JP6225699B2 - Control unit for direct injection engine - Google Patents

Description

  The present invention relates to a control device for a direct injection engine.

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

  In compression self-ignition combustion, it is necessary to raise the temperature of the air-fuel mixture in the cylinder to a temperature at which self-ignition is possible. Therefore, normally, when performing compression auto-ignition combustion, the air-fuel mixture is heated by performing internal EGR to leave high-temperature burned gas (internal EGR gas) in the combustion chamber. As shown in Patent Document 1, in the region where this compression auto-ignition combustion is performed, the temperature of the air-fuel mixture tends to be lower as the engine load is lower, so the internal EGR rate is set higher as the engine load is lower. Has been.

JP 2012-172665 A

  When the internal EGR rate is lower on the high load side than on the low load side, such as when the internal EGR rate is set higher as the engine load is lower as in Patent Document 1, the engine is operated between these operating conditions. If the vehicle is decelerated in the direction of decreasing the load, combustion noise may increase.

  Specifically, the higher the engine load, the greater the amount of heat generation, and the higher the exhaust gas temperature and the combustion chamber wall temperature. And when engine load is reduced from operating conditions where engine load is high and these temperatures are high, these temperatures do not fall immediately. Therefore, at the time of deceleration, the temperature of the wall surface of the combustion chamber and the temperature of the internal EGR gas become higher than the temperature corresponding to the operating condition of the deceleration destination, and as a result, the EGR gas having a temperature higher than this appropriate temperature is increased. The combustion noise increases as the temperature and the combustion temperature increase excessively. In particular, when compression auto-ignition combustion is performed at the deceleration destination, it is highly possible that combustion noise increases beyond an allowable range as the combustion temperature rises excessively.

  Further, at the time of deceleration, there is a problem that the air-fuel ratio of the air-fuel mixture deviates from an appropriate value and proper engine performance cannot be obtained.

  The present invention has been made in view of the above points, and more reliably performs combustion while appropriately controlling the air-fuel ratio of the air-fuel mixture during deceleration to an operating condition in which the engine load is low and the internal EGR rate is set high. A control device for a direct injection engine capable of suppressing an increase in noise is provided.

In order to solve the above problems, the present invention includes a fuel injection device capable of injecting fuel into a combustion chamber formed in a cylinder, and reciprocating a piston by combusting a mixture of fuel and air in the combustion chamber. A control device for a direct-injection engine to be moved, a determination unit for determining an operating state of the engine, an internal EGR unit for performing an internal EGR for leaving an internal EGR gas which is a burned gas in the combustion chamber, an engine load When the determination means determines that the engine is operating in a specific high load region that is greater than or equal to a predetermined load, each part of the engine including the fuel injection device is controlled so that the excess air ratio of the air-fuel mixture becomes 1 or less The combustion control means for causing the internal EGR means to perform internal EGR at least at a part on the low load side in the specific high load region, and With decreasing down load, from said first operating condition of a specific high-load region, the internal EGR ratio is a ratio of the internal EGR gas amount to the total gas amount in the combustion chamber at a specific high-load region the At the time of deceleration to shift to the second operating condition set higher than the first operating condition, an amount of fuel smaller than the original injection amount corresponding to the second operating condition is at least one of the intake stroke and the compression stroke. A first injection to be injected, and a second injection for injecting fuel in an amount equal to or greater than an original injection amount corresponding to the second operating condition to an amount obtained by subtracting an injection amount by the first injection in the first half of the expansion stroke, A control device for a direct injection engine characterized in that the injection device is temporarily implemented (claim 1).

  According to the present invention, when decelerating to the second operating condition with a high internal EGR rate, the injection amount of the first injection injected at least one of the compression stroke and the intake stroke is the original injection corresponding to the second operating condition. Because the amount is reduced below the amount, it is generated near the compression top dead center against the excessive temperature rise of the air-fuel mixture accompanying the increase in the internal EGR gas, the follow-up delay of the internal EGR gas temperature and the combustion chamber wall temperature. The amount of heat generation can be reduced to suppress an excessive increase in the combustion temperature, and an increase in combustion noise can be suppressed more reliably. In addition, in the present invention, during the deceleration, the second injection is performed in which the temperature rise due to the thermal energy generated in the first half of the expansion stroke is small, and the injection amount of the second injection is set to the first injection amount. It is set to be more than the reduction amount of the injection amount. For this reason, as described above, while suppressing an increase in combustion noise by reducing the injection amount of the first injection, the total injection amount is set to be equal to or higher than the original total injection amount corresponding to the operating condition of the deceleration destination during deceleration. The excess air ratio can be set to a value of 1 or less set in advance for the operating condition of the deceleration destination, and more appropriate engine performance can be ensured.

  In the present invention, the combustion control means causes the fuel injection device to perform the first injection and the second injection over a plurality of cycles during deceleration when shifting from the first operating condition to the second operating condition. It is preferable to increase the injection amount of the first injection and decrease the injection amount of the second injection as the number of elapsed cycles from immediately after the deceleration increases.

  According to this configuration, the combustion temperature can be suppressed to be low over a plurality of cycles, and an increase in combustion noise can be more reliably suppressed, and the temperature of the air-fuel mixture can be brought to an appropriate temperature in the second operating condition earlier. In addition, since the injection amount of the first injection increases and the injection amount of the second injection decreases as the number of elapsed cycles increases, the transition can be smoothly made depending on the original injection state corresponding to the second operating condition. it can.

  In the present invention, the combustion control means may be configured such that the internal EGR rate is temporarily higher than the value of the main body corresponding to the second operating condition at the time of deceleration when shifting from the first operating condition to the second operating condition. The internal EGR means is controlled so that the value is low, and the injection amount of the second injection is subtracted from the original injection amount corresponding to the second operating condition. The amount is preferably larger than the amount (claim 3).

  In this way, it is possible to suppress an increase in combustion temperature due to an increase in internal EGR rate and introduction of high-temperature internal EGR gas during the deceleration, and more reliably suppress an increase in combustion noise. Can do. Here, when the increase in the internal EGR rate is suppressed, the fresh air amount increases. However, in this configuration, the injection amount of the second injection is made smaller than the decrease amount of the first injection while suppressing the increase in the internal EGR rate. Therefore, the excess air ratio of the air-fuel mixture can be more reliably suppressed to 1 or less.

In the above-described configuration, the combustion control means is configured such that the internal EGR rate is greater than an original value corresponding to the second operating condition over a plurality of cycles during deceleration when shifting from the first operating condition to the second operating condition. The internal EGR means is controlled so as to be a low value, and the fuel injection device performs the first injection and the second injection, and as the number of cycles elapsed immediately after the deceleration increases, It is preferable to increase the internal EGR rate, increase the injection amount of the first injection, and decrease the injection amount of the second injection.

  In this way, at the time of deceleration, the excess air ratio of the air-fuel mixture is suppressed to 1 or less over a plurality of cycles, and the increase in the combustion temperature accompanying the increase in the internal EGR rate is suppressed, thereby further increasing the combustion noise. While being able to suppress reliably, it can transfer to the original internal EGR rate and injection state corresponding to a 2nd driving | running condition smoothly.

