JP2015124738A - Control device of direct injection engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To surely suppress increase in combustion noise in deceleration to an operation condition where an inside EGR rate of a compression self-ignition combustion region is high.SOLUTION: A control device of a direct injection engine includes a fuel injection device 15 capable of directly injecting fuel into a combustion chamber 11, and make an injection amount injected by the fuel injection device 15 smaller than an original injection amount corresponding to a second operation condition after deceleration, in deceleration of shifting from a first condition where an engine load is higher than a predetermined load to the second condition where the engine load is lower than the predetermined load, a hot EGR rate is higher than that in the first operation condition, and compression self-ignition is executed.

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 EGR to leave high-temperature burned gas in the combustion chamber. As shown in Patent Document 1, in the region where the compression ignition combustion is performed, the temperature of the air-fuel mixture tends to be lower as the engine load is lower. Therefore, the EGR rate is set higher as the engine load is lower. ing.

JP 2012-172665 A

  When the engine load is lower as in Patent Document 1, the EGR rate, that is, the ratio of EGR gas that is high-temperature burned gas to the total gas amount in the combustion chamber is set higher, and the higher the load side is, the lower the load is. When the EGR rate is lower than that on the load side, the combustion noise may increase when the engine is decelerated in the direction of lowering the engine load between these operating conditions and when compression auto-ignition combustion is performed at the deceleration destination. .

  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 wall temperature of the combustion chamber and the temperature of the EGR gas become higher than the temperature corresponding to the operation condition of the deceleration destination, and the EGR gas having a temperature higher than the appropriate temperature is increased. The temperature and combustion temperature are excessively raised, and combustion noise increases. In particular, when compression auto-ignition combustion is performed at the deceleration destination, the combustion noise is likely to increase beyond the allowable range as the combustion temperature rises excessively.

  The present invention has been made in view of the above points, and provides a control device for a direct injection engine that can more reliably suppress an increase in combustion noise when decelerating to an operating condition in which the EGR rate in a compression auto-ignition combustion region is high. provide.

  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 that determines an operating state of the engine, and an EGR unit that performs EGR to mix the burned gas into the air-fuel mixture without forcibly cooling the burned gas; When the determination means determines that the operation is in the compression auto-ignition combustion region set in the operation region including the low load region of the engine, the EGR device is driven and maintained at the high temperature. While mixing the hot EGR gas, which is a fuel gas, in the mixture, the mixture of fuel and air injected from the fuel injection device is combusted by self-ignition in the combustion chamber, thereby realizing compression self-ignition combustion. The combustion control means, and the combustion control means has a hot EGR rate that is a ratio of the hot EGR gas amount in the total gas amount in the combustion chamber during operation in the compression auto-ignition combustion region. The EGR means is controlled so as to increase as the engine speed decreases, and the hot EGR rate is increased in the compression auto-ignition combustion region from the first operating condition higher than a predetermined load as the engine load decreases. At the time of deceleration to shift to the second operating condition set higher than the operating condition, the amount of fuel injected by the fuel injection device is temporarily set to be smaller than the original injection amount corresponding to the second operating condition. There is provided a control device for a direct injection engine characterized in that it is reduced (claim 1).

  According to the present invention, at the time of deceleration to the second operating condition where the EGR rate is set high and compression auto-ignition combustion is performed, the injection amount is reduced below the original amount corresponding to the second operating condition. Therefore, against the excessive temperature rise of the air-fuel mixture accompanying the increase in the hot EGR gas, the temperature of the EGR gas and the wall temperature of the combustion chamber, the heat generation amount is suppressed to a low level and the combustion temperature is overheated. It is possible to suppress the increase in combustion noise more reliably. Moreover, since the combustion chamber wall temperature and the temperature of the EGR gas can be suppressed low by suppressing the combustion temperature, the temperature state and thus the temperature of the air-fuel mixture is set to an appropriate temperature in the second operating condition earlier. More appropriate compression auto-ignition combustion can be realized.

  In the present invention, the combustion control means is configured to reduce the injection amount over a plurality of cycles from the original injection amount corresponding to the second operating condition during deceleration when shifting from the first operating condition to the second operating condition. It is preferable to increase the injection amount as the number of elapsed cycles from immediately after the deceleration increases.

