JP6244882B2 - 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

  As described above, when the internal EGR rate is set higher as the engine load is lower in the compression ignition combustion, the operating condition is changed from the operating condition where the engine load is low and the internal EGR rate is high to the operating condition on the high load side where the internal EGR rate is low. As a result, the amount of internal EGR gas, which is a high-temperature burned gas, decreases, and the wall surface temperature in the combustion chamber does not immediately rise to a temperature suitable for operating conditions on the high load side. However, it is not possible to raise the temperature appropriately to a temperature at which self-ignition is possible, and there is a problem that misfire occurs.

  The present invention has been made in view of the above points, and can realize more stable and stable compression auto-ignition combustion at the time of acceleration to an operating condition with a low internal EGR rate in the compression auto-ignition combustion region. Provided is a control device for a direct injection engine.

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 that moves, a determination unit that determines an operating state of the engine, an internal EGR unit that performs an internal EGR that causes an internal EGR gas that is a burned gas to remain in the combustion chamber, and at least the engine During operation in a compression auto-ignition combustion region where the load is set to be lower than a predetermined load, the internal EGR means is driven to perform internal EGR, while the fuel and air injected from the fuel injection device Combustion control means for burning the air-fuel mixture by self-ignition in the combustion chamber and performing compression auto-ignition combustion, and in the compression auto-ignition combustion region, the total gas amount in the combustion chamber Internal EGR rate is the percentage of Mel the internal EGR gas amount is set so that the engine load becomes higher side lower, the combustion control means, during normal operation in the compression ignition combustion region, the fuel The fuel is injected into the injection device during the intake stroke, and the second operation in the compression auto-ignition combustion region in which the internal EGR rate is set lower than the first operation condition from the first operation condition in the compression auto-ignition combustion region . When accelerating to the condition, the fuel injection device is subjected to main injection for injecting fuel into the combustion chamber before compression top dead center and post injection for injecting fuel into the combustion chamber during the expansion stroke. And the total injection amount including the injection amount of the main injection and the injection amount of the post injection is made larger than the total injection amount during steady operation under the second operating condition. To provide a control apparatus for an engine (claim 1).

  According to the present invention, even at the time of acceleration when shifting to the second operating condition with a low internal EGR rate, even when the mixture temperature decreases due to the decrease in the internal EGR rate and the follow-up delay of the wall surface temperature of the combustion chamber occurs. Thus, it is possible to realize more stable and stable compression auto-ignition combustion.

  Specifically, in the present invention, at the time of acceleration, the total injection amount is made larger than the total injection amount during steady operation to increase the thermal energy generated in the combustion chamber more than during steady operation. This thermal energy compensates for the temperature drop of the air-fuel mixture accompanying the decrease in the internal EGR rate and the delay in tracking the wall temperature of the combustion chamber, so that the temperature of the air-fuel mixture can be raised, and stable self-ignition combustion that is stable early after acceleration Can be realized. In particular, since the post injection is performed during the expansion stroke, it is possible to avoid the occurrence of excessive engine torque due to excessive effective engine torque and the occurrence of torque shock while increasing the total injection amount, as well as the exhaust temperature and the internal EGR gas. The ignitability can be further enhanced by raising the temperature of the slag.

  In the present invention, the combustion control means performs the post-injection over a plurality of cycles during acceleration from the first operating condition to the second operating condition, and as the number of elapsed cycles from immediately after the acceleration increases, It is preferable that the injection amount of the main injection is increased while the injection amount of the post injection is decreased, and the injection timing of the post injection is retarded (Claim 2).

  In this way, it is possible to shift to an appropriate steady state while avoiding the occurrence of torque shock during acceleration. Specifically, the combustion energy generated by the post-injection is easily converted into a more effective engine torque as the wall surface temperature of the combustion chamber and the temperature of the air-fuel mixture rise until the transition to the steady state. To go. In contrast, in this configuration, the post-injection amount is decreased and the injection timing is retarded as the elapsed cycle immediately after acceleration increases and the wall temperature of the combustion chamber rises. Since the combustion energy generated by the injection is suppressed from being converted into an effective engine torque, it is more certain that an effective engine torque will be generated excessively and a torque shock will occur before the steady state is reached. Can be avoided.

