阅读说明：本技术 压燃式发动机及压燃式发动机的控制方法 (Compression ignition engine and control method thereof ) 是由 人见光夫 山本博之 山本寿英 藤本英史 于 2018-05-29 设计创作，主要内容包括：压燃式发动机包括发动机机体(柴油发动机(1))、第一燃料供给部(石脑油用喷油器(19))、第二燃料供给部(柴油燃料用喷油器(18))以及控制部(PCM(10)),该第二燃料供给部供给第二燃料,该第二燃料的实现压燃的压力和温度中的至少一者比第一燃料低,且该第二燃料比第一燃料难以气化,该控制部(PCM(10))分别向第一、第二燃料供给部输出信号。控制部向第一燃料供给部输出信号,以使所供给的第一燃料的重量比所供给的第二燃料的重量大,然后,控制部向第二燃料供给部输出信号,以使第二燃料供往燃烧室。所形成的混合气被压缩而自动着火。(A compression ignition engine is provided with an engine body (diesel engine (1)), a first fuel supply unit (naphtha injector (19)), a second fuel supply unit (diesel fuel injector (18)) that supplies a second fuel that has at least one of a pressure and a temperature that are lower than those of the first fuel and that is less likely to vaporize than the first fuel, and a control unit (PCM (10)) that outputs signals to the first and second fuel supply units, respectively. The control unit outputs a signal to the first fuel supply unit so that the weight of the first fuel supplied is greater than the weight of the second fuel supplied, and then outputs a signal to the second fuel supply unit so that the second fuel is supplied to the combustion chamber. The resulting mixture is compressed and automatically ignited.)

1. A compression ignition engine characterized by:

the compression ignition engine includes an engine body, a first fuel supply portion, a second fuel supply portion, and a control portion,

the engine body is provided with a combustion chamber,

the first fuel supply unit is configured to supply a first fuel to the combustion chamber,

the second fuel supply unit is configured to supply a second fuel to the combustion chamber, the second fuel having at least one of a pressure and a temperature at which compression ignition is achieved that is lower than the first fuel and being less likely to vaporize than the first fuel,

the control unit is configured to output signals to the first fuel supply unit and the second fuel supply unit,

the control unit outputs a signal to the first fuel supply unit so that the weight of the first fuel supplied to the combustion chamber is larger than the weight of the second fuel supplied to the combustion chamber, and then outputs a signal to the second fuel supply unit so that the second fuel is supplied to the combustion chamber,

the control unit outputs a signal to the first fuel supply unit and the second fuel supply unit to form an air-fuel mixture in the combustion chamber, and the air-fuel mixture is compressed to automatically ignite.

2. The compression ignition engine of claim 1, wherein:

the first fuel is a fuel having a lower boiling point than the second fuel.

3. The compression ignition engine according to claim 1 or 2, characterized in that:

the control unit outputs a signal to the first fuel supply unit and the second fuel supply unit so that the weight of the second fuel supplied to the combustion chamber is 10% or less of the weight of the total fuel supplied to the combustion chamber.

4. The compression ignition engine as claimed in any one of claims 1 to 3, wherein:

naphtha is included in the first fuel and diesel fuel is included in the second fuel.

5. The compression ignition engine as claimed in any one of claims 1 to 4, wherein:

gasoline is included in the first fuel and diesel fuel is included in the second fuel.

6. The compression ignition engine as claimed in any one of claims 1 to 5, wherein:

the control unit outputs signals to the first fuel supply unit and the second fuel supply unit so that the first fuel and the second fuel are supplied to the combustion chamber and are combusted, thereby ensuring that the air-fuel ratio of exhaust gas discharged from the combustion chamber is within a range of 14.5 to 15.0.

7. The compression ignition engine of claim 6, wherein:

the control unit outputs signals to the first fuel supply unit and the second fuel supply unit so that the first fuel and the second fuel are supplied to the combustion chamber to ensure that an air-fuel ratio of an air-fuel mixture in the combustion chamber is within a range of 14.5 to 15.0.

8. The compression ignition engine as claimed in any one of claims 1 to 5, wherein:

a three-way catalyst configured to purify an exhaust gas discharged from the combustion chamber is provided in an exhaust passage of the engine body,

the control unit outputs signals to the first fuel supply unit and the second fuel supply unit so that the first fuel and the second fuel are supplied to the combustion chamber and are combusted, thereby ensuring that an air-fuel ratio of an exhaust gas upstream of the three-way catalyst in the exhaust passage is a stoichiometric air-fuel ratio.

9. The compression ignition engine as claimed in any one of claims 1 to 8, wherein:

the first fuel supply portion is provided at a position where the first fuel is injected to an intake passage of the engine body,

the second fuel supply portion is provided at a position where the second fuel is injected into the combustion chamber.

10. The compression ignition engine as claimed in any one of claims 1 to 9, wherein:

the control portion outputs a signal to the first fuel supply portion to supply the first fuel to the combustion chamber during an intake stroke,

the control unit outputs a signal to the second fuel supply unit so that the second fuel is supplied to the combustion chamber during a compression stroke after an intake stroke.

11. The compression ignition engine of claim 9, wherein:

the control portion outputs a signal to the first fuel supply portion to cause the first fuel to be injected into the intake passage during an intake stroke,

the control unit outputs a signal to the second fuel supply unit so that the second fuel is injected into the combustion chamber during a compression stroke after an intake stroke.

12. A control method of a compression ignition engine, characterized in that:

the control unit outputs a signal to the first fuel supply unit to supply the first fuel to a combustion chamber of the engine,

the control portion outputs a signal to a second fuel supply portion such that a second fuel, at least one of which pressure and temperature at which compression ignition is achieved is lower than that of the first fuel and which is less likely to vaporize than the first fuel, is supplied to the combustion chamber after the end of the supply of the first fuel,

after the second fuel is supplied to the combustion chamber, an air-fuel mixture formed in the combustion chamber is compressed to automatically ignite,

the control unit outputs a signal to the first fuel supply unit so that the weight of the first fuel supplied to the combustion chamber is greater than the weight of the second fuel supplied to the combustion chamber.

Technical Field

The technology disclosed herein relates to a compression ignition engine and a control method of the compression ignition engine.

