JP5748822B2 - Combustion control device for compression self-ignition internal combustion engine - Google Patents

Description

  The present invention relates to a combustion control apparatus for a compression self-ignition internal combustion engine that self-ignites a mixture of air and fuel formed in a combustion chamber under a compression action of a piston under at least some operating conditions.

  In a compression self-ignition internal combustion engine, the air-fuel mixture formed by mixing air and fuel reaches the self-ignition temperature by being compressed by the piston, and occurs simultaneously at multiple locations in the combustion chamber space. Combustion starts. That is, the spark discharge used for ignition in the conventional gasoline combustion internal combustion engine is unnecessary in the compression self-ignition internal combustion engine.

  Here, the rise in the mixture temperature due to the compression of the piston is due to the adiabatic compression action. Therefore, in order to obtain a stronger adiabatic compression action and make the mixture temperature reach the self-ignition temperature, the compression self-ignition internal combustion engine is generally set at a higher compression ratio than the conventional spark ignition internal combustion engine. ing. However, it is difficult to raise the air-fuel mixture containing a large amount of room temperature air to the self-ignition temperature only by adiabatic compression. Therefore, by adjusting the exhaust valve closing timing to the advance side, the temperature of the air-fuel mixture is controlled by conventional spark ignition using means such as leaving a part of the high-temperature exhaust gas (hereinafter referred to as internal EGR gas). The invention has been devised to make it higher than in the case of a gasoline combustion internal combustion engine.

  In addition, raising the temperature of the gas mixture itself is relatively easily realized as described above. However, if the mixture temperature becomes higher than necessary, the combustion rate increases, combustion with vibration similar to knocking occurs, and if the mixture temperature becomes lower than necessary, ignition becomes unstable. Misfire occurs. For this reason, it is necessary to always control the mixture temperature to an appropriate temperature.

  As a first conventional technique for controlling the mixture temperature, the combustion timing estimated from the measured in-cylinder pressure and the appropriate combustion timing (target combustion timing) under the current operating conditions (operating state) determined in advance The internal EGR gas amount is adjusted by changing and controlling the exhaust valve closing timing and the intake valve opening timing in accordance with the deviation of the exhaust gas, and then the fresh air (air) and fuel supplied into the combustion chamber are converted into the internal EGR. There is one that adjusts the gas mixture temperature by mixing with gas (see, for example, Patent Document 1).

  In Patent Document 1, specifically, when the combustion timing estimated from the in-cylinder pressure is advanced from the appropriate combustion timing under the current operating conditions, the exhaust valve closing is performed to lower the mixture temperature. The internal EGR gas is reduced by changing and controlling the timing to the retard side or the intake valve opening timing to the advance side. On the other hand, if the combustion timing estimated from the in-cylinder pressure is retarded from the proper combustion timing under the current operating conditions, the exhaust valve closing timing should be advanced or increased to increase the mixture temperature. The internal EGR gas is increased by changing the intake valve opening timing to the retard side.

  Further, as a second conventional technique for controlling the air-fuel mixture temperature, in order to change the air-fuel mixture temperature without changing the internal EGR gas amount, a part of the internal EGR gas is transferred to the intake pipe having a relatively low temperature. There is what cools by making it reverse flow (for example, refer to patent documents 2).

  In Patent Document 2, specifically, after the internal EGR gas that has flowed back to the intake pipe is cooled by fresh air, it returns to the combustion chamber again, thereby suppressing an increase in the mixture temperature. Here, if there is no movement of heat to other than fresh air, fresh air superheated by the internal EGR gas flowing back to the intake pipe is also supplied to the combustion chamber at the same time. Therefore, there is no change in the amount of heat in the combustion chamber, and there is no change in the mixture temperature. However, in practice, as the heat transfer to other than fresh air, the heat transfer from the internal EGR gas flowing back to the intake pipe to the intake pipe wall, that is, the internal EGR gas by the intake pipe wall whose temperature is lower than the wall surface of the combustion chamber. Cooling occurs. Since the cooled internal EGR gas returns to the combustion chamber together with new air in the intake stroke, the internal EGR gas amount does not change from the initial value, and only the mixture temperature decreases.

JP-A-2005-139985 JP 2009-197740 A

However, the prior art has the following problems.
In the prior art described in Patent Document 1, when the combustion timing estimated from the in-cylinder pressure is advanced with respect to the appropriate combustion timing under the current operating conditions, the internal EGR gas amount is reduced and retarded. In this case, control is performed to increase the amount of internal EGR gas, and the control amount is corrected according to the deviation between the combustion timing estimated from the in-cylinder pressure and the appropriate combustion timing under the current operating conditions. Has been. However, a method for specifically determining the control amount is not disclosed.

  Furthermore, although the combustion timing is retarded by reducing the amount of internal EGR gas, the combustion rate is lowered due to a decrease in the mixture temperature, and the combustion rate is improved by reducing the amount of internal EGR gas that is an inert gas. Act simultaneously. Therefore, in practice, there is a problem that it is difficult to control the combustion timing to become the target combustion timing.

  Further, the prior art described in Patent Document 2 discloses that when the operating condition of the internal combustion engine is a high rotational speed and a high load, the intake valve opening timing is set according to the rotational speed or the load. Yes. However, it does not disclose a method for specifically setting the intake valve opening timing.