  In the present invention, at the time of acceleration when the combustion control means shifts from the first operating condition to the second operating condition, the internal EGR means corresponds to the second operating condition immediately after the acceleration. It is preferable to control the internal EGR means so as to obtain a low value.

  As described above, according to the present invention, stable compression auto-ignition combustion can be realized even during acceleration accompanied by a decrease in the internal EGR rate. Therefore, as in this configuration, at the time of acceleration when shifting from the first operating condition to the second operating condition, the internal EGR means is controlled so that the internal EGR rate becomes the value set in the second operating condition immediately after the acceleration. For example, the internal EGR rate can be quickly set to an appropriate value set in the second operating condition without complicating the control of the internal EGR.

  In the above configuration, the internal EGR means drives at least one of the exhaust valve and the intake valve, and the combustion control means changes the opening / closing timing of at least one of the exhaust valve and the intake valve. It is preferable to change the internal EGR rate by changing the rate.

  In this way, the internal EGR rate can be changed earlier.

  The present invention includes an external EGR means for recirculating the exhaust gas discharged from the combustion chamber to the intake side, and the combustion control means is provided when the engine load of the second operating condition is equal to or higher than a preset reference load. Preferably, during acceleration to shift to the second operating condition, the exhaust gas is recirculated to the intake side by driving the external EGR means while reducing the internal EGR rate (Claim 5).

  That is, in the present invention, as described above, the main injection and the post-injection performed during the expansion stroke are performed during acceleration, and heat generation can be divided and generated to keep the combustion temperature low. Therefore, even when the exhaust gas recirculated to the intake side by the external EGR means is not introduced into the cylinder immediately after acceleration due to a response delay, it is possible to suppress an excessive increase in combustion temperature and combustion noise. it can.

  As described above, according to the present invention, it is possible to realize more reliable and stable compression auto-ignition combustion at the time of acceleration when shifting to an operating condition where the internal EGR rate in the compression auto-ignition combustion region is low.

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 relationship between an engine load and an EGR rate. It is the figure which showed the injection pattern of intake stroke batch injection. It is a figure for demonstrating the problem which arises at the time of acceleration. It is a figure for demonstrating the problem which arises at the time of acceleration. (A) It is the figure which showed the injection pattern which concerns on embodiment of this invention. (B) It is the figure which showed the injection pattern which concerns on embodiment of this invention. It is the figure which showed the change of each parameter at the time of acceleration at the time of implementing control which concerns on embodiment of this invention. It is the figure which showed the change of each parameter at the time of acceleration at the time of implementing control which concerns on 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). At this time, the valve lift of the exhaust valve 22 gradually increases after the valve is opened, 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. Further, the air-fuel ratio of the air-fuel mixture is implemented by changing the internal EGR gas rate so that the pumping loss can be kept small and the air-fuel ratio of the air-fuel mixture can be changed earlier. Specifically, the closing timing of the exhaust valve 22 is changed by the exhaust VVT 23b, whereby the air-fuel ratio of the air-fuel mixture is controlled appropriately.

  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, 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.

  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. In this embodiment, the PCM 60 changes the internal EGR rate by controlling the exhaust VVT 23b and the exhaust VVL 23a by determining means for determining the operating state of the engine from the engine speed, the engine load, etc., and the injector 15 and spark plug 16 Are controlled to control the combustion mode of the air-fuel mixture in the combustion chamber 11 and function as combustion control means for realizing compression auto-ignition combustion in the CI combustion region, as will be described later. The PCM 60 also controls the EGR valve 53 to change the external 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. In the engine system 100, in a high load region (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 when performing compression auto-ignition combustion. Spark ignition combustion is performed in which the air-fuel mixture is burned by ignition by the plug 25. On the other hand, in the CI (Compression Ignition) combustion region (compression auto-ignition combustion region) where the engine load that suppresses the combustion noise is set to a low load region less than the combustion switching load T1, the air-fuel mixture is not ignited by the spark plug 25. Compressed self-ignition combustion is performed in which self-ignition is performed.

