JP2015098800A - Control device for compression ignition type engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a control device for realizing a compression self-ignition combustion over a wider range while keeping an exhaust performance high.SOLUTION: There is provided a catalyst device containing a three-way catalyst. In a low load side compression self-ignition area A1 having an engine load less than a specific load T2, an internal EGR is executed, and the air/fuel ratio of the mixture is made leaner than the stoichiometric air fuel ratio. In a high load side compression spontaneous ignition range A2 of a specific load T2 or higher, the air/fuel ratio of the mixture is set at a stoichiometric air/fuel ratio, and the specific load T2 is made lower for the higher engine speed.

Description

  The present invention relates to a control device for a compression ignition engine.

  Conventionally, for the purpose of improving fuel efficiency and exhaust performance, an air-fuel mixture in which the air-fuel ratio is set leaner than the stoichiometric air-fuel ratio in a combustion chamber formed in the engine body has been subjected to compression self-ignition combustion. ing.

  For example, Patent Document 1 discloses an engine system that performs compression auto-ignition combustion of a lean air-fuel mixture, and performs compression auto-ignition combustion only in a predetermined operation region of low rotation and low load, and sparks in other operation regions. An apparatus for performing ignition combustion is disclosed.

  JP 2009-091994 A

  Here, when the compression auto-ignition combustion of the air-fuel ratio lean air-fuel mixture is performed under the operating condition where the combustion temperature becomes high, there is a problem that the amount of NOx generated by the combustion increases and the exhaust performance deteriorates. On the other hand, in the apparatus disclosed in Patent Document 1, since compression auto-ignition combustion is performed only in a low load / low rotation region where the combustion temperature is relatively low, an increase in the amount of NOx can be avoided. it can. However, this apparatus has a problem that the fuel consumption performance cannot be sufficiently improved because the compression self-ignition combustion is performed only in a part of the region.

  The present invention has been made in view of such a point, and provides a control device for a compression ignition engine capable of realizing compression auto-ignition combustion in a wider range while maintaining high exhaust performance.

  In order to solve the above-mentioned problems, the present invention provides a cylinder in which a combustion chamber in which an air-fuel mixture containing at least fuel and air is burned is formed, an intake port for introducing intake air into the cylinder, Provided in an exhaust passage connected to the exhaust port, an engine body having an exhaust port for exhausting exhaust gas from the engine, an intake valve capable of opening and closing the intake port, and an exhaust valve capable of opening and closing the exhaust port. A catalyst device including an original catalyst, and a control means for controlling the combustion mode of the air-fuel mixture and the air-fuel ratio of the air-fuel mixture in the combustion chamber, wherein the control means is a low load region where at least the engine load is lower than a predetermined load Then, in the compression auto-ignition region in which the combustion mode is compression auto-ignition combustion and the compression auto-ignition combustion is performed, the combustion is performed in the low-load side compression auto-ignition region where the engine load is less than a specific load. The internal EGR is performed so that a part of the burned gas generated in the step remains in the combustion chamber and the air-fuel ratio of the air-fuel mixture in the combustion chamber is made leaner than the stoichiometric air-fuel ratio, while the engine load is higher than the specific load. In the high load side compression auto-ignition region, the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio, and the specific load is set to a lower value as the engine speed is higher. A control device is provided.

  According to the present invention, even in an operation region where the combustion temperature is high, compression auto-ignition combustion can be realized while suppressing the amount of NOx finally discharged to the outside, and in a wider range while maintaining high exhaust performance. Compressed self-ignition combustion can be realized.

  Specifically, in the low-load side compression auto-ignition region where the engine load and combustion temperature are relatively low and the amount of NOx produced by combustion is low, compression auto-ignition combustion is performed while the air-fuel ratio is leaner than the stoichiometric air-fuel ratio. Therefore, fuel efficiency can be improved in this region. In the high load side compression auto-ignition region where the engine load and the combustion temperature are relatively high, the amount of NOx generated by combustion is relatively large. In the present invention, the air-fuel ratio is reduced in this high load side compression auto-ignition region. Compressed self-ignition combustion is carried out with a stoichiometric air-fuel ratio capable of purifying NOx by a three-way catalyst. Therefore, the amount of NOx that is finally emitted to the outside is purified by the three-way catalyst. It is possible to obtain the fuel efficiency improvement effect associated with the execution of the compression auto-ignition combustion while keeping the exhaust performance high while suppressing the amount low.

  In particular, the temperature of the air-fuel mixture, the combustion temperature, and the amount of NOx generated by the combustion increase as the engine speed increases and the engine load increases. On the other hand, the load that becomes the boundary between the low load side compression autoignition combustion and the high load side compression autoignition combustion, that is, the specific load that serves as a reference for switching the air-fuel ratio between lean and stoichiometric air-fuel ratio, Since the higher the value, the lower the value, the higher the rotation speed and the higher load region where the amount of NOx increases, the more reliably the NOx discharged to the outside with the air-fuel ratio as the stoichiometric air-fuel ratio can be reliably reduced.

  In addition, the internal EGR is performed in the low load side compression self-ignition region so that high-temperature burned gas remains in the combustion chamber. Therefore, in the low-load side compression auto-ignition region where the engine load is low and the ignitability of the mixture tends to deteriorate, the air-fuel ratio of the mixture is leaner than the stoichiometric air-fuel ratio, and a large amount of low-temperature fresh air is introduced. Temperature can be raised appropriately and stable compression auto-ignition combustion can be realized.

  Here, since the air-fuel mixture and the combustion temperature are low in a low engine speed range where the engine speed is sufficiently low, the amount of NOx generated can be suppressed to a low level even when the air-fuel ratio lean compression auto-ignition combustion is performed. Therefore, in the present invention, the control means includes the burned gas generated in the combustion chamber in the entire region of the compression auto-ignition region in the low-rotation side compression auto-ignition region where the engine speed is less than the specific engine speed. It is preferable to carry out internal EGR that causes a part of the air to remain in the combustion chamber and to make the air-fuel ratio of the air-fuel mixture in the combustion chamber leaner than the stoichiometric air-fuel ratio (claim 2).

