JP4274060B2 - diesel engine - Google Patents

Description

  The present invention relates to a diesel engine, and more particularly to a diesel engine that controls the combustion state of each cylinder in order to improve fuel efficiency and emissions in a multi-cylinder diesel engine.

  Conventionally, in a diesel engine, in order to improve fuel efficiency while improving emissions, preliminary injection is performed before main injection of fuel to uniformly diffuse the preliminary injection fuel into the combustion chamber, and the main injection fuel is In general, a technique for burning this uniformly diffused pre-injected fuel at once, a technique related to so-called premixed combustion is generally known. For example, as shown in Patent Document 1, preliminary injection for injecting fuel at the beginning of the intake stroke is executed, and main injection into the lean mixture by the pre-injected fuel is performed twice at a predetermined timing at the end of the compression stroke and at the beginning of the expansion stroke. There is known a diesel engine that performs combustion at once.

Further, in the diesel engine described in Patent Document 1, an external EGR passage (exhaust gas recirculation passage) is provided to increase the intake air temperature, and stable combustion is performed by changing the external EGR amount in accordance with the operating state of the engine. Various ways to improve emissions while improving fuel efficiency, such as improving emissions and adjusting fuel injection nozzles to prevent fuel from adhering to the wall surface of the cylinder liner to reduce HC emissions. Ingenuity has been proposed.
JP 2002-322922 A

  Here, in order to improve the emission property while ensuring the output, in particular, in order to suppress the generation of nitrogen oxide (NOx), it is preferable to burn the injected fuel at once by simultaneous multi-point ignition. Therefore, the premixed combustion as disclosed in Patent Document 1, more preferably the premixed combustion disclosed in Patent Document 1, rather than the so-called diffusion combustion in which fuel is injected in the latter half of the compression stroke to sequentially burn the injected fuel. However, it is preferable to adopt a combustion mode in which all the fuel is injected at a time corresponding to the preliminary injection time, and the fuel is uniformly dispersed and burned.

  However, since a fuel having a high vaporization temperature such as light oil is generally used in a diesel engine, a diesel engine such as that disclosed in Patent Document 1 is particularly operated at a low load side depending on the operating state of the engine. Since the in-cylinder temperature is low in the region, if all the fuel is injected during the fuel pre-injection period, all the fuel will not vaporize properly, and it will adhere to the cylinder wall surface, resulting in poor fuel consumption and soot. Produce. This problem cannot be sufficiently improved even by improving the external EGR and the fuel injection nozzle as disclosed in Patent Document 1.

  The present invention is based on such a technique, and provides a diesel engine capable of improving emission performance by reducing NOx while maintaining good fuel efficiency particularly in the low load side operation region of the engine. is there.

  Here, as a result of earnest research, the applicant of the present application has already applied for the technology related to the spark ignition engine control device (Japanese Patent Laid-Open No. 2003-227377), that is, the exhaust stroke and the intake stroke overlap in the partial load region of the engine. The burned gas discharged from the preceding cylinder in the exhaust stroke between the pair of cylinders is in a two-cylinder connection state in which the burned gas is directly introduced into the succeeding cylinder in the intake stroke through the inter-cylinder gas passage. Is set to a lean air-fuel ratio larger than the stoichiometric air-fuel ratio, combustion is performed by forced ignition, and fuel is supplied to the burned gas of the lean air-fuel ratio introduced from the preceding cylinder in the subsequent cylinder, and combustion is performed by compression self-ignition To find out how to apply the technology related to the control system of the spark ignition type engine, and the fuel of the diesel engine with high vaporization temperature. Further improvements added to prompt reduction, without deteriorating the fuel economy by once burned homogeneously to distribute the fuel, is obtained so as to improve the emission property.

  That is, according to the first aspect of the present invention, in a multi-cylinder diesel engine in which the combustion cycle of each cylinder is performed with a predetermined phase difference, the exhaust stroke and the intake stroke overlap in a partial load region of the engine. A two-cylinder connection state in which the burned gas discharged from the preceding cylinder in the exhaust stroke between the pair of cylinders is directly introduced into the subsequent cylinder in the intake stroke, and the gas discharged from the subsequent cylinder is guided to the exhaust passage The operation mode control means for performing the control of the special operation mode in which the combustion is performed in the preceding cylinder, the fuel is newly supplied to the burned gas in the preceding cylinder and the subsequent cylinder is combusted, and the special operation Fuel injection control means for controlling the fuel injection for the subsequent cylinders in the mode, and this fuel injection control means is an operating range for the special operation mode. At least in the low load side operation region, the preceding cylinder is controlled to inject fuel for the subsequent cylinder, and the injection start timing is set to a predetermined timing in the latter half of the expansion stroke of the preceding cylinder. Is.

  According to the present invention, fuel injection for the subsequent cylinder is started with respect to the burned gas in the latter half of the expansion stroke after completion of combustion in the preceding cylinder. The burnt gas promotes the vaporization of the supplied fuel. Further, even if the fuel injected in the latter half of the expansion stroke adheres to the wall surface of the cylinder, there is sufficient vaporization time before the transition from the expansion stroke to the exhaust stroke. Is scraped off by the piston, and wasteful fuel consumption that does not contribute to output can be suppressed to avoid deterioration of fuel consumption, and soot generation can be suppressed as much as possible. This air-fuel mixture is introduced from the preceding cylinder to the succeeding cylinder via the inter-cylinder gas passage, and the fuel is sufficiently vaporized and mixed with air during the intake stroke in the succeeding cylinder, and the combustion chamber of the succeeding cylinder Can be burned at once by the simultaneous multi-point compression auto-ignition, thereby avoiding the reaction between oxygen and nitrogen as much as possible to generate nitrogen oxides (NOx) Can be suppressed. Further, in the succeeding cylinder, the burned gas from the preceding cylinder is introduced, so that a large amount of EGR (exhaust gas recirculation) is performed. This will contribute to gas purification.

  The burned gas discharged from the preceding cylinder uses the exhaust passage of the preceding cylinder to connect the exhaust passage to the intake passage of the succeeding cylinder, and provides route switching means in the exhaust passage of the preceding cylinder. By switching the means, the burned gas may be configured to be switched between a path through which the burned gas is exhausted through the exhaust passage and a path connected to the intake passage of the subsequent cylinder through the exhaust passage. It is preferable that the burnt gas of the preceding cylinder is introduced into the succeeding cylinder via the inter-cylinder gas passage.