  According to this configuration, since the injection amount is reduced over a plurality of cycles, the combustion temperature can be suppressed low over a plurality of cycles, an increase in combustion noise can be suppressed more reliably, and the temperature of the air-fuel mixture can be controlled more quickly in the second operation. Appropriate compression auto-ignition combustion can be realized at an early stage by setting an appropriate temperature under conditions. Further, since the injection amount increases as the number of elapsed cycles increases, it is possible to smoothly shift by the original injection amount corresponding to the second operating condition.

  In the present invention, the combustion control means is such that the EGR rate is temporarily lower than the original value corresponding to the second operating condition at the time of deceleration when shifting from the first operating condition to the second operating condition. It is preferable to control the EGR means so as to be a value (Claim 3).

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

  In the above configuration, the combustion control means is configured such that the 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. It is preferable to control the EGR means so that the EGR rate increases as the number of elapsed cycles from immediately after the deceleration increases.

  In this way, the EGR rate is suppressed to be small over a plurality of cycles, and the introduction of high-temperature EGR gas is suppressed, whereby the combustion temperature is suppressed to a low level, so that an increase in combustion noise can be suppressed more reliably and earlier. Appropriate compression auto-ignition combustion can be realized earlier by setting the temperature of the air-fuel mixture to an appropriate temperature under the second operating condition. In addition, since the EGR rate increases as the number of elapsed cycles increases, the transition can be smoothly made with the original EGR rate corresponding to the second operating condition.

The present invention also includes a fuel injection device capable of injecting fuel into a combustion chamber formed in a cylinder, and a direct injection engine that reciprocates a piston by burning a mixture of fuel and air in the combustion chamber. A control device for determining the operating state of the engine; EGR means for performing EGR to mix the burned gas into the mixture in the combustion chamber without forcibly cooling the burned gas;
When the determination means determines that the operation is in the compression auto-ignition combustion region set in the operation region including the low load region of the engine, the burned fuel that is maintained at the high temperature by driving the EGR unit While mixing hot EGR gas, which is a gas, in the mixture, the mixture of fuel and air injected from the fuel injection device is combusted by self-ignition in the combustion chamber to perform compression auto-ignition combustion Combustion control means, and the combustion control means has a hot EGR rate that is a ratio of the hot EGR gas amount to the total gas amount in the combustion chamber during operation in the compression auto-ignition combustion region. The EGR means is controlled so as to decrease as the engine speed decreases, and as the engine load decreases, from the first operating condition higher than a predetermined load, the The hot EGR rate is temporarily lower than the original value corresponding to the second operating condition at the time of deceleration when the EGR rate shifts to the second operating condition set higher than the first operating condition. A control device for a direct injection engine characterized by controlling the EGR means (claim 5).

  According to this device, since the EGR rate is reduced from the original value corresponding to the second operating condition at the time of deceleration to the second operating condition where the EGR rate is high and compression auto-ignition combustion is performed, EGR Against the delay in following the temperature of the gas and the wall surface temperature of the combustion chamber, the temperature of the air-fuel mixture can be kept low to suppress overheating of the combustion temperature, and the increase in combustion noise can be suppressed more reliably. it can. Moreover, since the combustion chamber wall temperature and the temperature of the EGR gas can be suppressed low by suppressing the combustion temperature, the temperature state and thus the temperature of the air-fuel mixture is set to an appropriate temperature in the second operating condition earlier. More appropriate compression auto-ignition combustion can be realized.

  Here, when the EGR means performs the internal EGR in which the burned gas is mixed into the mixture in the combustion chamber by causing the burned gas to remain in the combustion chamber, the temperature of the EGR gas is high, and the EGR gas has a high temperature during deceleration. Combustion noise may increase as the hot EGR gas increases. For this reason, if the present invention is applied to the case where the EGR means is applied to the case where internal EGR is performed in which the burned gas remains in the combustion chamber to mix the burned gas into the air-fuel mixture in the combustion chamber, The increase in noise can be more effectively suppressed (claim 6).