Background

Patent document 1 describes a diesel engine. The diesel engine aims to omit a high-cost selective reduction catalyst system including an exhaust gas purification system using a three-way catalyst. In order to purify exhaust gas using a three-way catalyst, the diesel engine adjusts the size and injection pressure of an injection hole that injects diesel fuel into a combustion chamber. In this way, the diesel fuel is diffused throughout the entire combustion chamber to form an air-fuel mixture having a stoichiometric air-fuel ratio, and the air-fuel mixture is combusted by compression ignition.

Patent document 2 describes a diesel engine in which gasoline as a secondary fuel is introduced into an intake passage through a carburetor and diesel fuel is injected into a combustion chamber. Patent document 2 describes the following: the ratio of diesel fuel to gasoline is such that the diesel fuel accounts for more than 50% of the total fuel quantity.

Patent document 3 describes a diesel engine in which vaporized naphtha is supplied into a combustion chamber through an intake passage, and liquid naphtha is injected into the combustion chamber. Patent document 3 describes the following: the amount of naphtha supplied to the combustion chamber through the intake passage is made not to exceed 50% of the total fuel amount.

Patent document 1: japanese patent No. 5620715

Patent document 2: british patent specification No. 714672

Patent document 3: british patent specification No. 821725

Disclosure of Invention

Technical problems to be solved by the invention

The diesel engine described in patent document 1 forms and burns an air-fuel mixture having a theoretical air-fuel ratio by spreading diesel fuel throughout a combustion chamber. However, the diesel engine described in patent document 1 has the following problems because the diesel fuel is difficult to vaporize: a portion where the fuel concentration locally becomes large occurs in the combustion chamber. If the fuel concentration becomes locally large, soot particles and carbon monoxide (CO) are generated in the combustion chamber.

The technology disclosed herein is intended to solve the above-mentioned problems, and has an object to: a compression ignition engine is provided which is capable of suppressing the generation of soot particulates and CO.

Technical solution for solving technical problem

More specifically, the technology disclosed herein relates to a compression ignition engine. The compression ignition engine is characterized in that: the engine body has a combustion chamber, the first fuel supply unit is configured to supply a first fuel to the combustion chamber, the second fuel supply unit is configured to supply a second fuel to the combustion chamber, at least one of a pressure and a temperature at which compression ignition is achieved for the second fuel is lower than the first fuel, and the second fuel is less likely to vaporize than the first fuel, the control unit is configured to output a signal to each of the first fuel supply unit and the second fuel supply unit, the control unit outputs a signal to the first fuel supply unit such that the weight of the first fuel supplied to the combustion chamber is greater than the weight of the second fuel supplied to the combustion chamber, and then the control unit outputs a signal to the second fuel supply unit, and a control unit for outputting a signal to the first fuel supply unit and the second fuel supply unit to form an air-fuel mixture in the combustion chamber, and the air-fuel mixture is compressed to automatically ignite the air-fuel mixture.

According to this configuration, the compression ignition engine includes the first fuel supply portion and the second fuel supply portion. Two fuels, a first fuel and a second fuel, are supplied to the combustion chamber. At least one of the pressure and temperature of the first fuel to effect compression ignition is higher than the second fuel, and the first fuel is susceptible to vaporization than the second fuel. At least one of the pressure and temperature of the second fuel to achieve compression ignition is lower than the first fuel, and the second fuel is more difficult to vaporize than the first fuel.

The first fuel supplied to the combustion chamber has a greater weight than the second fuel supplied to the combustion chamber. The readily vaporized first fuel primarily contributes to torque generation by the compression ignition engine. The second fuel, which is readily available for compression ignition, primarily helps to ignite the mixture.

The first fuel supply portion supplies the first fuel to the combustion chamber at a relatively early timing by receiving a signal from the control portion. If the supply timing is made earlier, the time from when the first fuel is supplied to the combustion chamber until the air-fuel mixture ignites and burns is long, and therefore the first fuel that is easily vaporized forms a homogeneous air-fuel mixture. This can suppress the generation of soot particles and CO during combustion. The emission performance of the compression ignition engine is improved.

The second fuel supply unit supplies the second fuel to the combustion chamber at a relatively later timing by receiving a signal from the control unit. The second fuel is used for igniting the air-fuel mixture, and the compression ignition timing and the combustion timing of the air-fuel mixture can be adjusted by adjusting the supply timing of the second fuel. The timing of supplying the second fuel to the combustion chamber is adjusted so that the air-fuel mixture is compressed to automatically ignite and combust at an appropriate timing, and thus the thermal efficiency of the compression ignition engine is improved.

Therefore, the compression ignition engine of the above configuration can suppress the generation of soot particles and CO, and can achieve an improvement in torque and an improvement in fuel efficiency.

It can also be: the first fuel is a fuel having a lower boiling point than the second fuel.

In this way, the first fuel is vaporized under the condition that the pressure and temperature in the combustion chamber are low, and therefore, the fuel can be supplied from the intake stroke in which the pressure in the combustion chamber is low. Since the fuel supply timing can be made earlier and the vaporization performance can be made higher, a homogeneous mixture can be formed even if the fuel supply amount of the first fuel is made larger. As a result, the generation of soot particles and CO can be suppressed, and the torque and the fuel saving performance can be improved.

It can also be: the control unit outputs a signal to the first fuel supply unit and the second fuel supply unit so that the weight of the second fuel supplied to the combustion chamber is 10% or less of the weight of the total fuel supplied to the combustion chamber.

In this way, the second fuel can be used for compression ignition of the mixture, and the compression ignition timing and the combustion timing can be adjusted by adjusting the supply timing of the second fuel.

It can also be: naphtha is included in the first fuel and diesel fuel is included in the second fuel. Because naphtha is more easily gasified than diesel fuel, it is advantageous to form a homogeneous mixture in the combustion chamber. Since diesel fuel is more easily ignited than naphtha, the mixture can be compressed and automatically ignited at an appropriate timing. Further, since naphtha is relatively inexpensive, naphtha is economically excellent.

It can also be: gasoline is included in the first fuel and diesel fuel is included in the second fuel. As described above, a homogeneous air-fuel mixture can be formed in the combustion chamber, and the air-fuel mixture can be compressed to automatically ignite at an appropriate timing.

It can also be: the control unit outputs signals to the first fuel supply unit and the second fuel supply unit so that the first fuel and the second fuel are supplied to the combustion chamber and are combusted, thereby ensuring that the air-fuel ratio of exhaust gas discharged from the combustion chamber is within a range of 14.5 to 15.0.

