JP6249667B2 - Control device for internal combustion engine - Google Patents

Description

  The present invention relates to a control device for an internal combustion engine. More specifically, the homogeneous mixture formed in the cylinder is combusted by compression ignition, and internal EGR control in which a part of the burned gas remains in the cylinder as an internal EGR gas and the type of fire in the homogeneous mixture. TECHNICAL FIELD The present invention relates to a control device for a fire type auxiliary compression ignition type internal combustion engine that is used in combination with fire type combustion control for injecting and igniting a fuel for use.

  There has been proposed an internal combustion engine capable of premixed compression ignition (HCCI) combustion in which a lean and homogeneous mixture is formed in a cylinder and the homogeneous mixture is burned by compression ignition. This HCCI combustion enables high-efficiency operation with low NOx emissions and an increased effective compression ratio. However, in this HCCI combustion, for example, in a high-load operation region, it is difficult to burn a homogeneous mixture at an appropriate timing, and knocking, misfire, etc. are likely to occur. For this reason, in order to compensate for the high load operation region where HCCI combustion is difficult, switching between spark ignition (SI) combustion in which the air-fuel mixture is burned by the spark plug and HCCI combustion is performed according to the operation region.

In recent years, in order to realize HCCI combustion more stably, fuel is injected and ignited more locally in the homogeneous mixture formed in the cylinder, and this is used as a fire type to induce self-ignition of the homogeneous mixture. Fire-type auxiliary compression ignition technology has been proposed. In such a control apparatus for an internal combustion engine of a fire type auxiliary compression ignition type, the fire type is burned as intended, and the compression end temperature (temperature in the cylinder when the piston is located at the top dead center in the compression stroke ) is homogeneously mixed. In order to maintain an appropriate temperature at which self-ignition occurs, it is necessary to appropriately control from the initial compression temperature (temperature in the cylinder when the piston is located at the bottom dead center in the compression stroke ).

  The initial compression temperature can be controlled by increasing or decreasing the amount of burnt gas remaining in the cylinder, that is, the so-called internal EGR amount (see, for example, Patent Document 1). In the invention of Patent Document 1, a target compression end temperature that is a target for the compression end temperature is set, and an internal EGR amount that can obtain the target compression end temperature is calculated according to a required torque for the internal combustion engine, The cam phase of the cam phase variable mechanism that defines the closing timing of the exhaust valve is set according to the internal EGR amount.

JP 2011-220121 A

  However, if the closing timing of the exhaust valve is changed, the pump loss increases. Therefore, when the closing timing of the exhaust valve is changed to control the compression end temperature to an appropriate temperature as in the invention of Patent Document 1, Although it is necessary to change the closing timing of the intake valve, if the closing timing of the intake valve is changed, the effective compression ratio may change, and the compression end temperature may not be maintained at the target temperature.

  An object of the present invention is to provide a control device for an internal combustion engine that can control the temperature in a cylinder with high accuracy so as not to lower the combustion stability of the internal combustion engine of the fire type auxiliary compression ignition type.

(1) An internal combustion engine (for example, an engine 1 described later) of the present invention causes a part of burned gas to remain in a cylinder (for example, a cylinder 11a described later) as an internal EGR gas and is formed in the cylinder. This is a fire type auxiliary compression ignition type in which fuel for a fire type is injected and ignited in a homogeneous mixture, and the homogeneous mixture is burned by compression ignition using this as a fire type. The control device has a variable valve mechanism (for example, an IN side described later) of an exhaust valve (for example, an exhaust valve 14i described later) and an intake valve (for example, an intake valve 14e described later) capable of changing the internal EGR rate and the effective compression ratio. VTC 18i and EX side VTC 18e), target fire type ignition timing setting means for setting a target fire type ignition timing that is a target for the ignition timing of the mixture for the fire type , and target fire type that is a target of the in-cylinder temperature at the target fire type ignition timing Target fire type ignition timing temperature setting means (for example, ECU 5 described later) for setting the ignition timing temperature (Thidane_trg), and a target internal EGR rate (target internal EGR rate) so that the target fire type ignition timing temperature is realized. set REGRn_trg) and the target spark ignition timing effective compression ratio is the goal of effective compression ratio in the target spark ignition timing of the (εHidane_trg), these th And a control means for controlling the variable valve mechanism based on the internal EGR ratio and the target spark ignition timing effective compression ratio (e.g., ECU 5 will be described later), the.

(2) In this case, the control device includes an in-cylinder state detecting means (for example, an in-cylinder pressure sensor 61, an intake sensor 66, an exhaust gas temperature sensor 67, and an ECU 5 described later) that detects an in-cylinder state. Fire type ignition timing temperature predicting means for calculating a predicted fire type ignition timing temperature (Thidane_pre), which is a predicted value of the in-cylinder temperature at the target fire type ignition timing , based on the output of the in-cylinder state detecting means and the target fire type ignition timing effective compression ratio (For example, an ECU 5 described later), and difference value calculating means (for example, an ECU 5 described later) for calculating a difference value (ΔThidane) between the target fire type ignition timing temperature (Thidane_trg) and the predicted fire type ignition timing temperature And the control means preferably sets the target internal EGR rate by feedback control based on the difference value calculated during the previous combustion cycle. Yes.