  Further, in the present invention, it is possible to change a closing timing of an intake valve provided in the cylinder, and includes an intake valve closing timing changing means controlled by the combustion control means, wherein the combustion control means includes the first At the time of deceleration that shifts from the first operating condition to the second operating condition, the closing timing of the intake valve set by the intake valve closing timing changing means is set from the original closing timing corresponding to the second operating condition. It is also preferable to temporarily advance or retard to reduce the amount of fresh air introduced into the combustion chamber.

  In this way, the amount of fresh air can be suppressed to a small amount immediately after deceleration by changing the closing timing of the intake valve at the time of deceleration, so that the excess air ratio of the air-fuel mixture is more reliably reduced to 1 or less. can do.

  In the above-described configuration, the combustion control means is configured so that the intake valve closing timing corresponds to the second operating condition over a plurality of cycles at the time of deceleration from the first operating condition to the second operating condition. It is preferable to control the intake valve closing timing changing means so that it is on the more advanced side than the above timing, and to decrease the advanced angle amount of the intake valve as the number of elapsed cycles immediately after the deceleration increases. Item 6).

  In this way, during deceleration, the excess air ratio of the mixture can be more reliably suppressed to 1 or less over a plurality of cycles, and the closing timing of the intake valve is the original timing corresponding to the second operating condition. Can be smoothly transitioned to.

Further, the present invention has a fuel injection device capable of injecting fuel into a combustion chamber formed in a cylinder, and controls a direct injection engine that reciprocates a piston by burning a mixture of fuel and air in the combustion chamber. A determination means for determining an operating state of the engine, an internal EGR means for carrying out an internal EGR in which an internal EGR gas which is a burned gas remains in the combustion chamber, and an intake valve provided in the cylinder It is determined by the determination means that the valve closing timing can be changed and the operation is performed in an intake valve closing timing changing means controlled by the combustion control means and a specific high load region where the engine load is equal to or higher than a predetermined load. And a combustion control means for controlling each part of the engine including the fuel injection device and the intake valve closing timing changing means so that the excess air ratio of the air-fuel mixture becomes 1 or less. Means, at least part of the low-load side of the specific high load range, with it perform internal EGR in the internal EGR means, with decreasing engine load, from a first operating condition of the specific high-load region During deceleration when the internal EGR rate, which is the ratio of the internal EGR gas amount to the total gas amount in the combustion chamber within the specific high load region, shifts to the second operating condition set higher than the first operating condition , The fuel injection amount injected by the fuel injection device is temporarily made smaller than the original injection amount corresponding to the second operating condition, and is set by the intake valve closing timing changing means The intake valve closing timing is temporarily advanced or retarded from the original valve closing timing corresponding to the second operating condition to reduce the amount of fresh air introduced into the combustion chamber. To To provide a control device for injection engine (claim 7).

  According to the present invention, at the time of deceleration to the second operating condition with a high internal EGR rate, the injection amount is decreased from the original injection amount corresponding to the second operating condition. The overheated temperature of the air-fuel mixture accompanying the delay in following the EGR gas temperature and the wall surface temperature of the combustion chamber can be resisted, and the overheated temperature of the combustion can be suppressed by reducing the amount of heat generation. Can be more reliably suppressed. In addition, in the present invention, at the time of the deceleration, the intake valve closing timing is advanced or retarded from the original timing, so that the amount of fresh air can be suppressed to a low level. A more appropriate engine performance can be obtained by reliably reducing the value to 1 or less.

  In the present invention, the combustion control means sets the injection amount injected by the fuel injection device to the second operating condition over a plurality of cycles during deceleration when shifting from the first operating condition to the second operating condition. The intake valve closing timing changing means is set to an amount smaller than the corresponding original injection amount, and so that the closing timing of the intake valve is more advanced than the original timing corresponding to the second operating condition. It is preferable to increase the injection amount and decrease the advance amount of the intake valve as the number of elapsed cycles from immediately after the deceleration increases.

  This makes it possible to more reliably suppress the excess air ratio of the air-fuel mixture to 1 or less and more reliably suppress the increase in combustion noise over a plurality of cycles during the deceleration, as well as the injection amount and the intake air. The valve closing timing can be smoothly shifted to the original value corresponding to the second operating condition.

  In the present invention, the combustion control means may temporarily cause the internal EGR rate to be temporarily lower than the original value corresponding to the second operating condition during deceleration when shifting from the first operating condition to the second operating condition. It is preferable to control the internal EGR means so as to have a low value (claim 9).

  In this way, it is possible to suppress an increase in combustion temperature due to an increase in internal EGR rate and introduction of high-temperature internal EGR gas during the deceleration, and more reliably suppress an increase in combustion noise. Can do.

  Further, in the above-described configuration, the combustion control means is configured so that the internal EGR rate corresponds to the second operating condition over a plurality of cycles at the time of deceleration when shifting from the first operating condition to the second operating condition. It is preferable to control the internal EGR means so as to be lower than the value, and to increase the internal EGR rate as the number of elapsed cycles from immediately after the deceleration increases.

  In this way, at the time of deceleration, an increase in the combustion temperature accompanying an increase in the internal EGR rate over a plurality of cycles can be suppressed, and an increase in combustion noise can be suppressed more reliably, and the second operating condition can be satisfied. It is possible to smoothly shift to the corresponding original internal EGR rate.

  Here, the second operating condition includes an operating condition set so that the air-fuel mixture is spark-ignited and combusted by ignition by an ignition means for igniting the air-fuel mixture in the combustion chamber.

  The first operating condition includes an operating condition set so that the air-fuel mixture in the combustion chamber burns by self-ignition (claim 12).

  As described above, according to the present invention, an increase in combustion noise can be more reliably suppressed during deceleration to an operating condition where the engine load is low and the internal EGR rate is set high.

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 ratio of the gas contained in an engine load and air-fuel | gaseous mixture. It is the figure which showed the ratio of the gas contained in an engine load and air-fuel | gaseous mixture. It is a figure for demonstrating the problem which arises at the time of deceleration. It is the figure which showed the injection pattern and heat release rate at the time of the deceleration shown in FIG. It is the figure which showed the change of each parameter at the time of deceleration at the time of implementing the control which concerns on embodiment of this invention. It is the figure which showed the heat release rate at the time of the deceleration shown in FIG. It is the figure which showed the change of each parameter at the time of deceleration at the time of implementing control which concerns on other embodiment of this invention.

  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 to which a control device for a direct injection engine according to the present invention is applied. 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. As will be described later, the engine body 1 is a direct injection engine configured such that fuel can be directly injected into a cylinder 10 (combustion chamber 11) by an injector (fuel injection device) 15. The engine body 1 is an engine that realizes compression self-ignition combustion, as will be described later.

  The engine body 1 has a cylinder block 12 in which a cylinder 10 is formed and a cylinder head 13. In the present embodiment, four cylinders 10 are formed in series with the cylinder block 12. In each cylinder 10, a piston 14 connected to a crankshaft via a connecting rod is fitted so as to be able to reciprocate. In each cylinder 10, a combustion chamber 11 surrounded by the inner surface of the cylinder 10, the crown surface of the piston 14, and the cylinder head 13 is formed.

  Hereinafter, the reciprocating direction of the piston 14 is referred to as the vertical direction, and the side on which the cylinder head 13 is disposed with respect to the cylinder block 12 is referred to as the upper side.

  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.