  In this way, it is possible to further improve fuel efficiency while suppressing the amount of NOx discharged to the outside.

  In the present invention, there is provided an external EGR means for recirculating a part of the burned gas generated in the combustion chamber to the intake air, and the control means is configured to provide the external EGR means in the high load side compression auto-ignition region. It is preferable to introduce a part of the burned gas into the combustion chamber.

  In this way, in the high load side compression auto-ignition region, the combustion temperature can be kept low to reduce the amount of NOx generated during combustion, and the combustion can be suppressed more reliably and more reliably and stably. Combustion can be realized. This more reliably increases the exhaust performance and the fuel consumption performance. Further, by performing EGR, the unburned gas contained in the EGR gas is burned again, so that the fuel can be efficiently consumed, and the fuel efficiency can be improved even in the high load side compression auto-ignition region.

  Further, the control means implements the internal EGR while introducing a part of the burned gas into the combustion chamber by the external EGR means in the high load side compression auto-ignition region, and a part of the burned gas is concerned. While remaining in the combustion chamber, the ratio of the residual burned gas by the internal EGR in the high load side compression autoignition region to the total air-fuel mixture is determined as the mixture of the residual burned gas by the internal EGR in the low load side compression autoignition region. It is preferable to make it smaller than the ratio with respect to the whole (Claim 4).

  If it does in this way, in the high load side compression auto-ignition area | region, the temperature of air-fuel | gaseous mixture can be made into appropriate temperature by the combination of internal EGR and external EGR.

  Further, the control means opens the exhaust valve at least during the exhaust stroke and the intake stroke, and causes the burned gas discharged to the exhaust port side to flow back into the cylinder, thereby causing the burned gas into the cylinder. It is preferable to leave it (claim 5).

  In this way, compared with the case where the burned gas is not exhausted to the exhaust port side and remains in the cylinder, the cooling loss can be suppressed to be small and the fuel efficiency can be improved more reliably.

  Specifically, when a method of closing the exhaust valve during the exhaust stroke is used as a method for causing the burned gas to remain in the cylinder, the burned gas is compressed and heated after the exhaust valve is closed. Although a relatively large cooling loss occurs because the heated burned gas is cooled by the cylinder wall, the burned gas discharged to the exhaust port side by opening the exhaust valve during the exhaust stroke and the intake stroke In the method of causing the gas to flow back into the cylinder, the combustion loss is not compressed at high temperature, so that the cooling loss can be kept small.

  As described above, according to the present invention, compression auto-ignition combustion can be realized in a wider range while maintaining high exhaust performance.

1 is a schematic view showing an engine system according to an embodiment of the present invention. It is a block diagram which concerns on control of the engine system shown in FIG. It is sectional drawing which expands and shows the combustion chamber shown in FIG. (A) It is the figure which showed the lift characteristic of the exhaust valve in normal mode. (B) It is the figure which showed the lift characteristic of the exhaust valve in a special mode. It is a figure which illustrates the operation control map of an engine. It is the figure which showed the relationship between an engine load and the breakdown of the gas in a cylinder. It is the figure which showed the relationship between an engine load and the breakdown of the gas in a cylinder.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings.

  FIG. 1 is a schematic configuration diagram of an engine system 100 to which a control device for a compression ignition engine according to the present invention is applied. The engine system 100 is mounted on a vehicle and has an engine body 1.

  The engine body 1 is a gasoline engine to which a fuel containing at least gasoline is supplied, and is a four-cycle engine, that is, an engine in which an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke are sequentially performed. The engine body 1 is a compression ignition type engine in which compression self-ignition combustion is performed. The engine body 1 includes a cylinder block 11 provided with a cylinder 18 and a cylinder head 12 disposed on the cylinder block 11. The engine body 1 has, for example, four cylinders 18.

  A piston 14 connected to the crankshaft 15 via a connecting rod 142 is fitted in each cylinder 18 so as to be able to reciprocate. In each cylinder 18, a combustion chamber 19 surrounded by the inner surface of the cylinder 18 and the top surface of the piston 14 is formed.

  Although the specific structure of the piston 14 and the combustion chamber 19 is not specifically limited, For example, it has a structure as shown in FIG. In the example shown in FIG. 3, the center of the top surface of the piston 14 is a so-called reentrant type that is recessed in a direction away from the cylinder head 12 and becomes shallower after the depth becomes deeper from the center toward the radially outer side. The cavity 141 is formed.

  In the present embodiment, the geometric compression ratio of the engine body 1 is set to a relatively high value of 15 or more for the purpose of improving the thermal efficiency and stabilizing the compression ignition combustion. The geometric compression ratio of the engine body 1 is not limited to this, but a range of about 15 or more and 20 or less is preferable.

  In the cylinder head 12, an intake port 16 for introducing intake air into the cylinder 18 and an exhaust port 17 for discharging exhaust gas from the cylinder 18 are formed for each cylinder 18. The intake port 16 and the exhaust port 17 are provided with these ports, specifically, an intake valve 21 and an exhaust valve 22 that open and close the openings of the ports 16 and 17 formed in the cylinder head 12, respectively. Yes.

  The exhaust valve 22 is driven by an exhaust valve drive mechanism 70a. The exhaust valve drive mechanism 70 a includes an exhaust valve lift variable mechanism (hereinafter referred to as “exhaust VVL (variable valve lift)”) 71 and an exhaust phase variable mechanism (hereinafter referred to as “exhaust VVT (variable valve timing)”) 75.

