JP6131840B2 - Control device for compression ignition engine - Google Patents

Description

  The present invention relates to a control device for a compression ignition engine.

  Conventionally, compression self-ignition combustion has been performed in a combustion chamber formed in an engine body for the purpose of improving fuel efficiency.

  For example, Patent Document 1 describes an engine that performs compression ignition combustion of an air-fuel mixture in a cylinder when the operating state of the engine is in a predetermined region on a low rotation / low load side. In this engine, when performing compression ignition combustion, the exhaust valve is closed on the advance side with respect to the intake top dead center, and the intake valve is opened on the retard side with respect to the intake top dead center. A negative overlap period is provided in which both the exhaust valve and the intake valve are closed with the top dead center in between, and the burned gas remains in the cylinder, and internal EGR (Exhaust Gas Recirculation) is performed. The temperature inside is increased to improve the ignitability of the air-fuel mixture.

JP 2009-197740 A

  In Patent Document 1, there is a case where compression auto-ignition combustion cannot be stably realized in an operation region where the engine speed is low, and more stable realization of compression auto-ignition combustion is desired.

  The present invention has been made in view of the above points, and provides a control device for a compression ignition engine that can realize more stable and stable compression autoignition combustion in the entire region where compression autoignition combustion is performed. .

  As a result of earnest research to solve the above-mentioned problems, the inventors of the present invention, as a method of leaving high-temperature burned gas in the cylinder, that is, as a method of internal EGR, A method of providing a negative overlap period in which both the intake valve and the exhaust valve are closed, and the exhaust valve is opened at least during the exhaust stroke and the intake stroke, and the burned gas discharged to the exhaust port side flows back into the cylinder. Focusing on the relationship between each method and the cooling loss, in the region where the engine speed is low, the burned gas discharged to the exhaust port side is less than the method that provides the negative overlap period. The method of backflowing into the cylinder can reduce the temperature drop of the burned gas and increase the temperature of the air-fuel mixture more effectively, but in the region where the engine speed is high, it is negative over Tsu towards a method of providing a flop period has found that it is possible to increase more effectively the temperature of the small suppressed by the gas mixture the temperature drop of the combustion gases.

  This is considered to be due to the following reasons. In the method of providing a negative overlap period, the temperature of the burned gas is reduced by cooling the burned gas that has been compressed and heated between the exhaust valve closing and the intake top dead center by the cylinder wall, that is, engine cooling water. Is lowered. On the other hand, in the method of reversing the burnt gas, the compressed gas is not heated at a high temperature, so the temperature drop of the burnt gas due to engine cooling water can be kept small, but the burnt gas is discharged to the exhaust port side. Therefore, the temperature drop of the burnt gas due to cooling at the exhaust port or the exhaust valve becomes large.

  When the engine speed is low, the calorific value of the air-fuel mixture in the cylinder per unit time is small and the wall surface temperature of the cylinder does not become too high, so the temperature of the burned gas after expansion can be kept relatively low. As a result, the temperature difference between the expanded burned gas and the exhaust port or exhaust valve becomes small, so that the temperature drop of the burned gas due to cooling with these can be suppressed to a small level. Since the time until the top dead center is long and the burned gas is cooled by the engine coolant for a relatively long time, the temperature drop of the burned gas by the engine coolant becomes relatively large. On the other hand, when the engine speed is high, the time from the exhaust valve closing to the intake top dead center is short, so the temperature drop of the burnt gas due to engine cooling water can be kept small, but the burnt gas after expansion Since the temperature is maintained relatively high and the temperature difference between the burned gas and the exhaust port or exhaust valve becomes large, the burned gas reciprocates between the exhaust port and the cylinder at a high speed and contacts the exhaust port over a wider range. As a result, the burned gas is cooled by the exhaust port, and the cooling loss due to cooling by the exhaust port and the exhaust valve becomes relatively large.

  Therefore, in the present invention, a combustion chamber in which an air-fuel mixture containing at least fuel and air burns is formed on the inner side in order to achieve more reliable and stable compression auto-ignition combustion in the entire region where compression auto-ignition combustion is performed. A cylinder, an intake port that introduces intake air into the cylinder, an exhaust port that discharges exhaust from the cylinder, an intake valve that can open and close the intake port, and an exhaust valve that can open and close the exhaust port And a control means for controlling the combustion mode of the air-fuel mixture and the open states of the intake valve and the exhaust valve, and the control means is a low load region where at least the engine load is lower than a predetermined load Then, the combustion mode in the combustion chamber is changed to compression auto-ignition combustion in the entire region, and in the compression auto-ignition region where the compression auto-ignition combustion is performed, the engine speed is higher than a specific speed. In the side compression auto-ignition region, the exhaust valve is closed on the advance side with respect to the intake top dead center, and the intake valve is opened on the retard side with respect to the intake top dead center, thereby sandwiching the intake top dead center. In the low-rotation side compression ignition region where the engine speed is less than the specific engine speed, a burn-up gas remains in the cylinder by providing a negative overlap period in which both the exhaust valve and the intake valve are closed. Then, the exhaust valve is opened at least during the exhaust stroke and the intake stroke, and the burned gas discharged to the exhaust port is caused to flow back into the cylinder, so that the burned gas remains in the cylinder.

  According to the present invention, in the high rotation side compression auto-ignition region where the engine speed is high, the cylinder is provided by providing a negative overlap period in which both the exhaust valve and the intake valve are closed across the intake top dead center. In the low-rotation-side compression auto-ignition region where the engine speed is low, the exhaust valve is opened at least during the exhaust stroke and the intake stroke, while the burned gas remains inside, that is, internal EGR is performed. Since internal EGR is performed by the method of causing the discharged burned gas to flow back into the cylinder, the temperature of the air-fuel mixture can be increased more effectively by suppressing the temperature drop of the burned gas in each rotational speed region. It is possible to improve the ignitability of the air-fuel mixture and to realize more stable and stable compression ignition combustion over the entire region where compression ignition combustion is performed.