  The piston 14 and the combustion chamber 11 have a configuration as shown in FIG. A cavity 11 a having a concave shape that is recessed in a direction away from the ceiling of the cylinder head 13 and the combustion chamber 11, that is, in the downward direction, is formed in the center in the radial direction of the piston 14. The cavity 11a has an opening facing the ceiling of the combustion chamber 11, that is, upward, at the upper end thereof. The opening area of the opening of the cavity 11a is set to be smaller than the maximum value of the area of the horizontal cross section at each height position inside the cavity 11a. That is, the cavity 11a is constricted so that the inner diameter becomes narrower in the range from the opening to a predetermined depth.

  An injector 15 that directly injects fuel into the combustion chamber 11 is attached to the cylinder head 13 for each cylinder 10. As shown in FIG. 3, the injector 15 is disposed so as to face the inside of the combustion chamber 11 from the radial center of the ceiling surface of the combustion chamber 11 so that the nozzle hole faces the radial center of the cavity 11 a. . In the present embodiment, the injector 15 is a multi-hole type having a plurality of nozzle holes, and the fuel spray injected from the injector 15 spreads radially downward from the radial center of the combustion chamber 11. The injector 15 injects relatively high-pressure fuel pumped from a fuel tank (not shown) by a fuel supply system (not shown) into the combustion chamber 11.

  A spark plug 16 that forcibly ignites the air-fuel mixture in the combustion chamber 11 is attached to the cylinder head 13 for each cylinder 10. Each spark plug 16 is arranged such that an electrode portion provided at the tip thereof is positioned in the vicinity of the nozzle hole of the injector 15.

  The cylinder head 13 is provided with an intake port 17 for introducing intake air into the cylinder 10 and an exhaust port 18 for discharging exhaust gas from the cylinder 10 for each cylinder 10. The intake port 17 and the exhaust port 18 are respectively provided with these ports 17, 18, and more specifically, an intake valve 21 and an exhaust valve 22 that open and close the openings of the ports 17, 18 formed in the cylinder head 13, respectively. It is installed.

  The exhaust valve 22 is driven by an exhaust valve drive mechanism (internal EGR means) 23. The exhaust valve drive mechanism 23 includes an exhaust valve lift variable mechanism (hereinafter referred to as an exhaust VVL (Variable Valve Lift)) 23a and an exhaust phase variable mechanism (hereinafter referred to as an exhaust VVT (Variable Valve Timing)) 23b.

  The exhaust VVL 23a switches the operation mode of the exhaust valve 22 between a normal mode shown in FIG. 4 (a) and a special mode shown in FIG. 4 (b). In the normal mode, the exhaust valve 22 is opened mainly during the exhaust stroke (the majority of the period from the start of valve opening to the valve closing overlaps with the exhaust stroke). The valve lift of the exhaust valve 22 gradually increases after opening, and when reaching the maximum lift, it gradually decreases again and reaches 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, a lift lower than the maximum lift in the first valve opening period t_1 is maintained to zero after maintaining the second valve opening period t_2. The exhaust VVL 23a 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 shown in FIG. 4A and has one cam crest. The second cam has a shape corresponding to the lift characteristics shown in FIG. 4B, and has two cam peaks. The exhaust VVL 23 a 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 23a is, for example, hydraulically operated.

  The exhaust VVT 23b changes the opening timing and closing timing of the exhaust valve 22 by changing the rotational phase of the exhaust camshaft with respect to the crankshaft. The exhaust VVT 23b changes the valve opening timing and the valve closing timing of the exhaust valve 22 while keeping the valve opening period of the exhaust valve 22 constant in each of the normal mode and the special mode. The exhaust VVT 23b may employ a hydraulic, electromagnetic, or mechanical known structure as appropriate, and a detailed description thereof is omitted.

  The exhaust VVT 23b sets the rotation phase of the exhaust camshaft so that the exhaust valve 22 opens in the intake stroke in addition to the exhaust stroke when the exhaust valve 22 is in the special mode. Further, when the exhaust valve 22 is in the special mode, the exhaust VVT 23b has an exhaust top dead center (TDC in FIG. 4B) during the second valve opening period t_2, that is, the exhaust VVT 23b is exhausted. The rotational phase of the exhaust camshaft is set so that the valve lift of the exhaust valve 22 at the top dead center 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 exhaust top dead center being sandwiched without closing in the middle. Here, when the exhaust valve 22 is continuously opened across the exhaust top dead center, the exhaust valve 22 and the piston 14 may interfere with each other. In contrast, in the present embodiment, as described above, the valve lift amount of the exhaust valve 22 near the exhaust 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 so-called internal EGR by leaving the high-temperature burned gas, that is, the internal EGR gas in the combustion chamber 11. 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 18 during the exhaust stroke is in the combustion chamber 11 during the intake stroke. The exhaust gas, that is, the high-temperature burned gas remains in the combustion chamber 11.

  In this embodiment, the exhaust VVT 23b changes the valve closing timing of the exhaust valve 22 in a state where the exhaust is opened twice, so that the internal EGR gas amount occupying the total gas amount in the combustion chamber 11 can be reduced. The internal EGR rate, which is a percentage, is changed. Specifically, when it is desired to reduce the internal EGR rate, the exhaust VVT 23b advances the valve closing timing of the exhaust valve 22 and closes the exhaust valve 22 at an earlier stage of the intake stroke. Thus, in this embodiment, since the internal EGR rate is changed by changing the valve closing timing of the exhaust valve 22 by the exhaust VVT 23b, the internal EGR rate is changed early.

  The intake valve 21 is driven by an intake valve drive mechanism 24. The intake valve drive mechanism 24 includes an intake VVT (intake valve closing timing changing means) 24b that changes the valve opening timing and the valve closing timing while keeping the valve opening period of the intake valve 21 constant. The intake VVT 24b may adopt a known hydraulic, electromagnetic or mechanical structure as appropriate, and a detailed description thereof will be omitted. The intake valve drive mechanism 24 may further include an intake VVL 24a that can change the valve lift of the intake valve 21.

  An intake passage 30 is connected to each intake port 17. Specifically, branch passages that branch in correspondence with the cylinders 10 are formed at the downstream end of the intake passage 30, and these branch passages are connected to the intake ports 17.

  In the intake passage 30, an air cleaner 31, a throttle valve 32, and a surge tank 33 are arranged in this order from the upstream side.

  An exhaust passage 40 is connected to each exhaust port 18. Specifically, similarly to the intake passage 30, branch passages that branch to correspond to the cylinders 10 are formed at the upstream end of the exhaust passage 40, and these branch passages are connected to the intake ports 17. .

  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 a three-way catalyst.

  Between the intake passage 30 and the exhaust passage 40, an EGR device (external EGR means) 50 for returning a part of the exhaust gas to the intake air, that is, performing external EGR is provided. The EGR device 50 includes an EGR passage 51 and an EGR cooler 52. The EGR passage 51 connects a portion of the intake passage 30 between the surge tank 33 and the throttle valve 32 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. In the present embodiment, the gas passing through the EGR passage 51 is necessarily cooled by the EGR cooler 52. The EGR passage 51 is provided with an EGR valve 53 that adjusts the flow rate of the exhaust gas that passes through the EGR passage 51. 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. The external EGR rate, which is the ratio of the external EGR gas amount to the total gas amount in the combustion chamber 11, is changed by changing the opening degree of the EGR valve 53.

  Each device is controlled by a powertrain control module (control means, hereinafter referred to as PCM) 60. The PCM 60 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> 11 are input to the PCM 60.