  In particular, when decelerating from an operating condition in which the excess air ratio of the air-fuel mixture is set to 1, that is, an operating condition in which the combustion temperature is high such that the exhaust gas generated by combustion needs to be purified by a three-way catalyst, Combustion noise is likely to occur as the combustion temperature rises excessively. Therefore, if this invention is applied to such an operating condition, it will be more effective (Claim 7).

  As described above, according to the present invention, it is possible to more reliably suppress an increase in combustion noise during deceleration to an operating condition in which the EGR rate in the compression auto-ignition combustion region is 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 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 the deceleration at the time of implementing the control which concerns on this embodiment. It is the figure which showed the heat release rate at the time of the deceleration shown in FIG.

  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, 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 rotation phase of the exhaust camshaft relative to the crankshaft 15 to change the opening timing and closing timing of the exhaust valve 22. 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 specific configuration of the intake valve drive mechanism 24b is not particularly limited. For example, an intake VVL 24a that can change the valve lift of the intake valve 21 and an intake VVT 24b that changes the opening timing and closing timing of the intake valve 21 are provided. Including.

  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, like the intake passage 30, branch passages branching corresponding 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 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 a three-way catalyst.

  Between the intake passage 30 and the exhaust passage 40, there is provided an external EGR device 50 for returning a part of the exhaust gas to the intake air, that is, for performing external EGR. External 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 forcibly cooled by the EGR cooler 52 without fail. Thus, in the present embodiment, the exhaust valve drive mechanism 23 functions as an EGR means for mixing the EGR gas, which is a burned gas, into the air-fuel mixture in the combustion chamber at the temperature without forcibly cooling it. The EGR device does not function as this EGR means. In this embodiment, the hot EGR gas, which is a burned gas mixed in the air-fuel mixture while being maintained at a high temperature by the EGR means, is composed only of the internal EGR gas, and is a hot occupying the total gas amount in the combustion chamber. The hot EGR rate, which is the ratio of the EGR gas amount, is equivalent to the internal EGR rate.

  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 external EGR device 50 is referred to as external EGR, and the exhaust gas recirculated to the intake air by the external EGR device 50 may be referred to as external EGR gas.

  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. Thus, in this embodiment, the PCM 60 controls the determination means for determining the operating state of the engine from the engine speed, the engine load, and the like, the injector 15, the spark plug 16, and the like in the combustion chamber 11. The combustion mode of the air-fuel mixture is controlled, and, as will be described later, compression self-ignition combustion is realized in the CI combustion region, and also functions as combustion control means for controlling the exhaust VVT 23b and changing the internal EGR rate.

  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 where the engine load is equal to or higher than a 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.

  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, which is the proportion of the internal EGR gas amount in the air-fuel mixture, and the external EGR rate, which is the proportion of the external EGR gas amount in the air-fuel mixture, are controlled to appropriate values according to the engine load. The 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, if the ratio of the high-temperature internal EGR gas is increased as the engine load is lower in this way, 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 increases from this load T1 to the specific load T10 almost in proportion to the 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 in proportion to the engine load from 0 toward 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 a smaller value as the engine load is higher. In the present embodiment, in a region where the engine load is lower than the specific load T20, the external EGR rate is kept constant with respect to the engine load. In a region where the engine load is higher than the load T20, the engine load is proportional to the engine load. Will be reduced.

  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. Here, since the amount of heat generated in the combustion chamber 11 decreases as the engine load decreases, the wall surface temperature of the combustion chamber 11 during steady operation decreases. Therefore, when the engine is decelerating 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 the deceleration and compression auto-ignition combustion is performed. If the control during the steady operation is performed as it is, the temperature of the wall of the combustion chamber 11 does not immediately decrease to the temperature during the steady operation under the operation condition of the deceleration destination, and the combustion temperature becomes excessive, resulting in combustion noise. There is a problem that increases.

  This point will be described in detail with reference to FIG.