  As a specific control configuration of the exhaust valve, the lift characteristic of the exhaust valve has a first characteristic that closes on the advance side of the intake top dead center, and a valve opening period that is longer than the first characteristic. Exhaust valve lift characteristic changing means that can be changed to a second characteristic that continues to open until halfway through the intake stroke, the second characteristic is a predetermined first valve opening period, and the exhaust valve has a predetermined lift amount. After the first valve opening period, the exhaust valve is in the first valve opening period after the end of the first valve opening period until the predetermined timing of the intake stroke. The control means is configured to open the valve with a lift amount smaller than the maximum lift amount, and the control means controls the exhaust valve lift characteristic changing means so that the lift characteristic of the exhaust valve is compressed on the high rotation side. In the auto-ignition region, the first characteristic is set, while in the low-rotation side compression auto-ignition region, Include those to the second characteristic (claim 2).

  In this way, with a simple configuration in which the lift characteristics of the exhaust valve are made different, the exhaust valve is opened during the exhaust stroke and during the intake stroke, and more advanced than the intake top dead center. It can be changed appropriately when the valve is closed on the corner side.

  In the present invention, it is preferable that the control means sets the opening timing of the intake valve to be retarded from the intake top dead center in the low rotation side compression auto-ignition region.

  In this way, in the low-rotation side compression auto-ignition region, the burned gas can be properly flown back from the exhaust port side into the cylinder during the intake stroke, and the burned gas can be reliably left in the cylinder. Can do.

  An intake valve lift characteristic changing means capable of changing the lift characteristic of the intake valve to a predetermined small lift characteristic and a large lift characteristic in which the maximum lift amount is set larger than the maximum lift amount of the small lift characteristic; The control means controls the intake valve lift characteristic changing means so that the lift characteristic of the intake valve becomes the large lift characteristic in the high-rotation side compression auto-ignition area, while the low-rotation side compression auto-ignition area Then, it is preferable to set it as the said small lift characteristic (Claim 4).

  In this way, it is possible to appropriately secure the intake air amount introduced into the cylinder in the high-rotation side compression auto-ignition region.

  In the above configuration, the opening timing of the exhaust valve in the high-rotation side compression auto-ignition region is controlled to be advanced from the opening timing of the exhaust valve in the low-rotation side compression auto-ignition region. (Claim 5).

  In this way, the exhaust valve is closed at a relatively advanced position before the intake top dead center in the high-rotation side compression auto-ignition region, so that more burned gas remains in the cylinder. In addition, the exhaust valve is closed at a relatively retarded position after intake top dead center in the low-rotation side compression auto-ignition region, so that more burned gas flows back into the cylinder. The amount of burnt gas can be ensured in the entire compression auto-ignition region, the temperature of the mixture can be increased, and the ignitability of the mixture can be improved.

  Further, in the present invention, the control means includes a retard amount amount from an intake top dead center of the intake valve opening timing and an intake valve timing of the exhaust valve closing timing in the high rotation side compression auto-ignition region. It is preferable that the amount of advance from the dead point is substantially the same (claim 6).

  In this way, the pumping loss caused by the compression / expansion of burnt gas can be kept small in the high-rotation side compression auto-ignition region where the intake valve and the exhaust valve are closed with the intake top dead center in between. Can perform good driving.

  Specifically, if the amount of delay from the intake top dead center at the opening timing of the intake valve is greater than the amount of advance from the intake top dead center at the closing timing of the exhaust valve, the burned gas is reduced. Therefore, the pumping loss must be expanded more than the amount of compression up to the intake top dead center, while the retard amount of the intake valve opening timing from the intake top dead center is If the advance amount from the dead center is made smaller, the burned gas must be compressed more than the expansion work from the intake top dead center, and the pumping loss increases. On the other hand, as described above, if the retard amount from the intake top dead center at the opening timing of the intake valve and the advance amount from the intake top dead center at the close timing of the exhaust valve are substantially the same. The pumping loss can be kept small.

  As described above, according to the present invention, it is possible to realize more stable and stable compression auto-ignition combustion in the entire region where compression auto-ignition combustion is performed.

1 is a schematic view showing an engine system according to an embodiment of the present invention. It is a block diagram which concerns on control of the engine system shown in FIG. It is sectional drawing which expands and shows the combustion chamber shown in FIG. (A) It is the figure which showed the lift characteristic of the exhaust valve in normal mode. (B) It is the figure which showed the lift characteristic of the exhaust valve in a special mode. It is a figure which illustrates the operation control map of an engine. It is the figure which showed the valve opening state of the exhaust valve in a 1st area | region. It is the figure which showed the valve opening state of the exhaust valve in a 2nd area | region. It is a figure which illustrates the driving | operation control map of the engine which concerns on other embodiment. It is the figure which showed the valve opening state of the exhaust valve in the 2nd area | region which concerns on other embodiment.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings.

  FIG. 1 is a schematic configuration diagram of an engine system 100 of a control device for a compression ignition engine according to the present invention. The engine system 100 is mounted on a vehicle and has an engine body 1.

  The engine body 1 is a gasoline engine to which a fuel containing at least gasoline is supplied, and is a four-cycle engine, that is, an engine in which an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke are sequentially performed. The engine body 1 is a compression ignition type engine in which compression self-ignition combustion is performed. The engine body 1 includes a cylinder block 11 provided with a cylinder 18 and a cylinder head 12 disposed on the cylinder block 11. The engine body 1 has, for example, four cylinders 18.

  A piston 14 connected to the crankshaft 15 via a connecting rod 142 is fitted in each cylinder 18 so as to be able to reciprocate. In each cylinder 18, a combustion chamber 19 surrounded by the inner surface of the cylinder 18 and the top surface of the piston 14 is formed.

  Although the specific structure of the piston 14 and the combustion chamber 19 is not specifically limited, For example, it has a structure as shown in FIG. In the example shown in FIG. 3, the center of the top surface of the piston 14 is a so-called reentrant type that is recessed in a direction away from the cylinder head 12 and becomes shallower after the depth becomes deeper from the center toward the radially outer side. The cavity 141 is formed.