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 that detects the temperature of fresh air. The air flow sensor SW1 and the intake air temperature sensor SW2 are disposed on the downstream side of the air cleaner 31 in the intake passage 30. The sensor SW3 is an EGR gas temperature sensor for detecting the temperature of the external EGR gas. The EGR gas temperature sensor SW3 is disposed in the vicinity of the connection portion with the intake passage 30 in the EGR passage 51. The sensor SW4 is an intake port temperature sensor that detects the temperature of intake air immediately before flowing into the cylinder 10. The intake port temperature sensor SW4 is attached to the intake port 17. The sensor SW5 is an exhaust temperature sensor that detects the exhaust temperature. The sensor SW6 is an exhaust pressure sensor that detects the exhaust pressure. The exhaust temperature sensor SW5 and the exhaust pressure sensor SW6 are disposed in the vicinity of the connection portion of the EGR passage 51 in the exhaust passage 40. The sensor SW7 is a linear O ₂ sensor that detects the oxygen concentration in the exhaust gas. The linear O ₂ sensor SW 7 is disposed on the upstream side of the direct catalyst 41 in the exhaust passage 40. The sensor SW8 is a lambda O ₂ sensor that detects the oxygen concentration in the exhaust gas. The lambda O ₂ sensor SW8 is disposed between the direct catalyst 41 and the underfoot catalyst 42. The sensor SW9 is a water temperature sensor that detects the temperature of the engine cooling water. The sensor SW10 is a crank angle sensor that detects the rotation angle of the crankshaft. The sensor SW11 is an accelerator opening sensor that detects an accelerator opening corresponding to an operation amount of an accelerator pedal (not shown) of the vehicle.

  The PCM 60 performs various calculations based on the detection signals of the sensors SW1 to SW11. The PCM 60 determines the operating conditions of the engine body 1 and the vehicle based on these detection signals. The PCM 60 outputs control signals to the actuators of the injector 15, spark plug 16, fuel supply system, intake VVT 24b, intake VVL 24a, exhaust VVT 23b, exhaust VVL 23a, and various valves (throttle valve 32, EGR valve 53) according to operating conditions. And control these. As described above, in this embodiment, the PCM 60 functions as a determination unit that determines the operating state of the engine based on the engine speed, the engine load, and the like, and controls the injector 15 and the like, and the injection amount injected from the injector 15. Combustion control means that changes the injection timing and injection pattern, controls the exhaust VVT 23b to change the internal EGR rate and the excess air ratio of the air-fuel mixture, and controls the intake VVT 24b to change the valve closing timing IVC of the intake valve 21 Function as.

  The contents of control by the PCM 60 will be described next.

  FIG. 5 is a control map of the engine speed on the horizontal axis and the engine load on the vertical axis. When compression auto-ignition combustion is performed in an SI (Spark Ignition) combustion region in which the engine load is equal to or higher than a preset combustion switching load T1, combustion noise becomes a problem. In the engine system 100, in this SI combustion region, Spark ignition combustion is performed in which the air-fuel mixture is burned by ignition by the spark plug 16. On the other hand, in the CI (Compression Ignition) combustion region (compression ignition combustion region) in which the engine load that suppresses the combustion noise is set to be less than the combustion switching load T1, the air-fuel mixture is not ignited by the spark plug 16. Compressed self-ignition combustion is performed by self-ignition and combustion. In the SI combustion region, fuel injection is performed in the compression stroke, and the air-fuel mixture burns by being ignited near the compression top dead center. On the other hand, in the CI combustion region, fuel injection is performed in the intake stroke, and an air-fuel mixture in which fuel and air are sufficiently mixed self-ignites near the compression top dead center.

In a high load region (specific high load region) in which the engine load is equal to or higher than the preset stoichiometric switching load T2, it is difficult to sufficiently reduce the amount of NOx generated by combustion. Therefore, in the engine system 100, the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio, that is, the excess air ratio λ of the air-fuel mixture is set to λ = 1 in the region of the stoichiometric switching load T2 or higher. On the other hand, in the region below the stoichiometric switching load T2, the air-fuel ratio of the mixture is made leaner than the stoichiometric air-fuel ratio, that is, the excess air ratio λ of the mixture is λ> 1.

  In the present embodiment, the stoichiometric switching load T2 is set smaller than the combustion switching load T1. Therefore, the excess air ratio λ of the air-fuel mixture is λ = 1 in the third CI combustion region CI_3 that is the region on the high load side of the CI combustion region in addition to the entire SI combustion region.

(1) EGR control in the CI combustion region In order to achieve stable compression auto-ignition combustion in the CI combustion region, it is necessary to raise the temperature of the air-fuel mixture to a temperature at which self-ignition is possible. Therefore, in the CI combustion region, internal EGR is performed to cause the internal EGR gas, which is a high-temperature burned gas, to remain in the combustion chamber 11, thereby increasing the temperature in the combustion chamber 11 and the air-fuel mixture. Specifically, in the CI combustion region, the exhaust state is opened twice with the operating state of the exhaust valve 22 as a special mode.

  However, in the operating region on the high load side in which the engine load is higher than the EGR switching load T3 in the CI combustion region, the temperature of the air-fuel mixture is sufficiently high, and heat is generated abruptly when the internal EGR gas is excessively introduced. The maximum value of the rate of occurrence is excessive, and combustion noise may exceed the allowable range. Therefore, in the present embodiment, only the internal EGR is performed in the first CI combustion region CI_1 in which the engine load is less than the EGR switching load T3. On the other hand, in the region where the engine load is set to be equal to or greater than the EGR switching load T3 and less than the combustion switching load T1, external EGR is performed in addition to the internal EGR, and the external EGR gas cooled by the EGR cooler 52 is introduced. Suppresses an excessive increase in the temperature of the gas mixture. That is, in an area where the engine load is set to be equal to or greater than the EGR switching load T3 and less than the combustion switching load T1, the exhaust valve 22 is opened twice with the operating state of the exhaust valve 22 being in a special mode and the EGR valve 53 is opened. .

  In the present embodiment, the EGR switching load T3 is set to a value smaller than the stoichiometric switching load T2. Therefore, in the second CI region CI_2 in which the engine load is set to be equal to or greater than the EGR switching load T3 and less than the stoichiometric switching load T2, the internal EGR and the external EGR are performed, and the excess air ratio λ of the mixture is λ> 1 Compressed auto-ignition combustion is performed while being lean. In the third CI region CI_3 in which the engine load is set to be not less than the stoichiometric switching load T2 and less than the combustion switching load T1, the internal EGR and the external EGR are performed and the excess air ratio λ of the air-fuel mixture is λ = 1. Compressed self-ignition combustion is performed while being performed.

  In the CI combustion region, the internal EGR rate and the external EGR rate, which is the ratio of the external EGR gas amount in the air-fuel mixture, are controlled to appropriate values according to the engine load. Changes in the gas ratio with respect to the engine load at the engine speed NE1 in FIG. 5 are shown in FIGS. FIG. 6 shows the ratio of internal EGR gas, external EGR gas, and fresh air, where the horizontal axis is the engine load and the vertical axis is the maximum amount of gas that can be introduced into the combustion chamber 11. is there. On the other hand, in FIG. 7, the horizontal axis represents the engine load, and the vertical axis represents the ratio of these gases to the total amount of gas introduced into the combustion chamber 11, that is, the total amount of internal EGR gas, external EGR gas, and fresh air. It is shown. Here, the internal EGR rate and the external EGR rate are the ratios of the respective EGR gases to the total gas amount introduced into the combustion chamber 11, and are values shown in FIG.