  FIG. 8 shows an operation condition (second operation condition) in which compression self-ignition combustion is performed at a time t1 from an operation condition (first operation condition) with a low internal EGR rate during steady operation to a high internal EGR rate during steady operation. When the control during steady operation is performed as it is at the time of deceleration to shift to (1), that is, the time change of each parameter when the control is simply switched according to the change of the operation condition is shown. In FIG. 8, as an example, the operating condition before deceleration is a point X1 in the SI region shown in FIGS. 5, 6 and 7, the air-fuel ratio of the air-fuel mixture during steady operation is the stoichiometric air-fuel ratio, 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 first CI region shown in FIGS. Thus, the air-fuel ratio of the air-fuel mixture during steady operation is leaner than the stoichiometric air-fuel ratio, the external EGR rate is set to 0, and the internal EGR rate is set to a value greater than 0, and the compression autoignition combustion is performed. Shows the case. FIG. 8 shows, in order from the top, the injection amount, the EGR rate, the temperature of the combustion chamber wall (combustion chamber wall temperature), the temperature before combustion, that is, the temperature of the air-fuel mixture in the combustion chamber at a predetermined timing before ignition, Shows the change 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 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 becomes an appropriate value set for the operating condition of the deceleration destination immediately after the 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 the 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 high engine load and exhaust temperature is high. Therefore, as the internal EGR rate increases, the pre-ignition temperature becomes higher by introducing a large amount of this high-temperature internal EGR gas. Here, the external EGR rate remains at the drive delay of the EGR valve 53 or on the intake side, although the EGR valve 53 is closed at time t1 along with the deceleration so as to become 0 as the target value of the deceleration destination. By introducing the external EGR gas, after the time t1, it gradually decreases. For this reason, the increase in the pre-ignition temperature is suppressed to some extent due to the delay of the external EGR gas, but the influence of the wall surface temperature of the combustion chamber and the internal EGR rate is strong, and the pre-ignition temperature excessively increases.

  In this way, the pre-ignition temperature is excessively increased to a value higher than the steady-state operation temperature at the deceleration destination and the temperature before deceleration, and the wall surface temperature of the combustion chamber is high, resulting in heat loss to the combustion chamber wall surface. As the temperature decreases, the combustion temperature also increases excessively. Specifically, for a while after deceleration, the combustion temperature becomes higher than the steady-state operation temperature of the deceleration destination and the temperature before deceleration, and further, the combustion noise upper limit temperature at which the combustion noise is above the allowable range, For a while after deceleration, the combustion noise increases beyond the allowable range.

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

  FIG. 9 shows the heat generation rate at time t1 (immediately after deceleration) in the example shown in FIG. FIG. 9 also shows an injection pattern. In FIG. 9, the solid line indicates the heat generation rate immediately after the deceleration, and the broken line indicates the heat generation rate during steady operation at the deceleration destination. As shown in FIG. 9, when the steady state control shown in FIG. 8 is performed as it is, the pre-ignition temperature is high immediately after deceleration, so that the premixed air-fuel mixture injected in the intake stroke is compressed. It burns violently near the dead center, and the combustion noise increases accordingly. Immediately after deceleration, the amount of heat generation (integrated amount of heat generation rate) increases and the center of gravity of the heat generation rate advances to generate high engine torque.

  In order to solve the above problem, in the present engine system 100, an operation in which the engine load is set low and the internal EGR rate is set high from the operating condition where the engine load is high and the internal EGR rate is low, and compression auto-ignition combustion is performed. When decelerating to the condition, the injection amount is temporarily larger than the original amount corresponding to the operation condition of the deceleration destination, that is, the injection amount in the steady operation of the operation condition of the deceleration destination (hereinafter sometimes referred to as the steady-state injection amount). Reduce it. Furthermore, the increase in the internal EGR rate is delayed. That is, the internal EGR rate is temporarily lower than the original value corresponding to the deceleration destination operating condition, that is, the internal EGR rate during steady operation under the deceleration destination operating condition (hereinafter sometimes referred to as the steady-state internal EGR rate). To do. And this injection quantity reduction | decrease control is implemented over several cycles after deceleration. At this time, the injection amount increases toward the amount during steady operation as the number of cycles elapsed immediately after deceleration increases. The injection is performed in the intake stroke as in the steady operation, and the injection start timing is the same as in the steady operation.

  In the present embodiment, the amount of reduction in the injection amount relative to the amount during steady operation is set to a value that does not generate combustion noise based on the engine speed, engine load, combustion chamber wall temperature, pre-ignition temperature, and the like. The reduced injection amount is preferably set so that the excess air ratio λ of the air-fuel mixture is λ> 2 at least immediately after deceleration in order to suppress the generation of NOx and maintain good exhaust performance. . 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.