Further, it may be: the control unit outputs signals to the first fuel supply unit and the second fuel supply unit so that the first fuel and the second fuel are supplied to the combustion chamber to ensure that an air-fuel ratio of an air-fuel mixture in the combustion chamber is within a range of 14.5 to 15.0.

The air-fuel ratio of the air-fuel mixture in the combustion chamber is a ratio of the total weight of the fuel supplied to the combustion chamber to the weight of the air filled in the combustion chamber.

Thus, the air-fuel ratio of the exhaust gas can be set within a range of 14.5 to 15.0. Further, the conventional diesel engine is operated in a state where the air-fuel ratio of the mixture is made larger (leaner) than the stoichiometric air-fuel ratio, and the torque is improved by making the air-fuel ratio of the mixture substantially the stoichiometric air-fuel ratio as compared with the conventional diesel engine.

It can also be: and a control unit configured to output signals to the first fuel supply unit and the second fuel supply unit so that the first fuel and the second fuel are supplied to the combustion chamber and are combusted, thereby ensuring that an air-fuel ratio of an exhaust gas upstream of the three-way catalyst in the exhaust passage is a theoretical air-fuel ratio.

When the air-fuel ratio of the exhaust gas is set to the stoichiometric air-fuel ratio, CO, HC, and NOx in the exhaust gas can be purified by the three-way catalyst. The emission performance of the compression ignition engine is further improved. It should be noted that the air-fuel ratio range of 14.5 to 15.0 corresponds to a purge window (window) of the three-way catalyst, and if the air-fuel ratio is set to the stoichiometric air-fuel ratio, purging of the three-way catalyst is more reliable.

In the conventional diesel engine, it is necessary to increase the supercharging capability and make the air-fuel ratio at the time of combustion larger to reduce the soot particles, CO, and NOx, but in the present configuration, the first fuel is supplied to make the air-fuel ratio of the exhaust gas the stoichiometric air-fuel ratio, and the three-way catalyst is used in combination to reduce the soot particles, CO, and NOx without relying on supercharging as in the conventional diesel engine. Therefore, an inexpensive engine without a supercharger can be provided.

It can also be: the first fuel supply portion is provided at a position where the first fuel is injected into an intake passage of the engine body, and the second fuel supply portion is provided at a position where the second fuel is injected into the combustion chamber.

If the first fuel is injected into the intake passage, the first fuel is diffused in the combustion chamber by the flow of the intake air, and therefore a homogeneous mixture can be formed. This is advantageous in suppressing the generation of soot particles and the generation of CO.

The second fuel supply unit injects the second fuel into the combustion chamber, and therefore the second fuel can be supplied into the combustion chamber at an appropriate timing before compression top dead center. The mixture is compressed to automatically ignite and combust at an appropriate timing.

It can also be: the control unit outputs a signal to the first fuel supply unit to supply the first fuel to the combustion chamber during an intake stroke, and the control unit outputs a signal to the second fuel supply unit to supply the second fuel to the combustion chamber during a compression stroke after the intake stroke.

It can also be: the control portion outputs a signal to the first fuel supply portion to cause the first fuel to be injected into the intake passage during an intake stroke, and the control portion outputs a signal to the second fuel supply portion to cause the second fuel to be injected into the combustion chamber during a compression stroke following the intake stroke.

In the method of controlling a compression ignition engine disclosed herein, a control unit outputs a signal to a first fuel supply unit to supply a first fuel to a combustion chamber of the engine, the control unit outputs a signal to a second fuel supply unit to supply a second fuel to the combustion chamber after the first fuel is supplied, at least one of a pressure and a temperature at which compression ignition is achieved for the second fuel is lower than that of the first fuel, and the second fuel is less likely to be vaporized than the first fuel, and after the second fuel is supplied to the combustion chamber, an air-fuel mixture formed in the combustion chamber is compressed to automatically ignite, and the control unit outputs a signal to the first fuel supply unit to make a weight of the first fuel supplied to the combustion chamber greater than a weight of the second fuel supplied to the combustion chamber.

In this way, the generation of soot particles and CO can be suppressed, and thus the emission performance of the compression ignition engine is improved. Further, since the second fuel is supplied, the air-fuel mixture can be compressed and automatically ignited and combusted at an appropriate timing, the torque of the compression ignition engine and the fuel saving performance can be improved.

Effects of the invention

As described above, according to the compression ignition engine and the control method of the compression ignition engine described above, the generation of soot and CO can be suppressed, and therefore the emission performance of the compression ignition engine can be improved.

Drawings

FIG. 1 is a schematic diagram showing a configuration example of an engine system;

fig. 2 is a block diagram showing an example of a configuration related to control of the engine system;

fig. 3 is a diagram showing an example of fuel injection timings;

FIG. 4 is a diagram showing an example of a preferred operating section of the engine system;

FIG. 5 is an explanatory diagram of intake late-closure control;

fig. 6 is a flowchart showing a specific example of control of the engine system;

fig. 7 is a diagram showing the main elements of the engine system;

fig. 8 is a diagram showing an example of a relationship between Indicated Mean Effective Pressure (IMEP) and indicated fuel consumption rate (gross) ISFC) according to the embodiment;

fig. 9 is a diagram showing an example of the relationship between Indicated Mean Effective Pressure (IMEP) and NOx emission amount according to the embodiment.

Fig. 10 is a diagram showing an example of a change in the in-cylinder pressure with respect to the crank angle in the embodiment.

Detailed Description

Embodiments of the present invention will be described below with reference to the drawings. The following preferred embodiments are merely examples for essentially illustrating the present invention, and are not intended to limit the present invention, its application objects, or its uses.

Fig. 1 shows a schematic configuration of an engine system. Fig. 2 shows an example of the configuration related to control of the engine system. The engine system is mounted on a four-wheel vehicle. The engine system disclosed herein is suitable for use in large vehicles such as large trucks, for example. However, the engine system disclosed herein can be widely applied to various four-wheel vehicles regardless of the size of the vehicle.

The engine system includes a diesel engine 1 as a compression ignition engine. The vehicle is advanced by the operation of the diesel engine 1.