  As shown in FIGS. 6 and 7, the internal EGR rate is increased as the engine load is reduced throughout the CI combustion region. The lower the engine load, the lower the temperature of the mixture. Therefore, in this way, if the proportion of the hot internal EGR gas is increased as the engine load is lower, the temperature of the air-fuel mixture can be increased to ensure good ignitability. In the present embodiment, the internal EGR rate is set to 0 at the combustion switching load T1, and is increased from the load T1 to the specific load T10 at a substantially constant rate with respect to a decrease in the engine load. In the region where the engine load is lower than the specific load T10, the internal EGR rate is kept constant at a high value.

  The external EGR rate is maintained substantially constant with respect to the engine load in the third CI region CI_3 in which the air-fuel ratio of the air-fuel mixture is lean while the external EGR is introduced. In the second CI area CI_2 where the engine load is lower than that in the third CI area CI_3, the external EGR rate is increased at a constant rate from 0 to the value of the external EGR rate in the third CI area CI_3. .

(2) EGR control in the SI combustion region In the SI combustion region, the operating state of the exhaust valve 22 is set to the normal mode, and the internal EGR is stopped. On the other hand, in the SI combustion region, external EGR is performed in other regions except for the full load.

  As shown in FIG. 7, in the SI combustion region, the external EGR rate is set to be a smaller value when the engine load is higher. In the present embodiment, when the engine load is lower than the specific load T20, the external EGR rate is kept constant with respect to the engine load. When the engine load is higher than the load T20, the engine load increases. Is reduced at a constant rate.

  Here, as described above, in the SI combustion region, the excess air ratio λ of the mixture is set to λ = 1, and the air-fuel ratio of the mixture is set to the stoichiometric air-fuel ratio. For this reason, in the SI combustion region where the external EGR rate is lower than the specific load T20 and is maintained almost constant regardless of the engine load, the total gas amount is reduced according to the decrease in the engine load, as shown in FIG. Reduce the amount of fresh air to an appropriate value according to the load. Specifically, the fresh air amount and the total gas amount are reduced by throttle the throttle valve 32.

  The above control is designed so that the combustion state and the engine performance are suitable in a state where the steady operation is performed under each operation condition and the wall surface temperature of the combustion chamber 11 is stable. The lower the engine load, the lower the amount of heat generated in the combustion chamber 11, and the lower the engine load, the lower the wall surface temperature of the combustion chamber 11 during steady operation. Therefore, when the engine is decelerated to a low engine load and the operating condition of the deceleration destination is a condition in which the value during steady operation of the internal EGR rate is higher than that before deceleration, the control during steady operation is left as it is. If implemented, there is a problem that the combustion temperature becomes excessive and the combustion noise increases because the wall surface temperature of the combustion chamber 11 does not immediately decrease to the temperature at the time of steady operation under the operating conditions of the deceleration destination. In particular, when the deceleration destination is a compression auto-ignition region where compression auto-ignition combustion is performed, there is a high possibility that the combustion noise exceeds the allowable range.

  Here, as a method of suppressing the increase in the combustion temperature and combustion noise, for example, a method of simply reducing the injection amount during deceleration is conceivable. However, in this method, the air-fuel ratio of the air-fuel mixture becomes lean, that is, the excess air ratio becomes 1 or more. Therefore, in a region where the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio or rich, that is, the air excess ratio of the air-fuel mixture is set to 1 or less, appropriate engine performance cannot be obtained as the air-fuel ratio deviates from an appropriate value.

  From this, the present inventors examined the deceleration in the region where the excess air ratio of the air-fuel mixture is set to 1 or less, particularly at the time of deceleration to an operating condition with a higher internal EGR rate. Has invented an apparatus capable of suppressing combustion noise while setting the excess air ratio of the air-fuel mixture to an appropriate value.

  First, the problem will be described with reference to FIG.

  FIG. 8 shows that at time t1, the internal EGR rate during steady operation is low and the air-fuel mixture has a high internal EGR rate during steady operation from the operation condition (first operation condition) where the excess air ratio of the mixture is set to 1 or less. When the control during normal operation is performed as it is at the time of deceleration to shift to the operation condition (second operation condition) in which the excess air ratio is 1 or less and compression self-ignition combustion is performed, that is, the change in the operation condition The time change of each parameter when the control is simply switched in response is shown. In FIG. 8, the operating condition before deceleration is a point X1 in the SI region shown in FIG. 5, the excess air ratio λ of the air-fuel mixture during steady operation is set to λ = 1, and the internal EGR rate is 0, the external EGR rate is set to a value greater than 0, and the spark ignition combustion is performed. The post-deceleration operation condition is a point X2 in the third CI region CI_3 shown in FIG. The excess air ratio λ during operation is set to λ = 1, the external EGR rate is larger than the value at the point X1, and the internal EGR rate is set to a value larger than 0, so that the compression ignition combustion is performed. The case where the operating conditions were implemented was shown. In FIG. 8, in order from the top, the injection amount, the injection pulse, the EGR rate, the wall temperature of the combustion chamber (combustion chamber wall temperature), the pre-ignition temperature, that is, the temperature of the air-fuel mixture in the combustion chamber at a predetermined timing before ignition, and the combustion temperature (Maximum temperature during combustion) and changes in engine torque. In FIG. 8, the change in the internal EGR rate and the change in the external EGR rate are shown together as the change in the EGR rate. In the injection pulse graph, the vertical axis represents the crank angle and the hatched portion represents the injection pulse. That is, the leading edge of the hatched portion is the injection start timing, and the retarding edge is the injection stop timing, and the vertical length of this portion represents the pulse width. Therefore, the larger the pulse width in the vertical direction, the larger the injection amount.

  In FIG. 8, when the vehicle is decelerated at time t1, the injection amount is decreased and the internal EGR rate is increased. As described above, in this embodiment, the internal EGR rate is changed early by changing the closing timing of the exhaust valve 22 by the exhaust VVT 23b. Therefore, the internal EGR rate is increased immediately after deceleration. The wall surface temperature of the combustion chamber 11 decreases only slowly after deceleration, and becomes a temperature higher than the appropriate temperature (steady-state temperature) in the steady state of the operating conditions of the deceleration destination for a while after deceleration. Thus, since the wall surface temperature of the combustion chamber 11 is higher than the appropriate temperature, the temperature before ignition becomes higher after the deceleration than the appropriate temperature (steady-state temperature) in the steady state of the operating conditions of the deceleration destination. Furthermore, as described above, since the internal EGR rate increases immediately after deceleration and high-temperature internal EGR gas is introduced, the pre-ignition temperature becomes higher than the pre-ignition temperature before deceleration. Here, the operating condition before deceleration is a condition where the engine load is high and the exhaust temperature is high. Therefore, as the internal EGR rate increases, a large amount of this high-temperature internal EGR gas is introduced, so that the pre-ignition temperature becomes particularly high. In the example shown in FIG. 8, the external EGR rate is the total EGR rate (external EGR gas and internal EGR gas occupying the total gas amount in the combustion chamber) due to the delay in driving the EGR valve 53 and the exhaust gas recirculation delay. The deviation from the appropriate value of the combined exhaust ratio) is shown by covering the internal EGR gas by increasing the internal EGR gas. The internal EGR gas becomes higher than the appropriate value immediately after deceleration, and this also causes the pre-ignition temperature. Becomes hot.

