JP5585528B2 - In-cylinder injection spark ignition internal combustion engine having a variable compression ratio mechanism - Google Patents

Description

  The present invention is a direct injection spark ignition internal combustion engine including a fuel injection valve that directly injects fuel into a cylinder (inside the cylinder) of the internal combustion engine and an ignition plug that ignites an air-fuel mixture in the cylinder. The present invention relates to a direct injection spark ignition internal combustion engine having a variable compression ratio mechanism.

  In order to improve the fuel consumption performance and output performance of the internal combustion engine, an internal combustion engine having a variable compression ratio mechanism that can change the mechanical compression ratio of the internal combustion engine has been proposed. For example, a cam mechanism is used. A variable compression ratio for changing the mechanical compression ratio of the internal combustion engine by changing the volume of the combustion chamber by moving the cylinder block that is an engine element constituting the combustion chamber of the internal combustion engine relative to the crankcase An internal combustion engine having a mechanism is known.

  By the way, in an internal combustion engine, a phenomenon in which residual heat of an ignition plug or in-cylinder deposit becomes a hot spot and a mixed gas spontaneously ignites during the compression stroke is called preignition. Or rotation failure. In particular, in an in-cylinder injection spark ignition internal combustion engine, pre-ignition may be caused if fuel leaks from the fuel injection valve while the engine is stopped and accumulates in the cylinder and restarts under a high engine temperature. It is possible. Such pre-ignition may lead to the occurrence of a knocking sound such as “cuckling”, which can greatly impair the merchantability.

  As a means for suppressing such pre-ignition, self-ignition due to pre-ignition can occur in a direct injection spark ignition internal combustion engine due to fuel leakage from the fuel injection valve when the engine is stopped into the combustion chamber. When cranking a specific cylinder, the fuel is supplied from the fuel injection valve of the specific cylinder through fuel injection, thereby increasing the fuel concentration in the combustion chamber of the specific cylinder and performing self-ignition during cranking. Proposals have been made to suppress the occurrence of self-ignition by making the condition not satisfied (see Patent Document 1). In addition, as another means for suppressing pre-ignition, knocking even when starting from idling while the vehicle is stopped by reducing the mechanical compression ratio when the driver's request for starting the vehicle is predicted. And suppression of occurrence of pre-ignition have been proposed (see Patent Document 2).

JP-A-2005-42677 JP 2010-185416 A

  However, in Patent Document 1, no proposal has been made for a measure for suppressing the occurrence of pre-ignition that effectively uses the variable compression ratio mechanism. Further, in the compression ratio control method disclosed in Patent Document 2, it may be necessary to greatly reduce the mechanical compression ratio in order to suppress the occurrence of pre-ignition. This may cause a situation such as a loss of combustion stability and a tendency to misfire. Furthermore, if the occurrence of pre-ignition is not suppressed even though the mechanical compression ratio is greatly reduced, a large amount of fuel is injected into the combustion chamber before the compression stroke is started so that the fuel chamber is over-rich. From the viewpoint of fuel economy, such as suppressing self-ignition, it may be necessary to take an unreasonable need.

  In view of the above-described problems, the present invention is an in-cylinder injection spark ignition internal combustion engine having a variable compression ratio mechanism, which effectively uses the variable compression ratio mechanism, and at the time of engine start, combustion stability, fuel consumption, etc. It is an object of the present invention to provide an in-cylinder injection spark ignition internal combustion engine capable of suppressing the occurrence of pre-ignition without impairing the engine.

According to the invention described in claim 1, capable of changing a fuel injection valve that performs direct fuel injection into a cylinder of an internal combustion engine, a spark plug to ignite an air-fuel mixture in the cylinder, the mechanical compression ratio A cylinder injection spark ignition internal combustion engine having a variable compression ratio mechanism , wherein the cylinder at the time of starting the engine is controlled in order to suppress preignition in which the air-fuel mixture in the cylinder ignites before the ignition timing of the spark plug. in direct injection spark ignition internal combustion engine that performs fuel injection control of the internal over-rich, comprises a preignition determining means for determining whether the establishment of the occurrence assumptions at the time of engine starting the flop les ignition, the When an engine stop request is made, the pre-ignition occurrence determination means specifies the respective threshold values of the fuel pressure, engine temperature, and atmospheric temperature with respect to the mechanical compression ratio when the engine stop request is made. If all of the fuel pressure, engine temperature, and atmospheric temperature at the time of the request exceed the respective threshold values, the pre-ignition occurrence precondition is established as the occurrence of pre-ignition at the next engine start. In the determination, when it is determined that the pre-ignition occurrence precondition at the next engine start is satisfied, when the engine stop request is made, the mechanical compression ratio is previously set by the variable compression ratio mechanism. An in-cylinder injection spark ignition internal combustion engine is provided in which the compression ratio is increased .