  Furthermore, the setting of the intake valve opening timing only takes into account the rotational speed or load of the internal combustion engine, control according to the amount of heat that requires cooling in the intake pipe, and internal EGR gas in the intake pipe or intake pipe. The control according to the heat transfer which changes with the state is not made. Accordingly, it is difficult to properly control the mixture temperature due to excessive or insufficient cooling of the internal EGR gas in the intake pipe, and as a result, premature combustion or misfire may occur. There is a problem that there is.

  The present invention has been made to solve the above-described problems. Combustion of a compression self-ignition internal combustion engine in which the mixture temperature is optimized and the combustion timing can be accurately controlled to a desired combustion timing. The object is to obtain a control device.

A combustion control apparatus for a compression self-ignition internal combustion engine according to the present invention is capable of self-igniting an air-fuel mixture formed in a combustion chamber by a compression action of a piston, a crank angle detection unit for detecting a crank angle, and a cylinder An in-cylinder pressure detection unit that detects an in-cylinder pressure that is an internal pressure, an intake pipe inner wall surface temperature detection unit that detects a wall surface temperature in the intake pipe, and an internal EGR gas that detects the temperature of the internal EGR gas remaining in the combustion chamber A combustion control device for executing combustion control on a compression self-ignition internal combustion engine having a temperature detection unit, an engine rotation number calculation unit that calculates an engine rotation number, and an engine load calculation unit that calculates an engine load The variable valve timing control unit and the crank angle detection unit that variably adjust the opening timing and closing timing of each of the intake valve and the exhaust valve are detected. Based on the rank angle and the in-cylinder pressure detected by the in-cylinder pressure detection unit, a combustion timing detection unit that detects a combustion timing, an engine speed and an engine load, and a first combustion timing associated with the target combustion timing A target combustion timing extraction unit that extracts the engine speed calculated by the engine rotation speed calculation unit and the engine load calculated by the engine load calculation unit from the data map, and the combustion timing detected by the combustion timing detection unit The combustion timing deviation calculating unit for calculating the combustion timing deviation, which is a deviation from the target combustion timing extracted by the target combustion timing extracting unit, and the combustion timing deviation calculated by the combustion timing deviation calculating unit so that the combustion timing deviation is zero. A calorific value estimation unit for estimating the amount of heat to be removed from the EGR gas, an internal EGR gas temperature detected by the internal EGR gas temperature detection unit, and an intake pipe inner wall surface temperature detection unit The intake pipe inner wall surface temperature difference calculation unit for calculating the intake pipe inner wall surface temperature difference, which is a temperature difference from the generated wall surface temperature, the intake pipe inner wall surface temperature difference calculation unit, and the heat quantity estimation unit estimated Based on the amount of heat, as the amount of internal EGR gas that flows back to the intake pipe, the backflow internal EGR gas amount estimation unit that estimates the amount of backflow internal EGR gas, the amount of backflow internal EGR gas, and the intake valve opening timing correction amount are associated An intake valve opening timing correction amount extraction unit that extracts an intake valve opening timing correction amount corresponding to the reverse flow internal EGR gas amount estimated by the reverse flow internal EGR gas amount estimation unit from the second data map thus obtained, and is variable When the absolute value of the combustion timing deviation calculated by the combustion timing deviation calculation unit is larger than a preset constant, the valve timing control unit compensates the intake valve opening timing extracted by the intake valve opening timing correction amount extraction unit. The intake valve opening timing is changed based on the positive amount.

  According to the present invention, a configuration is provided in which the intake valve opening timing is corrected and controlled in consideration of heat transfer characteristics, and an excessive amount of heat is generated from the air-fuel mixture without causing a change in combustion speed accompanying a change in the amount of inert gas. By controlling the intake valve opening timing so as to be removed, it is possible to obtain a combustion control device for a compression self-ignition internal combustion engine in which the mixture temperature is optimized and the combustion timing can be accurately controlled to a desired combustion timing. .

1 is a configuration diagram showing an outline of a compression self-ignition internal combustion engine in Embodiment 1 of the present invention. In Embodiment 1 of this invention, it is explanatory drawing which shows the change of the heat release rate with respect to a crank angle. In Embodiment 1 of this invention, it is explanatory drawing which shows the temperature change of the air-fuel | gaseous mixture in a combustion chamber with respect to a crank angle. In Embodiment 1 of this invention, it is explanatory drawing which shows the intake valve opening timing correction amount with respect to the internal EGR gas amount made to flow backward to the estimated intake pipe. 3 is a flowchart showing a series of operations of the compression self-ignition internal combustion engine in the first embodiment of the present invention. In Embodiment 2 of this invention, it is explanatory drawing which shows the heat transfer rate correction amount with respect to a pressure difference.

  Hereinafter, a combustion control apparatus for a compression self-ignition internal combustion engine according to the present invention will be described with reference to the drawings according to a preferred embodiment. In the description of the drawings, the same reference numerals are assigned to the same elements, and duplicate descriptions are omitted. In addition, the internal combustion engine may include an ignition plug, and may self-ignite the air-fuel mixture formed in the combustion chamber by a compression action of a piston under some operating conditions.

Embodiment 1 FIG.
FIG. 1 is a configuration diagram showing an outline of a compression self-ignition internal combustion engine according to Embodiment 1 of the present invention. An internal combustion engine used for driving a vehicle or the like generally has a plurality of combustion chambers. Here, in order to simplify the description of the operation, one of the plurality of combustion chambers is used. Only the configuration is shown.