  Next, control during steady operation in the CI combustion region will be described.

(1) EGR control 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 auto-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 valve 22 is opened twice with the operating state of the exhaust valve 22 as a special mode. However, in the CI combustion region where the engine load is relatively high (in the example shown in FIG. 5, the region where the engine load is equal to or higher than the reference load T2), the temperature of the air-fuel mixture is sufficiently high and excessive internal EGR gas is introduced. Then, heat generation occurs rapidly, that is, the maximum value of the heat generation rate becomes excessive, and the combustion noise may exceed the allowable range. Therefore, in the present embodiment, only the internal EGR is performed in the first region CI_1 where the engine load is less than the reference load T2, while the second region CI_2 where the engine load is set to be equal to or greater than the reference load T2 and less than the combustion switching load T1. Then, in addition to the internal EGR, the external EGR is performed to introduce the external EGR gas cooled by the EGR cooler 52, thereby suppressing the temperature of the air-fuel mixture from rising excessively. Specifically, in the second region CI_2, the exhaust valve 22 is opened twice while the operation state of the exhaust valve 22 is set to the special mode, and the EGR valve 53 is opened.

  In the CI fuel region, the internal EGR rate is controlled to an appropriate value according to the engine load. FIG. 6 shows a change in the EGR rate with respect to the engine load at the engine speed NE1 in FIG. FIG. 6 shows the ratios of internal EGR gas, external EGR gas, and fresh air, where the horizontal axis is the engine load and the vertical axis is the amount of gas that can be introduced into the combustion chamber 11 as 100%. As shown in FIG. 6, the internal EGR rate is set higher as the engine load is lower in the entire 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. Although FIG. 6 shows a case where the internal EGR rate is made constant when the engine load becomes lower than a predetermined value, the internal EGR rate may be increased as the engine load decreases over the entire engine load.

(2) Air-fuel ratio control In the second region CI_2, in a region on the relatively high load side where the engine load is equal to or higher than the preset stoichiometric switching load T3, the combustion temperature becomes high, so that it is generated by combustion (with catalyst) There is a possibility that NOx before purification (so-called RawNOx) cannot be sufficiently reduced. Therefore, in the present embodiment, in the high load side second region CI_2B in which the engine load is equal to or higher than the stoichiometric switching load T3 in the second region CI_2, the air-fuel ratio of the entire mixture is set to the stoichiometric air-fuel ratio. By controlling λ to λ = 1, it is possible to purify NOx with a three-way catalyst. On the other hand, in the region where the engine load is less than T3, RawNOx can be sufficiently reduced. Therefore, in the second region CI_2, the low load side second region CI_2A in which the engine load is set to be less than the stoichiometric switching load T3, and In the first region CI_1, the air-fuel ratio of the entire air-fuel mixture is made leaner than the stoichiometric air-fuel ratio and the air excess ratio λ of the entire air-fuel mixture is λ> 1 in order to further improve fuel efficiency.

(3) Fuel injection control In the CI combustion region, in order to obtain higher fuel consumption performance and exhaust performance, as shown in FIG. 7, intake stroke batch injection is performed in which fuel is injected collectively during the intake stroke. Implements homogeneous premixed compression auto-ignition combustion with lean fuel ratio. That is, fuel and air are sufficiently mixed in advance by the compression top dead center so that the air-fuel ratio of the entire mixture is substantially uniform (lean in the present embodiment), and self-ignition is performed in the vicinity of the compression top dead center. In FIG. 7, the heat generation rate dQ / dθ is also indicated by a broken line.

  As described above, in the CI combustion region, in steady operation, the lower the engine load, the smaller the internal EGR rate, and the intake stroke batch injection is performed.

  Here, the control at the time of steady operation as described above is set so that stable compression auto-ignition combustion can be realized in a state where the steady operation is performed under each operation condition and the wall surface temperature of the combustion chamber 11 is thereby stabilized. It is a thing. That is, the control is performed under a condition that the wall temperature in the combustion chamber is higher as the engine load is higher and the amount of heat generated in the combustion chamber is larger. Therefore, when accelerating to a region where the engine load is high, the wall temperature in the combustion chamber does not immediately rise to the steady state temperature under the operating conditions after acceleration. There is a problem that it becomes impossible to realize a proper compression ignition combustion.

