual mechanical compression ratio is gradually increased, the engine speed is made to decrease accordingly. When such reduction in the engine speed continues, eventually the engine speed at time t ₇ reaches the limit engine speed NEmin. The limit engine speed is, for example, a speed that makes it difficult to start the internal combustion engine even if the mechanical compression ratio or the like is appropriate if the engine speed further decreases. In the embodiment of the present invention, when the engine speed reaches the limit engine speed NEmin in this way, the mechanical compression ratio thereafter becomes the mechanical compression ratio εmx or less when the engine speed reaches the limit engine speed NEmin. Controlled within range. In the illustrated example, the after time t ₇ the engine speed reaches the limit engine speed NEmin, gradually so that to lower the mechanical compression ratio.

  Thereafter, as long as the engine speed does not reach the limit engine speed NEmin again, the mechanical compression ratio is repeatedly increased and decreased between the minimum limit mechanical compression ratio and the εmx. On the other hand, when the engine speed reaches the limit engine speed NEmin again, the mechanical compression ratio is the minimum limit mechanical compression ratio and the mechanical compression ratio when the engine speed reaches the limit machine speed NEmin again. It is repeatedly increased and decreased between.

  As described above, in the second embodiment of the present invention, when changing the mechanical compression ratio from the start of cranking to the start of the internal combustion engine, the engine speed becomes lower than the predetermined limit engine speed NEmin. By maintaining the mechanical compression ratio below the mechanical compression ratio at that time, it is possible to prevent the engine rotational speed from excessively decreasing due to excessively high mechanical compression ratio during cranking.

In the example shown in FIG. 13, the mechanical compression ratio is gradually decreased after the engine speed reaches the limit engine speed NEmin at time t ₇ , but as shown by the broken line in FIG. After the engine speed reaches the limit engine speed NEmin, the mechanical compression ratio may be maintained at the mechanical compression ratio at that time. Here, generally, if the engine speed at the time of cranking can be maintained at a predetermined speed or higher, the higher the mechanical compression ratio, the easier it is to start the internal combustion engine. Therefore, the internal combustion engine can be started quickly by maintaining the mechanical compression ratio at the mechanical compression ratio after the engine speed reaches the limit engine speed NEmin.

  FIG. 14 is a flowchart showing a control routine for target mechanical compression ratio calculation control in the second embodiment of the present invention. FIG. 14 shows a control routine in the case of performing the control indicated by the broken line in FIG. Also, steps S21 and S22 in FIG. 14 are the same as steps S11 and S12 in FIG.

  In step S23, it is determined whether or not the elapsed time t from the start of cranking or the like has exceeded a predetermined time Ta. If it is determined that the predetermined time Ta has not elapsed, the process proceeds to step S24. . In step S24, the target mechanical compression ratio is set to the starting mechanical compression ratio εmini.

  Thereafter, when the elapsed time t from the start of cranking or the like has passed the predetermined time Ta, the process proceeds from step S23 to step S25 in the next routine. In step S25, it is determined whether or not the current engine speed NE is greater than or equal to the limit engine speed NEmin. If it is determined that the current engine speed NE is greater than or equal to the limit engine speed NEmin, step S26 is performed. Proceed to In step S26, the target mechanical compression ratio εm (n) obtained by adding a predetermined value Δεm to the target mechanical compression ratio εm (n-1) in the previous routine is set as the current target mechanical compression ratio εm (n) (ε (n) = εm (N−1) + Δεm), the target mechanical compression ratio is increased.

  When the engine speed decreases and reaches the limit engine speed NEmin as the target mechanical compression ratio increases, the process proceeds from step S25 to step S27 in the next routine. In step S27, the current target mechanical compression ratio εm (n) is set to the previous target mechanical compression ratio εm (n-1), and the target mechanical compression ratio is kept constant.

  Thereafter, when the internal combustion engine is started and the engine speed becomes equal to or higher than the predetermined speed, in the next routine, it is determined in step S22 that the internal combustion engine has been started, and the process proceeds to step S28. In step S28, the target mechanical compression ratio is set so that normal control as shown in FIG. 9 is performed.

  In the above embodiment, parameters such as the predetermined time Ta, the predetermined value Δεm, the limit engine speed NEmin, etc. are always set to constant values. However, these parameters may be changed according to the temperature of the engine cooling water, the atmospheric temperature, the atmospheric pressure, the battery state, etc. immediately before the crankshaft drive and fuel supply are started.

  Next, a third embodiment of the present invention will be described with reference to FIG. The configuration of the spark ignition internal combustion engine of the third embodiment is basically the same as the configuration of the spark ignition internal combustion engine of the first embodiment or the second embodiment.

  By the way, as one of the factors that prevent the internal combustion engine from starting properly, liquid fuel adheres to the plug gap of the spark plug 6 and cannot be properly ignited by the spark plug 6 (that is, ignition). A fuel cover of the plug 6). Thus, when the fuel cover is generated in the spark plug 6, the fuel cover of the spark plug 6 can be eliminated even if the mechanical compression ratio is adjusted as in the first embodiment or the second embodiment. Therefore, the internal combustion engine cannot be started.

  Therefore, in the third embodiment of the present invention, when the internal combustion engine does not start for a predetermined time or more even when the control as shown in the first embodiment or the second embodiment is performed, the closing timing of the intake valve 7 is determined. Is advanced to approach the intake bottom dead center, and the mechanical compression ratio is lowered to approach the minimum critical mechanical compression ratio.