  The exhaust VVL 71 switches the operation mode of the exhaust valve 22 between a normal mode indicated by a solid line in FIG. 4A and a special mode indicated by a solid line in FIG. That is, the lift characteristic of the exhaust valve 22 is switched between a first characteristic indicated by a solid line in FIG. 4A and a second characteristic indicated by a solid line in FIG. In the normal mode, the valve lift of the exhaust valve 22 gradually increases after opening, and when it reaches the maximum lift, it gradually decreases again to zero. In the special mode, the valve lift of the exhaust valve 22 gradually increases after the valve opening and reaches the maximum lift during the first valve opening period t_1 as in the normal mode. Without reaching zero as it is, the lift amount, that is, the lift lower than the maximum lift in the first valve opening period t_1 is maintained for a predetermined period, and then reaches zero. As described above, in the special mode, the opening period of the exhaust valve 22, that is, the period t_3 from when the exhaust valve 22 is opened to when the exhaust valve 22 is finally closed (during the intake stroke in this embodiment) is a predetermined period. A first valve opening period t_1 that is the maximum lift of the second valve, and a second valve configured to be smaller than the maximum lift in the first valve opening period t_1 continuously from the first valve opening period t_1. It consists of a valve opening period t_2. In the special mode, the valve is opened from immediately before the closing timing in the normal mode to a predetermined timing that is retarded from the closing timing in the normal mode, and the opening period of the exhaust valve is longer in the special mode than in the normal mode. It is getting longer. The exhaust VVL 71 includes a first cam and a second cam having different cam shapes in order to realize these modes. The first cam has a shape corresponding to the lift characteristic indicated by the solid line in FIG. 4A and has one cam crest. The second cam has a shape corresponding to the lift characteristic indicated by the broken line in FIG. 4B, and has two cam peaks. The exhaust VVL 71 includes a lost motion mechanism that selectively transmits the operating state of the first cam and the second cam to the exhaust valve 22. By transmitting the operating state of the first cam to the exhaust valve 22, the exhaust valve 22. The operation state is set to the normal mode, and the operation state of the exhaust valve 22 is set to the special mode by transmitting the operation state of the second cam to the exhaust valve 22. The exhaust VVL 71 is, for example, a hydraulically operated type. When the operating state of the second cam is transmitted to the exhaust valve 22, the lift characteristic of the exhaust valve 22 has a shape shown by a broken line in FIG.

  The exhaust VVT 75 changes the opening timing and closing timing of the exhaust valve 22 by changing the rotation phase of the exhaust camshaft with respect to the crankshaft 15. The exhaust valve VVT75 changes the valve opening timing and the valve closing timing of the exhaust valve 22 while maintaining the valve opening period of the exhaust valve 22 constant in each mode of the normal mode and the special mode. The exhaust VVT 75 may employ a hydraulic, electromagnetic, or mechanical known structure as appropriate, and a detailed description thereof will be omitted.

  The exhaust VVT 75 sets the rotational phase of the exhaust camshaft so that the exhaust valve 22 opens in the intake stroke in addition to the exhaust stroke when the operating state of the exhaust valve 22 is in the special mode. Further, when the exhaust valve 22 is in the special mode, the exhaust VVT 75 has the intake top dead center during the second valve opening period t_2, that is, the valve of the exhaust valve 22 at the intake top dead center. The rotational phase of the exhaust camshaft is set so that the lift becomes a relatively small value realized during the second valve opening period t_2. Thus, in the present embodiment, the exhaust valve 22 is opened during the intake stroke in addition to the exhaust stroke when the operating state of the exhaust valve 22 is set to the special mode. In particular, in the present embodiment, the exhaust valve 22 is continuously opened in the exhaust stroke and the intake stroke with the intake top dead center being sandwiched without closing in the middle. Here, when the exhaust valve 22 is continuously opened across the intake top dead center, the exhaust valve 22 and the piston 14 may interfere with each other. On the other hand, in this embodiment, as described above, the valve lift amount of the exhaust valve 22 near the intake top dead center is suppressed to a small value, so that interference between the exhaust valve 22 and the piston 14 is avoided. Can do. The exhaust double opening, that is, the special mode is performed in order to perform a so-called internal EGR by leaving the high-temperature burned gas, that is, the internal EGR gas in the combustion chamber 19. Specifically, when the exhaust is opened twice and the exhaust valve 22 is opened even during the intake stroke, the exhaust once discharged to the exhaust port 17 during the exhaust stroke is in the combustion chamber 19 during the intake stroke. Then, exhaust gas, that is, high-temperature burned gas remains in the combustion chamber 19.

  Here, as a method of causing the high-temperature burned gas to remain in the combustion chamber 19, that is, a method of internal EGR, there is a method of closing the exhaust valve 22 during the exhaust stroke and confining the burned gas in the combustion chamber 19. However, as compared with this method, the internal EGR due to the double opening of the exhaust gas can reduce the cooling loss. For this reason, in this embodiment, internal EGR is performed by opening the exhaust twice in order to improve the thermal efficiency of the entire system, that is, the fuel efficiency.

  Specifically, when a negative overlap period is provided across the intake top dead center, the burnt gas is compressed and heated to a temperature after the exhaust valve 22 is closed until the intake top dead center. Due to the high temperature, a temperature difference is generated between the temperature of the burned gas and the wall surface of the cylinder. When the temperature difference occurs, the burned gas is cooled by the engine coolant that cools the cylinder wall surface, specifically, the cylinder wall surface. By this cooling, the thermal energy of the burned gas is reduced, and the temperature of the burned gas is lowered.

  In contrast, when the exhaust is opened twice, the exhaust valve 22 is opened across the intake top dead center, and the burnt gas is not compressed at high temperature. Therefore, when the exhaust is opened twice, the temperature drop of the burned gas accompanying the cooling of the cylinder wall surface, that is, the engine cooling water can be suppressed, and the cooling loss can be suppressed small.

  The intake valve 22 is driven by an intake valve drive mechanism 70b. Similarly to the exhaust valve drive mechanism 70a, the intake valve drive mechanism 70b changes the rotation phase of the intake camshaft with respect to the intake VVL 74 and the crankshaft 15 to switch the operation mode of the intake valve 21 in two modes, and opens the intake valve 21. Intake VVT72 which changes valve timing and valve closing timing is included.