  The compression self-ignition internal combustion engine in the first embodiment shown in FIG. 1 includes a crankshaft 1, a connecting rod 2, a piston 3, a cylinder 4, a combustion chamber 5, an intake valve 6, an intake pipe 7, an exhaust valve 8, and a valve operating valve. Mechanism 9 (variable lift mechanism 10 and variable phase mechanism 11), engine control unit (ECU) 12, valve mechanism control unit 13, fuel injection valve 14, fuel supply pipe 15, fuel injection control unit 16, exhaust pipe 17, crank Angle sensor (crank angle detection unit) 18, in-cylinder pressure sensor (in-cylinder pressure detection unit) 19, exhaust temperature sensor 20 (internal EGR gas temperature detection unit), intake pipe inner wall surface temperature sensor (intake pipe inner wall surface temperature detection unit) 21 The intake pressure sensor 22 (intake pressure detection unit) is provided.

  The engine control unit 12 performs overall control of the entire compression self-ignition internal combustion engine, and also includes an engine speed calculation unit 101, an engine load calculation unit 102, a variable valve timing control unit 103, a combustion timing detection unit 104, and a target combustion timing extraction unit 105. , Combustion timing deviation calculation unit 106, heat quantity estimation unit 107, intake pipe inner wall surface temperature difference calculation unit 108, backflow internal EGR gas amount estimation unit 109, intake valve opening timing correction amount extraction unit 110, heat transfer coefficient extraction unit 111, heat transfer amount An estimation unit 112 and a storage unit (not shown) are included.

  Further, in the storage unit of the engine control unit 12, the target combustion timing, fuel density, heat transfer coefficient, molecular weight of each substance, gas constant of each substance, specific heat ratio of each substance, air-fuel ratio, specific heat of each substance are estimated. A data map A in which an internal EGR gas amount to be backflowed and an intake valve opening timing correction amount are associated and a data map B in which a pressure difference and a heat transfer coefficient correction amount are associated are stored.

  Next, the operation of the compression self-ignition internal combustion engine in the first embodiment will be described. The compression self-ignition internal combustion engine in the first embodiment is a four-cycle internal combustion engine, and sequentially repeats four strokes of “intake”, “compression”, “expansion”, and “exhaust”. Further, the volume of the combustion chamber 5 changes as the piston 3 reciprocates along the cylinder 4 by the rotation of the crankshaft 1 and the action of the connecting rod 2.

  First, in the intake process, the intake valve 6 is gradually opened from the vicinity of the state where the piston 3 is most pushed in, and the piston 3 is pulled out, whereby air is sucked into the combustion chamber 5 through the intake pipe 7. In the compression self-ignition internal combustion engine, fuel is injected and supplied from the fuel injection valve 14 to the combustion chamber 5 during the intake stroke.

  The fuel is supplied after being pressurized to about 100 to 200 atm by a fuel boost pump (not shown) or the like via the fuel supply pipe 15. In the fuel injection supply, the fuel injection control unit 16 receives the control information of the injection amount and the injection timing determined by the arithmetic processing in the engine control unit 12 and supplies drive energy to the fuel injection valve 14. Be controlled.

  Here, the intake valve 6 and the exhaust valve 8 are driven in synchronization with the rotation of the crankshaft 1 by the camshaft of the valve mechanism 9. Further, the valve operating mechanism 9 includes a known lift variable mechanism 10 capable of continuously changing the valve lift amount on each of the intake side and the exhaust side, and a known lift mechanism capable of continuously changing the phase angle of the valve lift with respect to the crank rotation. The phase variable mechanism 11 is incorporated.

  Further, the valve lift control unit 13 receives the control information of the intake / exhaust valve opening / closing timing determined by the arithmetic processing in the variable valve timing control unit 103 in the engine control unit 12 for the valve lift amount and the phase angle, Control is performed by supplying drive energy to the variable lift mechanism 10 and variable phase mechanism 11. In other words, the valve mechanism control unit 13 supplies drive energy to the variable lift mechanism 10 and the variable phase mechanism 11 based on this control information, and operates the variable lift mechanism 10 and the variable phase mechanism 11. Accordingly, the valve mechanism control unit 13 can change the lift characteristics (that is, the valve lift amount and the phase angle) of the intake valve 6 and the exhaust valve 8, and the intake charge amount into the combustion chamber 5 and the internal EGR are changed. The amount of gas can be adjusted.

  Subsequently, in the compression process, when the intake valve 6 starts to close gradually from the middle of the extraction of the piston 3 and the intake valve 6 is completely closed in the vicinity where the piston 3 starts to be pushed again, the air taken into the combustion chamber 5 And the fuel are compressed while continuing mixing.

  In a conventional gasoline combustion internal combustion engine in which combustion is started by spark ignition, combustion of a mixture of air and fuel is started by an ignition plug or the like immediately before the piston 3 is most pushed. On the other hand, in a compression self-ignition internal combustion engine, combustion is started by adiabatic compression by strongly compressing the air-fuel mixture by the piston 3 until the temperature rises and self-ignition occurs.

  Subsequently, in the expansion stroke, when combustion is started, the pressure in the combustion chamber 5 suddenly rises and a force that pushes back the piston 3 acts. Therefore, the connecting rod 2 generates a rotational force on the crankshaft 1.

  Subsequently, in the exhaust process, the exhaust valve 8 is gradually opened from the vicinity where the piston 3 is most pushed back, and the high-temperature high-pressure combustion gas generated by the combustion is caused from the combustion chamber 5 through the exhaust pipe 17 by the piston 3 pushed back. Exhausted.

Next, the detection of combustion timing theta _a by combustion timing detecting section 104 of the engine control unit 12 will be described with reference to FIG. FIG. 2 is an explanatory diagram showing a change in the heat generation rate dQ / dθ with respect to the crank angle θ in the first embodiment of the present invention.