This engine system is configured to supply diesel fuel (that is, diesel oil or fuel containing diesel oil as a main component) and a different fuel having a different characteristic from that of the diesel fuel to the diesel engine 1. The heterogeneous fuel has the following characteristics: at least one of the pressure and temperature to effect compression ignition is higher and has a lower boiling point than diesel fuel. The fuel of the different fuel ratio is easily gasified than the diesel fuel, and is difficult to ignite than the diesel fuel. The heterogeneous fuel corresponds to the first fuel and the diesel fuel corresponds to the second fuel. The heterogeneous fuel is a fuel mainly used for generating torque. Diesel fuel is the fuel used primarily for fires.

Specifically, the heterogeneous fuel is naphtha. The naphtha that can be used by the engine system includes light naphtha, heavy naphtha, and full range naphtha. The boiling point ranges of light naphtha, heavy naphtha and full boiling range naphtha are different. Further, the engine system may use modified naphtha in which a small amount of crude oil or heavy oil is mixed in naphtha.

The heterogeneous fuel may be gasoline in addition to naphtha. The different fuel is not limited to one fuel, and may be a fuel in which two or more kinds of fuels are mixed. For example, a mixed fuel of naphtha and gasoline, a mixed fuel of naphtha and other fuel, or a mixed fuel of gasoline and other fuel may be used as the different fuel.

Next, an engine system for supplying diesel fuel and naphtha to the diesel engine 1 will be described.

< construction of Engine System >

The diesel engine 1 has a cylinder block 11 provided with a plurality of cylinders 11a (only one is shown in fig. 1), a cylinder head 12 provided on the upper side of the cylinder block 11, and an oil pan 13 provided on the lower side of the cylinder block 11 and storing lubricating oil. A piston 14 is inserted into each cylinder 11a of the diesel engine 1, and the piston 14 reciprocates along the center axis of the cylinder. The piston 14 is connected to a crankshaft 15 via a connecting rod 14 b. A cavity defining a combustion chamber 14a of a reduced type is formed in a top surface of the piston 14. The diesel engine 1 is configured such that the geometric compression ratio thereof is 13 or more and 18 or less.

On the cylinder head 12 of each cylinder 11a, an intake passage 16 and an exhaust passage 17 are formed. On the intake passage 16, an intake valve 21 that opens and closes an opening of the combustion chamber 14a is provided. On the exhaust passage 17, an exhaust valve 22 that opens and closes an opening of the combustion chamber 14a is provided.

The diesel engine 1 includes, as a Valve train that drives the intake Valve 21, a Sequential-Valve Timing (S-VT) 71 that makes the Valve Timing variable (see fig. 2). The intake air S-VT 71 can be formed in various known configurations such as a hydraulic type and an electric type. The diesel engine 1 changes the valve timing of the intake valve 21 according to the operating state.

An injector 19 for naphtha as a first fuel supply portion and an injector 18 for diesel fuel as a second fuel supply portion are mounted on the cylinder head 12.

The naphtha injector 19 is configured to inject naphtha into the intake passage 16. That is, the injector 19 for naphtha is provided so that an injection hole for injecting naphtha is directed to the intake passage 16 of each cylinder 11 a. The naphtha stored in the first fuel tank 191 is supplied to the naphtha injector 19 through a naphtha supply path, not shown.

The diesel fuel injector 18 is configured to directly inject diesel fuel into the combustion chamber 14 a. That is, the diesel fuel injector 18 is provided as an injection hole for injecting diesel fuel from the bottom surface of the cylinder head 12 toward the cylinder 11 a. In the illustrated example, the diesel fuel injector 18 is provided on the center axis of the cylinder 11a, but the diesel fuel injector 18 can be provided at an appropriate position. The diesel fuel stored in the second fuel tank 181 is supplied to the diesel fuel injector 18 through a diesel fuel supply path not shown.

An ignition assisting device is also mounted on the cylinder head 12. The ignition assisting means assists the mixture to ignite when the diesel engine 1 is in a specific operating state. Specifically, the ignition assisting device is an ignition device 20 that ignites the air-fuel mixture by spark ignition. The ignition device 20 is provided such that an electrode protrudes into the combustion chamber 14a, and detailed illustration thereof is omitted. The ignition assisting means may employ a glow plug, which improves the ignitability of the fuel by heating the air in the cylinder 11a, instead of the ignition means.

An intake passage 30 is connected to one side surface of the diesel engine 1. The intake passage 30 communicates with the intake passage 16 of each cylinder 11 a. The intake passage 30 introduces air and EGR gas into each cylinder 11 a. An exhaust passage 40 is connected to the other side surface of the diesel engine 1. The exhaust passage 40 communicates with the exhaust port 17 of each cylinder 11 a. The exhaust passage 40 discharges burned gas from each cylinder 11 a. A turbocharger 61 for supercharging air is provided in the intake passage 30 and the exhaust passage 40, as will be described in detail later.

An air cleaner 31 for filtering air is provided at an upstream end of the intake passage 30. A surge tank (purge tank)33 is provided near the downstream end of the intake passage 30. The intake passage 30 located on the downstream side of the surge tank 33 constitutes an independent passage that branches off for each cylinder 11 a. The downstream end of each independent passage is connected to the intake passage 16 of each cylinder 11 a.

A compressor 61a of a turbocharger 61, an intercooler 35, and a throttle valve 36 are provided between the air cleaner 31 and the surge tank 33 in the intake passage 30, the intercooler 35 cools air compressed by the compressor 61a, and the throttle valve 36 adjusts the amount of air. The intercooler 35 may be air-cooled or water-cooled. The throttle valve 36 is substantially in a fully open state, but when, for example, a large amount of EGR gas is to be recirculated to the intake passage 30, the throttle valve 36 is narrowed in order to generate negative pressure in the intake passage 30.

The upstream side portion of the exhaust passage 40 is constituted by an exhaust manifold. The exhaust manifold has a merging portion where a plurality of independent passages branched toward each cylinder 11a and connected to the outboard end of the exhaust passage 17 and a plurality of independent passages merge.

A turbine 61b of the turbocharger 61, an exhaust gas purifying device 41 for purifying harmful components in the exhaust gas, and a muffler 42 are provided in this order from the upstream side on the downstream side of the exhaust manifold in the exhaust passage 40.