  The intake VVL 74 selectively selects the large lift cam that relatively increases the valve lift of the intake valve 21, the small lift cam that relatively decreases the valve lift of the intake valve 21, and the operating state of one of these cams. And a lost motion mechanism that transmits to the intake valve 21. The intake VVL 74 transmits the operation state of the large lift cam to the intake valve 21, thereby setting the operation mode of the intake valve 21 to a mode in which the valve lift and the valve opening period are relatively large. The intake VVL 74 transmits the operation state of the small lift cam to the intake valve 21, thereby setting the operation mode of the intake valve 21 to a mode in which the valve lift and the valve opening period are relatively small. The large lift cam and the small lift cam are set to be switched at the same valve closing timing or valve opening timing.

  As with the exhaust VVT 75, the intake VVT 72 may employ a known hydraulic, electromagnetic, or mechanical structure as appropriate, and illustration of the detailed structure is omitted.

  An intake passage 30 is connected to each intake port 16. Specifically, a branch passage that branches in correspondence with the cylinder 18 is formed at the downstream end of the intake passage 30, and these branch passages are connected to the intake ports 16.

  In the intake passage 30, an air cleaner 31, a water-cooled intercooler / warmer 34, a throttle valve 36, and a surge tank 33 are arranged in this order from the upstream side.

  An intercooler bypass passage 35 that bypasses the intercooler / warmer 34 is connected to the intake passage 30. The intercooler bypass passage 35 is provided with an intercooler bypass valve 351 for adjusting the flow rate of air passing through the intercooler bypass passage 35 in order to adjust the temperature of fresh air flowing into the cylinder 18. Note that the intercooler / warmer 34 and its associated members may be omitted.

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

  The exhaust passage 40 is provided with an exhaust purification device that purifies harmful components in the exhaust gas. In the present embodiment, a direct catalyst 41 and an underfoot catalyst 42 are provided in order from the upstream side. Each of the direct catalyst 41 and the underfoot catalyst 42 includes a three-way catalyst that can purify CO, HC, and NOx under the condition that the air-fuel mixture has a stoichiometric air-fuel ratio.

  An EGR device 50 is provided between the intake passage 30 and the exhaust passage 40 to recirculate a part of the exhaust gas to the intake air, that is, to perform external EGR. The EGR device 50 includes an EGR passage 51, an EGR cooler 52, and an EGR cooler bypass passage 53. The EGR passage 51 connects a portion of the intake passage 30 between the surge tank 33 and the throttle valve 36 and a portion of the exhaust passage 40 upstream of the direct catalyst 41. The EGR cooler 52 is for cooling the gas passing through the EGR passage 51, and is interposed in the EGR passage 51. The EGR cooler bypass passage 53 is a passage that bypasses the EGR cooler 52, and connects the upstream and downstream portions of the EGR cooler 52 in the EGR passage 51. The EGR passage 51 and the EGR cooler bypass passage 53 are provided with an EGR valve 511 and an EGR cooler bypass valve 531 for adjusting the flow rate of exhaust gas passing through the passages 51 and 53, respectively. Hereinafter, recirculation of a part of the exhaust gas to the intake air using the EGR device 50 is referred to as external EGR, and the exhaust gas recirculated to the intake air by the EGR device 50 may be referred to as external EGR gas.

  An injector 67 that directly injects fuel into the combustion chamber 19 is attached to the cylinder head 12 for each cylinder 18. As shown in FIG. 3, the injector 67 is disposed so that its nozzle hole faces the inside of the combustion chamber 19 from the central portion of the ceiling surface of the combustion chamber 19, and is opposed to the cavity 141. In the present embodiment, the injector 67 is a multi-hole type having a plurality of nozzle holes. The fuel spray injected from the injector 67 spreads radially from the center position of the combustion chamber 19.

  Here, when the fuel is injected from the injector 67 at the timing when the piston 14 is positioned near the compression top dead center, the fuel spray flows along the wall surface of the cavity 141 as shown by the arrow in FIG. Therefore, in the engine system 100, when performing high-pressure retarded injection, which will be described later, the fuel spray can be diffused earlier and an air-fuel mixture can be formed earlier.

  Fuel is supplied to the injector 67 from a fuel tank (not shown) by the fuel supply system 62. The fuel supply system 62 includes a fuel pump 63 and a pressure accumulation rail 64. The fuel pump 63 pumps fuel from the fuel tank to the pressure accumulation rail 64. In the present embodiment, the fuel pump 63 is a plunger type pump driven by the engine 1. The pressure accumulation rail 64 stores the pumped fuel at a relatively high pressure. The injector 67 injects high-pressure fuel stored in the pressure accumulation rail 64 into the combustion chamber 19. Although the value of an injection pressure is not specifically limited, For example, it is set to 30 MPa or more and 120 MPa or less.

  A spark plug 25 that forcibly ignites the air-fuel mixture in the combustion chamber 19 is attached to the cylinder head 12. In the present embodiment, the spark plug 25 is disposed through the cylinder head 12 so as to extend obliquely downward from the exhaust side of the engine body 1. As shown in FIG. 3, the tip of the spark plug 25 faces the cavity 141 of the piston 14 located at the compression top dead center.

  Each device is controlled by a power train control module (control means, hereinafter referred to as PCM) 10. The PCM 10 includes a microprocessor having a CPU, a memory, a counter timer group, an interface, and a path connecting these units.

  As shown in FIGS. 1 and 2, detection signals from various sensors SW <b> 1 to SW <b> 16 are input to the PCM 10.