  In the present embodiment, the geometric compression ratio of the engine body 1 is set to a relatively high value of 15 or more for the purpose of improving the thermal efficiency and stabilizing the compression ignition combustion. The geometric compression ratio of the engine body 1 is not limited to this, but a range of about 15 or more and 20 or less is preferable.

  In the cylinder head 12, an intake port 16 for introducing intake air into the cylinder 18 and an exhaust port 17 for discharging exhaust gas from the cylinder 18 are formed for each cylinder 18. The intake port 16 and the exhaust port 17 are provided with these ports, specifically, an intake valve 21 and an exhaust valve 22 that open and close the openings of the ports 16 and 17 formed in the cylinder head 12, respectively. Yes.

  The exhaust valve 22 is driven by an exhaust valve drive mechanism 70a. The exhaust valve drive mechanism 70 a includes an exhaust valve lift variable mechanism (hereinafter referred to as “exhaust VVL (variable valve lift)”) 71 and an exhaust phase variable mechanism (hereinafter referred to as “exhaust VVT (variable valve timing)”) 75.

  The exhaust VVL 71 switches the operation mode of the exhaust valve 22 between a normal mode indicated by a solid line in FIG. 4A and a special mode indicated by a solid line in FIG. That is, the lift characteristic of the exhaust valve 22 is switched between a first characteristic indicated by a solid line in FIG. 4A and a second characteristic indicated by a solid line in FIG. In the normal mode, the valve lift of the exhaust valve 22 gradually increases after opening, and when it reaches the maximum lift, it gradually decreases again to zero. In the special mode, the valve lift of the exhaust valve 22 gradually increases after the valve opening and reaches the maximum lift during the first valve opening period t_1 as in the normal mode. Without reaching zero as it is, the lift amount, that is, the lift lower than the maximum lift in the first valve opening period t_1 is maintained for a predetermined period, and then reaches zero. As described above, in the special mode, the opening period of the exhaust valve 22, that is, the period t_3 from when the exhaust valve 22 is opened to when the exhaust valve 22 is finally closed (during the intake stroke in this embodiment) is a predetermined period. A first valve opening period t_1 that is the maximum lift of the second valve, and a second valve configured to be smaller than the maximum lift in the first valve opening period t_1 continuously from the first valve opening period t_1. It consists of a valve opening period t_2. In the special mode, the valve is opened from immediately before the closing timing in the normal mode to a predetermined timing that is retarded from the closing timing in the normal mode, and the opening period of the exhaust valve is longer in the special mode than in the normal mode. It is getting longer. The exhaust VVL 71 includes a first cam and a second cam having different cam shapes in order to realize these modes. The first cam has a shape corresponding to the lift characteristic indicated by the solid line in FIG. 4A and has one cam crest. The second cam has a shape corresponding to the lift characteristic indicated by the broken line in FIG. 4B, and has two cam peaks. The exhaust VVL 71 includes a lost motion mechanism that selectively transmits the operating state of the first cam and the second cam to the exhaust valve 22. By transmitting the operating state of the first cam to the exhaust valve 22, the exhaust valve 22. The operation state is set to the normal mode, and the operation state of the exhaust valve 22 is set to the special mode by transmitting the operation state of the second cam to the exhaust valve 22. The exhaust VVL 71 is, for example, a hydraulically operated type. When the operating state of the second cam is transmitted to the exhaust valve 22, the lift characteristic of the exhaust valve 22 has a shape shown by a broken line in FIG.

  The exhaust VVT 75 changes the opening timing and closing timing of the exhaust valve 22 by changing the rotation phase of the exhaust camshaft with respect to the crankshaft 15. The exhaust valve VVT75 changes the valve opening timing and the valve closing timing of the exhaust valve 22 while maintaining the valve opening period of the exhaust valve 22 constant in each mode of the normal mode and the special mode. The exhaust VVT 75 may employ a hydraulic, electromagnetic, or mechanical known structure as appropriate, and a detailed description thereof will be omitted.

  The exhaust VVT 75 sets the rotational phase of the exhaust camshaft so that the exhaust valve 22 opens in the intake stroke in addition to the exhaust stroke when the operating state of the exhaust valve 22 is in the special mode. Further, when the exhaust valve 22 is in the special mode, the exhaust VVT 75 has the intake top dead center during the second valve opening period t_2, that is, the valve of the exhaust valve 22 at the intake top dead center. The rotational phase of the exhaust camshaft is set so that the lift becomes a relatively small value realized during the second valve opening period t_2. Thus, in the present embodiment, the exhaust valve 22 is opened during the intake stroke in addition to the exhaust stroke when the operating state of the exhaust valve 22 is set to the special mode. In particular, in the present embodiment, the exhaust valve 22 is continuously opened in the exhaust stroke and the intake stroke with the intake top dead center being sandwiched without closing in the middle. Here, when the exhaust valve 22 is continuously opened across the intake top dead center, the exhaust valve 22 and the piston 14 may interfere with each other. On the other hand, in this embodiment, as described above, the valve lift amount of the exhaust valve 22 near the intake top dead center is suppressed to a small value, so that interference between the exhaust valve 22 and the piston 14 is avoided. Can do. The exhaust double opening, that is, the special mode is performed in order to perform a so-called internal EGR by leaving the high-temperature burned gas, that is, the internal EGR gas in the combustion chamber 19. Specifically, when the exhaust is opened twice and the exhaust valve 22 is opened even during the intake stroke, the exhaust once discharged to the exhaust port 17 during the exhaust stroke is in the combustion chamber 19 during the intake stroke. Then, exhaust gas, that is, high-temperature burned gas remains in the combustion chamber 19.

  The intake valve 22 is driven by an intake valve drive mechanism 70b. As with the exhaust valve drive mechanism 70a, the intake valve drive mechanism 70b changes the rotation phase of the intake jam shaft relative to the intake VVL 74 and the crankshaft 15 by switching the operation mode of the intake valve 21 in two modes, and opens the intake valve 21. Intake VVT72 which changes valve timing and valve closing timing is included.