  (3) In this case, the control means sets a target internal EGR rate so that the difference value converges to 0 by a feedback control algorithm including an integral term, and the pump loss is minimized based on the target internal EGR rate. Thus, it is preferable to set the target fire type ignition timing effective compression ratio.

(1) In the present invention, a target fire type ignition timing that is a target for the ignition timing of the mixture for the fire type is set , a target fire type ignition timing temperature that is a target of the in-cylinder temperature at the target fire type ignition timing is set, and this The target internal EGR rate and the target fire type ignition timing effective compression ratio are set so that the target fire type ignition timing temperature is realized, and the intake valve drive mechanism and the exhaust valve are based on the target internal EGR rate and the target fire type ignition timing effective compression ratio. Control the drive mechanism. As a result, the ignition type ignition timing temperature can be controlled with high accuracy, so that the ignition type can be burned as intended, and the homogeneous mixture can be self-ignited stably without knocking or misfiring. In addition, by setting both the target internal EGR rate and the target fire type ignition timing compression ratio from the target fire type ignition timing temperature, the pump loss can be reduced while maintaining the fire type ignition timing temperature at the target.

(2) In the present invention, the target internal EGR rate is determined by feedback control based on the difference value between the predicted fire type ignition timing temperature calculated based on the output of the in-cylinder state detection means and the target fire type timing effective compression ratio and the target fire type ignition timing temperature. And the intake valve drive mechanism and the exhaust valve drive mechanism are controlled based on the target internal EGR and the target fire type ignition timing effective compression ratio. Thereby, even when the operating state of the internal combustion engine changes transiently, the target fire type ignition timing temperature can be controlled with high accuracy.

  (3) In the present invention, the target internal EGR rate is set by the feedback control algorithm including the integral term so that the predicted fire type ignition timing temperature and the target fire type ignition timing temperature coincide with each other, and further, based on the target internal EGR rate. Set the target fire type ignition timing effective compression ratio so that the pump loss is minimized. This makes it possible to reliably reduce the pump loss while maintaining the target ignition timing temperature at the same time.

It is a figure which shows the structure of the engine which concerns on one Embodiment of this invention, and its control apparatus. It is a figure which shows the operation example of an exhaust valve and an intake valve in SI combustion mode. It is a figure which shows typically the breakdown of the means for raising self-ignition induction temperature and compression end temperature to this self-ignition induction temperature. It is a figure which shows the operation example of an exhaust valve and an intake valve in a fire type HCCI combustion mode. It is a figure which shows typically the control map referred when selecting a combustion mode. It is a flowchart which shows the specific procedure of a fire type ignition timing temperature control. It is a figure which shows the specific example of the map which determines a target combustion gravity center position. It is a figure which shows the specific example of the map which determines target fire type ignition timing temperature. It is a figure which shows the specific example of the map which determines the provisional value of target fire type ignition timing effective compression ratio. It is a figure which shows the specific example of the table which determines the provisional value with respect to the target phase angle of IN side VTC. It is a figure which shows the specific example of the table which determines a feedback gain. It is a figure which shows the specific example of the table which determines a target maximum stroke volume. It is a figure which shows the specific example of the map which determines the target phase angle of EX side VTC. It is a figure which shows the result by simulation of a fire type ignition timing temperature control. It is a figure which shows the result by the real machine of fire kind ignition timing temperature control. It is a figure which shows the result by the real machine of fire kind ignition timing temperature control.

Hereinafter, an embodiment of the present invention will be described with reference to the drawings.
FIG. 1 is a diagram showing a configuration of an internal combustion engine (hereinafter simply referred to as “engine”) 1 and a control device thereof according to the present embodiment.

  The engine 1 is a four-cylinder engine having a plurality of, for example, four cylinders 11a, and one of these is representatively shown in FIG. The engine 1 is configured by combining a cylinder block 11 in which a cylinder 11 a is formed and a cylinder head 12. The engine 1 is provided with an intake pipe 2 through which intake air flows, an exhaust pipe 3 through which exhaust flows, and an EGR pipe 4 that recirculates part of the exhaust gas in the exhaust pipe 3 to the intake pipe 2. ing.

  A piston 13a is slidably provided in the cylinder 11a, and a combustion chamber 12a of the engine 1 is formed by the top surface of the piston 13a and the surface of the cylinder head 12 on the cylinder 11a side. The piston 13a is connected to the crankshaft 13c via a connecting rod 13b. That is, the crankshaft 13c rotates according to the reciprocation of the piston 13a in the cylinder 11a.