  With this configuration, the inter-cylinder gas passage can be shortened by providing a dedicated passage called an inter-cylinder gas passage that directly connects the ports of the cylinder, so that the temperature of the burned gas can be lowered as much as possible. The burned gas can be introduced into the succeeding cylinder while being suppressed, and the supplied fuel can be appropriately vaporized.

  The fuel injection timing in the operation region on the high load side in the operation region set to the special operation mode may be the latter half of the expansion stroke, but is not particularly limited. For example, the fuel injection control means As the load increases and the operation region shifts from the low-load operation region to the high-load operation region in the operation region in which the special operation mode is set, the fuel injection start timing for the subsequent cylinder in the preceding cylinder is set to the exhaust stroke. It is good also as what delays to the predetermined time to the first half (Claim 3).

  That is, when the engine load increases to shift from the low load side operation region to the high load side operation region in the operation region in which the special operation mode is set, the fuel injection amount increases and the in-cylinder temperature rises. . When the in-cylinder temperature of the preceding cylinder rises, the fuel for the succeeding cylinder injected in the preceding cylinder is partially converted into thermal energy in the preceding cylinder, which may not contribute to the work of the succeeding cylinder and may deteriorate the fuel consumption. Therefore, with the above configuration, while promoting the vaporization of the fuel having a high vaporization temperature, it is possible to effectively suppress a part of the fuel from being converted into thermal energy (burned) in the preceding cylinder after the vaporization. As a result, emission performance can be improved while maintaining good fuel efficiency.

  Although the combustion mode and the fuel injection mode controlled by the operation mode control means in the operation region on the higher load side than the operation region set as the special operation mode are not particularly limited, the operation mode control means It is configured to execute control in a normal operation mode in which fresh air is introduced into each cylinder and combustion is performed independently in each cylinder in an operation region on a higher load side than the operation region that is set as a special operation mode, The fuel injection control means preferably controls to inject fuel into each cylinder in this normal operation mode and sets the fuel injection timing to the compression stroke of each cylinder.

  If comprised in this way, while improving a fuel consumption and an emission in the low load driving | running | working area | region, output performance is ensured in the high load side driving | running | working area | region.

  In the present invention, the operation mode control means opens a fresh air introduction intake valve that introduces fresh air into the succeeding cylinder in the operation region on the high load side of the operation region set to the special operation mode. It is preferable to control so that fresh air is introduced into the succeeding cylinder in addition to the burned gas derived from the preceding cylinder.

  That is, in the special operation mode, high-temperature burned gas in the preceding cylinder is introduced into the succeeding cylinder. Therefore, there is a concern that pre-ignition may occur in the succeeding cylinder in the operation region on the high load side among the operation regions in the special operation mode. The Accordingly, with the above configuration, a fresh air intake intake valve that introduces fresh air into the subsequent cylinder is further provided, and the intake valve is opened to supply fresh air to the subsequent cylinder in addition to the burned gas. It is possible to effectively prevent pre-ignition by introducing the gas into the gas and suppressing an increase in the burnt gas temperature.

  Further, the fuel injection control means may inject fuel for the succeeding cylinders in the operation region on the high load side in the operation region in the special operation mode. Is set to be divided into the first-stage injection in the preceding cylinder and the second-stage injection in the subsequent cylinder, and the later injection timing of the fuel at the time of this divided injection is set in the latter half of the compression stroke of the subsequent cylinder. It is preferable to set (Claim 6).

  As described above, in the operation region on the high load side among the operation regions that are set to the special operation mode, there is a concern about the occurrence of pre-ignition in the subsequent cylinders. However, if configured as described above, the fuel injection control means performs the subsequent operation. The fuel for each cylinder is set to be divided and injected, and the latter injection timing at the time of this divided injection is controlled to be injected in the second half of the compression stroke in the succeeding cylinder, so the latter injection is ensured while ensuring the engine output. It is possible to effectively prevent the occurrence of premature ignition by lowering the temperature and pressure in the succeeding cylinder by utilizing the latent heat of vaporization of the fuel for the minute, and suppressing the activation of the fuel for the late injection. .

  In the present invention, it is preferable to further include a supercharger for supercharging intake air to the preceding cylinder.

  If comprised in this way, air can fully be inhaled also to a preceding cylinder and generation | occurrence | production of the nitrogen oxide in a preceding cylinder can be suppressed effectively. Here, in the operation region on the low load side among the operation regions to be in the special operation mode, particularly when a supercharger is provided, the temperature of the burned gas is lower than the temperature on the high load side due to excess air. Therefore, in the present invention, the fuel for the subsequent cylinder is injected in the latter half of the expansion stroke of the preceding cylinder, in other words, immediately after the combustion of the preceding cylinder is completed, as described above. However, the supplied fuel is sufficiently mixed and can be burned at once in the succeeding cylinder, and the emission property can be improved by suppressing the generation of nitride oxide.

  According to the diesel engine of the present invention, it is possible to effectively vaporize a fuel whose vaporization temperature is relatively higher than that of gasoline in at least the low load side operation region of the operation region in the special operation mode. There is an advantage that the vaporized fuel is homogeneously diffused into the combustion chamber of the subsequent cylinder while being sufficiently mixed with air, and can be burned at once in the subsequent cylinder. Therefore, the reaction between nitrogen and oxygen can be suppressed as much as possible to suppress the generation of nitrogen oxides as much as possible, and emission properties can be improved. Moreover, even if fuel adheres to the cylinder wall of the preceding cylinder due to fuel injection for the subsequent cylinder, it can be sufficiently vaporized during the transition period from the expansion stroke to the exhaust stroke, and the fuel adhering to the wall surface can be removed. The fuel can be efficiently vaporized by being scraped up by the piston, so that the generation of soot can be effectively suppressed while maintaining good fuel efficiency.

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

(First embodiment)
FIG. 1 shows a schematic configuration of a diesel engine according to a first embodiment of the present invention, and FIG. 2 schematically shows a structure of one cylinder of a diesel engine main body and intake / exhaust valves provided for the cylinder. . In these drawings, the engine body 1 has a plurality of cylinders, and in the illustrated embodiment, has four cylinders 2A to 2D. A piston 3 is fitted into each of the cylinders 2 </ b> A to 2 </ b> D, and a combustion chamber 4 is formed above the piston 3.