  Here, the heat generation rate dQ / dθ (hereinafter simply referred to as the heat generation rate) shown in FIG. 2 is calculated by, for example, the following equation (1) using the in-cylinder pressure P that is the pressure in the cylinder.

  In the above equation (1), P is the cylinder pressure, θ is the crank angle, V is the combustion chamber volume, and κ is the specific heat ratio of the air-fuel mixture. The in-cylinder pressure P is detected by the in-cylinder pressure sensor 19, and the crank angle θ is detected by the crank angle sensor 18. Further, the specific heat ratio κ of the air-fuel mixture is extracted from a database in the engine control unit 12.

Further, as shown in FIG. 2, if the value of the crank angle theta is theta _a higher heat generation rate begins to increase. Such an increase in the heat generation rate indicates that combustion occurs in the combustion chamber 5 and combustion is expanding. That is, the crank angle theta _a heat generation rate starts to increase can be handled as a time at which combustion is initiated. Therefore, combustion timing detecting section 104, the crank angle theta _a heat generation rate starts to increase, it is detected as a combustion timing theta _a in the current operating conditions.

Further, with respect to combustion timing theta _a in the current operating conditions, the desired (appropriate) combustion timing in the same operating conditions as the target combustion timing theta _b. The target combustion timing theta _b is calculated based on engine speed and engine load at the current operating conditions. Specifically, the engine speed calculation unit 101 calculates the engine speed based on the detection value from the crank angle sensor 18, and the engine load calculation unit 102 calculates the fuel injection calculated in the engine control unit 12. The engine load is calculated based on the quantity. Then, the target combustion timing extraction unit 105, the data map provided in advance, the calculated engine speed and extracts the target combustion timing theta _b corresponding to the engine load. This data map is the engine speed and engine load, which is the target combustion timing theta _b associated, it can be set in advance by the experiment or the like.

  Next, estimation of the temperature change amount ΔT to be given to the air-fuel mixture by the heat quantity estimation unit 107 in the engine control unit 12 will be described with reference to FIG. FIG. 3 is an explanatory diagram showing a temperature change of the air-fuel mixture in the combustion chamber 5 with respect to the crank angle θ in the first embodiment of the present invention.

Here, heat estimating unit 107, the deviation between combustion timing theta _a and the target combustion timing theta _b for calculating the combustion timing deviation calculating section 106 (hereinafter, referred to as combustion timing deviation [Delta] [theta]) is to control so as to zero The amount of heat Q to be removed from the internal EGR gas is estimated. The heat quantity estimation unit 107 estimates a temperature change amount ΔT corresponding to the heat quantity Q as a process in the middle of this.

In FIG. 3, the characteristic curve (1) corresponds to the temperature change of the air-fuel mixture in the combustion chamber 5 under the current operating conditions. Further, the characteristic curve (1), the gas mixture temperature in the combustion chamber 5 corresponding to the current combustion timing theta _a and T _1.

Further, in FIG. 3, the characteristic curve (2) corresponds to the assumption of the adiabatic change of the air-fuel mixture in which the air-fuel mixture temperature in the combustion chamber 5 becomes T ₁ at the target combustion timing θ _b . Further, the characteristic curve (2), the gas mixture temperature in the combustion chamber 5 corresponding to the current combustion timing theta _a and T _2.

Here, in order to optimize the combustion timing for the current operating conditions, by shifting the combustion timing theta _a just combustion timing deviation [Delta] [theta], it is sufficient to combustion timing theta _a coincides with the target combustion timing theta _b. That is, in order characteristic curve (1) is intended to adjust so that the temperature change characteristic temperature of mixture in the combustion chamber 5 at the target combustion timing theta _b is T ₁ (curve (2)), T ₁ and T A temperature change amount ΔT that is a deviation from ₂ may be given to the air-fuel mixture.

Next, a procedure for estimating the temperature change amount ΔT will be described. First, the mixture temperature T ₁ in the combustion chamber 5 at the combustion timing θ _a under the current operating conditions is estimated. Incidentally, the gas mixture temperature T ₁ of the combustion chamber 5 is also a gas mixture temperature in the combustion chamber 5 at the target combustion timing theta _b, it becomes the target combustion timing theta _b and the gas mixture temperature T ₁ of the combustion chamber 5 mixture Qi has a temperature characteristic as shown in the characteristic curve (2).

Accordingly, the air-fuel mixture temperature of mixture in the combustion chamber 5 at the combustion timing theta _a is T ₂ are, mixture temperatures T ₁ and comprising air-fuel mixture in the combustion chamber 5 at the target combustion timing theta _b (i.e., characteristic curve ( 2), and T ₂ can be estimated using T ₁ from the relationship of adiabatic change. Thereby, it is possible to estimate a temperature change amount ΔT that is a deviation between T ₁ and T ₂ .

More specifically, first, the temperature characteristic of the air-fuel mixture in the combustion chamber 5 at the current operating conditions, estimates the gas mixture temperature T ₁ of the combustion chamber 5 at the combustion timing theta _a. T ₁ is cylinder pressure P _.theta.a in combustion timing theta _a, combustion chamber volume V _.theta.a in combustion timing theta _a molar number _{N egr} internal EGR gas, moles _{N air} of the intake air, the gas constant _{R (R egr} and If R _air ) is used, it is estimated as in the following equation (2).