The exhaust gas purifying device 41 has a three-way catalyst 41 a. The three-way catalyst 41a simultaneously purifies Hydrocarbons (HC), carbon monoxide (CO), and nitrogen oxides (NOx) in the exhaust gas. The three-way catalyst 41a oxidizes hydrocarbons into water and carbon dioxide, oxidizes carbon monoxide into carbon dioxide, and reduces nitrogen oxides into nitrogen. The three-way catalyst 41a can purify the exhaust gas more than enough when the air-fuel ratio (the weight ratio of air to fuel) of the exhaust gas is the stoichiometric air-fuel ratio, and the three-way catalyst 41a can purify the exhaust gas even when the air-fuel ratio is within a purification window of the stoichiometric air-fuel ratio of approximately 14.5 to 15.0.

The exhaust gas purification device 41 may further include a particulate filter that traps particulates such as soot particulates contained in the exhaust gas, in addition to the three-way catalyst 41 a.

An exhaust gas recirculation passage 51 is interposed between the intake passage 30 and the exhaust passage 40. The exhaust gas recirculation passage 51 recirculates a part of the exhaust gas to the intake passage 30. The upstream end of the exhaust gas recirculation passage 51 is connected to a portion of the exhaust passage 40 between the exhaust manifold and the turbine 61b (i.e., an upstream portion of the turbine 61 b). The downstream end of the exhaust gas recirculation passage 51 is connected to a portion of the intake passage 30 between the surge tank 33 and the throttle valve 36 (i.e., a downstream portion of the compressor 61 a). The exhaust gas recirculation passage 51 is provided with an EGR valve 51a for adjusting the amount of recirculation of exhaust gas into the intake passage 30, and an EGR cooler 52 for cooling the exhaust gas with engine coolant.

The turbocharger 61 has a compressor 61a provided in the intake passage 30 and a turbine 61b provided in the exhaust passage 40. The compressor 61a and the turbine 61b are coupled to each other, and the compressor 61a and the turbine 61b rotate integrally. The compressor 61a is provided on the intake passage 30 between the air cleaner 31 and the intercooler 35. The turbine 61b is provided between the exhaust manifold on the exhaust passage 40 and the exhaust gas purification device 41. The turbine 61b is rotated by the exhaust gas flow, thereby rotating the compressor 61a to compress air.

The exhaust passage 40 is connected to an exhaust bypass passage 65 that bypasses the turbine 61 b. An exhaust gas bypass valve (exhaust gate valve)65a for adjusting the amount of exhaust gas flowing through the exhaust gas bypass passage 65 is provided in the exhaust gas bypass passage 65. The exhaust bypass valve 65a is configured to be in a fully Open state (Normal Open) when not energized.

< construction of control device for engine >

As shown in fig. 1 and 2, the diesel engine 1 is controlled by a powertrain control module (hereinafter referred to as PCM) 10. The PCM10 is constituted by a microprocessor having a CPU, a memory, a counter timer group, an interface, and a path connecting these units. The PCM10 constitutes a control device (and a control unit). As shown in fig. 2, detection signals of various sensors are input to the PCM 10. The sensor herein includes: a liquid temperature sensor SW1 that detects the temperature of the engine coolant; a boost pressure sensor SW2 attached to the surge tank 33 and detecting the pressure of air supplied to the combustion chamber 14 a; an intake air temperature sensor SW3 that detects an air temperature; a crank angle sensor SW4 that detects the angle of rotation of the crankshaft 15; an accelerator opening sensor SW5 that detects an accelerator opening corresponding to an operation amount of an accelerator pedal (not shown) of the vehicle; o is₂Sensors SW6 that are attached to the exhaust passage on the upstream side and the downstream side of the three-way catalyst 41a and that detect the oxygen concentration in the exhaust gas; exhaust pressure sensor SW7 aligned withThe exhaust pressure on the upstream side of the turbine 61b in the air passage 40 is detected; an air flow sensor SW8 that detects the flow rate of intake air entering the intake passage 30; an EGR valve opening degree sensor SW9 that detects the opening degree of the EGR valve 51 a; an intake valve phase angle sensor SW10 that detects the phase angle of the intake valve 21; and an exhaust gas bypass valve opening degree sensor SW11 that detects the opening degree of the exhaust gas bypass valve 65 a.

The PCM10 determines the states of the diesel engine 1 and the vehicle by performing various calculations based on the detection signals of the sensors SW1 to SW11, and outputs control signals to the respective actuators of the diesel fuel injector 18, the naphtha injector 19, the ignition device 20, the intake air S-VT 71, the throttle valve 36, the EGR valve 51a, and the tail gas bypass valve 65 a.

< control of engine >

The PCM10 basically controls the diesel engine 1 by determining a target torque mainly based on the accelerator opening degree and causing the diesel fuel injector 18 and the naphtha injector 19 to inject fuel corresponding to the target torque.

The PCM10 also adjusts the amount of air introduced into the cylinder 11a in accordance with the operating state of the diesel engine 1. Specifically, the PCM10 regulates the amount of air by controlling the opening degrees of the throttle valve 36 and the EGR valve 51a (i.e., EGR control) and/or controlling the valve timing of the intake valve 21 using the intake air S-VT 71 (i.e., intake late-close control). If the late-close control is performed, that is, if the intake valve 21 is closed (defined as the closing timing when the lift height of the intake valve 21 is 0.4 mm) in the range of 60 ° to 120 ° after the intake bottom dead center in the middle stage of the compression stroke (the middle stage when the crank angle of the compression stroke is trisected by 180 ° into the early stage, the middle stage, and the late stage), the amount of air taken into the cylinder 11a can be adjusted without increasing the pumping loss. Further, if the EGR gas is recirculated, not only the amount of air taken into the cylinder 11a can be adjusted, but also the ignitability of the air-fuel mixture can be improved by increasing the temperature in the cylinder 11a (complementary to the insufficient increase in the temperature in the cylinder 11a near the compression top dead center as the effective compression ratio is lowered in the intake late-close control). In addition, in a high load section where the temperature in the cylinder 11a is high, if the EGR gas is recirculated, the low-temperature inert gas flowing through the EGR cooler 52 is recirculated to the combustion chamber 14a, so that the mixture (naphtha) can be suppressed from being ignited early, and the mixture can be ignited at an appropriate ignition timing that can generate a high engine torque.