  Thus, the temperature before ignition is excessively increased to a value higher than the steady-state temperature at which the vehicle is decelerated and the temperature before deceleration, and the wall surface temperature of the combustion chamber is high and heat loss to the wall surface of the combustion chamber is small. As a result, the combustion temperature also rises excessively. Specifically, for a while after deceleration, the combustion temperature is higher than the steady-state temperature of the deceleration destination and the temperature before deceleration, and further, the combustion noise upper limit temperature that is the upper limit temperature on the combustion noise where combustion noise increases. As a result, the combustion noise increases beyond the allowable range for a while after deceleration.

  Further, as the ignition delay is shortened and the heat loss is reduced due to the excessive temperature rise before ignition, the engine torque also becomes a torque that is equal to or higher than the torque in the steady operation of the deceleration destination, that is, the target torque. Therefore, there also arises a problem that an appropriate feeling of deceleration cannot be obtained.

  FIG. 9 shows an injection pulse and a heat generation rate at time t1 (immediately after deceleration) in the example shown in FIG. In FIG. 9, the heat generation rate in the example shown in FIG. 8 is indicated by a solid line, and the heat generation rate at the steady state under the operating condition of the deceleration destination, that is, the appropriate heat generation rate is indicated by a broken line. As shown in FIG. 9, in the example shown in FIG. 8, immediately after deceleration, the heat generation rate rises very rapidly and rises to a very high value compared to the steady state, and the temperature before ignition is high. The air-fuel mixture injected and premixed in the intake stroke burns violently near the compression top dead center. As shown in FIG. 9, in the case of the example shown in FIG. 8, that is, when the control during steady operation is performed as it is, the maximum value of the heat generation rate becomes very high and the combustion noise increases. Further, the heat generation amount (integral amount of the heat generation rate) is increased, and the center of gravity of the heat generation rate is advanced, so that an engine torque higher than an appropriate torque is generated.

  In order to solve the above problem, in the engine system 100, the engine load is low due to the operating condition in which the engine load is high and the internal EGR rate is set low and the excess air ratio of the mixture is set to 1 or less. At the time of deceleration where the internal EGR rate is set high and the excess air ratio of the air-fuel mixture is set to the operating condition set to 1 or less, the original injection amount corresponding to the operating condition of the deceleration destination, that is, the operating condition of the deceleration destination A first injection that injects an amount of fuel that is smaller than an injection amount during steady operation (hereinafter sometimes referred to simply as a steady-state injection amount) before the compression top dead center, and an injection amount of the first injection. The second injection in which the amount subtracted from the constant injection amount, that is, the amount of fuel equal to or greater than these differences, is injected in the first half of the expansion stroke. Further, in the present embodiment, the increase in the internal EGR rate is delayed, that is, the internal EGR rate corresponding to the deceleration destination operating condition, that is, the internal EGR rate during steady operation under the deceleration destination operating condition (hereinafter referred to as the internal EGR rate). In some cases, the internal EGR rate may be simply referred to as a steady-state internal EGR rate), and the injection amount of the second injection is made larger than the difference in accordance with the delay processing of the internal EGR rate. Further, in this embodiment, the fuel injection control and the delay control of the internal EGR rate are performed over a plurality of cycles after deceleration. At this time, the injection amount of the first injection increases toward the steady-state injection amount as the number of cycles elapsed immediately after deceleration increases, and the injection amount of the second injection decreases toward zero. It decreases as the number of elapsed cycles from immediately after increases. Further, as the number of elapsed cycles increases, the injection start timing of the second injection is delayed. The injection start timing of the first injection is the same as that during steady operation.

  In the present embodiment, the amount of reduction of the injection amount relative to the steady injection amount of the first injection is an amount that does not generate combustion noise based on the engine speed, engine load, combustion chamber wall temperature, pre-ignition temperature, and the like. Thus, the air-fuel ratio of the air-fuel mixture is set to be rich, that is, the excess air ratio is 1 or less. The wall surface temperature and the pre-ignition temperature of the combustion chamber may be estimated from the engine cooling water temperature, the intake air temperature, the exhaust gas temperature, and the like, the engine speed, the engine load, and the like.

  Further, in the present embodiment, the internal EGR rate is realized by increasing by a predetermined value for each cycle from the value during steady operation before deceleration to the value during steady operation before deceleration.

  FIG. 10 shows temporal changes in the injection amount, EGR rate, wall surface temperature of the combustion chamber 11 (combustion chamber wall surface temperature), and pre-ignition temperature when the above control is performed. FIG. 10 shows changes under the same deceleration conditions as in the example shown in FIG. 8, and changes when the above control is performed during deceleration from the point X1 in the SI area to the point X2 in the third CI area CI_3 shown in FIG. It is. In FIG. 10, the corresponding changes in FIG. 8 are indicated by broken lines.

  In FIG. 10, the first injection and the second injection are performed along with the deceleration at time t1. In this example, the deceleration destination is the compression auto-ignition region, and collective injection is performed in the intake stroke at the steady state. Therefore, the first injection is injected in the intake stroke as in the steady operation. Further, the internal EGR rate is set to a value substantially before deceleration (0 in this example) at time t1 by performing the delay process, and is sufficiently lower than the steady-state internal EGR rate. At this time, the injection amount of the first injection is set to be smaller than the steady-state injection amount of the deceleration destination, and the second injection amount is set to an amount larger than the decrease amount. Along with this, the total injection amount becomes larger than the total injection amount during steady operation (total injection amount during steady state). In the present embodiment, at time t1, the injection amount of the first injection is made smaller than the injection amount of the second injection.

  Thus, while the internal EGR rate is set to a value lower than the steady-state internal EGR rate, the excess air ratio λ of the air-fuel mixture is maintained at 1 at time t1 as the total injection amount increases. . That is, the amount of fresh air increases by reducing the internal EGR rate, but the total injection amount is increased, so the excess air ratio λ of the air-fuel mixture is maintained at 1.

  As described above, since the increase in the internal EGR rate is suppressed to substantially zero, the temperature before ignition is suppressed to approximately the same as the temperature before deceleration at time t1. Along with this, the fuel injected by the first injection near the compression top dead center burns slowly, and the combustion temperature is substantially the same as the temperature during steady operation before deceleration and the deceleration noise upper limit temperature. It can be held at a lower temperature. And a combustion noise is suppressed in an allowable range. In addition, as described above, the heat loss to the combustion chamber wall surface is reduced with the delay in following the wall surface temperature of the combustion chamber 11, but the injection amount of the first injection injected in the intake stroke is sufficiently reduced. And, by realizing the gradual combustion, the engine torque is suppressed to a sufficiently small value and decreases to the torque during steady operation at the deceleration destination.

  FIG. 11 shows an injection pulse and a heat generation rate at time t1 (immediately after deceleration) in the example shown in FIG. As shown in FIG. 11, heat generation occurs moderately immediately after deceleration.

  Here, as described above, at time t1, in addition to the first injection in the intake stroke, the second injection is performed in the expansion stroke, but the fuel injected in the second injection is more than the compression top dead center. It burns at a sufficiently retarded timing. For this reason, the maximum value of the combustion temperature is determined by the combustion generated by the first injection that burns near the compression top dead center, and the combustion temperature that affects the combustion noise because the injection amount of the first injection is suppressed to be small. It can be kept low. In addition, the amount of combustion energy generated by the fuel injected by the second injection and generated at a timing that is sufficiently retarded from the compression top dead center is converted into an effective engine torque. As described above, the engine torque can be kept small by reducing the injection amount of the first injection.