The sensor SW1 is an air flow sensor SW1 that detects the flow rate of fresh air. The sensor SW2 is an intake air temperature sensor SW2 that detects the temperature of fresh air. The air flow sensor SW <b> 1 and the intake air temperature sensor SW <b> 2 are disposed on the downstream side of the air cleaner 31 in the intake passage 20. The sensor SW3 is a second intake air temperature sensor SW3 that detects the temperature of fresh air after passing through the intercooler / warmer 34, and is disposed on the downstream side of the intercooler / warmer 34. The sensor SW4 is an EGR gas temperature sensor SW4 for detecting the temperature of the external EGR gas, and is disposed in the vicinity of the connection portion with the intake passage 30 in the EGR passage 50. The sensor SW 5 is an intake port temperature sensor SW 5 that detects the temperature of intake air immediately before flowing into the cylinder 18, and is attached to the intake port 16. The sensor SW 6 is an in-cylinder pressure sensor SW 6 that detects the pressure in the cylinder 18 and is attached to the cylinder head 12. The sensor SW7 is an exhaust temperature sensor SW7 that detects the exhaust temperature. The sensor SW8 is an exhaust pressure sensor SW8 that detects the exhaust pressure. The exhaust temperature sensor SW7 and the exhaust pressure sensor SW8 are arranged in the vicinity of the connection portion of the EGR passage 50 in the exhaust passage 40. The sensor SW9 is a linear O ₂ sensor SW9 that detects the oxygen concentration in the exhaust gas, and is arranged on the upstream side of the direct catalyst 41 in the exhaust passage 40. The sensor SW 10 is a lambda O ₂ sensor SW 10 that detects the oxygen concentration in the exhaust gas, and is disposed between the direct catalyst 41 and the underfoot catalyst 42. The sensor SW11 is a water temperature sensor SW11 that detects the temperature of the engine cooling water. The sensor SW12 is a crank angle sensor SW12 that detects the rotation angle of the crankshaft 15. The sensor SW13 is an accelerator opening sensor SW13 that detects an accelerator opening corresponding to an operation amount of an accelerator pedal (not shown) of the vehicle. Sensors SW14 and SW15 are intake-side and exhaust-side cam angle sensors SW14 and SW15, respectively. The sensor SW 16 is a fuel pressure sensor SW 16 that detects the pressure of the fuel supplied to the injector 67, and is attached to the common rail 64.

  The PCM 10 performs various calculations based on the detection signals of the sensors SW1 to SW16. The PCM 10 determines the operating conditions of the engine body 1 and the vehicle based on these detection signals. The PCM 10 controls the actuators of the injector 67, the spark plug 25, the fuel supply system 62, and various valves (throttle valve 36, intercooler bypass valve 351, EGR valve 511, EGR cooler bypass valve 531) according to the operating conditions. Output signals to control them. The PCM 10 outputs control signals to the intake VVT 72, the intake VVL 74, the exhaust VVT 75, and the exhaust VVL 71 in accordance with operating conditions, and controls these, the intake valve 21, and the exhaust valve 22.

  FIG. 5 shows a control map of the engine speed on the horizontal axis and the engine load on the vertical axis. As described above, the engine body 1 performs the compression auto-ignition combustion in which the air-fuel mixture is self-ignited and burned without ignition by the spark plug 25. However, when compression auto-ignition combustion is performed in an operation region where the engine load is high, the temperature of the air-fuel mixture is high, so that combustion becomes steep and problems such as combustion noise occur. Therefore, in the engine system 100 according to the present embodiment, compression auto-ignition combustion is performed only in a low load region where the engine load is less than the predetermined first load T1, and ignition is performed in a high load region where the engine load is equal to or higher than the first load T1. Spark ignition combustion in which the air-fuel mixture is forcibly ignited by the plug 25 is performed. That is, in this engine system 100, the low load region is set to the CI (Compression Ignition) combustion region, and the high load region is set to the SI (Spark Ignition) combustion region. Note that the boundary lines of these combustion regions are not limited to the illustrated examples.

  The CI combustion region is further divided into two regions depending on the engine load. Below, the detailed control content of each area | region is demonstrated.

(1) First region (low load side compression auto-ignition region)
In the first region A1 in which the engine load is set to a region below the predetermined second load T2 in the CI fuel region, the air-fuel ratio of the air-fuel mixture becomes leaner than the stoichiometric air-fuel ratio, that is, the excess air ratio λ The G / F, which is the ratio of the total amount of the air-fuel mixture including the EGR gas to the fuel amount, is controlled to be 35 or more.

  In the first area A1, the external EGR is not performed and only the internal EGR is performed. That is, in the first region A1, the EGR valve 511 and the EGR cooler bypass valve 531 are closed, while the exhaust VVL 71 sets the operating state of the exhaust valve 22 to a special mode, and the exhaust is opened twice. For example, the internal EGR rate is controlled to 20% or more.

  In the first region A1, as shown in FIGS. 6 and 7, control is performed such that the internal EGR rate increases as the engine load decreases. Specifically, as the engine load is lower, the closing timing of the exhaust valve 22 is retarded, and the amount of burned gas that flows back to the combustion chamber 19 in the intake stroke is increased. 6 and 7 show changes in the breakdown of the gas in the cylinder 18 with respect to the engine load under operating conditions with different engine speeds. The horizontal axis in FIGS. 6 and 7 represents the engine load, and the vertical axis represents the amount of gas that can flow into the cylinder 18 as 100% and each gas (fresh air, internal EGR gas) with respect to this flowable gas amount. , External EGR gas) ratio. FIG. 6 shows a breakdown of the air-fuel mixture at the rotational speed NE1 in FIG. 5, and FIG. 7 shows a breakdown of the air-fuel mixture at the rotational speed NE2 in FIG. 5 at a higher speed than the rotational speed NE1. Yes. As shown in FIGS. 6 and 7, in the first region A1, the internal EGR rate is maintained at a maximum value below a predetermined load Tmin. In a region where the engine load is larger than the load Tmin, the engine load decreases. Thus, the internal EGR rate is gradually increased until it reaches the maximum value.

  The broken line m in FIGS. 6 and 7 is a line indicating the fresh air amount at which the excess air ratio λ = 1, and is a portion above the broken line m (a portion sandwiched between the broken line m and the 100% line). ) Is the amount of excess air ratio λ = 1 with respect to the fuel injected at each engine load. 6 and 7, in the first region A1, it is shown that the fresh air is located below the broken line m and the excess air ratio λ is at least 1 or more.

  Further, in the first region A1, intake stroke injection in which injection is performed by the injector 67 during the intake stroke is performed.

  In the present embodiment, as shown in FIG. 5, the second load T2 is set to a different value with respect to the engine speed. Specifically, the second load T2 is set to the same value as the first load T1 when the engine speed is equal to or less than a preset reference speed NE1, and the first load is increased when the engine speed is higher than the reference speed NE1. It is set to decrease as the engine speed increases from T1.