PCM10 also performs air-fuel ratio feedback control, i.e., according to O₂The air amount and the fuel amount are adjusted by the oxygen concentration in the exhaust gas detected by the sensor SW6 and the intake air flow rate detected by the air flow sensor SW 8. The PCM10 sets the air-fuel ratio of the air-fuel mixture in the combustion chamber 14a, that is, the weight ratio (a/F) of the air (a) to the fuel (F) in the combustion chamber 14a, to the stoichiometric air-fuel ratio, thereby setting the air-fuel ratio of the exhaust gas discharged from the combustion chamber 14a to the stoichiometric air-fuel ratio.

Since the a/F is 14.5 to 15.0, which is an air-fuel ratio corresponding to the purge window of the three-way catalyst 41a, the air-fuel ratio in the combustion chamber 14a may be set to substantially the stoichiometric air-fuel ratio (14.5 to 15.0), so that the air-fuel ratio of the exhaust gas discharged from the combustion chamber 14a may be set to 14.5 to 15.0. The fuel amount referred to herein is the amount of the total fuel including both diesel fuel and naphtha. This engine system performs air-fuel ratio feedback control over the entire operating range of the diesel engine 1. In this way, the engine system purifies the exhaust gas by the three-way catalyst 41a throughout the entire operation section of the diesel engine 1.

(Fuel injection control)

The fuel injection control performed by the PCM10 is explained below. As described above, this engine system supplies the diesel engine 1 with naphtha mainly for generating torque and diesel fuel mainly for ignition. Comparing the weight of naphtha supplied to the weight of diesel fuel supplied, the weight of naphtha is greater than the weight of diesel fuel. The diesel fuel is made to be 10% by weight or less based on the total amount of fuel supplied to the combustion chamber 14 a. It is also possible to have diesel fuel, for example, 5% by weight of the total fuel.

Because naphtha has a lower boiling point than diesel fuel, it is easily vaporized in the combustion chamber 14 a. Then, a homogeneous mixture close to the stoichiometric air-fuel ratio is formed in the combustion chamber 14a by the naphtha. Thereby suppressing soot particle generation and suppressing CO generation.

On the other hand, at least one of the pressure and temperature for achieving compression ignition of naphtha is higher than that of diesel fuel. That is, the ignition of naphtha is low. As described above, the diesel engine 1 is configured to have a low compression ratio with a geometric compression ratio of 13 to 18, which is disadvantageous in terms of fuel ignition.

In this engine system, diesel fuel having excellent ignitability is supplied into the combustion chamber 14 a. Since the diesel fuel functions as an ignition fuel, the air-fuel mixture is compressed and can be automatically ignited reliably at a predetermined timing. A mixture comprising naphtha and diesel fuel is combusted.

Fig. 3 shows an example of the injection timing of naphtha and diesel fuel at a predetermined engine speed. The naphtha injector 19 for naphtha installed in the intake passage 16 injects naphtha into the intake passage 16 during the intake stroke in which the intake valve 21 is opened. The injection timing of naphtha may be set in a period from the middle stage to the early stage of the intake stroke. Here, the early stage and the middle stage of the intake stroke may be set to the early stage and the middle stage when the intake stroke is divided into three periods of the early stage, the middle stage, and the late stage by three, respectively. The intake air in the cylinder 11a flows relatively quickly during the period from the middle stage to the early stage of the intake stroke. By injecting the naphtha during this period, the naphtha can be diffused throughout the combustion chamber 14a by the intake air flow, and the mixture can be homogenized.

A diesel fuel injector 18 installed to project into the combustion chamber 14a injects diesel fuel into the combustion chamber 14a during a compression stroke. The injection timing of the diesel fuel may be set to a timing near compression top dead center, specifically, within a period of 30 to 10 ° CA before compression top dead center. In this way, the mixture can be compressed to automatically ignite near the compression top dead center, and combustion is started. If the combustion center of gravity of the combustion is located 5 to 10 CA degrees after the compression top dead center, the thermal efficiency of the diesel engine 1 is improved. As described above, since the geometric compression ratio of the diesel engine 1 is low, it is possible to avoid the mixture containing naphtha from being ignited prematurely before the diesel fuel is injected. By adjusting the injection timing of the diesel fuel, the timing at which the mixture is compressed and automatically ignited can be adjusted.

In the present embodiment, diesel fuel is used as the ignition fuel in the middle load section (S1 section) and the high load section (S2 section) of the operation section of the engine 1 shown in fig. 4. In the low load range (P range) and the engine 1 cold or forced start (CS range), the naphtha fuel is set to 100% without supplying the diesel fuel, and the fuel is ignited by the ignition assisting device.

When the engine load is low or when the engine is cold, the combustion chamber temperature is low, and therefore it is difficult to obtain desired ignitability even if diesel fuel is supplied. Further, the intake late-close control reduces the effective compression ratio of the engine, and deteriorates the ignitability of the fuel.

Then, in the low load section (P section) and the cold or forced start of the engine 1 (CS section), the fuel is ignited by the ignition assisting device without using the diesel fuel. It is also possible to supply diesel fuel and operate the ignition-assisting device.

(EGR control)

As described above, in order to make the air-fuel ratio a/F of the combustion chamber 14a substantially stoichiometric and improve the ignitability of the fuel, the PCM10 controls the EGR valve 51a to recirculate a part of the exhaust gas from the exhaust passage 40 to the intake passage 30(EGR) at least in the low load side operating region when both naphtha and diesel fuel are supplied to the combustion chamber 14 a.

In the operating section of the engine 1 shown in fig. 4, the middle load section (S1 section) and the high load section (S2 section) are operating sections in which both naphtha and diesel fuel are supplied to the combustion chamber 14a, and EGR is performed in at least the middle load section (S1 section), which is an operating section on the low load side.

In the present embodiment, the PCM10 executes EGR in a low load section (P section), an intermediate load section (S1 section), and a high load section (S2 section) of the engine 1 shown in fig. 4. The EGR rate (the ratio of the amount of returned exhaust gas to the total amount of the amount of intake air) in the operating section with a high engine load is made lower than that in the operating section with a low engine load. Specifically, in the low load section (P section) and the medium load section (S1 section), the EGR valve 51a is controlled so that the EGR rate becomes 40%, and in the high load section (S2 section), the EGR valve 51a is controlled so that the EGR rate is in the range of 30% to 0% and the EGR rate becomes lower as the load becomes higher.