  The cylinder head 12 is formed with an intake port 12 i that connects the combustion chamber 12 a and the intake pipe 2, and an exhaust port 12 e that connects the combustion chamber 12 a and the exhaust pipe 3. The cylinder head 12 is provided with an intake valve 14i for opening and closing an intake opening facing the combustion chamber 12a in the intake port 12i, and an exhaust valve 14e for opening and closing an exhaust opening facing the combustion chamber 12a among the exhaust ports 12e. It has been.

  The cylinder head 12 is provided with a spark plug 15 that faces the combustion chamber 12a. The spark plug 15 is connected to an electronic control unit (hereinafter referred to as “ECU”) 5 via an igniter and a driver (not shown). The spark plug 15 generates a spark at a timing determined by ignition control (not shown) executed by the ECU 5, and ignites the air-fuel mixture formed in the cylinder 11a.

  The cylinder head 12 is provided with an intake camshaft 16i and an exhaust camshaft 16e that are connected to a crankshaft 13c via a timing belt (not shown) and rotate according to the rotation of the crankshaft 13c. More specifically, when the crankshaft 13c rotates twice, the camshafts 16i and 16e rotate once. The intake camshaft 16i is provided with an intake cam 17i for driving the intake valve 14i to open and close, and the exhaust camshaft 16e is provided with an exhaust cam 17e for opening and closing the exhaust valve 14e. Accordingly, when the camshafts 16i and 16e rotate, the valves 14i and 14e advance and retract in a manner corresponding to the profiles of the cams 17i and 17e.

  At one end of the intake camshaft 16i, an intake-side cam phase variable mechanism (hereinafter referred to as "IN-side VTC") 18i for changing the cam phase of the intake cam 17i with respect to the crankshaft 13c, and the maximum opening of the intake valve 14i This is a mechanism called a variable valve lift on the intake side (so-called VTEC (registered trademark) 19i) that changes the valve opening period (that is, the valve opening angle width) (ie, the maximum lift amount), and is hereinafter referred to as “IN Side VTEC "). Similarly, at one end of the exhaust camshaft 16e, an exhaust-side cam phase variable mechanism (hereinafter referred to as “EX-side VTC”) 18e for changing the cam phase of the exhaust cam 17e with respect to the crankshaft 13c, and an exhaust valve 14e. There is provided an exhaust-side valve lift variable mechanism (hereinafter referred to as “EX-side VTEC”) that changes the maximum lift amount and the valve opening angle width.

  The IN side VTC 18i advances or delays the opening / closing timing of the intake valve 14i (that is, the opening timing (IVO) and the closing timing (IVC)) by advancing or retarding the cam phase of the intake camshaft 16i steplessly. it can. In the control device of the present embodiment, the operation in the Atkinson cycle (Miller cycle) in which the effective compression ratio of the combustion chamber 12a of the engine 1 can be variably controlled by the IN side VTC 18i is realized. That is, the effective compression ratio can be lowered by correcting the closing timing of the intake valve 14i to the advance side or the retard side with respect to the bottom dead center by the IN side VTC 18i and reducing the intake amount.

  The EX-side VTC 18e advances or delays the opening / closing timing of the exhaust valve 14e (that is, the opening timing (EVO) and the closing timing (EVC)) by almost the same mechanism as the IN-side VTC 18i. In the control device of the present embodiment, the EX-side VTC 18i changes the closing timing of the exhaust valve 14e to change the internal EGR rate (the total amount of gas existing in the combustion chamber 12a after the intake valve is closed (fresh air)). And the ratio of the amount of burned gas to the amount of gas, which is the sum of the burned gas and the burned gas). More specifically, when the closing timing of the exhaust valve 14e is corrected to the advance side with respect to the top dead center, the amount of burned gas confined in the combustion chamber 12a increases, so the internal EGR rate increases.

  As described above, in the control device of the present embodiment, the variable valve mechanisms of the intake valve 14i and the exhaust valve 14e that can change the internal EGR rate and the effective compression ratio are configured by the IN side VTC 18i and the EX side VTC 18e.

  As these VTCs 18i and 18e, for example, those that change the cam phase by hydraulic pressure are used. The ECU 5 sets the cam phase target (target phase angle) of the intake cam 17i and the exhaust cam 17e with respect to the crankshaft 13c for each combustion cycle (see, for example, FIG. 6 described later), and the VTCs 18i and 18e are set by the ECU 5. In order to achieve the target phase angle, the cam phases of the intake cam 17i and the exhaust cam 17e are advanced or retarded steplessly.

  The intake cam 17i and the exhaust cam 17e include at least two types of cams (not shown), a low-speed cam and a high-speed cam, each having a different cam nose. The IN side VTEC 19i selectively switches the intake cam 17i between the low speed cam and the high speed cam to switch the maximum lift amount and the valve opening angle width of the intake valve 14i between low speed and high speed. The EX side VTEC 19e uses the same mechanism as the IN side VTEC 19i to selectively switch the exhaust cam 17e between a low speed cam and a high speed cam, thereby reducing the maximum lift amount and the opening angle width of the exhaust valve 14e for low speed and high speed. It is switched between for use. As these VTECs 19i and 19e, those which perform the switching operation by hydraulic pressure, for example, are used in the same manner as the VTCs 18i and 18e. That is, these VTECs 19i and 19e switch the cam that drives the valves 14i and 14e between the low-speed cam and the high-speed cam at a timing determined by a process (not shown) in the ECU 5.