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

  FIGS. 8 and 9 are both during acceleration, and at time t1, operating conditions (first operating condition) in which the internal EGR rate during steady operation is high to operating conditions where the internal EGR rate during steady operation is low (first operating condition). This shows the time change of each parameter when the control at the time of steady operation is carried out as it is and the control is simply switched according to the change of the operation condition. The total injection amount, injection pulse, EGR rate, combustion chamber wall temperature (combustion chamber wall temperature), pre-ignition temperature, that is, the temperature of the air-fuel mixture in the combustion chamber at a predetermined timing before ignition are shown in order. FIG. 9 also shows changes in combustion temperature (maximum temperature during combustion). FIG. 8 shows a change when the vehicle is accelerated in the first CI region CI_1, that is, when the vehicle is accelerated in a region where only the internal EGR is performed. FIG. 9 shows an area where acceleration is performed from the first CI area CI_1 to the high load side second CI area CI_2A, that is, an area where the external EGR is performed from the area where the external EGR is not performed and the excess air ratio λ is set to λ = 1. It is a change when accelerating to. 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, the total injection amount increases with the acceleration at time t1. Specifically, the intake stroke batch injection is performed before and after the acceleration, and the injection pulse width is increased with the same injection start timing with the acceleration. Then, the internal EGR rate is lowered at time t1. 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 acceleration destination immediately after acceleration. On the other hand, the wall surface temperature of the combustion chamber rises only slowly after acceleration, and immediately after acceleration, the temperature remains almost the same as that before acceleration and is lower than the appropriate temperature in the steady state of the operating conditions after acceleration. As described above, the wall surface temperature of the combustion chamber is lower than the appropriate temperature, and further, the internal EGR rate is reduced early, so that immediately after acceleration, the temperature before ignition becomes the appropriate temperature in the steady state of the operating conditions after acceleration. The air-fuel mixture falls below the temperature at which self-ignition is possible (temperature at which self-ignition is possible).

  As described above, when the control during the steady operation is performed as it is, the temperature before ignition after the acceleration is lower than the temperature at which the air-fuel mixture can self-ignite, and accordingly, stable compression auto-ignition combustion cannot be realized. . Specifically, the air-fuel mixture does not self-ignite and misfires, or a large amount of unburned fuel generated due to misfiring remains as internal EGR gas and burns rapidly in the next cycle, resulting in combustion fluctuations.

  In FIG. 9 as well, the internal EGR rate decreases with acceleration, while the increase in the combustion chamber wall temperature is delayed, so that the temperature before arrival falls below the temperature at which self-ignition is possible at time t3, and misfires occur. In the case shown in FIG. 9, there is a further problem that combustion noise increases because the external EGR is set to be performed under the operating condition of the acceleration destination. Specifically, with acceleration at time t1, the EGR valve 53 is opened and external EGR is started. However, the external EGR gas is introduced into the combustion chamber 11 only after passing through the exhaust passage 40, the EGR passage 51 and the intake passage 30. Therefore, there is a delay in the introduction of the external EGR gas, and the external EGR rate increases only gradually after acceleration. Therefore, as shown in FIG. 9, the combustion temperature becomes very high until time t2 and for a while thereafter, and the combustion noise exceeds the allowable value. In the example shown in FIG. 9, the excess air ratio λ is set to λ = 1 in the operating condition of the acceleration destination. As described above, in the present embodiment, the air-fuel ratio of the air-fuel mixture is controlled by changing the internal EGR rate. Therefore, in this example, a large amount of internal EGR gas is introduced during the period from time t1 to time t3 as the introduction of external EGR gas is delayed, and accordingly, the temperature before ignition reaches the time of time t2. The temperature is maintained above the temperature at which self-ignition is possible, and misfire occurs after time t2. The wall surface temperature of the combustion chamber 11 and the temperature of the internal EGR gas gradually increase from about time t3 when the internal EGR rate becomes the original value in the operating condition of the acceleration destination, but until time t4, the pre-ignition temperature remains. It is not high enough to cause misfire or combustion fluctuation.