The gas constant R is extracted from a database in the engine control unit 12. The number of moles N _egr of the internal EGR gas is calculated based on the cylinder pressure at the exhaust valve closing timing, the combustion chamber volume, the exhaust temperature detected by the exhaust temperature sensor 20, and the gas constant R. Further, the number of moles N _air of the intake _air is calculated based on the air-fuel ratio, the molecular weight of the air and the fuel density, the engine speed, and the fuel injection amount extracted from the database provided in advance in the engine control unit 12. The

  Further, from the Poisson's law, regarding the combustion chamber volume V and the mixture temperature T in the adiabatic change, if the specific heat ratio of the mixture is κ, the relationship of the following equation (3) is established.

Further, when the above equation (3) is used, T ₂ is estimated as the following equation (4). V _θb is the combustion chamber volume at the target combustion timing θ _b .

  Further, when the above equation (2) and the above equation (4) are used, the temperature change amount ΔT to be given to the air-fuel mixture is estimated as the following equation (5).

Next, the amount of heat Q to be removed from the internal EGR gas by the heat amount estimation unit 107 and the internal EGR gas amount N _re to be back flowed by the back flow internal EGR gas amount estimation unit 109 (hereinafter referred to as back flow internal EGR gas amount N _re ) Will be described.

Here, the engine control unit 12 removes the heat quantity Q from the internal EGR gas so as to give the air-fuel mixture a temperature change amount ΔT. Further, the engine control unit 12, to remove the heat quantity Q from the internal EGR gas to flow back internal EGR gas by backflow internal EGR gas amount _{N re} estimated intake pipe 7.

The amount of heat Q to be removed from the internal EGR gas includes the specific heat C (C _egr and C _air ) of each substance, the number of moles N _egr of the internal EGR gas, the number of moles of intake air N _air , and the temperature change ΔT to be given to the mixture. When used, it is estimated as in the following formula (6). The specific heat C of each substance is extracted from a database in the engine control unit 12.

Further, the heat quantity Q to be removed from the internal EGR gas can be removed by the heat moving from the internal EGR gas flowing back into the intake pipe 7 to the intake pipe wall surface. The backflow internal EGR gas amount _Nre is estimated as shown in the following equation (7) when the heat amount Q to be removed from the internal EGR gas is used.

In addition, the intake pipe inner wall surface temperature difference ΔT _{in in} the above equation (7) corresponds to the exhaust temperature T _evc at the exhaust valve closing timing detected by the exhaust temperature sensor 20 (the temperature of the internal EGR gas remaining in the combustion chamber 5). ) and a temperature difference between the intake pipe wall temperature T _in is detected by the intake pipe wall temperature sensor 21, it is calculated by the intake pipe wall temperature difference calculation section 108.

Subsequently, FIG. 4, in the first embodiment of the present invention, is an explanatory diagram showing an intake valve opening timing correction amount for the estimated reverse flow internal EGR gas amount N _re. In other words, FIG. 4 corresponds to the data map A stored in the aforementioned database. This data map A has a heat transfer amount dQ backflow internal EGR gas amount N _re and below, which is an intake valve opening timing correction amount associated, be set in advance by performing experiments Can do.

The amount of heat transferred from the backflow internal EGR gas amount _Nre to the intake pipe wall surface varies depending on the state in the intake pipe 7. Therefore, the intake valve opening timing is corrected by further considering not only the estimated backflow internal EGR gas amount _Nre but also the heat transfer amount dQ moving from the internal EGR gas backflowed into the intake pipe 7 to the intake pipe wall surface. It is preferable to set to.

Here, the amount of heat transfer dQ that moves from the internal EGR gas that has flowed back into the intake pipe 7 per unit time to the intake pipe wall surface, when the intake pipe inner wall surface effective area A, the heat transfer coefficient h, and the temperature difference ΔT _in are used, It is estimated as in equation (8).

Since the intake pipe inner wall surface effective area A is set according to the shape of the intake pipe 7, it is constant regardless of the operating conditions, but the heat transfer coefficient h and the temperature difference ΔT _in vary depending on the operating conditions.

Further, the heat transfer coefficient extraction unit 111 calculates the engine speed calculated by the engine speed calculation unit 101 and the intake pipe inner wall surface temperature difference ΔT _in calculated by the intake pipe inner wall surface temperature difference calculation unit 108 from a data map provided in advance. The corresponding heat transfer coefficient h is extracted. This data map is obtained by associating the engine speed, the intake pipe inner wall surface temperature difference ΔT _in and the heat transfer coefficient h, and can be set in advance by performing an experiment or the like.

Further, the heat transfer amount estimation unit 112 is based on the heat transfer rate h extracted by the heat transfer rate extraction unit 111 and the intake pipe inner wall surface temperature difference ΔT _in calculated by the intake pipe inner wall surface temperature difference calculation unit 108 based on the above formula ( The heat transfer amount dQ is estimated according to 8). Characteristic curve of FIG. 4 shows the intake valve opening timing correction amount for the reverse flow internal EGR gas amount N _re in each heat transfer amount. In FIG. 4, the larger the heat transfer amount dQ, the greater the heat transfer amount per unit time of the backflow internal EGR gas.

Therefore, if the value of the estimated reverse flow internal EGR gas amount N _re is positive (increase side), the larger the amount of heat transfer dQ, for values of the same reverse flow internal EGR gas amount N _re, the intake valve opening timing Correction control is performed so that the amount of correction to the advance side is set as a smaller value. Further, when the value of the estimated reverse flow internal EGR gas amount N _re is negative (loss side), the larger the amount of heat transfer dQ, for values of the same reverse flow internal EGR gas amount N _re, the intake valve opening timing Correction control is performed so that the amount of correction to the retard side is set to a smaller value.