  That is, in the CI region, the low speed region where the engine rotational speed is lower than the reference rotational speed NE1 is entirely included in the first region A1, and in the operating region where the engine rotational speed is lower than the reference rotational speed NE1. The first region (= second load) is divided into a first region A1 and an SI combustion region.

  The above-described control is performed in the first area A1 for the following reason.

  In the first region A1, where the engine load is less than the second load T2, the amount of heat generated by the air-fuel mixture is small and the combustion temperature is also relatively low, so that NOx generated by combustion, that is, ROW NOx, is suppressed to a low level. Therefore, in the first region A1, it is not necessary to purify NOx with a three-way catalyst. That is, the air-fuel ratio does not have to be a stoichiometric air-fuel ratio that allows NOx purification by the three-way catalyst.

  Therefore, in the first region A1, the air-fuel ratio is made lean, that is, the excess air ratio λ> 1 and the fuel ratio G / F is 35 or more in order to improve fuel efficiency. Note that fuel efficiency is improved by making the air-fuel ratio lean because the amount of fresh air at a low temperature increases and the temperature of the air-fuel mixture and the combustion temperature are kept low, resulting in a reduction in cooling loss and exhaust loss. .

  Further, as described above, in order to avoid combustion noise and the like, it is preferable to lower the temperature in the combustion chamber 19, but if the temperature in the combustion chamber 19 is too low, the temperature of the air-fuel mixture is now reduced. There is a possibility that a stable compression self-ignition combustion cannot be realized due to a misfire or the like that does not increase to a temperature at which ignition is possible.

  Therefore, the engine load is a relatively low load less than the first load T1, and the heat generation amount of the air-fuel mixture is small. The first region A1 in which the temperature of the combustion chamber 19 and the air-fuel mixture cannot be sufficiently increased only by this heat generation amount. Then, in order to avoid misfire and the like, internal EGR is performed to leave high-temperature burned gas (internal EGR gas) in the combustion chamber 19, thereby increasing the temperature of the combustion chamber 19 and the air-fuel mixture.

  The temperature in the combustion chamber 19 and the air-fuel mixture becomes lower as the engine load is lower and the heat generation amount of the air-fuel mixture becomes smaller.

Therefore, in the first region A1, the internal EGR rate is increased as the engine load decreases, and the temperature of the air-fuel mixture is appropriately increased by the internal EGR according to the engine load, thereby mixing in the entire first region A1. Increase the ignitability of qi.

As a result of studies by the inventors of the present application, it has been found that if the EGR rate is increased to a predetermined value or more, the compression end temperature may be decreased. This is considered to be due to the following reason. If the internal EGR rate is increased to increase the ratio of the high-temperature internal EGR gas, the pre-compression temperature of the air-fuel mixture in the combustion chamber 19 is reliably increased. However, the internal EGR gas, that is, the burned gas, contains a large amount of triatomic molecules such as CO ₂ and H ₂ O, and is introduced into the cylinder (that is, nitrogen (N ₂ ) and oxygen (O ₂ )). The specific heat ratio is higher than Therefore, when the ratio of burnt gas introduced into the cylinder is increased by increasing the EGR rate excessively, the temperature of the air-fuel mixture before the compression starts increases, but the temperature of the air-fuel mixture does not increase so much even after compression. As a result, it is considered that the compression end temperature is lowered.

  Therefore, in the present embodiment, the internal EGR rate is increased as the engine load decreases as described above. However, the compression end temperature does not increase even if the internal EGR rate is further increased. When the value is reached, this predetermined value is maintained without further increasing the internal EGR rate. That is, as described above, the internal EGR rate is maintained at the maximum value below the predetermined load Tmin, thereby reliably increasing the compression end temperature.

  Here, as described above, the compression auto-ignition combustion can be performed with the air-fuel ratio lean while suppressing the generation of NOx in a region where the combustion temperature is low. The combustion temperature increases as the engine speed increases and as the engine load increases. Therefore, the second load, which is the upper limit load of the first region A1 in which this control is performed, is set to be lower as the engine speed increases. Further, in the operation region where the engine speed is sufficiently low, the combustion temperature can be kept low. In particular, in an operation region where the engine speed is sufficiently low, the amount of NOx produced can be reduced even if the engine is not capable of combustion noise exceeding an allowable value. Therefore, in the region where the engine speed is equal to or lower than the reference speed NE1, it is possible to reduce the amount of NOx generated even if the air-fuel ratio is lean over the entire region where the CI combustion is performed. Therefore, in the present embodiment, the value of the engine speed of the second load that is equal to or lower than the reference engine speed NE1 is set to the same value as the first load that is the upper limit load of the CI combustion, and from the reference engine speed NE1 in the CI combustion region. However, in the region where the engine speed is low, the compression auto-ignition combustion with the air-fuel ratio lean is performed in the entire region.

  Further, if the air and the fuel are sufficiently mixed, the air-fuel mixture can be self-ignited appropriately, that is, with high exhaust performance and high thermal efficiency. In the first region A1, since the temperature of the air-fuel mixture is relatively low, there is no possibility that the air-fuel mixture will ignite prematurely.

  Therefore, in the first region A1, the air-fuel mixture is appropriately self-ignited in the vicinity of the compression top dead center by injecting fuel during the intake stroke and mixing it with air in advance.

(2) Second region (high load side compression auto-ignition region)
In the second region A2, in which the engine load is set to a region equal to or greater than the second load T2 in the CI fuel region, unlike the first region, the air-fuel ratio of the air-fuel mixture becomes the stoichiometric air-fuel ratio, that is, the excess air ratio. Control is performed so that λ = 1. As shown in FIGS. 6 and 7, the excess air ratio λ = 1 is not realized by reducing the amount of fresh air by restricting the throttle valve 36, but introduces a large amount of EGR gas into the combustion chamber 19. It is realized by doing. That is, in the second region A2, all of the amount of gas that can flow into the cylinder 18 other than fresh air with an excess air ratio λ = 1 is made EGR gas.