  In order to solve the above-described problem, in the engine system 100, when accelerating from an operation region in which the internal EGR rate is high in the CI combustion region to an operation region in which the internal EGR rate is low, the total injection amount to be injected into the combustion chamber 11 While making it larger than the total injection amount in the steady operation under the operating conditions, as shown in FIG. 10A, only a part of this total injection amount is injected in the intake stroke, and the rest is injected in the expansion stroke. That is, the main injection is performed in the intake stroke and the post injection is performed in the expansion stroke. In FIG. 10A, the injection pulse is indicated by a solid line and the heat generation rate is indicated by a broken line.

  Then, the total injection amount is decreased as the number of elapsed cycles after acceleration increases. At this time, as shown in FIG. 10B, the injection amount of the post injection is decreased while the injection amount of the main injection is increased. In FIG. 10B, the solid line represents the injection pulse and the broken line represents the heat generation rate.

  Specifically, in the present embodiment, the injection amount of the main injection is made larger than the total injection amount in the operating condition before acceleration. The injection start timing of the main injection is the same as the timing at the time of steady operation of the operating conditions after acceleration. Here, in this embodiment, the injection start timing of the injection in the intake stroke is fixed, and the injection start timing of the main injection is maintained at the same timing before and after acceleration. As the number of elapsed cycles increases, the injection pulse width of the main injection is increased.

  Further, the injection timing of the post injection is set to a timing at which the post-injected fuel surely burns. For example, the injection start timing of the post injection is set to about 20 ° after compression top dead center to about 50 ° CA after compression top dead center immediately after acceleration. As the number of elapsed cycles increases, the injection start timing of post injection is retarded and the injection pulse width of post injection is decreased. In the present embodiment, the injection end timing of the post injection is maintained constant.

  By performing such main injection and post injection, heat generation divided into two occurs as shown in FIGS.

  FIGS. 11 and 12 show temporal changes in the total injection amount, the injection pulse, the EGR rate, the combustion chamber wall temperature, and the pre-ignition temperature when the above control is performed. FIG. 11 shows changes under the same acceleration conditions as in the example shown in FIG. 8. When the acceleration is performed in the first CI region CI_1, that is, when the acceleration is performed in a region where only the internal EGR is performed, the control is performed. This is a change when implemented. FIG. 12 shows changes under the same acceleration conditions as the example shown in FIG. 9. When the acceleration is performed from the first CI area CI_1 to the high load side second CI area CI_2A, that is, the external EGR is performed from the area where the external EGR is not performed. This is a change when the excess air ratio λ is accelerated to a region where λ = 1 is set. In FIGS. 11 and 12, the corresponding changes in FIGS. 8 and 9 are indicated by broken lines.

  In FIG. 11, the total injection amount is increased with the acceleration at time t1, and the main injection during the intake stroke and the post injection during the expansion stroke are performed. At this time, the total injection amount is increased more than the total injection amount in the steady operation, that is, the total injection amount in the example shown in FIG. Specifically, the main injection amount is larger than the pre-acceleration condition, but smaller than the amount during steady operation, and the post injection amount is larger than the main injection amount. The injection start timing of the main injection is maintained at the same value as the timing at the time of steady operation before acceleration and after acceleration. As the time elapses, that is, as the number of elapsed cycles immediately after acceleration increases, the main injection amount gradually increases and the post injection amount gradually decreases. In the example shown in FIG. 11, the injection amount is increased or decreased by a certain amount every time one cycle is advanced. At this time, the injection pulse of the main injection is increased in a state where the injection start timing is maintained constant. On the other hand, in the post-injection, the injection pulse is decreased while the injection end timing is maintained constant. That is, the start timing of the post injection is gradually delayed. In the example shown in FIG. 11, the injection start timing of post injection, that is, the injection timing is increased or decreased by a certain amount every time one cycle is advanced. When a predetermined time elapses, main injection and post injection are stopped, and injection control during steady operation, that is, intake air batch injection is performed. In the example shown in FIG. 11, when the wall surface temperature of the combustion chamber is close to the temperature during steady operation, switching to intake stroke batch injection is performed. The timing of switching to the intake stroke batch injection is not limited to this, and this injection control may be switched according to the elapsed time or the number of elapsed cycles from immediately after acceleration.