(air intake delay closing control)

The PCM10 executes intake late-closing control using the intake air S-VT 71 in the engine low load section (P section) in addition to the EGR control so that the air-fuel ratio a/F of the combustion chamber 14a becomes substantially the stoichiometric air-fuel ratio.

Here, basically, the throttle valve 36 is controlled to change in the closing direction in order to obtain an intake negative pressure for EGR. That is, as a method of making the air-fuel ratio a/F substantially the stoichiometric air-fuel ratio (a method of reducing the amount of fresh air introduced), control of the throttle valve can be used, but this control of the throttle valve causes a large pumping loss.

In the present embodiment, the air-fuel ratio control performs a late closing control of the intake valve 21 (this control increases the opening period of the intake valve 21 in the compression stroke).

In fig. 5, the imaginary contour line shows the reference valve timing of the intake valve 21, and in the present embodiment, the closing timing is 30 ° CA after intake bottom dead center. The PCM10 delays the closing timing of the intake valve 21 so that the lower the engine load, the smaller the amount of intake air. The solid line of fig. 5 shows the valve timing at which the closing timing of the intake valve 21 is retarded to 90 ° CA after intake bottom dead center. Note that the closing timing of the intake valve 21 is defined as when the lift amount of the intake valve 21 is reduced to 0.4 mm.

< specific example of Engine control >

As shown in fig. 6, the detection signals of the sensors SW1 to SW11 are read to determine whether the operating state of the engine 1 is in the CS zone (cold or forced start) (S1, S2).

When the operating state of the engine 1 is in the CS range, the routine proceeds to step S3, where the exhaust bypass valve 65a is opened. Thus, the exhaust gas bypasses the turbine 61b and is sent to the three-way catalyst 41 a. Therefore, the heat of the exhaust gas can be prevented from being taken away by the turbine 61b, which is advantageous for the early temperature rise of the three-way catalyst 41a by the heat of the exhaust gas. In the next step S4, the naphtha injector 19 is driven at a predetermined timing in the intake stroke so that the fuel supplied to the combustion chamber 14a is 100% naphtha and the air-fuel ratio is equal to or less than the stoichiometric air-fuel ratio (fuel-rich fuel with a/F of 15 or less). In the next step S5, the ignition device 20 is operated so that the fuel is ignited at a predetermined timing near compression top dead center.

If the operating state of the engine 1 is not in the CS range in step S2, the routine proceeds to step S6, where it is determined whether or not the operating state of the engine 1 is in the P range (low load range).

When the operating state of the engine 1 is in the P range, the process proceeds to step S7, where the opening degree of the EGR valve 51a is controlled so that the EGR rate becomes 40%. In the next step S8, the intake air S-VT 71 is driven so that the closing timing of the intake valve 21 becomes a predetermined late closing timing. In the next step S9, the naphtha injector 19 is driven at a predetermined timing in the intake stroke so that the fuel supplied to the combustion chamber 14a is 100% naphtha and the air-fuel ratio is near the stoichiometric air-fuel ratio (a/F is about 14.7). In the next step S10, the ignition device 20 is operated so that the fuel is ignited at a predetermined timing near compression top dead center.

If the operating state of the engine 1 is not in the P section in step S6, the routine proceeds to step S11, where it is determined whether or not the operating state of the engine 1 is in the S1 section (intermediate load section).

When the operating state of the engine 1 is in the section S1, the routine proceeds to step S12, where the opening degree of the EGR valve 51a is controlled so that the EGR rate becomes 40%. Then, control is performed so that the valve timing of the intake valve 21 is set as a reference timing (a virtual contour line of fig. 5). In the next step S13, the naphtha injector 19 is driven at a predetermined timing in the intake stroke so that the naphtha accounts for 95% of the total amount of fuel supplied to the combustion chamber 14a and the air-fuel ratio becomes the stoichiometric air-fuel ratio. In the next step S14, the diesel fuel injector 18 is driven at a predetermined timing in the second half of the compression stroke so that the diesel fuel occupies 5% of the total fuel supplied to the combustion chamber 14a and the air-fuel ratio becomes the stoichiometric air-fuel ratio.

In step S11, the operating state of the engine 1 is not in the S1 interval or is in the S2 interval (high load interval) when the operating state of the engine 1 is in the S2 interval. In this case, the process proceeds to step S15, where the opening degree of the EGR valve 51a is controlled so that the EGR rate becomes 30% or less. Then, control is performed so that the valve timing of the intake valve 21 is set as a reference timing (a virtual contour line of fig. 5). In the next step S16, the naphtha injector 19 is driven at a predetermined timing in the intake stroke so that the naphtha accounts for 95% of the total amount of fuel supplied to the combustion chamber 14a and the air-fuel ratio becomes the stoichiometric air-fuel ratio. In the next step S17, the diesel fuel injector 18 is driven at a predetermined timing in the second half of the compression stroke so that the diesel fuel occupies 5% of the total fuel supplied to the combustion chamber 14a and the air-fuel ratio becomes the stoichiometric air-fuel ratio.

< control example >

Fig. 7 shows an example of main elements related to combustion control in the engine 1 having the geometric compression ratio of 16 in a low load range (P range), a medium load range (S1 range), and a high load range (S2 range) at an engine speed of 1500 rpm. The numerical values shown here are merely examples, and can be changed in accordance with specifications. Each numerical value is a reference value, and may include a plurality of deviations in practical use.

In the low load region, the EGR rate is 40%, and a large amount of EGR gas is introduced into the combustion chamber. Intake late-closing control is performed at the closing timing (IVC) of the intake valve, which is 90 ° CA after intake bottom dead center. In addition, since stable compression ignition is difficult to achieve by reducing the effective compression ratio by the intake late-closure control, the mixture is forcibly ignited by the ignition assisting device, and the fuel is only naphtha, which is not only inexpensive but also capable of forming a homogeneous mixture and advantageous for reducing emissions.

In the middle load region, the EGR rate is 40% as in the low load region, and a large amount of EGR gas is introduced into the combustion chamber. The intake valve closing timing (IVC) is returned to the reference setting, which is 30 ° CA after intake bottom dead center. Since stable compression ignition can be achieved, the mixture gas is combusted by compression ignition without using an ignition assisting device. Stable compression ignition can be achieved by adding 5% diesel fuel to the main fuel, i.e., naphtha. Further, since the inert gas (EGR gas) having a low temperature cooled by the EGR cooler 52 is introduced into the combustion chamber, rapid combustion of the air-fuel mixture after ignition can be suppressed, and an increase in combustion noise and an increase in thermal load can be suppressed.