  A port injector PI for injecting fuel toward the combustion chamber 12a is provided upstream of the intake opening in the intake port 12i, and direct injection for injecting fuel directly into the combustion chamber 12a is provided in the cylinder head 12. Injector DI is provided. The fuel injection amount (injection period) from these injectors PI and DI and the fuel injection timing thereof are controlled by fuel injection control (not shown) executed in the ECU 5.

  The intake pipe 2 is provided with a throttle valve 22 that controls the amount of air supplied to the combustion chamber 12a of the engine 1 (that is, the intake amount). The throttle valve 22 is connected to the ECU 5 via a driver (not shown). That is, the throttle valve 22 is called a so-called DBW (Drive By Wire) throttle, which is mechanically disconnected from an accelerator pedal (not shown) operated by the driver. The throttle valve 22 is controlled to an appropriate opening degree by an intake air amount control (not shown) executed in the ECU 5.

  The exhaust pipe 3 is provided with an exhaust purification catalyst 31 for purifying the exhaust. The exhaust purification catalyst 31 is, for example, a three-way catalyst, and purifies HC, CO, NOx, etc. in the exhaust.

  The EGR pipe 4 connects the downstream side of the exhaust purification catalyst 31 in the exhaust pipe 3 and the downstream side of the throttle valve 22 in the intake pipe 2 to recirculate part of the exhaust discharged from the engine 1. The EGR pipe 4 is provided with an EGR cooler 41 that cools the recirculated exhaust gas and an EGR valve 42 that controls the flow rate of the recirculated exhaust gas. The EGR valve 42 is connected to the ECU 5 via a driver (not shown). The EGR valve 42 is controlled to an appropriate opening degree by intake air amount control (not shown) executed in the ECU 5.

  The ECU 5 is an electronic control unit that electronically controls the engine 1 and its auxiliary devices, and includes an electronic circuit such as a CPU, a ROM, a RAM, and various interfaces. A plurality of sensors 61 to 67 are connected to the ECU 5 in order to grasp the state of the engine 1 and the state of the vehicle on which the engine 1 is mounted.

  One in-cylinder pressure sensor 61 is provided for each cylinder 11a. FIG. 1 representatively shows only one of the plurality of in-cylinder pressure sensors 61. The in-cylinder pressure sensor 61 detects the pressure in the cylinder 11a (in-cylinder pressure) and transmits it to the ECU 5. The combustion barycentric position and the in-cylinder maximum pressure, which are one of the plurality of parameters, are calculated by a process (not shown) in the ECU 5 based on the output of the in-cylinder pressure sensor 61.

  The intake side cam sensor 62i transmits a pulse signal to the ECU 5 at every predetermined cam angle as the intake camshaft 16i rotates. The exhaust cam sensor 62e transmits a pulse signal to the ECU 5 at every predetermined cam angle as the exhaust camshaft 16e rotates. The ECU 5 grasps the actual cam phase of the camshafts 16i and 16e based on the pulse signals transmitted from the cam sensors 62i and 62e.

  The crank angle sensor 63 transmits a pulse signal to the ECU 5 at every predetermined crank angle in accordance with the rotation of the pulser fixed to the crankshaft 13c. The ECU 5 grasps the actual engine speed based on the pulse signal from the crank angle sensor 63.

  The accelerator pedal sensor 65 detects the amount of depression of the accelerator pedal operated by the driver, and transmits a detection signal corresponding to the depression amount to the ECU 5. The required torque corresponding to the torque required for the engine 1 from the driver is calculated by a process (not shown) in the ECU 5 based on the detection signal transmitted from the accelerator pedal sensor 65, the engine speed, and the like.

  The intake sensor 66 is a sensor that detects the state of intake air in the intake pipe 2. More specifically, the intake sensor 66 includes an intake air temperature sensor that transmits a detection signal substantially proportional to the intake air temperature at the target location (hereinafter referred to as “intake air temperature”) to the ECU 5, and the intake air pressure at the target location ( Hereinafter, it is constituted by an intake pressure sensor or the like that transmits a detection signal substantially proportional to “intake pressure”) to the ECU 5.

  The exhaust temperature sensor 67 is a sensor that detects the temperature of the exhaust gas in the exhaust pipe 3. More specifically, the exhaust temperature sensor 67 transmits to the ECU 5 a detection signal that is substantially proportional to the exhaust temperature at the target location (hereinafter referred to as “exhaust temperature”).

  The engine 1 configured as described above has a combustion mode according to an operating state, which is a spark ignition combustion mode (hereinafter referred to as “SI (Spark Ignition) combustion mode”) and a fire type auxiliary premixed compression ignition combustion mode ( Hereinafter, it can be switched between “fire type HCCI (Homogeneous Charge Compression Ignition) combustion mode”).