  Normally, when the engine is stopped, there is some engine rotation after the end of fuel injection / ignition control, and the cylinder is not completely sealed, so the burned gas concentration in the cylinder will increase with time while the engine is stopped. Therefore, it is considered that the amount of fresh air in the cylinder at the start of the engine is determined by the stroke volume and the combustion chamber volume. That is, the amount of fresh air in the cylinder at the start of the engine changes according to the mechanical compression ratio. For example, the higher the mechanical compression ratio, the smaller the combustion chamber volume, so the amount of fresh air in the cylinder at the start of the engine decreases. . Therefore, the higher the mechanical compression ratio at the time of engine start, the more the inside of the cylinder at the time of engine start can be over-rich with a smaller fuel injection amount, the self-ignition condition is not satisfied, and the occurrence of pre-ignition at the time of engine start Can be avoided more reliably.

  Based on this, in the invention described in claim 1, when it is determined that the pre-ignition generation precondition at the time of starting the engine is satisfied, the mechanical compression ratio is increased by the variable compression ratio mechanism. Configured to be According to the present invention having such a configuration, the amount of fresh air in the cylinder at the time of starting the engine can be reduced, and the injected fuel necessary for suppressing the occurrence of pre-ignition by making the inside of the cylinder excessively rich. The amount can be reduced, and fuel consumption can be improved and the time required for fuel injection can be shortened. Also, if the amount of fresh air in the cylinder at the start of the engine is small, the possibility of reverse engine rotation due to explosive force is small even if pre-ignition occurs, and the generated abnormal noise is also small. Become. Thus, according to the present invention, an in-cylinder injection spark ignition internal combustion engine having a variable compression ratio mechanism that effectively uses the variable compression ratio mechanism, such as combustion stability and fuel consumption at the time of engine start-up. It is possible to provide an in-cylinder injection spark ignition internal combustion engine that can suppress the occurrence of pre-ignition without impairing the engine.

According to the second aspect of the present invention, when it is determined by the pre-ignition occurrence determination means that the pre-ignition occurrence precondition at the next engine start is satisfied, when an engine stop request is made. 2. The direct injection spark ignition internal combustion engine according to claim 1 , wherein a mechanical compression ratio is increased in advance by the variable compression ratio mechanism to a maximum mechanical compression ratio that can be varied by the variable compression ratio mechanism. Is provided.

According to the third aspect of the present invention, the pre-ignition occurrence determination means determines that the pre-ignition occurrence precondition at the next engine start is satisfied, and The in-cylinder injection spark ignition internal combustion engine according to claim 1 or 2 , wherein when the engine temperature is lower than a predetermined value, the engine is started after the mechanical compression ratio is lowered by the variable compression ratio mechanism. Is provided.

  According to the invention described in each claim, in a direct injection spark ignition internal combustion engine having a variable compression ratio mechanism, the variable compression ratio mechanism is effectively used, and combustion stability, fuel consumption, etc. are impaired at the time of engine start. It is possible to suppress the occurrence of pre-ignition without causing a common effect.

1 is an overall view of a direct injection spark ignition internal combustion engine that includes a variable compression ratio mechanism. 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 flowchart which shows one Embodiment of the control routine of the mechanical compression ratio at the time of an engine stop. It is a flowchart which shows one Embodiment of the control routine at the time of engine starting. It is a flowchart which shows one Embodiment of the control routine of the starting pattern 3 in FIG.

  FIG. 1 shows a side sectional view of a direct injection spark ignition internal combustion engine equipped with 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 the surge tank 12 via the intake branch pipe 11. A fuel injection valve (in-cylinder fuel injection device) 13 that directly injects fuel into the cylinder (inside the cylinder) is disposed around the inner wall surface of the cylinder head 3.