  A fuel injection valve 9 as fuel injection means for directly injecting fuel into the combustion chamber 4 is provided at a side portion of the combustion chamber 4. This fuel injection valve 9 incorporates a needle valve and a solenoid (not shown). When a pulse signal described later is input, the fuel injection valve 9 is driven for a time corresponding to the pulse width at the pulse input timing to open the valve. An amount of fuel corresponding to the valve time is injected. The fuel injection valve 9 is supplied with fuel by a fuel pump (not shown) via a fuel supply passage and the like, and a fuel supply system is provided so that a fuel pressure higher than the pressure in the combustion chamber in the compression stroke can be applied. It is configured.

  Further, intake ports 11, 11a, 11b and exhaust ports 12, 12a, 12b are opened to the combustion chambers 4 of the respective cylinders 2A to 2D, and an intake passage 15 and an exhaust passage 20 are connected to these ports. Each port is opened and closed by intake valves 31, 31a, 31b and exhaust valves 32, 32a, 32b.

  The cylinders 2A to 2D perform a combustion cycle including intake, compression, expansion, and exhaust strokes with a predetermined phase difference. In the case of a four-cylinder engine, the first cylinder from one end in the cylinder row direction 2A, 2nd cylinder 2B, 3rd cylinder 2C, 4th cylinder 2D, as shown in FIG. 5, the cycle is in the order of 1st cylinder 2A, 3rd cylinder 2C, 4th cylinder 2D, 2nd cylinder 2B. The combustion cycle is performed with a phase difference of 180 ° in crank angle. In FIG. 5, EX represents an exhaust stroke, IN represents an intake stroke, F represents fuel injection, and a star mark in the drawing represents that compression ignition is performed. Further, for the fuel injection F in the figure, the fuel for the preceding cylinder is indicated by F1, and the fuel for the succeeding cylinder is indicated by F2.

  Between a pair of cylinders in which the exhaust stroke and the intake stroke overlap, a cylinder on the intake stroke side (referred to herein as a preceding cylinder) from the cylinder on the exhaust stroke side when the exhaust stroke and the intake stroke overlap (this specification) The inter-cylinder gas passage 22 is provided so that the burned gas can be directly introduced to the subsequent cylinder). In this embodiment, as shown in FIG. 5, the exhaust stroke (EX) of the first cylinder 2A and the intake stroke (IN) of the second cylinder 2B overlap, and the exhaust stroke (EX) of the fourth cylinder 2D and the third stroke Since the intake stroke (IN) of the cylinder 2C overlaps, the first cylinder 2A and the second cylinder 2B, and the fourth cylinder 2D and the third cylinder 2C form a pair, respectively, and the first cylinder 2A and the fourth cylinder 2D are the preceding cylinder. The second cylinder 2B and the third cylinder 2C are the subsequent cylinders.

  The intake / exhaust port of each cylinder and the intake passage, exhaust passage, and inter-cylinder gas passage connected to the cylinder are specifically configured as follows.

  The first cylinder 2A and the fourth cylinder 2D, which are the preceding cylinders, respectively include an intake port 11 for introducing fresh air, and a first exhaust port 12a for sending burned gas (exhaust gas) to the exhaust passage. A second exhaust port 12b for leading the burned gas to the subsequent cylinder is provided. The second cylinder 2B and the third cylinder 2C, which are the subsequent cylinders, respectively, have a first intake port 11a for introducing fresh air and a second intake port for introducing burned gas from the preceding cylinder. 11b and an exhaust port 12 for sending burned gas to the exhaust passage.

  In the example shown in FIG. 1, one intake port 11 in each of the first and fourth cylinders 2A and 2D and two first intake ports 11a in each of the second and third cylinders 2B and 2C are provided in one side of the combustion chamber 4. The first exhaust port 12a and the second exhaust port 12b in the first and fourth cylinders 2A and 2D and the second intake port 11b and the exhaust in the second and third cylinders 2B and 2C are provided in parallel in the side half. A port 12 is provided in parallel with the other half of the combustion chamber 4.

  The intake port 11 in the first and fourth cylinders 2A and 2D and the first intake port 11a in the second and third cylinders 2B and 2C are connected to the downstream ends of the branch intake passages 16 for each cylinder in the intake passage 15. Yes.

  An upstream end of a branch exhaust passage 21 for each cylinder in the exhaust passage 20 is connected to the first exhaust port 12a in the first and fourth cylinders 2A and 2D and the exhaust port 12 in the second and third cylinders 2B and 2C. Yes. Further, an inter-cylinder gas passage 22 is provided between the first cylinder 2A and the second cylinder 2B and between the third cylinder 2C and the fourth cylinder 2D, respectively, and the first and fourth cylinders which are the preceding cylinders. The upstream end of the inter-cylinder gas passage 22 is connected to the second exhaust ports 12b of 2A and 2D, and the inter-cylinder gas passage 22 is connected to the second intake ports 11b of the second and third cylinders 2B and 2C as the subsequent cylinders. The downstream end is connected.

  The inter-cylinder gas passage 22 is provided with a burned gas temperature sensor 25 for detecting the burned gas temperature introduced into the succeeding cylinders 2B and 2C, and the detection result of the sensor 25 is transmitted to an ECU 40 described later. It has become so.

  The branch exhaust passages 21 in the exhaust passage 20 gather downstream, and an oxidation catalyst 24 is provided in the exhaust passage 20 further downstream. As generally known, the oxidation catalyst 24 is a catalyst for oxidizing hydrocarbons and carbon monoxide in the exhaust gas exhausted from the cylinders 2A to 2D at the catalyst activation temperature. .

  A turbine 28 of the supercharger 27 is disposed between the collecting portion of the branch exhaust passage 21 and the oxidation catalyst 24. The supercharger 27 assists intake of the preceding cylinders 2A and 2D, and in this embodiment, a turbocharger is provided. In the present embodiment, the supercharger 27 is driven so as to assist the intake of the preceding cylinders 2A and 2D in a special operation mode and a normal operation mode, which will be described later. The salary may be limited.