  The intake VVL 74 selectively selects the large lift cam that relatively increases the valve lift of the intake valve 21, the small lift cam that relatively decreases the valve lift of the intake valve 21, and the operating state of one of these cams. And a lost motion mechanism that transmits to the intake valve 21. The intake VVL 74 transmits the operation state of the large lift cam to the intake valve 21, thereby setting the operation mode of the intake valve 21 to a mode in which the valve lift and the valve opening period are relatively large. The intake VVL 74 transmits the operation state of the small lift cam to the intake valve 21, thereby setting the operation mode of the intake valve 21 to a mode in which the valve lift and the valve opening period are relatively small. The large lift cam and the small lift cam are set to be switched at the same valve closing timing or valve opening timing.

  As with the exhaust VVT 75, the intake VVT 72 may employ a known hydraulic, electromagnetic, or mechanical structure as appropriate, and illustration of the detailed structure is omitted.

  An intake passage 30 is connected to each intake port 16. Specifically, a branch passage that branches in correspondence with the cylinder 18 is formed at the downstream end of the intake passage 30, and these branch passages are connected to the intake ports 16.

  In the intake passage 30, an air cleaner 31, a water-cooled intercooler / warmer 34, a throttle valve 36, and a surge tank 33 are arranged in this order from the upstream side.

  An intercooler bypass passage 35 that bypasses the intercooler / warmer 34 is connected to the intake passage 30. The intercooler bypass passage 35 is provided with an intercooler bypass valve 351 for adjusting the flow rate of air passing through the intercooler bypass passage 35 in order to adjust the temperature of fresh air flowing into the cylinder 18. Note that the intercooler / warmer 34 and its associated members may be omitted.

  An exhaust passage 40 is connected to each exhaust port 17. Specifically, similarly to the intake passage 30, branch passages that branch in correspondence with the cylinders 18 are formed at the upstream end of the exhaust passage 40, and these branch passages are connected to the intake ports 18. .

  The exhaust passage 40 is provided with an exhaust purification device that purifies harmful components in the exhaust gas. In the present embodiment, a direct catalyst 41 and an underfoot catalyst 42 are provided in order from the upstream side. The direct catalyst 41 and the underfoot catalyst 42 include, for example, a three-way catalyst.

  An EGR device 50 is provided between the intake passage 30 and the exhaust passage 40 to recirculate a part of the exhaust gas to the intake air, that is, to perform external EGR. The EGR device 50 includes an EGR passage 51, an EGR cooler 52, and an EGR cooler bypass passage 53. The EGR passage 51 connects a portion of the intake passage 30 between the surge tank 33 and the throttle valve 36 and a portion of the exhaust passage 40 upstream of the direct catalyst 41. The EGR cooler 52 is for cooling the gas passing through the EGR passage 51, and is interposed in the EGR passage 51. The EGR cooler bypass passage 53 is a passage that bypasses the EGR cooler 52, and connects the upstream and downstream portions of the EGR cooler 52 in the EGR passage 51. The EGR passage 51 and the EGR cooler bypass passage 53 are provided with an EGR valve 511 and an EGR cooler bypass valve 531 for adjusting the flow rate of exhaust gas passing through the passages 51 and 53, respectively. Hereinafter, recirculation of a part of the exhaust gas to the intake air using the EGR device 50 is referred to as external EGR, and the exhaust gas recirculated to the intake air by the EGR device 50 may be referred to as external EGR gas.

  An injector 67 that directly injects fuel into the combustion chamber 19 is attached to the cylinder head 12 for each cylinder 18. As shown in FIG. 3, the injector 67 is disposed so that its nozzle hole faces the inside of the combustion chamber 19 from the central portion of the ceiling surface of the combustion chamber 19, and is opposed to the cavity 141. In the present embodiment, the injector 67 is a multi-hole type having a plurality of nozzle holes. The fuel spray injected from the injector 67 spreads radially from the center position of the combustion chamber 19.

  Here, when the fuel is injected from the injector 67 at the timing when the piston 14 is positioned near the compression top dead center, the fuel spray flows along the wall surface of the cavity 141 as shown by the arrow in FIG. Therefore, in the engine system 100, when high pressure retarded injection described later is performed, the fuel spray can be diffused early to form an air-fuel mixture early.

  Fuel is supplied to the injector 67 from a fuel tank (not shown) by the fuel supply system 62. The fuel supply system 62 includes a fuel pump 63 and a pressure accumulation rail 64. The fuel pump 63 pumps fuel from the fuel tank to the pressure accumulation rail 64. In the present embodiment, the fuel pump 63 is a plunger type pump driven by the engine 1. The pressure accumulation rail 64 stores the pumped fuel at a relatively high pressure. The injector 67 injects high-pressure fuel stored in the pressure accumulation rail 64 into the combustion chamber 19. Although the value of an injection pressure is not specifically limited, For example, it is set to 30 MPa or more and 120 MPa or less.

  A spark plug 25 that forcibly ignites the air-fuel mixture in the combustion chamber 19 is attached to the cylinder head 12. In the present embodiment, the spark plug 25 is disposed through the cylinder head 12 so as to extend obliquely downward from the exhaust side of the engine body 1. As shown in FIG. 3, the tip of the spark plug 25 faces the cavity 141 of the piston 14 located at the compression top dead center.

  Each device is controlled by a power train control module (control means, hereinafter referred to as PCM) 10. The PCM 10 includes a microprocessor having a CPU, a memory, a counter timer group, an interface, and a path connecting these units.

  As shown in FIGS. 1 and 2, detection signals from various sensors SW <b> 1 to SW <b> 16 are input to the PCM 10.