  After time t1, the injection amount of the first injection is gradually increased at a constant injection start timing, the injection amount of the second injection is gradually decreased at a constant injection end timing, and the total injection amount is gradually decreased. Also, the internal EGR rate is gradually increased. Thus, since the decrease in the total injection amount and the increase in the internal EGR rate are performed at the same time, the excess air ratio λ of the air-fuel mixture is maintained at 1 even after time t1. In addition, although the internal EGR rate is gradually increased and the injection amount of the first injection is gradually increased, for a while after time t1, the internal EGR rate is set to a value smaller than the steady-state internal EGR rate. The injection amount for one injection is set to be smaller than the steady-state injection amount. Therefore, even after time t1, the wall temperature of the combustion chamber 11 is higher than the temperature during steady operation, but the pre-ignition temperature and combustion temperature are kept lower than the values in FIG. 8 and the combustion noise upper limit temperature. The combustion noise and engine torque are within the proper ranges. In addition, since the combustion temperature in the previous cycle is suppressed to be low, the wall surface temperature of the combustion chamber 11 and the temperature of the internal EGR gas are decreased, and thereby the pre-ignition temperature and the combustion temperature are also suppressed to a low level.

  In this way, the operating condition is shifted to the low load side while avoiding the generation of combustion noise and excessive engine torque. Then, at time t2, the injection pattern and the injection amount are substantially the same as the temperature during steady operation when the wall surface temperature of the combustion chamber 11 is the temperature during steady operation, and the steady operation pattern (collective injection during the intake stroke) and steady injection. The internal EGR rate is the steady-state internal EGR rate. Here, as described above, in this control, as the combustion temperature is kept low, the wall surface temperature of the combustion chamber 11 is promoted to be lowered. Therefore, the wall surface temperature of the combustion chamber 11 is as shown in FIG. It is possible to lower the temperature to an appropriate steady state temperature at an earlier time t2, and it is possible to shift to an appropriate control at the time of steady operation earlier.

  Further, as described above, as the number of cycles is increased, the injection amount of the first injection is increased, the injection amount of the second injection is decreased, and the internal EGR rate is increased. Until the shift to the hour pattern or the like, it is possible to smoothly shift to the upper body during steady operation while suppressing the occurrence of torque fluctuation or the like.

  As described above, in the engine system 100 according to the present embodiment, the internal EGR rate is reduced from the operating condition in which the air excess ratio λ of the air-fuel mixture is set within the region where the air excess rate λ is set to 1 or less and the internal EGR rate is low. When decelerating to higher operating conditions, it is possible to suppress an excessive increase in combustion temperature and suppress an increase in combustion noise. Further, the engine torque can be reduced to the deceleration destination value earlier, and high driving operability can be obtained.

  Here, in the said embodiment, the control which implements 1st injection and 2nd injection, and the control which suppresses the increase in an internal EGR rate by making an internal EGR rate into a value lower than a steady-state internal EGR rate were implemented over multiple cycles. Although the case has been described, these controls may be performed immediately after deceleration. However, the implementation over a plurality of cycles can more reliably suppress an increase in combustion noise, for example, when the follow-up delay of the wall surface temperature of the combustion chamber 11 is long.

  Moreover, although the said embodiment demonstrated the case where the control which implements 1st injection and 2nd injection, and the control which suppresses the increase in an internal EGR rate were performed simultaneously, the control which suppresses the increase in an internal EGR rate is demonstrated. May be omitted. In this case, since an increase in the amount of fresh air that accompanies an increase in the internal EGR rate is avoided, the injection amount of the second injection is the difference between the injection amount of the first injection and the steady-state injection amount. What is necessary is just to set equally. However, it is possible to more reliably suppress an increase in combustion noise by performing these controls simultaneously.

  Moreover, although the said embodiment demonstrated the time of the deceleration from SI combustion area | region, the driving | running condition before deceleration is not restricted to this. Further, the deceleration destination is not limited to the CI combustion region. However, in the CI combustion region, the combustion noise tends to exceed the allowable range as the wall surface temperature of the combustion chamber 11 rises. Therefore, when applied to deceleration to the CI combustion region, the increase in the combustion noise is effectively suppressed. be able to.

  In the above example, the case where the excess air ratio λ of the air-fuel mixture is set to 1 before and after deceleration is described. However, the excess air ratio λ of the air-fuel mixture is set to less than 1 before and after deceleration or after deceleration. May be. Even in this case, by increasing the second injection amount, it is possible to make the engine torque and the combustion noise within appropriate ranges while making the excess air ratio λ less than 1.

  Further, at the time of deceleration, that is, deceleration within a region where the excess air ratio λ of the air-fuel mixture is set to 1 or less, the deceleration that shifts from an operating condition with a low internal EGR rate to an operating condition with a high internal EGR rate. At this time, the intake air VVT 23b causes the fresh air amount to be reduced from the original timing corresponding to the operating condition of the deceleration destination, that is, the timing of steady operation of the operating condition of the deceleration destination, by the closing timing IVC of the intake valve 21. You may temporarily advance or retard.

  As described above, if the valve closing timing IVC of the intake valve 21 is temporarily advanced or retarded to reduce the amount of fresh air, the excess air ratio λ of the air-fuel mixture can be more reliably suppressed to 1 or less. Further, if the amount of fresh air can be reduced, the fuel excess performance can be improved by reducing the amount of injection into the combustion chamber 11 while keeping the excess air ratio λ to 1 or less. For example, the injection amount of the second injection can be made smaller than when the closing timing IVC of the intake valve 21 is maintained at the original timing.

  Furthermore, during deceleration in the region where the excess air ratio λ of the air-fuel mixture is set to 1 or less as described above, and when shifting from an operating condition with a low internal EGR rate to an operating condition with a high internal EGR rate, By the intake VVT 23b, the closing timing IVC of the intake valve 21 is advanced or retarded from the timing at the time of steady operation under the operating conditions of the deceleration destination (hereinafter sometimes referred to as the steady timing) to reduce the amount of fresh air. In that case, the second injection may be stopped.

  For example, as shown in FIG. 12, at the time of deceleration, the valve closing timing IVC of the intake valve 21 is temporarily advanced from the steady-state timing so that the injection amount of the first injection is less than the steady-state injection amount, Further, the second injection may be stopped while the internal EGR rate is delayed. Even in this case, an increase in combustion noise can be suppressed by reducing the injection amount of the first injection and delaying the internal EGR rate, as in the above embodiment. Further, the excess air ratio λ of the air-fuel mixture can be suppressed to 1 or less due to a decrease in the amount of fresh air that accompanies the advance of the valve closing timing IVC of the intake valve 21.

  In addition, as shown in FIG. 12, at time t1 (immediately after deceleration), the closing timing IVC of the intake valve 21 is advanced from the steady-state timing, and toward the steady-state timing for a while after time t1. By gradually retarding the angle, it is possible to avoid the occurrence of a torque shock or the like until the steady time, and to more smoothly shift to the steady time.

  In FIG. 12, the fresh air amount may be decreased by retarding the closing timing IVC of the intake valve 21 from the steady timing.