In the high load range, the EGR rate is 30%, and the air amount is relatively increased to achieve efficient combustion. As in the intermediate load range, the closing timing (IVC) of the intake valve is 30 ° CA after intake bottom dead center, and stable compression ignition can be achieved. As in the medium load compartment, the fuel was 5% diesel fuel and 95% naphtha. Further, since the inert gas (EGR gas) having a low temperature cooled by the EGR cooler 52 is introduced into the combustion chamber, early ignition of the air-fuel mixture (naphtha) can be suppressed, and the ignition timing can be set to the ignition timing at which a high engine torque can be generated.

Similarly, in the high engine speed range, the EGR rate is 30%, and the air amount is relatively increased to achieve efficient combustion. The intake valve closing timing (IVC) is set to a timing at which the intake charge amount can be made large in the high speed section, and is about 45 ° CA after intake bottom dead center. In the high speed range, the elapsed time from the intake stroke to the compression stroke is shorter than that in the low speed range, and therefore, the naphtha feed period for feeding naphtha through the intake port 16 is long in terms of crank angle, while the time interval from the end of the naphtha feed to the vicinity of the compression top dead center is particularly short, and the resulting naphtha mixture becomes inhomogeneous, but the recirculation of EGR gas promotes the vaporization of naphtha, so that deterioration of homogenization is suppressed, soot particles are not generated, and the engine torque can be increased. In the high speed range, 5% diesel fuel and 95% naphtha are used, but when the optimum ignition timing cannot be obtained due to the relationship between the engine speed and the time interval from the start of naphtha supply to the vicinity of the compression top dead center, 100% naphtha may be supplied and the mixture may be forcibly ignited by the ignition assisting device.

As described above, in the case where the EGR gas is recirculated in the high speed range, when the main fuel is diesel fuel, the amount of soot particles increases, so that the EGR gas cannot be recirculated.

Fig. 8 shows a relationship between Indicated Mean Effective Pressure (IMEP) and indicated fuel consumption rate (gross) ISFC in the example and the conventional example (100% diesel fuel) relating to the control elements shown in fig. 7. Since the air-fuel ratio is set to substantially the stoichiometric air-fuel ratio in the embodiment, the command fuel consumption rate is reduced at each of the low load, the medium load, and the high load as compared with the conventional example in which the lean operation is performed. That is, the engine torque and fuel economy performance of the engine system disclosed herein are improved as compared to existing diesel engine systems.

Fig. 9 shows an example of the relationship between Indicated Mean Effective Pressure (IMEP) and NOx emission amount in the above-described embodiment and conventional example. In the conventional example, if the engine load becomes high, the NOx emission amount from the combustion chamber increases. In contrast, in the embodiment, the amount of exhaust from the tail pipe downstream of the three-way catalyst 41a is shown, and the air-fuel ratio of the exhaust gas discharged from the combustion chamber 14a is made the stoichiometric air-fuel ratio and NOx is purified by the three-way catalyst 41a, so the amount of NOx discharged is substantially zero. That is, the emissions performance of the engine system disclosed herein is improved as compared to existing diesel engine systems.

Fig. 10 shows changes in the in-cylinder pressure with respect to the crank angle in the above embodiment. According to this map, in both the case where diesel fuel is used as the ignition fuel with IMEP 852 (intermediate load region) and the case where IMEP 1440 (high load region), a peak of the in-cylinder pressure occurs before around 20 ° CA after compression top dead center. Therefore, the following steps are carried out: the fuels (naphtha and diesel fuel) are burned at the time when the diesel fuel ignites to increase the thermal efficiency.

As described above, this engine system supplies the diesel engine 1 with naphtha for generating torque and diesel fuel for ignition. By using naphtha having excellent gasification performance, a homogeneous air-fuel mixture close to the stoichiometric air-fuel ratio is formed in the entire combustion chamber 14a, whereby generation of soot particles and CO can be suppressed. Further, the exhaust gas can be purified by the three-way catalyst 41a provided in the exhaust passage 40 by setting the weight ratio (a/F) of the fuel including both naphtha and diesel fuel and air to substantially the stoichiometric air-fuel ratio and setting the air-fuel ratio of the exhaust gas discharged from the combustion chamber 14a to the stoichiometric air-fuel ratio with respect to the air-fuel mixture in the combustion chamber 14 a. An aftertreatment system for purifying NOx, which is required in a conventional diesel engine, can be omitted, and simplification and cost reduction of the engine system can be achieved. In the engine system, the air-fuel ratio of the mixture is set to substantially the stoichiometric air-fuel ratio, as compared with a conventional diesel engine that operates with a lean mixture, and therefore, the engine torque can be increased.

The invention disclosed herein is not limited to the above configuration. For example, the air-fuel ratio of the mixture may be made considerably larger (lean) than the stoichiometric air-fuel ratio (for example, 30 to 45 a/F) between a low load range and a light load range in which the total fuel injection amount is small. If the air-fuel ratio is set to about 30 to 45, the generation of NOx in the combustion chamber 14a can be suppressed. Further, naphtha (first fuel) can be directly injected into the combustion chamber.

In the above configuration, the turbocharger 61 is mounted, but the turbocharger is not necessarily mounted. That is, in the conventional diesel engine, it is necessary to install a supercharger to make the air-fuel ratio at the time of combustion large to reduce soot and CO and to reduce NOx by using a selective reduction catalyst at high cost, or it is necessary to install a plurality of superchargers to make the air-fuel ratio at the time of combustion large by remarkably increasing the supercharging pressure and to reduce the combustion temperature by lowering the compression ratio of the engine body, but in the present invention, the air-fuel ratio of the mixture can be made to be in the range of 14.5 to 15.0 by supplying the first fuel, so that by using the three-way catalyst 41a in combination, the soot and CO can be reduced without depending on the supercharging pressure and NOx can be purified more than enough, and as a result, an inexpensive engine without installing a supercharger can be provided.

-description of symbols-

1 Diesel engine (Engine body)

10 PCM (control part)

14a combustion chamber

16 air inlet channel

18 Diesel fuel injector (second fuel supply part)

19 naphtha injector (first fuel supply part)

40 exhaust passage

41a three-way catalyst