The sensor SW1 is an air flow sensor SW1 that detects the flow rate of fresh air. The sensor SW2 is an intake air temperature sensor SW2 that detects the temperature of fresh air. The air flow sensor SW <b> 1 and the intake air temperature sensor SW <b> 2 are disposed on the downstream side of the air cleaner 31 in the intake passage 20. The sensor SW3 is a second intake air temperature sensor SW3 that detects the temperature of fresh air after passing through the intercooler / warmer 34, and is disposed on the downstream side of the intercooler / warmer 34. The sensor SW4 is an EGR gas temperature sensor SW4 for detecting the temperature of the external EGR gas, and is disposed in the vicinity of the connection portion with the intake passage 30 in the EGR passage 50. The sensor SW 5 is an intake port temperature sensor SW 5 that detects the temperature of intake air immediately before flowing into the cylinder 18, and is attached to the intake port 16. The sensor SW 6 is an in-cylinder pressure sensor SW 6 that detects the pressure in the cylinder 18 and is attached to the cylinder head 12. The sensor SW7 is an exhaust temperature sensor SW7 that detects the exhaust temperature. The sensor SW8 is an exhaust pressure sensor SW8 that detects the exhaust pressure. The exhaust temperature sensor SW7 and the exhaust pressure sensor SW8 are arranged in the vicinity of the connection portion of the EGR passage 50 in the exhaust passage 40. The sensor SW9 is a linear O ₂ sensor SW9 that detects the oxygen concentration in the exhaust gas, and is arranged on the upstream side of the direct catalyst 41 in the exhaust passage 40. The sensor SW 10 is a lambda O ₂ sensor SW 10 that detects the oxygen concentration in the exhaust gas, and is disposed between the direct catalyst 41 and the underfoot catalyst 42. The sensor SW11 is a water temperature sensor SW11 that detects the temperature of the engine cooling water. The sensor SW12 is a crank angle sensor SW12 that detects the rotation angle of the crankshaft 15. The sensor SW13 is an accelerator opening sensor SW13 that detects an accelerator opening corresponding to an operation amount of an accelerator pedal (not shown) of the vehicle. Sensors SW14 and SW15 are intake-side and exhaust-side cam angle sensors SW14 and SW15, respectively. The sensor SW 16 is a fuel pressure sensor SW 16 that detects the pressure of the fuel supplied to the injector 67, and is attached to the common rail 64.

  The PCM 10 performs various calculations based on the detection signals of the sensors SW1 to SW16. The PCM 10 determines the operating conditions of the engine body 1 and the vehicle based on these detection signals. The PCM 10 controls the injector 67, the spark plug 25, the fuel supply system 62, and the actuators of various valves (throttle valve 36, intercooler bypass valve 351, EGR valve 511, EGR cooler bypass valve 531) according to operating conditions. Output signals to control them. The PCM 10 outputs control signals to the intake VVT 72, the intake VVL 74, the exhaust VVT 75, and the exhaust VVL 71 in accordance with operating conditions, and controls these, the intake valve 21, and the exhaust valve 22.

  FIG. 5 shows a control map of the engine speed on the horizontal axis and the engine load on the vertical axis. As described above, the engine body 1 performs the compression auto-ignition combustion in which the air-fuel mixture is self-ignited and burned without ignition by the spark plug 25. However, when compression auto-ignition combustion is performed in an operation region where the engine load is high, the temperature of the air-fuel mixture is high, so that combustion becomes steep and problems such as combustion noise occur. Therefore, in the engine system 100 according to the present embodiment, compression auto-ignition combustion is performed only in a low load region where the engine load is less than the predetermined first load T1, and ignition is performed in a high load region where the engine load is equal to or higher than the first load T1. Spark ignition combustion in which the air-fuel mixture is forcibly ignited by the plug 25 is performed. That is, in this engine system 100, the low load region is set to the CI (Compression Ignition) combustion region, and the high load region is set to the SI (Spark Ignition) combustion region. Note that the boundary lines of these combustion regions are not limited to the illustrated examples.

  The CI combustion region is further divided into two regions according to the engine speed. Below, the detailed control content of each area | region is demonstrated.

(1) First region (low-rotation side compression auto-ignition region)
In the first region A1 in which the engine speed is set to a low speed region less than the predetermined reference speed NE1 in the CI combustion region, the external EGR is not performed and the internal EGR, that is, the burned gas is returned to the intake air. Instead of remaining directly in the cylinder. And in 1st area | region A1, as shown in FIG. 6, internal EGR is performed by exhaust_gas | exhaustion twice opening. Specifically, in the first region A1, the EGR valve 511 and the EGR cooler bypass valve 531 are closed. On the other hand, the operating state of the exhaust valve 22 is set to the special mode by the exhaust VVL 71. By setting the special mode, the exhaust valve 22 opens during the intake stroke in addition to the exhaust stroke, and the exhaust once discharged to the exhaust port 17 during the exhaust stroke flows back into the combustion chamber 19 during the intake stroke. Therefore, high-temperature burned gas can be left in the combustion chamber 19. In FIG. 6, the solid line indicates the valve lift of the exhaust valve 22, and the broken line indicates the valve lift of the intake valve 21.

  In the first region A1, the intake valve 21 is slightly lifted by the intake VVL 54, and the opening timing of the intake valve 21 is retarded from the intake top dead center by the intake VVT 72.

  Further, in the first region A1, intake stroke injection in which injection is performed by the injector 67 during the intake stroke is performed.

  The reference rotational speed NE1 is set to about 4000 rpm, for example.