  Specifically, the supercharger 27 communicates with the turbine 28 provided in the exhaust passage 20 that communicates with the exhaust ports 12, 12 a, 12 b via the branch exhaust passage 21 and the intake ports 11, 11 a, 11 b. And a compressor 29 that is linked to the turbine 28 and is connected to the turbine 28. The turbine 28 is rotated by the energy of the exhaust gas flowing through the exhaust passage 20. The compressor 29 is also rotated in conjunction with the rotation of the turbine 28. In addition, the intake air is supercharged by the rotation of the compressor 29. The turbocharger is not limited to the turbocharger 27 but may be a mechanical supercharger or the like. However, the provision of the turbocharger 27 increases the exhaust pressure. Therefore, it is easy for the burned gas to remain in the preceding cylinders 2A and 2D, the burned gas temperature in the preceding cylinders 2A and 2D becomes higher due to the remaining burned gas, and the fuel (for example, light oil) having a higher vaporization temperature than gasoline. ) Is advantageous in that vaporization is promoted. An intercooler 26 for lowering the intake air temperature supercharged by the supercharger 27 is provided in the intake passage 15 on the downstream side of the compressor 29.

  The intake / exhaust valves for opening and closing the intake / exhaust ports of each cylinder and the valve operating mechanism for these valves are as follows. That is, the intake port 11, the first exhaust port 12a and the second exhaust port 12b in the first and fourth cylinders 2A and 2D are provided with the intake valve 31, the first exhaust valve 32a and the second exhaust valve 32b, respectively. A first intake valve 31a, a second intake valve 31b, and an exhaust valve 32 are provided in the first intake port 11a, the second intake port 11b, and the exhaust port 12 in the second and third cylinders 2B, 2C, respectively. These intake / exhaust valves are opened and closed at predetermined timings by the valve mechanisms comprising the camshafts 33, 34, etc. so that the intake stroke and exhaust stroke of each cylinder are performed with the predetermined phase difference as described above. To be driven.

  Further, among these intake / exhaust valves, the first exhaust valve 32a and the second exhaust valve 32b are provided with a valve stop mechanism 35a for switching each valve between an operating state and a stopped state. Since the valve stop mechanism 35a is conventionally known, detailed illustration thereof is omitted. For example, the hydraulic oil is supplied to and discharged from the tappet interposed between the cams of the camshafts 33 and 34 and the valve shaft. In the state where hydraulic oil is provided and hydraulic oil is supplied to the hydraulic chamber, the operation of the cam is transmitted to the valve and the valve is opened and closed.When the hydraulic oil is discharged from the hydraulic chamber, the cam operation is not performed. The valve is stopped when it cannot be transmitted to the valve.

  On the other hand, among the intake / exhaust valves, the first intake valve 31a and the second intake valve 31b are not only switched between the operation state and the stop state, but also in a plurality of stages (2 in the present embodiment). Valve timing switching mechanisms 35b and 35c for switching the valve timing over the stage) are provided. That is, the valve timing switching mechanisms 35b and 35c have a high load in the operation region A in the special operation mode to be described later in addition to the normal valve opening operation that opens in the intake and exhaust strokes of the subsequent cylinders 2B and 2C. In the operation region A2 on the side, the first intake valve 31a is configured to open at the beginning of the intake stroke of the succeeding cylinders 2B and 2C, the valve opening timing of the second intake valve 31b is delayed, and the valve opening period (See the dotted line in FIG. 6 or the valve operation in FIG. 12). Accordingly, the cams for the intake valves 31a and 31b are provided in three types: a cam for stopping the valve, a cam for the operating region A2 on the high load side, and a cam for operating the normal valve. Note that these cam switching mechanisms are well known in the art, and are not shown here.

  The first control valve 37 and the second exhaust valve 32b are stopped in the hydraulic oil supply / discharge passage 36 for the valve stop mechanism 35a of the first exhaust valve 32a and the valve stop mechanism 35b of the first intake valve 31a. A second control valve 39 is provided in each of the hydraulic oil supply / discharge passages 38 to the mechanism 35a and the valve timing switching mechanism 35c of the second intake valve 31b (see FIG. 3). In particular, the valve timing switching mechanisms 35b and 35c of the first intake valve 31a and the second intake valve 31b are also connected to passages 36 and 38 for supplying and discharging hydraulic oil.

  FIG. 3 shows the configuration of the engine drive and control system. In this figure, a signal from the burned gas temperature sensor 25 is inputted to an engine control ECU (control unit) 40 composed of a microcomputer or the like, and from a water temperature sensor 51 for detecting the cooling water temperature of the engine. A signal is input, and further, a signal from an engine speed sensor 52 for detecting the engine speed and an accelerator position sensor 53 for detecting an accelerator position (accelerator pedal depression amount) for determining the operating state is also input. . In addition, control signals are output from the ECU 40 to the fuel injection valves 9 and the first and second control valves 37 and 39.

  The ECU 40 includes an operating state determination unit 41, a mode setting unit 42, a valve opening / closing control unit 43, a fuel injection control unit 44, and the like as functional components.

  The operating state discriminating means 41 has a control map in which the engine operating region is divided into a region A on the low speed and low load side and a region B on the high speed side or the high load side as shown in FIG. It is determined whether the engine operating state (engine speed and engine load), which is examined from signals from the sensor 52, the accelerator opening sensor 53, or the like, is in the above-described region A or B. Further, when the operation state is in the special operation mode region A, the operation state determination means 41 determines whether the region A is in the low load side operation region A1 or the high load side operation region A2. It has become.

  Based on the determination by the operation state determination unit 41, the mode setting unit 42 is configured to follow the combustion gas discharged from the preceding cylinders 2A, 2D in the exhaust stroke as it is in the intake stroke in the operation region A on the low load low rotation side. A special operation mode in which the cylinders 2B and 2C are introduced and burned is selected, and a normal operation mode in which each cylinder is independently burned is selected in the high load side or high rotation side operation region B.

  In response to the mode setting by the mode setting means 42, the valve opening / closing control means 43 passes the burned gas of the preceding cylinders (first and fourth cylinders) 2A and 2D through the inter-cylinder gas passage 22 in the special operation mode. (2nd and 3rd cylinders) In order to change the intake / exhaust flow state so that the two cylinders connected to 2B and 2C are connected, and in the normal operation mode, each cylinder is brought into an independent state in which fresh air is introduced into each cylinder. It controls the valve stop mechanism 35a and the valve timing switching mechanisms 35b and 35c. Specifically, by controlling the control valves 37 and 39 depending on whether the operating state is in the region A or B. In principle, each valve stop mechanism 35a and the like are controlled as follows.