In the SI combustion mode, for example, a predetermined amount of combustion is injected into the intake port 12i by the injector PI during the intake stroke , and a predetermined amount of fuel is injected by the injector DI during the compression stroke. A spark is emitted from 15 to ignite and burn the air-fuel mixture formed in the cylinder 11a. Here, in the intake air amount control and the fuel injection control in the SI combustion mode, the air-fuel ratio of the air-fuel mixture formed in the cylinder 11a is different from the fire type HCCI combustion mode described later, and the stoichiometric air-fuel ratio (for example, A / F = 14.7). Degree).

FIG. 2 is a diagram illustrating an operation example of the exhaust valve and the intake valve in the SI combustion mode. In the SI combustion mode, the piston opens the exhaust valve over the exhaust stroke from the bottom dead center to the top dead center, and the piston opens the intake valve over the intake stroke from the top dead center to the bottom dead center. Further, as shown in FIG. 2, in the SI combustion mode, the intake valve and the exhaust valve are driven so that a positive overlap occurs between the opening timings of both.

  In the fire type HCCI combustion mode, unlike the SI combustion mode, the air-fuel mixture is combusted by compression ignition. For this reason, the in-cylinder temperature (especially, the compression end temperature, which is the in-cylinder temperature when the piston reaches top dead center), rises to a self-ignition induction temperature (approximately 1000 [K]) higher than the SI combustion mode. It is necessary to let If the compression end temperature is much lower than the self-ignition inducing temperature, misfire occurs, and if the compression end temperature is much higher than the self-ignition inducing temperature, pre-ignition occurs. For this reason, in order to stably realize the fire type HCCI combustion mode, it is important to control the compression end temperature close to the self-ignition induction temperature with high accuracy.

  FIG. 3 is a diagram schematically showing the breakdown of the self-ignition induction temperature and the means for raising the compression end temperature to the self-ignition induction temperature. In the control device of the present embodiment, the compression end temperature is combined by combining adiabatic compression, temperature rise by internal EGR control in which a part of burned gas remains in the cylinder as internal EGR gas, and temperature rise by fire type heat generation. Is raised to the autoignition induction temperature.

More specifically, in the fire type HCCI combustion mode, a lean and homogeneous air-fuel mixture is formed in the cylinder by first injecting a predetermined amount of fuel into the intake port 12i by the port injector PI during the intake stroke . Next, during the compression stroke , a small amount of fire type fuel is injected into the cylinder by the direct injection injector DI, and a fire type mixture is locally formed in the previously formed homogeneous mixture. Next, for example, a spark is generated at an appropriate timing by the spark plug 15, and this is used as an opportunity to ignite the mixture for fire type, which is used as a fire type to induce compression auto-ignition of the surrounding homogeneous mixture. In the fire type HCCI combustion mode, the locally formed fire mixture is burned, thereby raising the compression end temperature and inducing compression auto-ignition of the surrounding homogeneous mixture. In the intake air amount control and the fuel injection control in the fire type HCCI combustion mode, the air-fuel ratio of the air-fuel mixture formed in the cylinder 11a is leaner than the stoichiometric air-fuel ratio (for example, A / F = 25) unlike the SI combustion mode. Degree). It should be noted that the spark plug control is used to determine the amount and timing of fuel injection from the direct injector DI in the fire type HCCI combustion mode, and the spark plug is ignited so that the mixture for the spark is ignited at the target spark ignition timing described later. Detailed explanations such as ignition control for determining the ignition timing are omitted.

  FIG. 4 is a diagram illustrating an operation example of the exhaust valve and the intake valve in the fire type HCCI combustion mode. In FIG. 4, in order to clarify the difference from the SI combustion mode, an operation example of these valves in the SI combustion mode is indicated by a broken line.

  In the fire type HCCI combustion mode, the burned gas is reduced by increasing the closing timing of the exhaust valve so that there is a negative overlap (NOL (Negative OverLap)) between the opening timing of the exhaust valve and the opening timing of the intake valve. It is confined in the cylinder, and the compression end temperature is raised to the self-ignition induction temperature. In the HCCI combustion mode, the pump loss is minimized by advancing the closing timing of the exhaust valve and delaying the closing timing of the intake valve at the same time.

  FIG. 5 is a diagram schematically showing a control map that is referred to when the ECU 5 selects the combustion mode. The combustion mode is determined at a predetermined cycle in the ECU 5 based on an operating state parameter that specifies the operating state of the engine. Here, as the operating state parameter, specifically, as shown in FIG. 5, the engine speed and the required torque can be mentioned. According to the example shown in FIG. 5, the ignition type HCCI combustion mode is determined as the combustion mode only when the engine operating state is a low engine speed and a low load, and SI combustion is performed when the engine operating state is otherwise. The mode is determined as the combustion mode.