  In the second region A2, external EGR is performed in addition to internal EGR. In particular, the burnt gas (cooled EGR gas) cooled through the EGR cooler is returned to the intake air. That is, in the second region A2, the operating state of the exhaust valve 22 is set to a special mode by the exhaust VVL 71, the exhaust is opened twice, the EGR cooler bypass valve 531 is closed, and the EGR valve 511 is opened. To be spoken.

  Also in the second region A2, as shown in FIGS. 6 and 7, the internal EGR rate is controlled to increase as the engine load decreases. However, in the second region A2, the internal EGR rate is made smaller than that in the first region A1. More specifically, the internal EGR rate is increased by a substantially constant increase amount as the engine load decreases in a region of the first region A1 and the second region A2 that is equal to or greater than the predetermined load Tmin. On the other hand, the external EGR rate is controlled so as to decrease slightly as the engine load decreases.

  In the second region A2, as in the first region A1, intake stroke injection in which injection is performed by the injector 67 during the intake stroke is performed.

  The above-described control is performed in the second region A2 for the following reason.

  In the second region A2 where the engine load is equal to or higher than the second load T2 and the engine load and the engine speed are higher, the heat generation amount of the air-fuel mixture is large and the combustion temperature is relatively high. Therefore, NOx (ROW NOx) generated by combustion cannot be sufficiently suppressed. Therefore, in the second region A2, NOx generated by combustion is purified by a three-way catalyst, so that NOx can be purified by a three-way catalyst so as to finally reduce NOx discharged from the vehicle to the outside. Thus, the excess air ratio λ = 1 is set with the air-fuel ratio of the mixture as the stoichiometric air-fuel ratio.

  Here, if the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio, the amount of fresh air having a low temperature is reduced and the temperature of the air-fuel mixture is increased as compared with the case where the air-fuel ratio is lean, resulting in an increase in cooling loss and exhaust loss. However, as a result of detailed examination of the difference in fuel efficiency performance due to the difference in air-fuel ratio during compression auto-ignition combustion, the present inventors have found that even when the air-fuel ratio is the stoichiometric air-fuel ratio, If the amount of fresh air is reduced by introducing a large amount of EGR gas, the cooling loss and exhaust loss will increase as compared with the case of lean air-fuel ratio, but the unburned fuel in EGR gas will burn in the combustion chamber. As a result, it has been found that fuel is efficiently consumed and unburned loss is reduced, and as a result, deterioration of fuel consumption can be minimized as compared with the case of lean air-fuel ratio.

  Therefore, in the present embodiment, in the second region A2, the air-fuel ratio of the air-fuel mixture is controlled to the stoichiometric air-fuel ratio, that is, the excess air ratio λ of the air-fuel mixture to λ = 1 while performing EGR. In particular, in the present embodiment, fuel consumption performance is enhanced by setting all of the amount of gas that can flow into the cylinder 18 other than fresh air with an excess air ratio λ = 1 as EGR gas.

  Here, as described above, the second region A2 has a larger calorific value and a relatively higher combustion temperature than the first region A1. Therefore, in the second region A2, the internal EGR rate is made smaller than that in the first region A1, and the cooled EGR gas is introduced into the cylinder 18 so that the temperature of the air-fuel mixture can be self-ignited and an appropriate temperature that does not cause abnormal combustion. To control. In this way, if the cooled EGR gas is introduced into the cylinder 18, the combustion temperature can be kept low and the amount of NOx produced during combustion can be kept small. Further, if the cooled EGR gas is introduced into the cylinder 18, it can flow into the cylinder 18 as described above while suppressing the amount of high-temperature internal EGR gas to be small and suppressing the temperature of the mixture from becoming excessively high. Among the various gas amounts, all except fresh air with an excess air ratio λ = 1 can be used as EGR gas.

  The reason why the internal EGR rate is controlled to increase as the engine load is lower in the second region A2 is the same as in the first region A1. That is, also in the second region A2, the temperature of the air-fuel mixture is appropriately increased by increasing the internal EGR rate in accordance with the heat generation amount of the air-fuel mixture becoming smaller as the engine load is lower.

  Also in the second region A2, as in the first region A1, fuel is injected during the intake stroke and mixed with air in advance in order to cause the air-fuel mixture to self-ignite appropriately, that is, in a state of high exhaust performance and thermal efficiency. The air-fuel mixture is appropriately self-ignited near the compression top dead center.

  In the second region A2, since the temperature of the air-fuel mixture is relatively high, depending on the type of engine or the like, there is a possibility of pre-ignition by injecting fuel during the intake stroke as described above. Therefore, in such a system in which premature ignition occurs, high-pressure retarded injection in which fuel is injected into the combustion chamber 19 by the injector 67 is performed during the period from the late stage of the compression stroke to the early stage of the expansion stroke, and a homogeneous mixture is compared. It may be formed in a short time to realize combustion in the expansion stroke period.

(3) SI Combustion Region Although the specific control content in the SI combustion region is not particularly limited, in this engine system 100, the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio in the SI combustion region, and the pre- For the purpose of avoiding abnormal combustion such as ignition and knocking, suppressing NOx generation, and reducing cooling loss, high-pressure retarded injection is performed, the internal EGR is stopped, and an external EGR that introduces cooled EGR gas into the combustion chamber 19 is provided. To be implemented. In the SI combustion region, the throttle valve 36 is fully opened to reduce the pump loss, and the amount of fresh air introduced into the cylinder 18 is adjusted by adjusting the opening of the EGR valve 511.

  The intake valve 21 may be controlled so that the air-fuel ratio and the internal EGR rate of the air-fuel mixture are realized as described above, and the control content is not particularly limited, but for example, by the intake VVL 54 The CI combustion region is a small lift and the SI combustion region is a large lift.

