JP2006328999A - Premixed compression self-ignition type internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To restrain generation of NOx and smoke, by forming a uniform lean mixture, by promoting diffusion of fuel and mixing with air, while preventing an injected fuel spray from directly colliding with a cylinder liner, in a premixed compression self-ignition type internal combustion engine. <P>SOLUTION: A recessed head cavity 29 is formed on an explosive surface of a cylinder head 12 opposed to a top surface of a piston 11. A fuel injection nozzle 13 having a plurality of fuel injection ports 44 for injecting the fuel toward a peripheral wall surface 31 of the head cavity 29, is arranged in a bottom part of the head cavity 29. A bottom surface 30 of the head cavity is formed as a flat surface in substantially parallel to the explosive surface of the cylinder head 12. The peripheral wall surface 31 is inclined in an expansively opening shape toward the piston side. Fuel injection from the injection nozzle 13 is performed by an optional combination of main injection for starting in the vicinity of the top dead center of a compression stroke and pre-injection for performing the fuel injection in the initial stage or the intermediate stage of the compression stroke. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to a premixed compression self-ignition internal combustion engine.

  In recent years, in the field of internal combustion engines such as diesel engines, stricter exhaust regulations have been enforced year by year due to the growing environmental problems, reducing emissions of soot, smoke, CO, NOx and other air pollutants, and fuel consumption There is a strong demand for improvement.

  However, since diesel engines spray fuel on compressed air in a cylinder and burn it by self-ignition, light and dark mixing tends to occur in the mixture of air and fuel, and soot and smoke can be produced where the fuel concentration is high. It is easy to generate NOx near the theoretical ratio.

  In view of such circumstances, in recent years, a premixed compression self-ignition system has attracted attention. This premixed compression auto-ignition method injects fuel from the beginning to the middle of the compression stroke before the piston reaches top dead center, so that the fuel and air are uniformly mixed in advance in the combustion chamber. This is a technology that raises the temperature and raises the pressure by compression by the self-ignition at the ignition point of the fuel. According to this method, the generation of smoke can be suppressed by a high excess air ratio, and the generation of NOx can be suppressed because the flame temperature can be made uniform.

  However, in the premixed compression self-ignition system, fuel is injected when the piston is at a low position before reaching the top dead center, so that the fuel may directly collide with the cylinder liner. In the process of raising the piston, the temperature and pressure in the cylinder do not rise so much, so the fuel that has collided with the cylinder liner is likely to adhere as it is without evaporating. The adhering fuel may be scraped off by the piston to dilute the oil, or may be discharged as white smoke without contributing to the combustion as it is, and naturally causes a deterioration in fuel consumption.

  On the other hand, in Patent Document 1 below, a fuel injection port is formed in which a concave dome (head cavity) is formed on the explosion surface of the cylinder head facing the top surface of the piston and opens toward the peripheral wall surface of the dome at the bottom of the dome. A diesel engine with an injection nozzle having the following is disclosed.

  When premixed compression auto-ignition is performed using this technique, the fuel collides with the peripheral wall of the dome before reaching the cylinder liner, and the direction is changed and atomized. It becomes possible to prevent adhesion.

JP 10-141622 A

  In Patent Document 1, the dome of the cylinder head is formed in a hemispherical shape, and most of the fuel that collides with the peripheral wall surface of this dome diffuses downward, but part of it diffuses upward. However, since there is little space above the fuel collision portion, there is a drawback that a fuel rich mixture is generated and smoke is likely to be generated.

  In the premixed compression self-ignition internal combustion engine, the present invention prevents the injected fuel spray from directly colliding with the cylinder liner and promotes the diffusion of the fuel and the mixing with the air, thereby producing a uniform lean mixture. The purpose of this is to suppress the generation of NOx and smoke.

  According to the first aspect of the present invention, a concave head cavity is formed on the explosion surface of the cylinder head facing the top surface of the piston, and a plurality of fuels inject fuel toward the peripheral wall surface of the head cavity at the bottom of the head cavity. In a premixed compression self-ignition internal combustion engine provided with a fuel injection nozzle having an injection port, the bottom surface of the head cavity is formed on a flat surface substantially parallel to the explosion surface of the cylinder head, and the peripheral wall surface is a piston. A combination of main injection in which fuel injection from the injection nozzle starts near the top dead center of the compression stroke and pre-injection in which fuel injection is performed at the initial or middle stage of the compression stroke. It is performed by.

  According to a second aspect of the present invention, in the first aspect of the invention, the pre-injection is further performed in a plurality of times.

  A third aspect of the invention is characterized in that, in the first aspect of the invention, the fuel injection from the injection nozzle is further performed in the vicinity of the overlap TDC in the transition period from the exhaust stroke to the intake stroke.

  The invention according to claim 4 is the invention according to claim 1, wherein an oxidation catalyst for removing carbon monoxide or unburned hydrocarbon is provided in an exhaust passage communicating with an exhaust hole formed in the cylinder head. It is characterized by.