FIG. 6 is a flowchart showing a specific procedure of the ignition type ignition timing temperature control in the ignition type HCCI combustion mode. As described above, in the fire type HCCI combustion mode, after the ignition mixture is ignited at a predetermined timing during the compression stroke , the homogeneous mixture is compressed and ignited using the combustion of the ignition mixture. Here, the “fire type ignition timing temperature” corresponds to the in-cylinder temperature when the mixture for fire type ignites. The series of processes shown in FIG. 6 is executed in the ECU 5 for each combustion cycle while the fire type HCCI combustion mode is selected as the combustion mode.

  First, in S1, the engine speed NE (k) and the required torque TRQ (k) necessary for the following calculation are acquired, and the process proceeds to the next step S2. In the following description, a value “k” indicating discrete time is attached to a value calculated or acquired in the current combustion cycle. On the other hand, the sign “k-1” is attached to the value calculated or acquired in the previous combustion cycle.

  In S2, a plurality of parameters (actual internal EGR rate rEGR_act (k), in-cylinder intake gas temperature estimated value TA_cyl (k), internal EGR temperature estimated value Tex_cyl (k) indicating the in-cylinder state necessary for the following calculation are obtained. Etc.) and move to S4.

  Here, the actual internal EGR rate rEGR_act (k) is estimated by a known method, for example, based on the in-cylinder volume and the in-cylinder pressure at the closing timing of the exhaust valve. The in-cylinder intake gas temperature estimated value TA_cyl (k) corresponds to the temperature of the gas newly introduced into the cylinder, and is calculated by a known method, for example, based on the intake air temperature detected by the intake sensor.

  The internal EGR temperature estimated value Tex_cyl (k) corresponds to the temperature of burned gas confined in the cylinder by the internal EGR control, and is calculated by a known method based on, for example, the exhaust temperature detected by the exhaust temperature sensor. . For example, when the detection location of the exhaust temperature sensor is close to the exhaust port, the exhaust temperature detected by the exhaust temperature sensor can be approximated as the temperature of burned gas confined in the cylinder.

  In S4, a map illustrated in FIG. 7 is searched based on the rotational speed NE and the required torque TRQ acquired in S1, thereby obtaining a target combustion gravity center position MBF50_trg [deg.] Corresponding to a target with respect to the combustion gravity center position in the current combustion cycle. ] Is set. Here, the combustion center of gravity position corresponds to a position where the ratio of the mass of the burned portion to the mass of the entire air-fuel mixture in the cylinder is 50%. In addition, although this embodiment demonstrates the case where a target is set to a combustion gravity center position, this invention is not limited to this. For example, instead of the combustion center of gravity position, the target may be set at the in-cylinder maximum pressure position (Pmax position) corresponding to the position where the in-cylinder pressure becomes maximum. Since the combustion gravity center position and the in-cylinder maximum pressure position are substantially the same, it is considered that the same effect can be obtained even in this way.

In S5, a target fire type ignition timing AngleHidaneIGN_trg [deg.], Which is a target for the ignition timing of the mixture for fire type (hereinafter referred to as “fire type ignition timing”), is set. In other words, the target fire type ignition timing AngleHidaneIGN_trg corresponds to the timing at which the mixture for fire type ignites in the ignition control (not shown). More specifically, the target fire type ignition timing AngleHidaneIGN_trg is calculated from the target MBF50_trg calculated in S4, as shown in the following formula (1), the fire type combustion period AngleHidane [deg. ] And the main combustion period AngleMain [deg.] Corresponding to the time for which a predetermined amount of the homogeneous air-fuel mixture burns is subtracted. Here, the two periods AngleHidane and AngleMain are calculated for each combustion cycle by a process not shown.

  In S6, a target ignition type ignition timing temperature that is the target of the in-cylinder temperature in the target ignition type ignition timing AngleHidaneIGN_trg by searching a map as illustrated in FIG. 8 based on the rotational speed NE and the required torque TRQ acquired in S1. Set Thidane_trg [K].

  In S7, a map as illustrated in FIG. 9 is searched based on the rotational speed NE and the demanded torque TRQ acquired in S1, so that a temporary value εHidane_trg0 [− ] Is calculated. Here, the target fire type ignition timing effective compression ratio corresponds to the effective compression ratio at the fire type ignition timing.

In S8, the target fire type ignition timing stroke volume Vhidane_trg [cc] is calculated based on the following formula (2). In the following formula (2), “Offset” is the offset length, “Stroke” is the stroke length, “Boa” is the bore diameter, and “l” is the connecting rod length. Yes, “e” is the theoretical compression ratio, which is a fixed value. “Q (k)” is the unit of the target fire type ignition timing AngleHidaneIGN_trg set in S5 converted from [deg.] To [rad.], And “L (k)” is the above target fire type ignition. The piston movement amount calculated using the time q (k).

In S9, as shown in the following formula (3), the provisional maximum value eHidane_trg0 of the target ignition type effective ignition timing calculated in S7 is multiplied by the target ignition type ignition timing stroke volume Vhidane_trg calculated in S8 to obtain a target maximum described later. The provisional value Vmax_trg0 [cc] for the stroke volume Vmax_trg [cc] is calculated.