  In the present embodiment, the delay in the increase of the internal EGR rate is realized by performing so-called annealing processing.

  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 first CI area shown in FIG. is there. In FIG. 10, the corresponding changes in FIG. 8 are indicated by broken lines.

  In FIG. 10, the injection amount decreases with the deceleration at time t1. At this time, the injection amount is reduced from the steady-state injection amount. The internal EGR rate is set to a value before deceleration (0 in this example) that is lower than the steady-state internal EGR rate at time t1 by performing delay processing.

  Thus, the increase in the internal EGR rate is suppressed to substantially zero, so that the temperature before ignition is suppressed to substantially the same as the temperature before deceleration at time t1. As a result, more gradual combustion is realized, and the combustion temperature is almost the same as the temperature before deceleration, and is suppressed to a temperature lower than the combustion noise upper limit temperature at which combustion noise increases. It is suppressed in. Further, as described above, the heat loss to the combustion chamber wall surface decreases with the delay in following the wall surface temperature of the combustion chamber 11, but the injection amount is sufficiently reduced and gradual combustion is realized. As a result, the engine torque is suppressed to a sufficiently small value, and the torque is reduced to the torque during steady operation at the deceleration destination. FIG. 11 shows the 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. In addition, the injection pattern is shown in FIG.

  After time t1, the injection amount and the internal EGR rate gradually increase, but these values are set to values smaller than the values during steady operation (steady-state injection amount, steady-state internal EGR rate). Although the wall surface temperature of the combustion chamber 11 is higher than the steady-state temperature, the pre-ignition temperature and the combustion temperature are kept lower than the values in FIG. 8 and the combustion noise upper limit temperature, and the combustion noise and the engine torque are appropriate. Scope. 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 amount is set to the original injection amount, that is, the injection amount in the steady operation almost at the same time when the wall surface temperature of the combustion chamber 11 is set to the temperature in the steady operation. 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. As compared with the above, it is possible to lower the temperature to an appropriate steady state at an early stage, and it is possible to shift to an appropriate control at the time of steady operation at an earlier stage. In the present embodiment, the internal EGR rate is set to an original value, that is, a value during steady operation earlier than time t2.

  In particular, as described above, since the injection amount and the internal EGR rate are increased as the number of cycles is increased, torque fluctuations and the like are suppressed until these values are shifted to the values during steady operation. However, it is possible to smoothly shift to the state during steady operation.

  As described above, in the engine system 100 according to the present embodiment, the combustion temperature is excessively increased at the time of deceleration from the operating condition with the low internal EGR rate to the operating condition with the high internal EGR rate and performing the compression ignition combustion. The increase in combustion noise can be suppressed by suppressing the temperature. Further, the engine torque can be reduced to the deceleration destination value earlier, and high driving operability can be obtained.

  Here, although the case where the control for reducing the injection amount and the control for suppressing the increase in the internal EGR rate are performed over a plurality of cycles has been described in the above embodiment, these controls may be performed only 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 reduces injection amount and the control which suppresses the increase in an internal EGR rate were performed simultaneously, you may implement only these one control. 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 deceleration conditions are not restricted to this. That is, the deceleration is limited to the deceleration from the SI combustion region at the time of deceleration when the operation condition shifts from the high load side operation condition with a low internal EGR rate to the low load side operation condition with a high internal EGR rate and compression auto-ignition combustion is performed. Therefore, it may be applied at the time of deceleration because combustion noise may increase.

  In other words, it is referred to as “at the time of deceleration when the engine load decreases and the first operating condition higher than the predetermined load shifts to the second operating condition in which the internal EGR is performed in the CI combustion region”. The “predetermined load” in the definition may be a load in which the internal EGR rate is set to a value that is somewhat low (including zero), and the specific value can be set as appropriate. When the SI combustion region is set on the high load side of the CI combustion region as in the embodiment, the “predetermined load” may be a load (value less than T1) included in the CI combustion region, It may be a load (a value greater than or equal to T1) included in the SI combustion region on the higher load side than the CI combustion region.