  Further, advance or retard control of the closing timing IVC of the intake valve 21 may be performed only at time t1 (immediately after deceleration).

  Further, the advance control of the valve closing timing IVC of the intake valve 21 and the delay processing of the internal EGR rate are not used together, and only the advance or delay of the valve closing timing IVC of the intake valve 21 and the injection amount reduction control are performed. You may implement.

1 Engine (Engine body)
10 cylinders 11 combustion chambers 15 injectors (fuel injection devices)
23 Exhaust valve drive mechanism (internal EGR means)
60 PCM (combustion control means)

Claims (12)

A control device for a direct injection engine having a fuel injection device capable of injecting fuel into a combustion chamber formed in a cylinder, and reciprocating a piston by burning a mixture of fuel and air in the combustion chamber;
Determining means for determining the operating state of the engine;
An internal EGR means for performing an internal EGR that leaves an internal EGR gas that is a burned gas in the combustion chamber;
Each part of the engine including the fuel injection device so that the excess air ratio of the air-fuel mixture becomes 1 or less when it is determined by the determination means that the engine load is an operation in a specific high load region of a predetermined load or more. Combustion control means for controlling
The combustion control means includes
In the specific high load region, at least part of the low load side, the internal EGR means performs internal EGR,
With the reduction in the engine load, from said first operating condition of a specific high-load region, the internal EGR ratio is a ratio of the internal EGR gas amount to the total gas amount in the combustion chamber at a specific high-load region the At the time of deceleration to shift to the second operating condition set higher than the first operating condition, an amount of fuel smaller than the original injection amount corresponding to the second operating condition is at least one of the intake stroke and the compression stroke. A first injection to be injected, and a second injection for injecting fuel in an amount equal to or greater than an original injection amount corresponding to the second operating condition to an amount obtained by subtracting an injection amount by the first injection in the first half of the expansion stroke, A control apparatus for a direct injection engine, wherein the injection apparatus is temporarily implemented. The direct injection engine control device according to claim 1,
The combustion control means causes the fuel injection device to perform the first injection and the second injection over a plurality of cycles during deceleration when shifting from the first operating condition to the second operating condition. The control apparatus for a direct injection engine, characterized by increasing the injection amount of the first injection and decreasing the injection amount of the second injection as the number of elapsed cycles from immediately after the deceleration increases. A control device for a direct injection engine according to claim 1 or 2,
The combustion control means causes the internal EGR rate to be temporarily lower than the value of the main body corresponding to the second operating condition during deceleration when shifting from the first operating condition to the second operating condition. As described above, the internal EGR means is controlled, and the injection amount of the second injection is made larger than the amount obtained by subtracting the injection amount of the first injection from the original injection amount corresponding to the second operating condition. A control device for a direct injection engine. A control device for a direct injection engine according to claim 3,
The combustion control means is configured such that the internal EGR rate is lower than an original value corresponding to the second operating condition over a plurality of cycles at the time of deceleration when shifting from the first operating condition to the second operating condition. The internal EGR means is controlled so that the fuel injection device performs the first injection and the second injection, and as the number of cycles elapsed immediately after the deceleration increases, the internal EGR rate increases. the high, by increasing the injection quantity of the first injection, and the control apparatus for a direct injection engine, characterized in that to reduce the injection amount of the second injection. It is a control apparatus of the direct-injection engine in any one of Claims 1-4,
An intake valve closing timing changing means that can change the closing timing of the intake valve provided in the cylinder and is controlled by the combustion control means;
The combustion control means sets the intake valve closing timing set by the intake valve closing timing changing means to the second operating condition during deceleration when shifting from the first operating condition to the second operating condition. A control device for a direct injection engine, wherein the amount of fresh air introduced into the combustion chamber is reduced by temporarily advancing or retarding the corresponding original valve closing timing. A control device for a direct injection engine according to any one of claims 1 to 5,
The combustion control means is configured such that at the time of deceleration from the first operating condition to the second operating condition, the closing timing of the intake valve is more than the original timing corresponding to the second operating condition over a plurality of cycles. The direct valve engine is characterized in that the intake valve closing timing changing means is controlled so as to be on the advance side, and the advance amount of the intake valve is reduced as the number of cycles elapsed immediately after the deceleration increases. Control device. A control device for a direct injection engine having a fuel injection device capable of injecting fuel into a combustion chamber formed in a cylinder, and reciprocating a piston by burning a mixture of fuel and air in the combustion chamber;
Determining means for determining the operating state of the engine;
An internal EGR means for performing an internal EGR that leaves an internal EGR gas that is a burned gas in the combustion chamber;
An intake valve closing timing changing means that can change the closing timing of the intake valve provided in the cylinder and is controlled by the combustion control means;
When the determination means determines that the engine load is operating in a specific high load region where the engine load is equal to or greater than a predetermined load, the fuel injection device and the intake valve closing are performed so that the excess air ratio of the air-fuel mixture becomes 1 or less. Combustion control means for controlling each part of the engine including the timing change means,
The combustion control means includes
In the specific high load region, at least part of the low load side, the internal EGR means performs internal EGR,
With the reduction in the engine load, from said first operating condition of a specific high-load region, the internal EGR ratio is a ratio of the internal EGR gas amount to the total gas amount in the combustion chamber at a specific high-load region the At the time of deceleration that shifts to the second operating condition set higher than the first operating condition, the amount of fuel injected by the fuel injection device is set to be larger than the original injection amount corresponding to the second operating condition. The intake valve closing timing set by the intake valve closing timing changing means is temporarily reduced and temporarily advanced or delayed from the original valve closing timing corresponding to the second operating condition. An apparatus for controlling a direct injection engine, characterized in that the amount of fresh air introduced into the combustion chamber is reduced by being horned. A control device for a direct injection engine according to claim 7,
The combustion control means is configured to reduce the amount of fuel injected by the fuel injection device over a plurality of cycles at the time of deceleration shifting from the first operating condition to the second operating condition. The intake valve closing timing changing means is controlled so that the amount is smaller than the injection amount, and the closing timing of the intake valve is advanced from the original timing corresponding to the second operating condition, The control device for a direct injection engine, wherein the injection amount is increased and the advance amount of the intake valve is decreased as the number of elapsed cycles from immediately after the deceleration increases. A control device for a direct injection engine according to claim 7 or 8,
The combustion control means is configured such that the internal EGR rate is temporarily lower than an original value corresponding to the second operating condition when decelerating from the first operating condition to the second operating condition. Thus, the control device for the direct injection engine controls the internal EGR means. A control device for a direct injection engine according to claim 9,
The combustion control means is configured such that the internal EGR rate is lower than an original value corresponding to the second operating condition over a plurality of cycles at the time of deceleration when shifting from the first operating condition to the second operating condition. The control device for a direct injection engine, wherein the internal EGR means is controlled so that the internal EGR rate is increased as the number of elapsed cycles immediately after the deceleration is increased. A control device for a direct injection engine according to any one of claims 1 to 10,
Ignition means for igniting the air-fuel mixture in the combustion chamber;
The control apparatus for a direct injection engine, wherein the second operating condition is an operating condition set so that the air-fuel mixture is spark-ignited and combusted by ignition by the ignition means. A control device for a direct injection engine according to any one of claims 1 to 11,
The control device for a direct injection engine, wherein the first operating condition is an operating condition set so that the air-fuel mixture in the combustion chamber burns by self-ignition.

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