  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 (cylinder fuel injection device) 13, the throttle valve drive actuator 16, the variable compression ratio mechanism A, and the variable valve timing mechanism B through a corresponding drive circuit 38. Is done.

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

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

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

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

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

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

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

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

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

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

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

  In FIG. 5, the solid line shows the time when the cam phase of the intake valve driving camshaft 70 is advanced 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 this configuration 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.

  By the way, as described above, in an internal combustion engine, a phenomenon in which residual heat of an ignition plug or in-cylinder deposit becomes a hot spot and a mixed gas spontaneously ignites during a compression stroke is called pre-ignition. There may be a case where the output of the internal combustion engine is drastically reduced or the rotation is unstable. In particular, in an in-cylinder injection spark ignition internal combustion engine, pre-ignition is caused when the fuel leaks from the fuel injection valve while the engine is stopped and accumulates as a deposit in the cylinder and restarts under a high engine temperature condition. There is a case. Such pre-ignition may lead to the occurrence of a knocking sound such as “cuckling”, which can greatly impair the merchantability.

  The present invention makes effective use of the variable compression ratio mechanism as described above, and makes it possible to suppress the occurrence of pre-ignition at the time of starting the engine without impairing combustion stability or fuel consumption. Next, an embodiment of control over the mechanical compression ratio in a direct injection spark ignition internal combustion engine equipped with the variable compression ratio mechanism of the present invention will be described.

  Normally, when the engine is stopped, there is some engine rotation after the end of fuel injection / ignition control, and the cylinder is not completely sealed, so the burned gas concentration in the cylinder will increase with time while the engine is stopped. Therefore, it is considered that the amount of fresh air in the cylinder at the start of the engine is determined by the stroke volume and the combustion chamber volume. That is, the amount of fresh air in the cylinder at the start of the engine changes according to the mechanical compression ratio. For example, the higher the mechanical compression ratio, the smaller the combustion chamber volume, so the amount of fresh air in the cylinder at the start of the engine decreases. . Therefore, the higher the mechanical compression ratio at the time of engine start, the more the inside of the cylinder at the time of engine start can be over-rich with a smaller fuel injection amount, the self-ignition condition is not satisfied, and the occurrence of pre-ignition at the time of engine start Can be suppressed more reliably.

  Based on this, in the present invention, pre-ignition occurrence determination means for determining whether or not a precondition for occurrence at the time of engine start of the pre-ignition in which the air-fuel mixture in the cylinder ignites before the ignition timing of the spark plug is satisfied. And when the pre-ignition occurrence determining means determines that the pre-ignition generation preconditions at the time of engine start are satisfied, the variable compression ratio mechanism is configured to increase the mechanical compression ratio. The

  That is, the present invention realizes the suppression of the occurrence of pre-ignition by increasing the mechanical compression ratio at the time of engine start where the occurrence of pre-ignition is expected. The time to increase the mechanical compression ratio may be at least before the compression stroke at the start of the engine is started. Incidentally, in the embodiment of the present invention shown in FIG. 9 to be described later, in order to avoid the deterioration of the startability in consideration of the deterioration of the startability due to the time required to increase the mechanical compression ratio, It is assumed that the compression ratio is increased when an engine stop request is made.

  FIG. 9 shows an embodiment of the present invention in which the mechanical compression ratio is controlled in consideration of the deterioration of the startability due to the time required to increase the mechanical compression ratio, and when the engine is stopped. 5 is a flowchart showing an embodiment of a control routine for a mechanical compression ratio at the time. In the embodiment shown in FIG. 9, the pre-ignition occurrence determination means determines whether or not the pre-ignition occurrence precondition at the next engine start is satisfied when the engine stop request is made. When it is determined that the pre-ignition occurrence precondition at the time of starting the engine is satisfied, the mechanical compression ratio is increased by the variable compression ratio mechanism in advance when the engine stop request is made.