Area A: (Special operation mode)
Stop the first exhaust valve 32a and the first intake valve 31a
Operating the second exhaust valve 32b and the second intake valve 31b Region B: (Normal operation mode)
The first exhaust valve 32a and the first intake valve 31a are activated.
The second exhaust valve 32b and the second intake valve 31b are stopped. The valve opening / closing control means 43 is operated when the special operation mode is selected, and the engine operation is performed based on the determination of the operation state determination means 41. When the state is in the operation region A2 on the high load side, the valve timing is set so as to temporarily open the intake valve for introducing fresh air (first intake valve 31a) for introducing fresh air into the succeeding cylinders 2B and 2C. The switching mechanism 35b is controlled. Specifically, when the valve opening / closing control means 43 is in the operation region A2, the first intake valve 31a in the stopped state is moved from the vicinity of the intake top dead center of the succeeding cylinders 2B and 2C to the middle of the intake stroke. In order to open the valve, the valve timing switching mechanism 35b is controlled to switch the cam of the switching mechanism 35b, and the valve timing switching mechanism 35c is controlled to delay the opening timing of the second intake valve 31b in the operating state. The cam of the mechanism 35c is switched.

  The fuel injection control means 44 controls the fuel injection amount and the injection timing from the fuel injection valve 9 provided in each of the cylinders 2A to 2D in accordance with the operating state of the engine. The fuel control is changed based on the result.

  That is, when the special operation mode is set, the fuel injection amount is controlled based on the input from the accelerator opening sensor 53 for the preceding cylinders 2A and 2D, and the fuel is injected in the compression stroke. The injection timing is set so that diffusion combustion is performed. On the other hand, for the succeeding cylinders 2B and 2C, the fuel injection amount is controlled based on the input from the accelerator opening sensor 53, and this fuel is supplied by the preceding cylinders 2A and 2D. The fuel injection timing is set so that the amount of fuel is directly injected into the combustion chamber 4 by the fuel injection valve 9 of the preceding cylinders 2A and 2D, and the fuel injection valve 9 is selected. In the subsequent cylinders 2B and 2C, homogeneous compression ignition To do.

  Here, the specific fuel injection timing for the subsequent cylinders is based on the determination by the operation state determination means 41, in the operation region A1 on the low load side of the operation region A to be in the special operation mode, or on the high load side. It differs depending on whether it is in the operation region A2. That is, when the fuel injection control means 44 is in the operation region A1 on the low load side, as shown in FIG. 6, fuel injection for the subsequent cylinders is started from the latter half of the expansion stroke of the preceding cylinders 2A and 2D. (Start time point TFS), set so that the entire injection amount is injected in the latter half of the expansion stroke (end time point TFE), while not in the high load side operation region A2, although not shown, The fuel injection timing is delayed as compared with the case in the operation region A1, and the fuel for the subsequent cylinders is set to be injected in the first half of the exhaust stroke of the preceding cylinders 2A and 2D. Then, the fuel for the subsequent cylinders injected by the preceding cylinders 2A and 2D is introduced into the subsequent cylinders 2B and 2C through the inter-cylinder gas passage 22 together with the burned gas.

  Further, when the normal operation mode is selected, the fuel injection control means 44 controls the fuel injection amount based on the input from the accelerator opening sensor 53, and the fuel injection valves provided in the respective cylinders 2A to 2D. 9 is set to inject fuel in each cylinder 2A to 2D in the latter half of the compression stroke, so that diffusion combustion is performed in each cylinder 2A to 2D.

  Next, the operation of the apparatus according to the embodiment as described above will be described with reference to FIGS.

  In the operation region A on the low load and low rotation side, the special operation mode is set, and as described above, the first exhaust valve 32a and the first intake valve 31a are stopped, and the second exhaust valve 32b and the second intake valve 31b are operated. As a result, as shown in FIG. 7, the substantial fresh air and gas flow paths are such that the burned gas discharged from the preceding cylinders 2A, 2D is directly passed through the inter-cylinder gas passage 22 and the subsequent cylinders 2B, 2C. In addition, a two-cylinder connection state in which only burned gas discharged from the subsequent cylinders 2B and 2C is guided to the exhaust passage 20 is established.

  In this state, while the intake air is supercharged by the supercharger 27, fresh air is introduced into the preceding cylinders 2 </ b> A and 2 </ b> D from the intake passage 15 in the intake stroke (arrow a in FIG. 7). While the fuel injection amount is controlled based on the detection, fuel is injected in the latter half of the compression stroke (F1 in FIGS. 5 and 6), whereby diffusion combustion is performed in the preceding cylinders 2A and 2D (see FIG. 5).

  When this diffusion combustion is completed, that is, when the engine operating state is in the operating region A1, in the latter half of the expansion stroke of the preceding cylinders 2A and 2D, and in the one operating region A2, the first half of the exhaust stroke of the preceding cylinders 2A and 2D. Then, the fuel for the subsequent cylinder is directly injected into the combustion chamber 4 of the preceding cylinder 2A, 2D (F2 in FIGS. 5 and 6).

  During the period in which the exhaust strokes of the preceding cylinders 2A and 2D overlap with the intake strokes of the subsequent cylinders 2B and 2C, the burned gas containing this fuel passes through the inter-cylinder gas passage 22 while being discharged from the preceding cylinders 2A and 2D. Are introduced into the succeeding cylinders 2B and 2C (the white arrows in FIGS. 5 and 6 and the arrow b in FIG. 7). Then, in the succeeding cylinders 2B and 2C, compression self-ignition is performed by the increase in pressure and temperature in the combustion chamber 4 near the top dead center of the compression stroke.

  In this case, fuel having a relatively high vaporization temperature such as light oil is injected into the high-temperature burned gas immediately after being combusted in the preceding cylinders 2A and 2D, and after this injection, combustion is performed in the succeeding cylinders 2B and 2C. Since there is sufficient vaporization time, the fuel for the subsequent cylinders injected by the preceding cylinders 2A and 2D is appropriately vaporized, and the combustion in the subsequent cylinders 2B and 2C is almost entirely in the combustion chamber 4. It is carried out in a diffused homogeneous gas mixture and spreads all over the combustion chamber 4 by simultaneous multi-point ignition.

  Then, the burnt gas after combustion in the succeeding cylinders 2B and 2C is discharged to the exhaust passage 20 (arrow c in FIG. 7).