In S10, a table such as that illustrated in FIG. 10 is searched based on the provisional value Vmax_trg0 of the target maximum stroke volume calculated in S9, whereby a provisional value INVTC_trg0 for a target phase angle INVTC_trg [deg.] Of the IN-side VTC to be described later. [deg.] is calculated.

  In S11, based on the difference value ΔThidane (k-1) between the target fire type ignition timing temperature Thidane_trg (k-1) calculated in S19 described later during the previous combustion cycle and the predicted fire type ignition timing temperature Thidane_pre (k-1). Thus, the feedback gain rEGRn_gain (k) is calculated by searching the table illustrated in FIG.

In S12, a feedback correction value ΔrEGRn_fb [−] for a target internal EGR rate rEGRn_trg [−], which will be described later, is calculated by an integral calculation as shown in the following equation (4).

In S13, the difference value ΔThidane (k−) is obtained by adding the actual internal EGR rate rEGR_act (k) acquired in S2 and the feedback correction value ΔrEGRn_fb (k) calculated in S12 as shown in the following equation (5). The target internal EGR rate rEGRn_trg (k) [-], which is the target for the internal EGR rate, is set so that 1) converges to 0.

In S14, based on the target internal EGR rate rEGRn_trg determined in S13, a phase angle shift amount ΔINVTC [deg.] Corresponding to the correction amount for the target phase angle INVTC_trg of the IN side VTC is calculated so that the pump loss is minimized ( The target phase angle INVTC_trg of the IN-side VTC is set by adding the correction value ΔINVTC and the provisional value INVTC_trg0 (see the following formula (6-1)) (see the following formula (6-2)). In the following equation (6-1), the shift amount that minimizes the pump loss is assumed on the assumption that the phase angle 1 [deg.] Of the IN-side VTC corresponds to the internal EGR rate 1 [%]. Although ΔINVTC is set, the present invention is not limited to this.

In S15, a table as illustrated in FIG. 12 is searched based on the target phase angle INVTC_trg (k) of the IN-side VTC that is determined so that the target internal EGR rate rEGRn_trg is realized and the pump loss is minimized in S14. Thus, the target maximum stroke volume Vmax_trg (k) is set.

In S16, based on the target maximum stroke volume Vmax_trg (k) and the target fire type ignition timing stroke volume Vhidane_trg (k) determined in S15, the target fire type ignition timing effective compression ratio eHidane_trg is set by the following equation (7).

In S17, based on the in-cylinder intake gas temperature estimated value TA_cyl, the internal EGR temperature estimated value Tex_cyl acquired in S2, and the target internal EGR rate rEGRn_trg set in S13, the in-cylinder at the initial stage of compression is expressed by the following equation (8). A predicted compression initial temperature T1_pre corresponding to the predicted temperature value is calculated.

In S18, based on the predicted compression initial temperature T1_pre calculated in S17 and the target fire type ignition timing effective compression ratio εHidane_trg set in S16, the predicted fire type corresponding to the predicted value of the in-cylinder temperature at the fire type ignition timing by the following equation (9). Ignition timing temperature Thidane_pre [K] is calculated.

In S19, as shown in the following formula (10), a difference value ΔTHidane between the target fire type ignition timing temperature THidane_trg set in S6 and the predicted fire type ignition timing temperature THidane_pre calculated in S18 is calculated. As described above, the difference value ΔTHidane calculated here is used as an input for feedback control in the calculation at the next combustion cycle.

  In S20, by searching a map as illustrated in FIG. 13 based on the target internal EGR rate rEGRn_trg set in S13 and the target phase angle INVTC_trg of the IN side VTC, the target value for the phase angle of the EX side VTC is obtained. The corresponding target phase angle EXVTC_trg [deg.] Is set, and this process ends.

Next, the effect of the above-described fire type ignition timing temperature control will be described.
FIG. 14 is a diagram showing a result of simulation of the above-described fire type ignition timing temperature control. The upper part of FIG. 14 shows the change of the target value (thick broken line) of the ignition type ignition timing temperature and the actual value (thin solid line) in the simulation, and the middle part shows the target ignition type effective compression ratio (dashed line) and the target internal EGR rate ( The lower row shows changes in the target phase angle (broken line) of the IN side VTC and the target phase angle (solid line) of the EX side VTC.

  FIG. 14 shows a case where the required torque TRQ is increased or decreased within the range of the fire type HCCI combustion mode while the engine speed NE is kept constant (1500 [rpm]). More specifically, the required torque TRQ was increased at a constant rate between times t0 and t1, and thereafter the required torque TRQ was decreased at a constant rate between times t2 and t3.