  The pre-ignition occurrence determination means in the present embodiment can detect a fuel pressure that is a pressure applied to the fuel in the fuel injection valve, and an engine temperature represented by, for example, an engine cooling water temperature or an engine oil temperature. A sensor 44, a sensor 45 capable of detecting the atmospheric temperature, and a relative position sensor 22 capable of detecting the relative positional relationship between the crankcase 1 and the cylinder block 2 and indicating the current mechanical compression ratio, It is configured to determine whether or not a pre-ignition occurrence precondition is satisfied based on the fuel pressure, engine temperature, atmospheric temperature, and mechanical compression ratio detected from each sensor when a stop request is made. . The pre-ignition occurrence determination means is not limited to such a configuration. For example, the temperature condition may be determined only from the engine cooling water temperature, and other than the above parameters. It may be configured to determine whether or not pre-ignition has occurred at the time of the next engine start after the engine stop using the above parameters.

  In step 101 in FIG. 9, first, the presence or absence of an engine stop request is confirmed. Specifically, the presence / absence of an on / off operation of the ignition switch is confirmed. When the off operation is confirmed, it is determined that an engine stop request has been made.

  When it is confirmed that the ignition switch is turned off, the process proceeds to the next step 102, where the next ignition switch is turned on, that is, whether or not there is a possibility that pre-ignition will occur at the next engine start is determined as pre-ignition occurrence determination. By means. More specifically, whether or not the pre-ignition occurrence precondition is satisfied by the pre-ignition occurrence determination means is determined based on the mechanical compression ratio, the fuel pressure, the engine temperature, and the atmospheric temperature when the engine stop request is made. . For example, the amount of fuel leakage into the cylinder is estimated from the fuel pressure. Further, from the engine temperature represented by the engine cooling water temperature and the engine oil temperature and the atmospheric temperature, it is determined whether or not the engine has been stopped in a state where the engine has not been sufficiently warmed up. This is because when the internal combustion engine is stopped without being sufficiently warmed up, it can be determined that no pre-ignition occurs regardless of the amount of fuel leakage into the cylinder. It is.

  In the present embodiment, the threshold values of the fuel pressure, the engine cooling water temperature, the engine oil temperature, and the atmospheric temperature with respect to the mechanical compression ratio when the engine stop request is made are specified. These specified threshold values are used to determine whether or not pre-ignition has occurred at the next engine start. In the present embodiment, these threshold values are created based on a prior evaluation test or analysis. Identified using a map or the like. Then, it is determined whether the fuel pressure, the engine coolant temperature, the engine oil temperature, and the atmospheric temperature when the engine stop request is made exceed the respective threshold values. It is determined that pre-ignition has occurred when the engine is started, and the routine proceeds to step 103. If any one of the fuel pressure, engine cooling water temperature, engine oil temperature, and atmospheric temperature when the engine stop request is requested does not exceed the threshold, pre-ignition at the next engine start In step 104, the mechanical compression ratio is controlled during normal operation.

  The control in step 103 is performed when the pre-ignition occurrence determination means determines that the pre-ignition occurrence precondition at the next engine start is satisfied, and the variable compression is performed in advance when an engine stop request is made. The mechanical compression ratio is increased by the ratio mechanism. More specifically, when it is determined by the pre-ignition occurrence determination means that the pre-ignition occurrence precondition at the next engine start is satisfied, the variable compression ratio mechanism is used in advance when an engine stop request is made. The mechanical compression ratio is increased to the maximum mechanical compression ratio that can be varied by the variable compression ratio mechanism. Since the control content at the next engine start differs depending on whether the mechanical compression ratio is increased or not when the engine stop request is made, the mechanical compression ratio set in either case is the ECU. It is comprised so that it may be memorize | stored.

  According to such a control routine shown in FIG. 9, when an engine stop request is made, it is determined whether or not a pre-ignition occurrence precondition at the next engine start is satisfied, and at the next engine start. If it is determined that the pre-ignition occurrence preconditions are satisfied, the mechanical compression ratio is increased by the variable compression ratio mechanism in advance when an engine stop request is made. It may be possible to avoid deterioration in startability due to the time required for comparison.

  FIG. 10 is a flowchart showing an embodiment of a control routine at the time of engine start, and an embodiment of the control routine at the time of the next engine start after the mechanical compression ratio control at the time of engine stop shown in FIG. It is a flowchart which shows.