(2) Second region (high rotation side compression auto-ignition region)
Even in the second region A2 in which the engine speed in the CI combustion region is set to a high engine speed region that is greater than or equal to the reference engine speed NE1, the EGR valve 511 and the EGR cooler bypass valve 531 are closed, the external EGR is stopped, and the internal EGR is stopped. Only is implemented. However, in the second region A2, as shown in FIG. 7, the exhaust valve 21 and the exhaust valve 22 are not closed twice with the intake top dead center TDC sandwiched therebetween, as shown in FIG. The internal EGR is performed by providing the (NVO) period. Specifically, the exhaust VVL 71 sets the operation state of the exhaust valve 22 to the normal mode, and the exhaust VVT 75 sets the closing timing EVC of the exhaust valve 22 to the advance side with respect to the intake top dead center TDC. The valve opening timing IVO of the intake valve 21 is set to be retarded from the intake top dead center TDC. Thus, when the negative overlap period is provided across the intake top dead center TDC, a part of the burned gas remains in the cylinder 18 without being discharged to the exhaust port 17 side. In the present embodiment, by the intake VVT 72 and the exhaust VVT 75, the advance amount θ1 from the intake top dead center TDC of the exhaust valve 22 and the retard amount θ2 from the intake top dead center of the intake valve 21 are made the same. Further, due to the exhaust VVT 75, the valve opening timing EVO of the exhaust valve 22 is set to an advance side with respect to the valve opening timing EVO in the first region A1.

  In the present embodiment, also in the second region A2, the intake valve 21 is set to a small lift by the intake VVL 54.

  Note that, also in the second region A2, intake stroke injection in which injection is performed by the injector 67 during the intake stroke is performed.

  The above-described control is performed in the first area A1 and the second area A2 for the following reason.

  As described above, when the engine load is high, the temperature in the combustion chamber 19 is high, and thus, when compression ignition is performed, problems such as combustion noise occur. Therefore, in order to avoid combustion noise and the like, the temperature in the combustion chamber 19 is preferably low. On the other hand, when the temperature in the combustion chamber 19 is too low, the temperature of the air-fuel mixture does not rise to a temperature at which self-ignition is possible, and misfires occur, so that compression self-ignition combustion cannot be realized stably. Problems arise.

  Therefore, the engine load is a relatively low load less than the first load T1, and the heat generation amount of the air-fuel mixture in the combustion chamber 19 is small, and the temperature in the combustion chamber 19 cannot be sufficiently increased only by this heat generation amount. In the second region A2, in order to avoid misfire or the like, the internal EGR is performed as described above to leave the high-temperature burned gas (internal EGR gas) in the combustion chamber 19, thereby Increase temperature.

  Here, as a method of performing internal EGR, as described above, there are a method of opening exhaust twice and a method of providing a negative overlap period with intake top dead center interposed therebetween. As a result of research, in the region where the engine speed is low, the method of opening the exhaust twice reduces the temperature drop of the burned gas more effectively than the method of providing the negative overlap period, and the temperature of the mixture is more effective. On the other hand, in the region where the engine speed is high, it has been found that the method of providing a negative overlap period can more effectively increase the temperature of the air-fuel mixture by suppressing the temperature drop of the burned gas to a smaller level. It was.

  This is considered to be due to the following reasons.

  When a negative overlap period is provided across the intake top dead center, the burnt gas is compressed and heated up after the exhaust valve 22 is closed until the intake top dead center. Due to the high temperature, a temperature difference is generated between the temperature of the burned gas and the wall surface of the cylinder. When the temperature difference occurs, the burned gas is cooled by the engine coolant that cools the cylinder wall surface, specifically, the cylinder wall surface. By this cooling, the thermal energy of the burned gas is reduced, and the temperature of the burned gas is lowered.

  On the other hand, when the exhaust is opened twice, the exhaust valve 22 is opened across the intake top dead center. For this reason, the burnt gas is not compressed and heated up to the intake top dead center. Therefore, when the exhaust is opened twice, the temperature drop of the burned gas due to the cooling by the cylinder wall surface, that is, the engine cooling water, can be suppressed to a small level. However, when the exhaust is opened twice, the burnt gas flows into the cylinder 18 after being exhausted to the exhaust port 17 side. For this reason, when the exhaust is opened twice, the burned gas is cooled by the wall surface of the exhaust port 17 and the exhaust valve 22, thereby reducing the thermal energy of the burned gas and lowering its temperature.

  Here, when the engine speed is low, the calorific value of the air-fuel mixture in the cylinder per unit time is small and the wall surface temperature of the cylinder does not increase so much, so the temperature of the burned gas after expansion is relatively low. Accordingly, the temperature difference between the burned gas temperature and the exhaust port 17 and the exhaust valve 22 in a state where the compression is not increased is small. Therefore, when the engine speed is low, even if the exhaust is opened twice, the temperature drop of the burned gas due to cooling by the exhaust port 17 or the exhaust valve 22 can be suppressed to a small level. On the other hand, when the engine speed is low, the time from the closing of the exhaust valve 22 to the intake top dead center TDC (not the crank angle) is relatively long, and the intake top dead center TDC after the exhaust valve closes. This is the time until the temperature of the compression is increased and the time for cooling by the wall surface of the cylinder 18 becomes longer. Therefore, when the engine speed is low, if the negative overlap is performed, the temperature decrease of the burned gas due to the high temperature compression and cooling by the wall surface of the cylinder 18 increases.

  On the other hand, when the engine speed is high, the calorific value of the air-fuel mixture in the cylinder per unit time is large, and the wall surface temperature of the cylinder is relatively high, so the temperature of the burned gas after expansion is relatively high. Therefore, the temperature difference between the burned gas temperature and the exhaust port 17 or the exhaust valve 22 in a state where the compression is not increased is large. Therefore, when the engine speed is high, if the exhaust is opened twice, the temperature drop of the burnt gas due to cooling by the exhaust port 17 and the exhaust valve 22 becomes large. When the engine speed is high, the burned gas reciprocates between the exhaust port and the cylinder at a high speed, so that the burned gas and the exhaust port come into contact with each other over a wider range. Cooling is promoted. On the other hand, when the engine speed is high, the time from the closing of the exhaust valve 22 to the intake top dead center TDC, that is, the time during which the burned gas heated to the wall surface of the cylinder 18 is cooled is short. Therefore, when the engine speed is high, even if the negative overlap is performed, the temperature drop of the burned gas due to cooling of the burned gas compressed at a high temperature by the wall surface of the cylinder 18 is suppressed to a small level.