FIG. 15 is a time chart of the presence or absence of cranking, the presence or absence of fuel supply, the target closing timing of the intake valve 7 and the target mechanical compression ratio when the control in the third embodiment is performed. In the example shown in FIG. 15, the ignition is turned on at time t ₈ , and simultaneously, cranking by the starter motor 44 and fuel supply from the fuel injection valve 13 are started.

As can be seen from FIG. 15, in this embodiment as well, in the same way as in the above embodiment, after cranking or the like is started, the target mechanical compression ratio is varied, and the closing timing of the intake valve 7 is the target at the start. The valve closing time is maintained. Thus when start of the engine even at varying target mechanical compression ratio is not started, the allowed fuel supply stop from the fuel injection valve 13 at time t ₉ the predetermined time Tb has elapsed from the start of such cranking In addition, the target valve closing timing of the intake valve 7 is advanced to the vicinity of the intake bottom dead center. Also, when the target mechanical compression ratio becomes higher compression ratio than the minimum limit mechanical compression ratio at time t _9, the target mechanical compression ratio is made to drop to near the minimum critical mechanical compression ratio.

  As described above, when the target valve closing timing of the intake valve 7 is advanced to the vicinity of the intake bottom dead center, a large amount of air is supplied into the combustion chamber 5, and as a result, passes through the intake port 8 and the combustion chamber 5. The air flow rate increases. Therefore, the liquid fuel adhering to the wall surface in the combustion chamber 5 and the wall surface of the intake port 8 is easily discharged from the combustion chamber 5 to the exhaust port 10. Along with this, the fuel adhering to the plug gap of the spark plug 6 is also vaporized and discharged to the exhaust port 10, so that the fuel cover of the spark plug 6 can be eliminated.

  Further, when the closing timing of the intake valve 7 is advanced, a large amount of air flows into the combustion chamber 5, so that the energy required for the compression action by the piston increases, resulting in a decrease in engine speed. Will be invited. When the engine speed is thus reduced, the flow rate of air passing through the intake port 8 and the combustion chamber 5 per unit time is reduced. As a result, the wall surface in the combustion chamber 5 and the plug gap of the spark plug 6 are reduced. The vaporization of the fuel adhering to the fuel cannot be promoted. According to this embodiment, when the closing timing of the intake valve 7 is advanced, the mechanical compression ratio is lowered to the vicinity of the minimum limit mechanical compression ratio, so that an increase in energy necessary for the compression action by the piston is suppressed. As a result, a decrease in engine speed can be suppressed.

Then, at time t ₁₀ a certain time Tc has passed a predetermined from the time t _9, the fuel supply from the fuel injection valve 13 is resumed, the target closing timing, the original starting target closing of the intake valve 7 The valve timing is returned to and the fluctuation of the mechanical compression ratio is started.

  In addition, the control for eliminating such fuel covering may be performed not only once but a plurality of times. In this case, for example, the control for eliminating the fuel covering is performed again after the predetermined time Tb has elapsed since the control for eliminating the previous fuel covering was completed.

DESCRIPTION OF SYMBOLS 1 Crankcase 2 Cylinder block 3 Cylinder head 4 Piston 5 Combustion chamber 7 Intake valve 17 Throttle valve 44 Starter motor A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (7)

In a spark ignition internal combustion engine having a variable compression ratio mechanism capable of changing a mechanical compression ratio, wherein the variable compression ratio mechanism is controlled so that the mechanical compression ratio becomes a target mechanical compression ratio set according to an engine operating state ,
A spark ignition type internal combustion engine in which the target mechanical compression ratio is changed between the start of cranking and fuel supply and the start of start of the internal combustion engine when starting the internal combustion engine.   2. The spark according to claim 1, wherein the change of the target mechanical compression ratio from the start of cranking to the start of start of the internal combustion engine is continuously performed as time elapses from the start of cranking and fuel supply. Ignition internal combustion engine.   The mechanical compression ratio is a predetermined mechanical compression ratio lower than the mechanical compression ratio at the time of engine low load operation before the start of cranking and fuel supply, and the high compression ratio is first set immediately after the start of cranking and fuel supply. The spark ignition type internal combustion engine according to claim 1, wherein the spark ignition type internal combustion engine is changed to a side.   The change of the target mechanical compression ratio from the start of cranking and fuel supply is performed by repeatedly increasing and decreasing the target mechanical compression ratio with the passage of time from the start of cranking and fuel supply. The spark ignition internal combustion engine of any one of 1-3.   If the instantaneous engine speed falls below the reference engine speed as the mechanical compression ratio increases from the start of cranking and fuel supply to the start of the internal combustion engine, the target mechanical compression ratio is set to the instantaneous engine speed. The spark ignition internal combustion engine according to any one of claims 1 to 3, which is controlled to be equal to or less than a mechanical compression ratio when the engine speed becomes equal to or less than a reference rotational speed.   A variable valve timing mechanism capable of controlling the closing timing of the intake valve is further provided, and when the internal combustion engine is started, the target valve closing of the intake valve is performed between the start of cranking and fuel supply and the start of the internal combustion engine. The spark ignition internal combustion engine according to any one of claims 1 to 5, wherein the timing is maintained substantially constant. It further comprises an elapsed time measuring means for measuring an elapsed time from the start of cranking and fuel supply,
If the internal combustion engine is not started even if the time measured by the elapsed time measuring means exceeds a predetermined time, the closing timing of the intake valve is moved to the intake bottom dead center side and the target mechanical compression ratio is decreased. The spark ignition internal combustion engine according to claim 6.

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