  In the embodiment shown in FIG. 10, first, in step 201, the mechanical compression ratio when the engine is stopped is confirmed. That is, it is confirmed in step 103 shown in FIG. 9 whether or not the mechanical compression ratio has been increased. When the engine stop request is made, it is determined that pre-ignition is not generated at the next start, and if the high compression ratio of the mechanical compression ratio is not performed, the process proceeds to step 207, and the start control ( A start pattern 0) is made. On the other hand, when the engine stop request is made and it is determined that pre-ignition has occurred at the next start and a high compression ratio of the mechanical compression ratio is performed, the routine proceeds to step 202.

  In step 202, reconfirmation of the possibility of occurrence of pre-ignition at the time of engine start is performed. In the present embodiment, after the engine is stopped, the engine temperature is greatly reduced due to the passage of time or some other factor, and when the engine is actually started, the engine temperature may cause pre-ignition. Considering the case where the temperature has dropped to a temperature that cannot be controlled, the control for suppressing the occurrence of pre-ignition based on the engine coolant temperature at the time of engine start, that is, the cylinder is over-rich. It is determined whether it is necessary to perform such fuel injection control. Specifically, it is determined that it is necessary to perform start control for suppressing the occurrence of pre-ignition when the engine coolant temperature at the time of engine start is equal to or higher than a predetermined threshold (A) set in advance. Proceeding to step 203, start control of start pattern 3 is executed.

  FIG. 11 is a flowchart showing an embodiment of a control routine in the start pattern 3 executed in step 203 of FIG. In the start control according to the start pattern 3 shown in FIG. 11, a large amount of fuel is injected, and fuel injection control in which the fuel is injected during the compression stroke is executed. As a result, the inside of the cylinder can be over-rich, and the temperature in the cylinder can be lowered due to the latent heat of vaporization of the fuel, thereby making it possible to suppress the occurrence of pre-stage.

  As a matter of course, the engine speed starts from zero at the start of the start control by the start pattern 3 shown in FIG. In this embodiment, when the engine speed becomes greater than a predetermined threshold value (c), it is determined that the engine temperature has decreased to a temperature at which pre-ignition can be suppressed, and the fuel injection amount and fuel injection during normal operation are determined. It is switched to fuel injection control by time. That is, in this embodiment, at step 301, it is first determined whether or not the engine speed exceeds a predetermined threshold value (c) starting from zero, and the engine speed exceeds the predetermined threshold value. If it is determined that the fuel injection amount is determined, the routine proceeds to step 303, where fuel injection control is performed based on the fuel injection amount and fuel injection timing during normal operation. For example, the fuel injection amount is controlled to the normal fuel injection amount, and fuel is injected during the intake stroke Thus, the fuel injection timing is controlled. When it is determined in step 301 that the engine speed has exceeded a predetermined threshold value (c), it is not immediately switched to the fuel injection control during normal operation, but the occurrence of pre-ignition is more reliably suppressed. Therefore, the start control for suppressing the occurrence of pre-ignition in step 302, which will be described later, may be switched after performing several more times.

  On the other hand, in step 301, it is determined whether or not the engine speed exceeds a predetermined threshold starting from zero, and if it is determined that the engine speed does not exceed the predetermined threshold, the process proceeds to step 302. In order to suppress the occurrence of pre-ignition, a large amount of fuel is injected, and start control by fuel injection control that is performed during the compression stroke is executed. By the way, the start control that suppresses such pre-ignition is executed when the engine temperature is high enough to cause pre-ignition. Even if it is in a state, it is relatively easy to ignite and can be started by ignition with a spark plug.

  Here, referring again to FIG. 10, the start control in the present embodiment when it is confirmed in step 202 that the engine coolant temperature at the time of engine start is lower than a predetermined threshold (A) set in advance. This will be described below.

  When it is confirmed in step 202 of FIG. 10 that the engine cooling water temperature at the time of starting the engine is lower than a predetermined threshold value (A) set in advance, the routine proceeds to step 204. In step 204, after the engine is stopped, the engine temperature has further decreased due to the passage of time or some other factor, and cranking at a speed sufficient for starting the engine if the mechanical compression ratio remains high. In consideration of the case where the engine is not possible, it is determined whether or not it is necessary to reduce the mechanical compression ratio before starting the engine.

  In this embodiment, if it is confirmed in step 204 that the engine coolant temperature at the time of engine start is lower than a predetermined threshold (B) set in advance, the mechanical compression ratio is reduced to a low value before engine start. It is determined that there is a need for comparison, and the process proceeds to step 206 where start control of a start pattern 1 described later is executed.