  Therefore, in the present engine system 100, in order to suppress the decrease in the temperature of burned gas and to effectively increase the temperature of the air-fuel mixture, internal EGR is performed by opening the exhaust twice in the first region A1 where the engine speed is low. In the second region A2 where the engine speed is high, internal EGR is performed by providing a negative overlap period.

  In order to secure a larger amount of internal EGR when performing internal EGR by a method of providing a negative overlap period, it is preferable to make the valve closing timing EVC of the exhaust valve 22 more advanced. On the other hand, in order to secure a larger amount of internal EGR when performing internal EGR by opening the exhaust twice, the valve closing timing EVC of the exhaust valve 22 is set so as to extend the valve opening period during the intake stroke of the exhaust valve 22. It is preferable that the angle is retarded.

  Therefore, in the present embodiment, in order to satisfy this requirement, the opening timing EVO of the exhaust valve 22 in the second region A2 in which the negative overlap period is provided is relatively advanced, and the exhaust is opened twice. The valve opening timing EVO of the exhaust valve 22 in the first region A1 is relatively retarded. The amount of internal EGR is secured in each area.

  Further, in the case of performing internal EGR by opening the exhaust twice, in order to allow the burned gas to properly flow back into the cylinder from the exhaust port 17 side, the valve opening timing of the intake valve 21 is preferably on the retard side. Therefore, in the present embodiment, the intake valve 21 is opened on the retard side from the intake top dead center, and the internal EGR is more reliably ensured when the exhaust is opened twice.

  Further, when the negative overlap period is provided, the retard amount θ2 from the intake top dead center TDC at the valve opening timing IVO of the intake valve 21 is set to the advance angle from the intake top dead center TDC at the valve closing timing EVC of the exhaust valve 22. When the amount is larger than the amount θ1, the burned gas must be expanded more than the compression amount up to the intake top dead center TDC, the pumping loss increases, and the intake top dead at the valve opening timing IVO of the intake valve 21 increases. When the retard amount θ1 from the point TDC is smaller than the advance amount θ1 from the intake top dead center TDC at the closing timing EVC of the exhaust valve 22, it is more than the expansion work from the intake top dead center TDC. The fuel gas must be compressed and the pumping loss increases. Therefore, in the present embodiment, the retard amount θ2 from the intake top dead center TDC of the valve opening timing IVO of the intake valve 21 and the exhaust valve 22 in the second region A2 in which the negative overlap period is provided in order to suppress the pumping loss. The advance amount θ1 from the intake top dead center TDC at the valve closing timing EVC is made the same.

  Further, if the air and the fuel are sufficiently mixed, the air-fuel mixture can be self-ignited appropriately, that is, with high exhaust performance and high thermal efficiency. In the first region A1 and the second region A2, since the engine load is relatively low and the temperature of the air-fuel mixture is not so high, there is no possibility that the air-fuel mixture will ignite prematurely.

  Therefore, in the first region A1 and the second region, fuel is injected during the intake stroke and mixed with air in advance, and the air-fuel mixture is appropriately self-ignited near the compression top dead center.

(3) SI Combustion Area Although the specific control content in the SI combustion area is not particularly limited, in the engine system 100, in the SI combustion area, avoidance of abnormal combustion such as premature ignition and knocking, NOx generation For the purpose of suppression and reduction of cooling loss, high-pressure retarded injection is performed, the internal EGR is stopped, and external EGR for introducing cooled EGR gas into the combustion chamber 19 is performed. In the SI combustion region, the throttle valve 36 is fully opened to reduce the pump loss, and the amount of fresh air introduced into the cylinder 18 is adjusted by adjusting the opening of the EGR valve 511. Further, the intake valve 21 is made a large lift by the intake VVL 54.

  As described above, in the engine system 100, since the internal EGR is performed in the entire CI combustion region where the compression ignition combustion is performed, the temperature of the air-fuel mixture is appropriately increased to realize stable compression ignition combustion. In addition, in the second region A2 set on the high rotation side in the CI combustion region, a negative overlap period is provided across the intake top dead center TDC, while internal EGR is performed, while on the low rotation side Since the internal EGR is performed by opening twice in the set first region A1, the temperature drop of the burned gas is effectively suppressed in each region A1, A2, and the The temperature of the air-fuel mixture can be reliably increased.

  In particular, the exhaust VVL 71 selectively transmits the operating state of the first cam and the second cam to the exhaust valve 22 to switch the lift characteristic of the exhaust valve 22 between the special mode and the normal mode. The valve 22 can be switched between a state where a negative overlap period is provided and a state where the exhaust gas is opened twice, and the structure of the entire system can be simplified.

  Further, the valve opening timing EVO and the valve closing timing EVC of the exhaust valve 22 in the first region A1 in which the low rotation side exhaust double opening is performed are set to the advance side, and the high overlap negative overlap period Since the valve opening timing EVO and the valve closing timing EVC of the exhaust valve 22 in the second region A2 are provided on the retard side, the internal EGR amount can be reliably ensured in each of the regions A1 and A2. .

  Further, the negative overlap period is set so that the advance amount θ1 from the intake top dead center TDC of the exhaust valve 22 and the retard amount θ2 from the intake top dead center of the intake valve 21 are the same. In addition, it is possible to increase the fuel efficiency by suppressing the pumping loss while providing the negative overlap period and securing the internal EGR.

  In addition, this invention is not limited to embodiment mentioned above. For example, when fuel is injected during the intake stroke, the fuel may be injected into the intake port 16 by a port injector provided separately in the intake port 16 instead of the injector 67 provided in the cylinder 18.

  Further, regarding the valve train of the engine 1, instead of the VVL 74 of the intake valve 21, a CVVL (Continuously Variable Valve Lift) capable of continuously changing the lift amount may be provided.

  Further, the high-pressure retarded injection may be divided injection as necessary, and similarly, the intake stroke injection may also be divided injection as necessary. In these divided injections, fuel may be injected in each of the intake stroke and the compression stroke.