  Thus, although the preceding cylinders 2A and 2D perform combustion similar to that of a normal diesel engine, after the combustion in the preceding cylinders 2A and 2D ends, fuel injection for the succeeding cylinders starts with respect to the burned gas. Therefore, even the fuel having a high vaporization temperature such as light oil promotes the vaporization of the supplied fuel by the high-temperature burned gas immediately after the end of combustion. In particular, in the operating region A1 on the low load side, the in-cylinder temperature is often low, and according to a normal diesel engine, even if fuel is injected in advance for homogeneous combustion, most of the fuel injected at high pressure is the cylinder wall surface. The diesel engine of this embodiment is not only contributing to the combustion energy by being attached to the gas but also being discharged as a soot, whereas according to the diesel engine of the present embodiment, the burnt gas is extremely hot in the second half of the expansion stroke immediately after the end of combustion. Vaporization is promoted because the fuel is injected into the inside, and even if it adheres to the wall surface, the attached fuel is scraped as the piston 3 of the preceding cylinders 2A, 2D is shifted from the expansion stroke to the exhaust stroke. And there is sufficient vaporization time, which promotes the vaporization of the fuel. Therefore, it is possible to suppress wasteful fuel consumption that does not contribute to output, avoid deterioration of fuel consumption, and suppress soot generation as much as possible.

  Then, this air-fuel mixture is introduced from the preceding cylinders 2A, 2D to the succeeding cylinders 2B, 2C via the inter-cylinder gas passage 22, and the fuel is sufficiently vaporized during the intake stroke in the succeeding cylinders 2B, 2C. Since it is mixed with the combustion gas and diffused uniformly in the combustion chamber 4, it can be burned all over the combustion chamber 4 by multipoint compression self-ignition, thereby allowing the reaction between oxygen and nitrogen as much as possible. Therefore, generation of nitrogen oxides (NOx) can be suppressed. Further, in the succeeding cylinder, the burned gas from the preceding cylinder is introduced, so that a large amount of EGR (exhaust gas recirculation) is performed, and this also sufficiently suppresses the generation of NOx. This will contribute to the purification of exhaust gas.

  In addition, since the inter-cylinder gas passage 22 is disposed so as to directly connect the ports between the adjacent cylinders, the inter-cylinder gas passage 22 can also be formed short, thereby exhausting from the preceding cylinders 2A and 2D. The temperature decrease of the burnt gas can be suppressed as much as possible to effectively prevent re-liquefaction of the vaporized supply fuel, and the generation of NOx and soot can be effectively suppressed.

  Further, the fuel injection control means 44 starts fuel injection in the latter half of the expansion stroke after the combustion of the preceding cylinders 2A and 2D in the low load side operation region A1 where the in-cylinder temperature is often low, and promotes the vaporization thereof. In addition, when the engine load increases and shifts to the operation region A2 on the high load side, fuel is injected into an extremely high temperature burned gas, and a part of the fuel is heated in the preceding cylinders 2A and 2D. The injection timing is changed to the first half of the exhaust strokes of the preceding cylinders 2A and 2D in order to avoid deterioration in fuel consumption due to conversion to energy.

  In particular, in the present embodiment, since the supercharger 27 for supercharging intake air is provided in the preceding cylinders 2A and 2D, there is a concern about a decrease in the burnt gas temperature. Since the fuel is injected immediately after the end of the 2D combustion, it is possible to effectively prevent a situation in which the fuel for the subsequent cylinders is not vaporized and the fuel consumption is deteriorated. Further, the supercharger 27 sufficiently sucks air into the preceding cylinders 2A and 2D to increase the lean degree, thereby suppressing the generation of nitrogen oxides (NOx) in the preceding cylinders 2A and 2D. Can also be improved.

  On the other hand, in the operation region B on the high load side or high rotation side, the normal operation mode is set, and the first exhaust valve 32a and the first intake valve 31a are in the operating state as described above, and the second exhaust valve 32b and the second intake valve 31b. Is brought into a stopped state, the substantial new air and gas flow paths are as shown in FIG. 8, and the intake ports 11 and 11a and the exhaust ports 12a and 12 of each cylinder 2A to 2D are substantially independent. Then, fresh air is introduced from the intake passage 15 to the intake ports 11 and 11a of the cylinders 2A to 2D, and burned gas is discharged to the exhaust passage 20 from the exhaust ports 12 and 12a of the cylinders 2A to 2D.

(Second Embodiment)
FIG. 9 shows a schematic configuration of a diesel engine according to the second embodiment of the present invention. This second embodiment differs in the specific configuration of the passage for introducing the burned gas of the preceding cylinders 2A and 2D into the succeeding cylinders 2B and 2C, and the function of the ECU 40 is slightly different accordingly. Hereinafter, these different points will be described with emphasis, and other points will be denoted by the same reference numerals in the drawings, and description thereof will be omitted.

  In other words, in the first embodiment, the preceding cylinders 2A, 2D and the succeeding cylinders 2B, 2C are connected by the dedicated inter-cylinder gas passages 22, respectively, and the burned gas of the preceding cylinders 2A, 2D is connected to the inter-cylinder gas passages 22. However, in the second embodiment, the branch cylinders 2B and 2C branch in the merged exhaust passage 210a where the branch exhaust passages 210 of the preceding cylinders 2A and 2D merge. It is connected to the intake passage 161 via the connection passage 220, and the path of the burned gas discharged from the first and fourth cylinders 2A, 2D is switched to the connection passage 220 side in the special operation mode, and the normal mode Route switching means (switching valves in the present embodiment) 19a, 19b, 19c for switching to the exhaust passage 20 are sometimes provided.

  Specifically, the intake passage 15 branches in three directions at a predetermined portion up to the intake port 11 of each cylinder 2A to 2D, and the passages on both sides sandwiching the central communication passage 160a are branched intake passages for the preceding cylinders. The intake port 11 of the preceding cylinders 2A and 2D is connected to the downstream end of the preceding cylinder branch intake passage 160. On the other hand, the central communication passage 160a further branches in three directions on the downstream side, and the central passage is configured as a connection passage 220, and both sides sandwiching the connection passage 220 are branch intake passages for subsequent cylinders. The intake port 11 of the subsequent cylinders 2B and 2C is connected to the downstream end of the subsequent cylinder branch intake passage 161.

  On the other hand, the branch cylinder branch exhaust passage 210 whose upstream end is connected to the exhaust port 12 of the preceding cylinders 2A and 2D joins downstream to form a preceding cylinder joining exhaust passage 210a. A connection passage 220 is connected. Further, on the downstream side of the merged exhaust passage 210a, a branch cylinder branch exhaust passage 211 whose upstream end is connected to the exhaust port 12 of the subsequent cylinders 2B and 2C joins from both sides and communicates with the exhaust passage 20. Has been made.