However, if the backflow internal EGR gas amount _Nre is too small, the amount of heat of the backflow internal EGR gas itself is small, so the ratio of the amount of heat transferred to the intake pipe wall surface with respect to the amount of heat of the backflow internal EGR gas increases, and the backflow internal EGR The temperature of the gas is greatly reduced and the temperature difference ΔT _in is reduced.

When the temperature difference ΔT _in is decreased from the above equation (8), the heat transfer amount dQ is also decreased, so that all the heat amount Q that should be removed from the internal EGR gas while staying in the intake pipe 7 is transferred to the wall surface of the intake pipe. I can't move. To prevent this, the intake valve opening timing correction amount in Fig. 4, in accordance with the reverse flow internal EGR gas amount N _re a heat transfer amount dQ, so that decrease of the temperature difference [Delta] T _in is suppressed It is set in consideration of that.

Then, the intake valve opening timing correction amount extraction unit 110, from FIG. 4, corresponding to the amount of heat transfer dQ backflow internal EGR gas amount _{N re} and heat transfer amount estimating unit 112 backflow internal EGR gas amount estimation unit 109 estimates estimated The intake valve opening timing correction amount is extracted. The variable valve timing control unit 103 optimizes the combustion timing by changing (correcting) the intake valve opening timing with the phase variable mechanism 11 based on the intake valve opening timing correction amount.

As described above, when the intake valve opening timing is set to be corrected, the combustion timing can be optimized by considering not only the backflow internal EGR gas amount _Nre but also the heat transfer amount dQ. Therefore, it is preferable. However, the effect of correcting the intake valve opening timing can be obtained without considering the heat transfer amount dQ. In such a case, it is not necessary to calculate the heat transfer amount dQ, and the data map A may be any as long as the backflow internal EGR gas amount _Nre and the intake valve opening timing correction amount are associated with each other.

  Next, a series of operations of the combustion control device for the compression self-ignition internal combustion engine according to Embodiment 1 of the present invention will be described with reference to the flowchart of FIG. FIG. 5 is a flowchart showing a series of operations of the compression self-ignition internal combustion engine according to the first embodiment of the present invention.

In step S <b> 100, the combustion timing detection unit 104 calculates a heat generation rate based on the in-cylinder pressure P detected by the in-cylinder pressure sensor 19. The combustion timing detecting section 104, a timing calculated heat release rate begins to increase, is detected as the combustion timing theta _a, the process proceeds to step S101.

In step S101, the target combustion timing extraction unit 105, an engine rotational speed calculated and the target combustion timing theta _b corresponding to the engine load, and extracted from the data map of time combustion has been set in advance, the process proceeds to step S102.

In step S102, the combustion timing deviation calculating unit 106, and a combustion timing theta _a detected in step S100, based on the target combustion timing theta _b extracted in step S101, calculates a combustion timing deviation [Delta] [theta], step S103 Proceed to

  In step S103, the combustion timing deviation calculation unit 106 determines whether or not the absolute value | Δθ | of the combustion timing deviation Δθ is larger than a preset constant α, and the absolute value | Δθ | is larger than the constant α. In step S104, if the absolute value | Δθ | is less than the constant α, the process returns to step S100.

  Here, if the absolute value | Δθ | is very small, hunting may occur by performing correction control of the intake valve opening timing, and therefore a dead zone is provided in advance to prevent this. Then, the minimum value of the absolute value | Δθ | at which hunting does not occur is set as a constant α, and correction control of the intake valve opening timing is performed only when | Δθ | is larger than the constant α. For example, the constant α that is set in advance is used as a crank angle of 0.5 deg. If hunting occurs, the constant α may be changed within a range of 0 <α ≦ 3.0.

In step S104, the amount of heat estimating unit 107, based on the cylinder pressure P _.theta.a and combustion chamber volume V _.theta.a corresponding to combustion timing theta _a, and the combustion chamber volume V _.theta.b corresponding to the target combustion timing theta _b, the air-fuel mixture The temperature change amount ΔT to be given to is estimated, and the process proceeds to step S105.

  In step S105, the heat quantity estimation unit 107 estimates the heat quantity Q to be removed from the internal EGR gas based on the temperature change amount ΔT to be given to the air-fuel mixture estimated in step S104, and proceeds to step S106.

In step S106, the backflow internal EGR gas amount estimating section 109, based on the quantity of heat Q to remove from the internal EGR gas is estimated in step S105, to estimate the reverse flow internal EGR gas amount _{N re,} the process proceeds to step S107.

In step S107, the heat transfer coefficient extraction unit 111 extracts the heat transfer coefficient h corresponding to the engine speed and the intake pipe inner wall surface temperature difference ΔT _in from the database, and proceeds to step S108.

In step S108, the heat transfer amount estimation unit 112 estimates the heat transfer amount dQ based on the heat transfer coefficient h extracted in step S107, the intake pipe inner wall surface temperature difference ΔT _in, and the intake pipe inner wall surface effective area A, Proceed to step S109.

In step S109, the intake valve opening timing correction amount extraction unit 110 performs intake air in accordance with the heat transfer amount dQ with respect to the internal EGR gas amount N _re to flow back to the intake pipe 7 based on the heat transfer amount dQ estimated in step S108. The intake valve opening timing correction amount is extracted from the relationship of the valve opening timing correction amount (data map A), and the process proceeds to step S110.

  In step S110, the variable valve timing control unit 103 instructs the phase variable mechanism 11 to correct the intake valve opening timing based on the intake valve opening timing correction amount extracted in step S109, and then proceeds to step S100. And return.