  As described above, in the present engine system 100, the internal EGR is reduced in the first region A1 where the engine load is less than the specific load T2 and in which the temperature of the mixture and the combustion temperature are low and the amount of NOx generated by combustion is small. As a result, the ignitability of the air-fuel mixture can be improved and stable compression auto-ignition combustion can be realized, and the air-fuel ratio of the air-fuel mixture is made leaner than the stoichiometric air-fuel ratio. While improving fuel efficiency. Further, in the second region A2 where the engine load is equal to or higher than the specific load T2 and the region where the temperature of the air-fuel mixture and the combustion temperature are relatively high and the amount of NOx generated by combustion is large, the air-fuel ratio of the air-fuel mixture is equal to the stoichiometric air-fuel ratio. Since the compression auto-ignition combustion is performed, the NOx purification by the three-way catalyst, that is, the amount of NOx finally discharged to the outside can be reduced while improving the fuel efficiency. Therefore, according to the engine system 100, the exhaust performance and the fuel consumption performance can be reliably improved.

  In addition, the second load T2 is set to a lower value as the engine speed is lower, and the air-fuel ratio of the air-fuel mixture is reliably increased in a region where the temperature and combustion temperature of the air-fuel mixture and the amount of NOx generated by combustion increase. Since the stoichiometric air-fuel ratio is set, NOx discharged to the outside can be surely suppressed to a low level.

  In addition, since the external EGR is performed in the second region A2, the temperature of the air-fuel mixture and the combustion temperature are likely to be high. In this second region A2, the increase in the combustion temperature is suppressed and the amount of NOx produced during combustion is reduced. In addition to being able to suppress the abnormal combustion, the abnormal combustion can be suppressed, and the improvement of the exhaust performance and the improvement of the fuel consumption performance accompanying the realization of the stable compression auto-ignition combustion can be realized. Further, the unburned gas contained in the EGR gas can be burned again in the combustion chamber 19, and the unburned loss can be kept small in the second region A2 to improve the fuel efficiency.

  Further, in the second region A2, in addition to the external EGR, the internal EGR is also performed. Therefore, by combining these EGRs, the temperature of the air-fuel mixture and the combustion temperature are more reliably set to appropriate temperatures, and the compression is performed more reliably. Self-ignition combustion can be realized.

  In addition, the introduction of these EGR gases can suppress the introduction of a fresh air amount into the cylinder 18 without reducing the throttle, so that the air-fuel ratio of the mixture becomes the stoichiometric air-fuel ratio, and the pumping loss can be suppressed. . This further increases fuel efficiency.

  In addition, since a method by double exhaust opening is used as the internal EGR, it is possible to suppress the temperature drop of the burned gas accompanying the cooling by the cylinder wall surface, that is, the engine cooling water, that is, to suppress the cooling loss. . This enhances the overall system thermal efficiency, i.e. fuel efficiency.

  In addition, this invention is not limited to embodiment mentioned above. For example, when fuel is injected during the intake stroke, the fuel may be injected into the intake port 16 by a port injector provided separately in the intake port 16 instead of the injector 67 provided in the cylinder 18.

  Further, regarding the valve train of the engine 1, instead of the VVL 74 of the intake valve 21, a CVVL (Continuously Variable Valve Lift) capable of continuously changing the lift amount may be provided.

  Further, the high-pressure retarded injection may be divided injection as necessary, and similarly, the intake stroke injection may also be divided injection as necessary. In these divided injections, fuel may be injected in each of the intake stroke and the compression stroke.

1 Engine (Engine body)
10 PCM (control means)
18 cylinder 21 intake valve 22 exhaust valve

Claims (5)

A cylinder in which a combustion chamber in which an air-fuel mixture containing at least fuel and air burns is formed; an intake port for introducing intake air into the cylinder; an exhaust port for exhausting exhaust gas from the cylinder; and the intake port An engine body having an intake valve capable of opening and closing, and an exhaust valve capable of opening and closing the exhaust port;
A catalyst device including a three-way catalyst provided in an exhaust passage connected to the exhaust port;
Control means for controlling the combustion mode of the air-fuel mixture and the air-fuel ratio of the air-fuel mixture in the combustion chamber,
The control means includes
At least in the low load region where the engine load is lower than the predetermined load, the combustion mode is compression self-ignition combustion,
In the compression auto-ignition region where the compression auto-ignition combustion is performed, in the low load side compression auto-ignition region where the engine load is less than a specific load, a part of the burned gas generated in the combustion chamber remains in the combustion chamber. While the internal EGR is performed and the air-fuel ratio of the air-fuel mixture in the combustion chamber is made leaner than the stoichiometric air-fuel ratio, the air-fuel ratio of the air-fuel mixture is higher in the compression load auto-ignition region where the engine load is higher than the specific load. Is the theoretical air-fuel ratio,
The control apparatus for a compression ignition engine, wherein the specific load is set to a lower value as the engine speed is higher. The control apparatus for a compression ignition engine according to claim 1,
The control means combusts a part of the burned gas generated in the combustion chamber in the entire region of the compression auto-ignition region in the low-rotation side compression auto-ignition region where the engine speed is less than a specific speed. A control device for a compression ignition engine, wherein internal EGR to be left in a chamber is performed and an air-fuel ratio of an air-fuel mixture in the combustion chamber is made leaner than a stoichiometric air-fuel ratio. The control apparatus for a compression ignition engine according to claim 1 or 2,
An external EGR means for recirculating a part of the burned gas generated in the combustion chamber to the intake air;
The control means is a control apparatus for a compression ignition engine, wherein in the high load side compression auto-ignition region, a part of the burned gas is introduced into a combustion chamber by the external EGR means. The control apparatus for a compression ignition engine according to claim 3,
The control means implements the internal EGR while introducing a part of the burned gas into the combustion chamber by the external EGR means in the high load side compression auto-ignition region, and a part of the burned gas is transferred to the combustion chamber. And the ratio of the residual burned gas by the internal EGR in the high load side compression autoignition region to the total mixture of residual burned gas by the internal EGR in the low load side compression autoignition region A control device for a compression ignition engine, characterized in that the ratio is smaller than the ratio to the engine. In the control device for the compression ignition engine according to any one of claims 1 to 4,
The control means opens the exhaust valve at least during an exhaust stroke and an intake stroke, and causes the burned gas discharged to the exhaust port side to flow back into the cylinder, thereby leaving the burned gas in the cylinder. A control device for a compression ignition type engine.

Images