  The central communication passage 160a branched from the intake passage 15, the connection passage 220, and the converging exhaust passage 210a for the preceding cylinder are respectively connected to the first to third switching valves 19a for switching the gas passage by opening and closing each passage. 19b and 19c are provided. Actuators (not shown) of the switching valves 19a, 19b, 19c are connected to the ECU 40, and opening / closing control is executed by the valve opening / closing control means 43 of the ECU 40 (see FIG. 3). FIG. 3 illustrates the ECU 40 and the like according to the first embodiment, but the configuration diagram of the ECU 40 does not change in the second embodiment, and therefore this will be described with the aid of this.

  That is, the ECU 40 includes an operating state determination unit 41, a mode setting unit 42, a valve opening / closing control unit 43, and a fuel injection control unit 44, as in the first embodiment. Since the configuration excluding the valve opening / closing control means 43 is the same as that of the first embodiment, the description thereof is omitted here.

  In response to the mode setting by the mode setting means 42, the valve opening / closing control means 43 passes the burned gas of the preceding cylinders (first and fourth cylinders) 2A and 2D through the inter-cylinder gas passage 22 in the special operation mode. (2nd and 3rd cylinders) In order to change the intake / exhaust flow state so that the two cylinders connected to 2B and 2C are connected, and in the normal operation mode, each cylinder is brought into an independent state in which fresh air is introduced into each cylinder. The actuators of the switching valves 19a, 19b, and 19c are controlled. Specifically, the actuators are controlled as follows depending on whether the operating state is in the region A or B.

Area A: (Special operation mode)
The first switching valve 19a and the third switching valve 19c are closed.
The second switching valve 19b is opened. Region B: (normal operation mode)
Opening of the first switching valve 19a and the third switching valve 19c
The second switching valve 19b is in a closed state. Note that the special operation mode is selected for the valve opening / closing control means 43, and the engine operating state is determined based on the determination of the operation state determining means 41 on the high load side. When in the operation region A2, the first switching valve 19a may be opened before the second switching valve 19b is opened, and fresh air may be introduced into the succeeding cylinders 2B and 2C. By introducing fresh air into the succeeding cylinders 2B and 2C in this way, the temperature rise in the succeeding cylinders 2B and 2C can be suppressed, and unexpected early ignition in the succeeding cylinders 2B and 2C can be prevented. In addition, the amount of fuel for combustion in the succeeding cylinders 2B and 2C can be increased, and the output performance can be improved.

  Next, the operation of the apparatus of the embodiment as described above will be described with reference to FIGS.

  In the operation region A on the low load and low rotation side, the special operation mode is set, and as described above, the first switching valve 19a and the third switching valve 19c are opened, and the second switching valve 19b is closed. As shown in FIG. 10, the fresh air and gas flow paths are such that the burned gas discharged from the preceding cylinders 2A and 2D is directly introduced into the succeeding cylinders 2B and 2C through the connection passage 220, A two-cylinder connection state is established in which only the burned gas discharged from the cylinders 2B and 2C is guided to the exhaust passage 20.

  In this state, while the intake air is supercharged by the supercharger 27, fresh air is introduced into the preceding cylinders 2A and 2D by branching from the intake passage 15 to the preceding cylinder branch intake passage 160 in the intake stroke, respectively (in FIG. 10). Arrows d, e), fuel is injected in the latter half of the compression stroke while the fuel injection amount is controlled based on the detection from the accelerator opening sensor 53 (F1 in FIGS. 5 and 6), thereby leading cylinders 2A, 2D In FIG. 5, diffusion combustion is performed (see FIG. 5).

  When this diffusion combustion is completed, that is, when the engine operating state is in the operating region A1, in the latter half of the expansion stroke of the preceding cylinders 2A and 2D, and in the one operating region A2, the first half of the exhaust stroke of the preceding cylinders 2A and 2D. Then, the fuel for the subsequent cylinder is directly injected into the combustion chamber 4 of the preceding cylinder 2A, 2D (F2 in FIGS. 5 and 6).

  Then, during the period in which the exhaust strokes of the preceding cylinders 2A and 2D overlap with the intake strokes of the subsequent cylinders 2B and 2C, the burned gas containing this fuel is discharged from the preceding cylinders 2A and 2D, and the branched exhaust passage 210 for the preceding cylinders. Then, the air is introduced into the succeeding cylinders 2B and 2C through the connecting passage 220 and the succeeding cylinder branch intake passage 161 (the white arrows in FIGS. 5 and 6 and the arrows f, g and h in FIG. 10). In the succeeding cylinders 2B and 2C, compression self-ignition is performed near the top dead center of the compression stroke due to an increase in pressure and temperature in the combustion chamber.

  In this case, fuel having a relatively high vaporization temperature such as light oil is injected into the high-temperature burned gas immediately after being combusted in the preceding cylinders 2A and 2D, and after this injection, combustion is performed in the succeeding cylinders 2B and 2C. Since there is sufficient vaporization time, the fuel for the subsequent cylinders injected by the preceding cylinders 2A and 2D is appropriately vaporized, and the combustion in the subsequent cylinders 2B and 2C is almost entirely in the combustion chamber 4. It is performed in the diffused homogeneous mixture and spreads all over the combustion chamber 4 at once.

  Then, the burnt gas after combustion in the succeeding cylinders 2B and 2C is discharged to the exhaust passage 20 (arrows j and k in FIG. 10).

  On the other hand, in the operation region B on the high load side or the high rotation side, the normal operation mode is set, and the first switching valve 19a and the third switching valve 19c are opened and the second switching valve 19b is closed as described above. As a result, the flow path of substantial fresh air and gas is as shown in FIG. 11, and the intake port 11 and the exhaust port 12 of each cylinder 2 </ b> A to 2 </ b> D are substantially independent, and each cylinder 2 </ b> A to 2 </ b> A to Fresh air is introduced into the 2D intake port 11 and burned gas is discharged from the exhaust ports 12 of the cylinders 2A to 2D to the exhaust passage 20.

  The specific configuration of the apparatus of the present invention is not limited to the first and second embodiments described above, and can be variously modified. Examples thereof will be described below.