  As described above, according to the first embodiment, the amount of temperature change that should be given to the air-fuel mixture based on the deviation between the combustion timing estimated based on the in-cylinder pressure and the target combustion timing under the current operating conditions is determined. The amount of heat to be calculated and removed from the internal EGR gas is estimated from the relationship between the temperature change amount and the air-fuel mixture amount. Further, the amount of internal EGR gas that flows back to the intake pipe according to the estimated amount of heat to be removed from the internal EGR gas is estimated, and the amount of internal EGR gas that flows back and the heat transfer characteristics between the intake pipe wall and the internal EGR gas are calculated. In consideration, the intake valve opening timing is corrected and controlled. Therefore, the mixture temperature is optimized by correcting the intake valve opening timing so that excess heat is removed from the mixture without causing a change in the combustion speed due to the change in the amount of inert gas, and the combustion timing is adjusted. The desired combustion timing can be accurately controlled.

Embodiment 2. FIG.
In the first embodiment, the case where the heat transfer coefficient h is not changed (corrected) has been described. On the other hand, in the second embodiment, a case where the heat transfer coefficient h is changed (corrected) based on the pressure difference ΔP will be described. The configuration of the compression self-ignition internal combustion engine in the second embodiment is the same as that in the first embodiment, and a description thereof will be omitted.

Here, in the second embodiment, the in-cylinder pressure P _ivo at the intake valve opening timing detected by the crank angle sensor 18 and the in-cylinder pressure sensor 19 and the intake air detected by the crank angle sensor 18 and the intake pressure sensor 22. The heat transfer coefficient h is corrected based on the pressure difference ΔP (P _ivo -P in _ivo ) with _{respect to} the intake pressure (pressure in the intake pipe 7) P in _ivo at the valve opening timing. The second embodiment is technically characterized in that the intake valve opening timing is corrected based on the heat transfer amount dQ estimated according to the corrected heat transfer coefficient h. The heat transfer coefficient h is corrected by a heat transfer coefficient changing unit (not shown) in the engine control unit 12.

  FIG. 6 is an explanatory diagram showing the heat transfer coefficient correction amount with respect to the pressure difference ΔP in the second embodiment of the present invention. In other words, FIG. 6 corresponds to the data map B stored in the aforementioned database. In addition, this data map B associates the pressure difference ΔP with the heat transfer coefficient correction amount, and can be set in advance by performing an experiment or the like.

  Here, as the pressure difference ΔP increases, the amount of internal EGR gas that flows back into the intake pipe 7 per unit time increases, and the convection of the internal EGR gas increases in the intake pipe 7. In this case, since the heat transfer coefficient h increases as the convection of the fluid increases and the flow velocity increases, it is preferable to correct the heat transfer coefficient h according to the pressure difference ΔP. Specifically, the heat transfer coefficient h is corrected based on the heat transfer coefficient correction amount extracted from the relationship between the pressure difference ΔP and the heat transfer coefficient correction amount as shown in FIG. 6 based on the pressure difference ΔP. Do.

  Further, the combustion control device for the compression self-ignition internal combustion engine in the second embodiment differs from the series of operations of the combustion control device in the first embodiment after extracting the heat transfer coefficient h in step S107. Step S108 is performed after executing the process of correcting the extracted heat transfer coefficient h based on the pressure difference ΔP.

  That is, the combustion control apparatus for a compression self-ignition internal combustion engine according to the second embodiment executes steps S100 to S106 in FIG. 5 in the same manner as in the first embodiment. In step S107 in FIG. 5, the heat transfer coefficient extraction unit 111 extracts the heat transfer coefficient h, and before executing step S108, the heat transfer coefficient change unit performs the heat transfer extracted in step S107. The rate h is corrected based on the pressure difference ΔP.

Subsequently, when the correction of the heat transfer coefficient h by the heat transfer coefficient changing unit is completed, in step S108, the heat transfer amount estimating unit 112, based on the corrected heat transfer coefficient h, in the same manner as in the first embodiment. It estimates the heat transfer amount dQ, at step S109, the intake valve opening timing correction amount extraction unit 110, based on the back-flow amount of internal EGR gas _{N re} a heat transfer rate dQ extracts intake valve opening timing correction amount, in step S110 The variable valve timing control unit 103 instructs the phase variable mechanism 11 to correct the intake valve opening timing, and the entire control is executed in a flow of returning to step S100.

  As described above, by correcting the heat transfer coefficient h according to the heat transfer coefficient correction amount extracted based on the pressure difference ΔP, it is possible to improve the estimation accuracy of the heat transfer amount dQ. Therefore, the intake valve opening timing correction amount can be accurately extracted so as to remove an excessive amount of heat from the air-fuel mixture, so that the accuracy of controlling the combustion timing to a desired combustion timing can be improved.

  As described above, according to the second embodiment, by considering the pressure difference ΔP, the estimation accuracy for the heat transfer characteristic with the internal EGR gas is improved, and the effect of the first embodiment is further enhanced. Can do.