  In the above embodiment, the case where the external EGR is stopped in the entire CI combustion region has been described. For example, as shown in FIG. 7, in the region where the engine load is equal to or higher than the specific load T2 in the CI combustion region. External EGR may be performed in addition to internal EGR. In the example illustrated in FIG. 7, the specific load T2 is set to a different value depending on the engine speed, and is set so as to decrease as the engine speed increases from the first load T1.

  Specifically, in the CI fuel region where the engine load is relatively high, the temperature in the combustion chamber 19 tends to be high, and there is a risk that problems such as combustion noise may occur if compression self-ignition ignition combustion is performed. Therefore, in the CI combustion region, in the region where the engine load is equal to or higher than the specific load T2, in addition to the high-temperature internal EGR, the external EGR gas, in particular, the cooled EGR gas that has become low temperature by passing through the EGR cooler 52 is burned. The temperature inside the combustion chamber 19 may be set to an appropriate temperature. In other words, the EGR valve 511 may be opened while the EGR cooler bypass valve 531 is closed.

  Further, when the temperature in the combustion chamber 19 is high, if fuel is injected during the intake stroke, there is a risk of premature ignition. Therefore, in the CI combustion region where the engine load is equal to or greater than the specific load T2, high pressure retarded injection is performed in the combustion chamber 19 by the injector 67 during the period from the late compression stroke to the early expansion stroke to avoid premature ignition. By performing injection, combustion in the expansion stroke period may be realized while forming a homogeneous air-fuel mixture in a relatively short time.

  In the above embodiment, the intake valve 21 has a small lift characteristic in the entire CI combustion region. However, in order to ensure the intake amount more reliably, in the second region A2 on the high rotation side, The intake valve 21 may have a large lift characteristic. Furthermore, as shown in FIG. 9, the intake valve 21 may have a large lift characteristic while the valve opening timing and the valve closing timing may be set to the retard side compared to the case of the small lift characteristic. In this case, it is preferable to advance the exhaust valve 22 in accordance with the retardation of the intake valve 21 in order to suppress the pumping loss. In FIG. 9, the solid line is the valve lift of the intake valve 21 and the exhaust valve 22 corresponding to the large lift characteristic, and the broken line is the valve lift in the case of the small lift characteristic.

1 Engine (Engine body)
10 PCM (control means)
18 cylinder 21 intake valve 22 exhaust valve 71 VVL

Claims (6)

A cylinder in which a combustion chamber in which an air-fuel mixture containing at least fuel and air burns is formed; an intake port for introducing intake air into the cylinder; an exhaust port for exhausting exhaust gas from the cylinder; and the intake port An engine body having an intake valve capable of opening and closing, and an exhaust valve capable of opening and closing the exhaust port;
Control means for controlling the combustion mode of the air-fuel mixture and the open states of the intake valve and the exhaust valve;
The control means includes
At least in the low load region where the engine load is lower than the predetermined load, the combustion mode in the combustion chamber is set to compression auto-ignition combustion in the entire region,
In the compression auto-ignition region where the compression auto-ignition combustion is performed, in the high-rotation side compression auto-ignition region where the engine speed is equal to or higher than a specific speed, the exhaust valve is closed on the advance side from the intake top dead center, In addition, by opening the intake valve on the retard side from the intake top dead center, a negative overlap period in which both the exhaust valve and the intake valve are closed across the intake top dead center is provided in the cylinder. In the low-rotation side compression auto-ignition region where the engine speed is less than the specific speed while the burned gas remains, the exhaust valve is opened at least during the exhaust stroke and the intake stroke and discharged to the exhaust port side. A control apparatus for a compression ignition engine, wherein the burned gas is allowed to flow back into the cylinder to cause the burned gas to remain in the cylinder. The control apparatus for a compression ignition engine according to claim 1,
A first characteristic that closes the lift characteristic of the exhaust valve closer to the advance side than the intake top dead center, and a second characteristic that has a valve opening period longer than the first characteristic and continues to open halfway through the intake stroke. Exhaust valve lift characteristics change means that can be changed to characteristics,
The second characteristic is that after the exhaust valve is opened with a predetermined lift amount during a predetermined first valve opening period, the intake stroke is continued from the end of the first valve opening period continuously after the first valve opening period. The exhaust valve is configured to open with a lift amount smaller than the maximum lift amount during the first valve opening period during the second valve opening period until the predetermined timing of
The control means includes
The exhaust valve lift characteristic changing means is controlled so that the lift characteristic of the exhaust valve is the first characteristic in the high-rotation side compression auto-ignition area, while the second characteristic is in the low-rotation side compression auto-ignition area. A control device for a compression ignition engine characterized by comprising: The control apparatus for a compression ignition engine according to claim 1 or 2,
The control device for a compression ignition engine characterized in that the control means sets the valve opening timing of the intake valve to the retard side from the intake top dead center in the low rotation side compression auto-ignition region. In the control apparatus of the compression ignition type engine according to any one of claims 1 to 3,
An intake valve lift characteristic changing means capable of changing the lift characteristic of the intake valve to a predetermined small lift characteristic and a large lift characteristic in which the maximum lift amount is set larger than the maximum lift amount of the small lift characteristic;
The control means controls the intake valve lift characteristic changing means so that the lift characteristic of the intake valve becomes the large lift characteristic in the high-rotation side compression auto-ignition area, while the low-rotation side compression auto-ignition area Then, a control device for a compression ignition type engine having the small lift characteristic. In the control device for the compression ignition engine according to any one of claims 1 to 4,
The compression ignition is characterized in that the opening timing of the exhaust valve in the high-rotation side compression auto-ignition region is controlled to be advanced from the opening timing of the exhaust valve in the low-rotation side compression auto-ignition region. Type engine control device. In the control device for the compression ignition engine according to any one of claims 1 to 5,
In the high-rotation side compression auto-ignition region, the control means includes a retard amount from the intake top dead center of the intake valve opening timing and an advance angle from the intake top dead center of the exhaust valve closing timing. A control device for a compression ignition engine characterized in that the amount is substantially the same.

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