  However, particularly when the vehicle is decelerated from an operating condition where the combustion temperature is high so that the exhaust gas generated by combustion needs to be purified with a three-way catalyst, combustion noise is likely to occur due to an excessive increase in the combustion temperature. Therefore, it is effective if applied during such deceleration.

  In the above-described embodiment, the exhaust valve driving mechanism 23 performs the internal EGR by opening the exhaust gas twice. However, the means for performing the internal EGR is not limited thereto. Moreover, although the case where EGR gas which is burned gas is mixed into the air-fuel mixture in the combustion chamber at high temperature without forcibly cooling by the implementation of internal EGR has been shown, the exhaust gas which is burnt gas is forced By recirculating to the intake air without cooling, it may be mixed into the air-fuel mixture at a high temperature. For example, in the above-described embodiment, the high-temperature burned gas may be mixed into the air-fuel mixture while keeping the temperature high without being cooled by the EGR cooler 52.

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

Claims (7)

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;
EGR means for performing EGR to mix the burned gas into the mixture in the combustion chamber without forcibly cooling the burned gas;
When the determination means determines that the operation is in the compression auto-ignition combustion region set in the operation region including the low load region of the engine, the burned fuel that is maintained at the high temperature by driving the EGR unit While mixing hot EGR gas, which is a gas, in the mixture, the mixture of fuel and air injected from the fuel injection device is combusted by self-ignition in the combustion chamber to perform compression auto-ignition combustion Combustion control means,
The combustion control means includes
The EGR means is controlled so that the hot EGR rate, which is the ratio of the hot EGR gas amount to the total gas amount in the combustion chamber, increases as the engine load decreases during operation in the compression auto-ignition combustion region. With
As the engine load decreases, the first operating condition higher than the predetermined load shifts to the second operating condition in which the hot EGR rate is set higher than the first operating condition in the compression ignition combustion region. A control apparatus for a direct injection engine, characterized in that, during deceleration, an injection amount of fuel injected by the fuel injection device is temporarily reduced below an original injection amount corresponding to the second operating condition. The direct injection engine control device according to claim 1,
The combustion control means reduces the injection amount over a plurality of cycles from the original injection amount corresponding to the second operating condition during deceleration when shifting from the first operating condition to the second operating condition, The control apparatus for a direct injection engine, wherein the injection amount is increased as the number of elapsed cycles 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 hot rate to be temporarily lower than an original value corresponding to the second operating condition at the time of deceleration when shifting from the first operating condition to the second operating condition. And a control device for the direct injection engine, which controls the EGR means. A control device for a direct injection engine according to claim 3,
In the combustion control means, the hot EGR rate becomes a value 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. Thus, the control device for the direct injection engine controls the EGR means so that the hot EGR rate increases as the number of elapsed cycles from immediately after the deceleration increases. 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;
EGR means for performing EGR to mix the burned gas into the mixture in the combustion chamber without forcibly cooling the burned gas;
When the determination means determines that the operation is in the compression auto-ignition combustion region set in the operation region including the low load region of the engine, the burned fuel that is maintained at the high temperature by driving the EGR unit While mixing hot EGR gas, which is a gas, in the mixture, the mixture of fuel and air injected from the fuel injection device is combusted by self-ignition in the combustion chamber to perform compression auto-ignition combustion Combustion control means,
The combustion control means includes
The EGR means is controlled so that the hot EGR rate, which is the ratio of the hot EGR gas amount to the total gas amount in the combustion chamber, increases as the engine load decreases during operation in the compression auto-ignition combustion region. With
As the engine load decreases, the first operating condition higher than the predetermined load shifts to the second operating condition in which the hot EGR rate is set higher than the first operating condition in the compression ignition combustion region. A control apparatus for a direct injection engine, wherein the EGR means is controlled so that the hot EGR rate is temporarily lower than an original value corresponding to the second operating condition during deceleration. A control device for a direct injection engine according to any one of claims 1 to 5,
The control device for a direct injection engine, wherein the EGR means performs an internal EGR in which the burned gas remains in the combustion chamber to mix the burned gas into the air-fuel mixture in the combustion chamber. A control device for a direct injection engine according to any one of claims 1 to 6,
The control device for a direct injection engine, wherein the first operating condition is an operating condition in which an air excess ratio of the air-fuel mixture is set to 1.

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