  DESCRIPTION OF SYMBOLS 1 Crankshaft, 2 Connecting rod, 3 Piston, 4 Cylinder, 5 Combustion chamber, 6 Intake valve, 7 Intake pipe, 8 Exhaust valve, 9 Valve mechanism, 10 Lift variable mechanism, 11 Phase variable mechanism, 12 Engine control part, DESCRIPTION OF SYMBOLS 13 Valve mechanism control part, 14 Fuel injection valve, 15 Fuel supply pipe, 16 Fuel injection control part, 17 Exhaust pipe, 18 Crank angle sensor, 19 In-cylinder pressure sensor, 20 Exhaust temperature sensor, 21 Intake pipe inner wall surface temperature sensor, 22 intake pressure sensors, 101 engine speed calculation unit, 102 engine load calculation unit, 103 variable valve timing control unit, 104 combustion timing detection unit, 105 target combustion timing extraction unit, 106 combustion timing deviation calculation unit, 107 heat quantity estimation unit, 108 Intake pipe inner wall surface temperature difference calculation unit, 109 Backflow internal EGR gas amount estimation unit, 110 Intake air Valve opening timing correction amount extraction unit, 111 heat transfer coefficient extraction unit, 112 heat transfer amount estimation unit.

Claims (3)

The air-fuel mixture formed in the combustion chamber can be self-ignited by the compression action of the piston,
A crank angle detector for detecting the crank angle;
An in-cylinder pressure detector that detects an in-cylinder pressure that is a pressure in the cylinder;
An intake pipe inner wall surface temperature detection unit for detecting the wall surface temperature in the intake pipe;
An internal EGR gas temperature detector for detecting the temperature of the internal EGR gas remaining in the combustion chamber;
An engine speed calculator for calculating the engine speed;
An engine load calculation unit for calculating the engine load;
A combustion control device for performing combustion control on a compression self-ignition internal combustion engine having
A variable valve timing control unit that variably adjusts the opening timing and closing timing of each of the intake valve and the exhaust valve; the crank angle detected by the crank angle detection unit; and the in-cylinder detected by the in-cylinder pressure detection unit A combustion timing detector that detects the combustion timing based on the pressure;
The engine speed calculated by the engine speed calculator and the engine load calculated by the engine load calculator from a first data map in which the engine speed and the engine load are associated with a target combustion timing. A target combustion timing extraction unit that extracts a target combustion timing corresponding to
A combustion timing deviation calculating section that calculates a combustion timing deviation that is a deviation between the combustion timing detected by the combustion timing detecting section and the target combustion timing extracted by the target combustion timing extracting section;
A calorific value estimation unit that estimates the amount of heat to be removed from the internal EGR gas so that the combustion timing deviation calculated by the combustion timing deviation calculation unit becomes zero;
Intake pipe inner wall surface temperature difference for calculating an intake pipe inner wall surface temperature difference which is a temperature difference between the temperature of the internal EGR gas detected by the inner EGR gas temperature detector and the wall surface temperature detected by the intake pipe inner wall surface temperature detector. A calculation unit;
Based on the intake pipe inner wall surface temperature difference calculated by the intake pipe inner wall surface temperature difference calculation unit and the heat quantity estimated by the heat quantity estimation unit, an internal EGR gas amount that flows back to the intake pipe is used as a back flow internal EGR gas amount. A backflow internal EGR gas amount estimation unit for estimating
The intake valve opening corresponding to the reverse flow internal EGR gas amount estimated by the reverse flow internal EGR gas amount estimation unit from the second data map in which the reverse flow internal EGR gas amount and the intake valve opening timing correction amount are associated with each other. An intake valve opening timing correction amount extraction unit for extracting a timing correction amount;
With
The variable valve timing controller is
Based on the intake valve opening timing correction amount extracted by the intake valve opening timing correction amount extraction unit when the absolute value of the combustion timing deviation calculated by the combustion timing deviation calculation unit is larger than a preset constant, A combustion control apparatus for a compression self-ignition internal combustion engine that changes a valve opening timing of the intake valve. The combustion control apparatus for a compression self-ignition internal combustion engine according to claim 1,
The engine speed and the intake pipe inner wall surface temperature difference calculation unit calculated by the engine speed calculation unit from the third data map in which the engine speed and the intake pipe inner wall surface temperature difference are associated with the heat transfer coefficient. A heat transfer coefficient extraction unit that extracts a heat transfer coefficient corresponding to the intake pipe inner wall surface temperature difference calculated by:
A heat transfer rate estimating unit that estimates a heat transfer amount based on the heat transfer rate extracted by the heat transfer rate extracting unit and the intake pipe inner wall surface temperature difference calculated by the intake pipe inner wall surface temperature difference;
Further comprising
The intake valve opening timing correction amount extraction unit,
From the fourth data map in which the heat transfer amount and the backflow internal EGR gas amount are associated with the intake valve opening timing correction amount, the heat transfer amount and the backflow internal EGR gas amount estimated by the heat transfer amount estimation unit are estimated. A combustion control apparatus for a compression self-ignition internal combustion engine that extracts the intake valve opening timing correction amount corresponding to the backflow internal EGR gas amount estimated by the engine. The combustion control apparatus for a compression self-ignition internal combustion engine according to claim 2,
An intake pressure detector that detects an intake pressure that is a pressure in the intake pipe;
Based on the pressure difference between the in-cylinder pressure detected by the in-cylinder pressure detection unit at the opening timing of the intake valve and the intake pressure detected by the intake pressure detection unit at the opening timing of the intake valve, A heat transfer coefficient changing unit that changes the heat transfer coefficient extracted by the heat transfer coefficient extracting unit;
Further comprising
The heat transfer amount estimation unit
Combustion control of a compression self-ignition internal combustion engine that estimates the amount of heat transfer based on the heat transfer coefficient changed by the heat transfer coefficient changing unit and the intake pipe inner wall surface temperature difference calculated by the intake pipe inner wall surface temperature difference calculating unit apparatus.

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