  By performing the injection control described above, the wall surface temperature of the combustion chamber rapidly increases immediately after the acceleration even though the internal EGR rate decreases immediately after the acceleration, and the mixture before the ignition is It is maintained at a temperature that can be ignited. In this way, stable compression ignition combustion is continued before and after acceleration without misfire or the like.

  Thus, in the injection control of the engine system 100, the total injection amount is increased to increase the thermal energy generated in the combustion chamber 11, so that the wall surface temperature of the combustion chamber can be raised early. Although not shown in FIG. 11, the exhaust gas temperature can be increased by increasing the thermal energy to increase the temperature of the internal EGR gas, and the increase in the wall surface temperature of these combustion chambers and the internal EGR gas temperature. As a result of the increase, the temperature of the air-fuel mixture can be kept high, and stable compression auto-ignition combustion can be realized. In particular, in the engine system 100, a part of the fuel is injected by post injection in the expansion stroke while increasing the total injection amount. Therefore, it is possible to reliably avoid a situation in which the engine torque is excessively generated and the torque shock is generated while increasing the total injection amount. That is, the combustion energy of the fuel injected in the expansion stroke is not easily converted into effective engine torque. Therefore, the engine torque can be maintained at an appropriate value while increasing the total injection amount and increasing the combustion energy. Further, if post injection is performed in the expansion stroke, the temperature of the exhaust gas can be effectively increased, and therefore the ignitability of the air-fuel mixture can be more reliably increased in the engine system 100.

  Further, as the number of elapsed cycles increases, the total injection amount is decreased while increasing the injection amount of main injection and the injection amount of post injection, and as the number of elapsed cycles increases, post injection The injection timing (injection start timing) is retarded, so that the intake stroke batch injection can be shifted more smoothly and the occurrence of torque shock can be avoided. That is, after acceleration, as the wall surface temperature of the combustion chamber and the temperature of the air-fuel mixture rise, the combustion energy generated by post-injection becomes easier to be converted into more effective engine torque, but the control is performed as described above. Thus, it can be avoided that the combustion energy generated by the post injection is converted into an effective engine torque and excessive engine torque is generated.

  In FIG. 12, as in the case shown in FIG. 11, after the acceleration at time t1, the total injection amount is increased, the main injection in the intake stroke and the post injection in the expansion stroke are performed, and the injection of the main injection The amount is gradually increased, and the post injection amount is gradually decreased. Further, the injection timing of the post injection is gradually delayed. Even when the EGR rate is appropriately controlled near the time t3, the internal EGR rate is reduced and the external EGR rate is set to an appropriate value, and the wall surface temperature of the combustion chamber is not sufficiently increased, before ignition. The temperature is maintained above the temperature at which the air-fuel mixture can self-ignite and misfires are avoided.

  Further, in FIG. 12, the combustion temperature is suppressed to an allowable temperature or less, and the generation of combustion noise is avoided. That is, as described above with reference to FIG. 9, before and after acceleration when accelerating from the first CI area CI_1 to the high-load-side second CI area CI_2A, that is, when accelerating from an area where external EGR is not performed to an area where external EGR is performed. When the intake stroke batch injection is performed in this case, there is a problem that the combustion temperature is excessively increased due to the delay in the introduction of the external EGR gas, and the combustion noise exceeds the allowable value. However, this problem is solved in this engine system. The

  This is because, after acceleration, the main injection and post injection during the intake stroke are performed, so that the heat generation rate is divided as shown in FIGS. This is because the value was kept small.