  (1) In the above embodiment, the fuel for the expansion stroke is supplied all at once by the preceding cylinders 2A and 2D. However, a part of the operation region A in the special operation mode, particularly a high load, is used. The fuel injection control means may be set so that the fuel for the expansion stroke is divided and injected in the operation region A2 on the side. For example, as shown in FIG. 12, the early injection timing of fuel at the time of split injection is set to be injected in the latter half of the expansion stroke of the preceding cylinders 2A and 2D (F20 in FIG. 12), and the latter injection is set to the succeeding cylinder 2B. , 2C, the fuel injection valve 9 may be set, and this late injection timing may be set in the latter half of the compression stroke of the subsequent cylinder (F21 in FIG. 12).

  If comprised in this way, while ensuring the engine output, the temperature and pressure in the succeeding cylinders 2B and 2C are reduced by utilizing the latent heat of vaporization of the fuel for the late injection, and the fuel activity for the late injection is achieved. It is possible to effectively prevent the occurrence of premature ignition.

  Further, even when the fuel for the subsequent cylinders is injected collectively in the preceding cylinders 2A and 2D, the injection does not have to be completed in the latter half of the expansion stroke as in the first embodiment. For example, FIG. As shown, the fuel injection for the succeeding cylinder is started in the latter half of the expansion stroke of the preceding cylinders 2A and 2D (start time TFS in FIG. 13), and the injection completion (completion time TFE in FIG. 13) is exhausted. It may take the first half of the process.

  In FIG. 13, unlike the case of the first embodiment, only the burned gas is introduced into the succeeding cylinders 2B and 2C, and no fresh air is introduced.

  (2) In the first embodiment, when the control in the special operation mode is executed, the fuel injection timing for the subsequent cylinders is changed to two stages (multiple stages) according to the engine load. However, the present invention is not limited to this, and the fuel injection timing may be continuously delayed according to the engine load.

  (3) In the first embodiment, a mechanical valve opening / closing mechanism using a cam is used as the valve operating mechanism. Instead of the mechanical valve opening / closing mechanism using the cam or the like, for example, an electromagnetic valve is used. An electric valve opening / closing mechanism may be used.

1 is a schematic plan view of an entire diesel engine according to the present invention. It is a schematic sectional drawing, such as an engine main body. It is a block diagram of a control system. It is explanatory drawing which shows an operation area | region. It is a figure which shows the exhaust stroke of each cylinder, an intake stroke, fuel injection timing, ignition timing, etc. FIG. It is a figure which shows each stroke, fuel injection timing, ignition timing, etc. of a pair of preceding / following cylinders. It is explanatory drawing which shows the distribution path | route of substantial fresh air and gas at the time of low load and low rotation. It is explanatory drawing which shows the distribution path | route of a substantially new air and gas when it exists in the operating area | region of a high load and high and low rotation side. It is a schematic plan view which shows 2nd Embodiment of the diesel engine which concerns on this invention. It is explanatory drawing which shows the distribution path | route of substantial fresh air and gas at the time of low load and low rotation. It is explanatory drawing which shows the distribution path | route of a substantially new air and gas when it exists in the operating area | region of a high load and high and low rotation side. It is explanatory drawing which shows the distribution path | route of the substantial fresh air and gas at the time of the low load and low rotation in 2nd Embodiment. It is explanatory drawing which shows the distribution path | route of substantial fresh air and gas at the time of the high load high rotation in 2nd Embodiment.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Engine main body 2A-2D Cylinder 9 Fuel injection valve 11 Intake port 11a 1st intake port 11b 2nd intake port 12 Exhaust port 12a 1st exhaust port 12b 2nd exhaust port 15 Intake passage 20 Exhaust passage 22 Inter-cylinder gas passage 31 Intake Valve 31a First intake valve 31b Second intake valve 32 Exhaust valve 32a First exhaust valve 32b Second exhaust valve 35 Valve stop mechanism 40 ECU
41 Operating state discriminating means 43 Valve opening / closing mechanism control means 44 Fuel injection control means

Claims (7)

In a multi-cylinder diesel engine in which the combustion cycle of each cylinder is performed with a predetermined phase difference,
In the partial load region of the engine, the burned gas discharged from the preceding cylinder in the exhaust stroke between a pair of cylinders where the exhaust stroke and the intake stroke overlap is directly introduced into the succeeding cylinder in the intake stroke and discharged from this succeeding cylinder. Special operation mode in which combustion is performed in the preceding cylinder and fuel is newly supplied to the burned gas in the preceding cylinder and combustion in the subsequent cylinder is performed while the two cylinders are connected such that the gas to be guided to the exhaust passage An operation mode control means for executing the control of, and a fuel injection control means for controlling fuel injection for the subsequent cylinders in this special operation mode,
The fuel injection control means controls to inject fuel for the succeeding cylinder in the preceding cylinder at least in the operating area on the low load side among the operating areas in the special operation mode, and sets the injection start timing to the preceding start time. A diesel engine characterized by being set to a predetermined time in the latter half of the expansion stroke of the cylinder.   The diesel engine according to claim 1, wherein burned gas discharged from the preceding cylinder is introduced into the succeeding cylinder through an inter-cylinder gas passage.   The fuel injection control means is provided for the following cylinders in the preceding cylinder as the engine load increases and shifts from the low load side operation region to the high load side operation region in the operation region where the special operation mode is set. 3. The diesel engine according to claim 1, wherein the fuel injection start timing is delayed to a predetermined time until the first half of the exhaust stroke.   The operation mode control means performs control in a normal operation mode in which fresh air is introduced into each cylinder and each cylinder is burned in an independent state in an operation region on a higher load side than the operation region set as the special operation mode. The fuel injection control means controls to inject fuel to each cylinder in the normal operation mode, and sets the fuel injection timing to the compression stroke of each cylinder. The diesel engine according to any one of claims 1 to 3.   The operation mode control means opens the fresh air introduction intake valve that introduces fresh air into the subsequent cylinder in the high load side operation region of the operation region to be the special operation mode. The diesel engine according to any one of claims 1 to 4, wherein control is performed so that fresh air is introduced into the subsequent cylinder in addition to the burned gas derived from the preceding cylinder.   The fuel injection control means, in the operation region on the high load side of the operation region that is set to the special operation mode, the fuel injection for the subsequent cylinder is divided into the first injection in the preceding cylinder and the second injection in the subsequent cylinder. 6. The fuel injection system according to claim 1, wherein the fuel injection is set so as to be divided and the late injection timing of the fuel at the time of the divided injection is set in the latter half of the compression stroke of the subsequent cylinder. Diesel engine.   The diesel engine according to any one of claims 1 to 6, further comprising a supercharger for supercharging intake air to the preceding cylinder.

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