  As shown in FIG. 14, the target fire type ignition timing temperature increases when the required torque TRQ is increased while the rotation speed is kept constant, and decreases when the required torque TRQ is decreased. When the target fire type ignition timing temperature changes, the target fire type ignition timing compression ratio and the target internal EGR rate are set accordingly. Further, the target phase angle of the IN-side VTC and the target phase angle of the EX-side VTC are determined so as to realize the target fire type ignition temperature, the target fire type ignition timing compression ratio, and the target internal EGR rate.

  As described above, by simultaneously controlling the target internal EGR rate and the target fire type ignition timing compression ratio so that the target fire type ignition timing temperature is realized, the actual fire type ignition timing temperature is ± 10 [K with respect to the target. ] Can be controlled with high accuracy within the range.

  FIGS. 15 and 16 are diagrams showing the results of the above-mentioned fire type ignition timing temperature control by the actual machine. The upper part of FIG. 15 shows the change of torque, the middle part shows the change of the estimated ignition type ignition timing temperature (dark solid line) and the target ignition type ignition timing temperature (thin solid line), and the lower part shows the ignition type compression timing compression ratio (broken line) and the actual internal A change in the EGR rate (solid line) is shown. The upper part of FIG. 16 shows changes in the actual phase angle of the IN side VTC (solid line) and the actual phase angle of the EX side VTC (thick broken line), and the lower part shows the target in-cylinder maximum pressure position (thick solid line) and the actual in-cylinder maximum pressure. The change in position (thin solid line) is shown. As described above, the target in-cylinder maximum pressure position is a parameter that can be used instead of the target combustion gravity center position. 15 and 16 show the results when the required torque TRQ is intermittently changed up and down as shown in the upper part of FIG. 15 while keeping the engine speed NE constant (1500 [rpm]). .

  As shown in the middle of FIG. 15, according to the present invention, it is verified not only by simulation but also by an actual machine that the ignition type ignition timing temperature can be controlled to the target ignition type ignition timing temperature with high accuracy within a range of ± 10 [K]. It was done. In addition, as shown in the lower part of FIG. 16, according to the present invention, the ignition type ignition timing temperature is accurately controlled to the target, so that the ignition mixture is appropriately combusted, and the in-cylinder maximum pressure position is set to the target. It was verified that control can be performed with high accuracy within a range of ± 5 [deg.].

  Although one embodiment of the present invention has been described above, the present invention is not limited to this. For example, in the above-described embodiment, the case where the spark type mixture is ignited using the spark plug in the fire type HCCI combustion mode has been described, but the present invention is not limited thereto. In some cases, the air-fuel mixture for sparks is self-ignited by compression without using a spark plug, and the present invention can be applied to such a case.

1. Engine (internal combustion engine)
14i ... Intake valve 14e ... Exhaust valve 18i ... IN side VTC (Variable valve mechanism)
18e ... EX side VTC (Variable valve mechanism)
5 ... ECU (target fire type ignition timing temperature setting means, fire type ignition timing temperature prediction means, difference value calculation means, control means, in-cylinder state detection means)
61 ... In-cylinder pressure sensor (in-cylinder state detection means)
66. Intake sensor (in-cylinder state detection means)
67 ... Exhaust temperature sensor (in-cylinder state detection means)

Claims (3)

A part of the burned gas is left in the cylinder as internal EGR gas, and the fuel for the fire type is injected and ignited in the homogeneous mixture formed in the cylinder, and the homogeneous mixture is compressed and ignited using this as a fire type. A control device for a combustion type auxiliary compression ignition type internal combustion engine to be burned by
A variable valve mechanism for an exhaust valve and an intake valve capable of changing an internal EGR rate and an effective compression ratio;
A target fire type ignition timing setting means for setting a target fire type ignition timing which is a target for the ignition timing of the mixture for the fire type;
A target fire type ignition timing temperature setting means for setting a target fire type ignition timing temperature which is a target of the in-cylinder temperature at the target fire type ignition timing;
The way target spark ignition timing temperature is achieved, sets a target spark ignition timing effective compression ratio is the goal of effective compression ratio in the target internal EGR rate and the target spark ignition timing which is the target of the internal EGR rate, these And a control means for controlling the variable valve mechanism based on a target internal EGR rate and a target fire type ignition timing effective compression ratio. An in-cylinder state detecting means for detecting a state in the cylinder;
Fire type ignition timing temperature predicting means for calculating a predicted fire type ignition timing temperature which is a predicted value of the in-cylinder temperature at the target fire type ignition timing based on the output of the in-cylinder state detection means and the target fire type ignition timing effective compression ratio;
A difference value calculating means for calculating a difference value between the target fire type ignition timing temperature and the predicted fire type ignition timing temperature;
2. The control device for an internal combustion engine according to claim 1, wherein the control means sets the target internal EGR rate by feedback control based on the difference value calculated at the previous combustion cycle. The control means sets a target internal EGR rate so that the difference value converges to 0 by a feedback control algorithm including an integral term, and a target fire type ignition timing so as to minimize the pump loss based on the target internal EGR rate. The control apparatus for an internal combustion engine according to claim 2, wherein an effective compression ratio is set.