  As described above, in the engine system 100 according to the present embodiment, in the CI combustion region, when accelerating to an operating condition in which the internal EGR rate is set low, the main injection before the compression top dead center and the effective engine torque It is generated in the combustion chamber while avoiding excessive engine torque because the total injection amount is larger than the total injection amount during steady operation while performing post injection in the expansion stroke that is difficult to convert to The thermal energy generated can be increased more than that during steady operation, and this thermal energy can compensate for the decrease in internal EGR rate and the decrease in the temperature of the air-fuel mixture due to the delay in tracking the combustion chamber wall temperature, while avoiding torque shock. Thus, it is possible to avoid the misfire or the like and to realize more stable and stable compression autoignition combustion. In addition, when accelerating to operating conditions where external EGR is implemented, heat generation can be divided and generated to reduce the combustion temperature and suppress the increase in combustion noise caused by the delay of external EGR. can do.

  In particular, as the number of cycles elapsed immediately after acceleration increases, the total injection amount decreases, the main injection amount increases, the post injection amount decreases, and the post injection timing is retarded. Therefore, it is possible to smoothly shift to control during steady operation while avoiding torque shock during acceleration.

  Here, in the above-described embodiment, the case where the injection amount and the injection timing are increased or decreased by a certain amount each time the number of elapsed cycles increases is described, but the increase / decrease method is not limited thereto.

  In the above embodiment, the case where the internal EGR is performed by opening the exhaust gas twice is shown, but the specific means for performing the internal EGR is not limited to this. For example, the exhaust valve 22 and the intake valve 21 may be closed with the compression top dead center interposed therebetween, and a negative overlap period in which both the valves are closed in the vicinity of the compression top dead center may be provided. In this case, the internal EGR rate may be changed by changing the period of the negative overlap period. In this way, if the internal EGR rate is changed by changing the opening timing of the exhaust valve 22 or the intake valve 21, the internal EGR rate can be changed to an appropriate value earlier.

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

Claims (5)

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;
Fuel and air injected from the fuel injection device while driving the internal EGR means and performing internal EGR during operation in a compression auto-ignition combustion region where the engine load is set to a region lower than a predetermined load Combustion control means for burning the air-fuel mixture by self-ignition in the combustion chamber and performing compression self-ignition combustion,
In the compression auto-ignition combustion region, the internal EGR rate, which is the ratio of the internal EGR gas amount to the total gas amount in the combustion chamber, is set to be lower as the engine load is higher,
The combustion control means includes
During steady operation in the compression auto-ignition combustion region, the fuel injection device is made to inject fuel during the intake stroke,
The fuel injection is performed at the time of acceleration from the first operating condition in the compressed auto-ignition combustion region to the second operating condition in the compressed auto-ignition combustion region where the internal EGR rate is set lower than the first operating condition. The apparatus performs main injection for injecting fuel into the combustion chamber before compression top dead center and post-injection for injecting fuel into the combustion chamber during an expansion stroke, and the injection amount of the main injection and the post A control device for a direct injection engine, characterized in that a total injection amount including an injection amount of the injection is made larger than a total injection amount during steady operation under the second operating condition. The direct injection engine control device according to claim 1,
The combustion control means performs the post-injection over a plurality of cycles at the time of acceleration from the first operating condition to the second operating condition, and as the number of elapsed cycles from immediately after the acceleration increases, A control device for a direct injection engine, wherein the injection amount is increased while the injection amount of the post injection is decreased, and the injection timing of the post injection is set to the retard side. A control device for a direct injection engine according to claim 1 or 2,
The combustion control means sets the internal EGR means to a low value corresponding to the second operating condition immediately after the acceleration at the time of acceleration from the first operating condition to the second operating condition. A control device for a direct injection engine, which controls the internal EGR means. A control device for a direct injection engine according to claim 3,
The internal EGR means drives at least one of an exhaust valve and an intake valve,
The control device for a direct injection engine, wherein the combustion control means changes the internal EGR rate by changing an opening / closing timing of at least one of an exhaust valve and an intake valve. It is a control apparatus of the direct-injection engine in any one of Claims 1-4,
An external EGR means for recirculating the exhaust gas discharged from the combustion chamber to the intake side;
When the engine load of the second operating condition is greater than or equal to a preset reference load, the combustion control means reduces the internal EGR rate while accelerating the transition to the second operating condition, and the external EGR means To control the exhaust gas to return to the intake side.

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