  On the other hand, when it is confirmed in step 204 that the engine coolant temperature at the time of engine startup is higher than a predetermined threshold value (B) set in advance, that is, the engine coolant temperature at the time of engine start is set to the threshold value (A ) And lower than the threshold (B), it is determined that there is no possibility of pre-ignition and it is not necessary to lower the mechanical compression ratio before starting the engine. Then, the process proceeds to step 205 where the start control of the start pattern 2 for executing the start control during the normal operation similar to the start pattern 0 is executed at the mechanical compression ratio that remains high.

  Further, when the engine coolant temperature at the time of starting the engine is lower than a predetermined threshold value (B) set in advance in step 204, the engine temperature is extremely low, and the mechanical compression ratio remains at a high compression ratio. It is determined that cranking rotation necessary for starting cannot be obtained due to high friction and high air density, and cranking at a speed sufficient for starting is impossible. Proceeding to step 206, mechanical compression is performed before starting. A start pattern 1 is executed in which the start control during normal operation is executed after the ratio is reduced.

  As described above, according to the control at the time of engine stop and engine start in the cylinder injection type spark ignition internal combustion engine of the present invention shown in FIGS. 9 to 11, the amount of fresh air in the cylinder at the time of engine start can be reduced. This makes it possible to reduce the amount of injected fuel required to avoid the occurrence of pre-ignition in an over-rich state in the cylinder, thereby improving fuel efficiency and shortening the time required for fuel injection. Also, if the amount of fresh air in the cylinder at the start of the engine is small, the possibility of reverse engine rotation due to explosive force is small even if pre-ignition occurs, and the generated abnormal noise is also small. Become. In other words, according to the present invention, in a direct injection spark ignition internal combustion engine having a variable compression ratio mechanism, the variable compression ratio mechanism is effectively used without losing combustion stability, fuel consumption, or the like when the engine is started. It becomes possible to suppress the occurrence of pre-ignition.

DESCRIPTION OF SYMBOLS 1 Crankcase 2 Cylinder block 3 Cylinder head 4 Piston 5 Combustion chamber 7 Intake valve 70 Camshaft for intake valve drive A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (3)

Fuel injection valves to perform direct fuel injection into a cylinder of an internal combustion engine and a spark plug which performs the ignition of the mixture in a cylinder, a cylinder injection type comprising a variable compression ratio mechanism able to change a mechanical compression ratio In a spark ignition internal combustion engine, when suppressing preignition in which the air-fuel mixture in the cylinder ignites before the ignition timing of the spark plug, fuel injection control is performed to make the cylinder excessively rich when the engine is started. In- cylinder injection spark ignition internal combustion engine
Comprises a preignition determining means for determining whether the establishment of the occurrence prerequisites for flops les ignition engine during starting,
The pre-ignition occurrence determination means, when an engine stop request is made, identifies the respective threshold values of the fuel pressure, the engine temperature and the atmospheric temperature with respect to the mechanical compression ratio when the engine stop request is made, Preconditions for pre-ignition are established when there is a pre-ignition at the next engine start when all of the fuel pressure, engine temperature, and atmospheric temperature at the time of the stop request exceed the respective thresholds. In this determination, if it is determined that the pre-ignition occurrence precondition at the next engine start is satisfied, the mechanical compression is performed in advance by the variable compression ratio mechanism when an engine stop request is made. An in-cylinder spark-ignition internal combustion engine with a high compression ratio . When it is determined by the pre-ignition occurrence determination means that the pre-ignition occurrence precondition at the next engine start is satisfied, the mechanical compression ratio is previously set by the variable compression ratio mechanism when the engine stop request is made. The in-cylinder injection spark ignition internal combustion engine according to claim 1, wherein the compression ratio is increased to a maximum mechanical compression ratio that can be varied by the variable compression ratio mechanism. When it is determined by the pre-ignition occurrence determination means that the pre-ignition occurrence precondition at the next engine start is satisfied, and the engine temperature at the next engine start is lower than a predetermined value The in-cylinder injection spark ignition internal combustion engine according to claim 1 or 2, wherein the engine is started after the mechanical compression ratio is lowered by the variable compression ratio mechanism.

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