  According to a fifth aspect of the invention, in the first aspect of the invention, an EGR device that circulates EGR gas into the cylinder, a target EGR rate map in which a target EGR rate corresponding to the engine speed and the fuel injection amount is recorded, The EGR device is controlled so that the EGR gas amount obtained from the target EGR rate map is circulated into the cylinder according to the engine speed and the fuel injection amount.

  The invention according to claim 6 is the invention according to claim 1, further comprising an intake valve and an exhaust valve for opening and closing the intake hole and the exhaust hole of the cylinder head, respectively, and variably setting the opening and closing timing of the intake valve and / or the exhaust valve. A variable valve device is provided, and the variable valve device is controlled so as to obtain an effective compression ratio according to the engine operating state.

  A seventh aspect of the invention is characterized in that, in the first aspect of the invention, a variable swirl generating device is provided that generates the swirl strength in a variable manner according to the engine operating state.

  According to the first aspect of the present invention, the fuel injected from the fuel injection port is prevented from directly adhering to the cylinder liner by colliding with the peripheral wall surface of the head cavity, and the fuel is atomized by the collision. Therefore, mixing with air is promoted. In addition, since the bottom surface of the head cavity is formed on a flat surface substantially parallel to the explosion surface of the cylinder head, it is possible to form a wider space on the bottom surface side of the head cavity than the collision portion of the fuel. And mixing of fuel and air in the head cavity can be promoted.

  Further, by performing the fuel injection in a plurality of stages of pre-injection and main injection, the output that is insufficient only by the combustion by the pre-injection can be supplemented by the combustion by the main injection, and the HC and CO generated by the combustion by the pre-injection Can be incinerated and removed by combustion by main injection.

  According to the invention of claim 2, by performing the pre-injection in a plurality of times, the fuel injection period in each pre-injection can be shortened and the spray reach distance can be shortened. Therefore, the possibility that the fuel adheres directly to the cylinder liner can be further reduced.

  According to the invention of claim 3, the fuel injected in the vicinity of the overlap TDC does not adhere to the cylinder liner, and the fuel can be evaporated during the intake stroke by contacting the high-temperature piston. Moreover, it becomes possible to cool a piston with a fuel.

  According to the invention of claim 4, HC and CO generated by the combustion of the premixed gas accompanying the pre-injection are incinerated and removed by the combustion accompanying the main injection, but further, by providing an oxidation catalyst in the exhaust pipe, It becomes possible to suppress the emission of HC and CO.

  According to the invention of claim 5, by performing the exhaust gas recirculation according to the engine operating condition, the ignition delay of the premixed gas by the pre-injection is extended, the mixture is made uniform and diluted, and the main injection is performed. The high temperature combustion area generated by the accompanying combustion can be reduced.

  According to the sixth aspect of the present invention, when the fuel injection amount is high at high load or the like, the effective compression ratio is lowered to extend the ignition delay, and when the fuel injection amount at low load is small, the effective compression is reduced. The ratio can be increased to improve ignitability, and knocking and misfire can be prevented.

  According to the invention of claim 7, when the injection amount of the pre-injection is large at the time of high load or the like, the swirl flow is strengthened to promote uniform dilution of the air-fuel mixture, and the ignition delay is extended to prevent knocking. When the pre-injection amount at low load is small, the swirl flow is weakened to form a fuel rich region and a lean region in layers, and the ignition is actively promoted in the overconcentrated region to misfire. Can be prevented.

Hereinafter, embodiments of the present invention will be described with reference to the drawings.
The internal combustion engine of the present embodiment is a premixed compression auto-ignition type diesel engine, and a basic operation process will be briefly described before the details of the structure are described. FIG. 1 is an explanatory diagram showing an outline of the operation process of the diesel engine of the present embodiment. In this diesel engine, a piston 11 is slidably fitted to a cylinder 10, and a fuel injection nozzle 13, an intake hole 14, and an exhaust hole 15 are provided in a cylinder head 12, and the intake hole 14 and the exhaust hole 15 are provided. Each of them can be opened and closed by an intake valve 17A and an exhaust valve 17B. A combustion chamber 18 is formed between the top surface of the piston 11, the side surface of the cylinder 10, and the bottom surface of the cylinder head 12.

  In FIG. 1, (a) is an intake stroke, (b) is a compression stroke, (c) is an expansion stroke, and (d) is an exhaust stroke. In the intake stroke (a), the piston 11 moves from the top dead center to the bottom dead center, and air is taken into the combustion chamber 18 from the intake hole 14 during that time. In the compression stroke of (b), the intake air is compressed and fuel S is injected (pre-injected) from the fuel injection nozzle 13 from the initial stage to the middle stage of compression to form a mixture of air and fuel ( Premix). When the piston 11 is in the vicinity of the top dead center, the premixed gas is heated and pressurized by compression and self-ignited. Further, near the top dead center, the fuel S is injected again (main injection; shown in (c)), the fuel S immediately diffuses while evaporating, and self-ignites in a high-temperature and high-pressure atmosphere. In the expansion stroke (c), the piston 11 is lowered, and in the exhaust stroke (d), the exhaust valve 17B is opened, and the exhaust in the combustion chamber 18 is discharged from the exhaust hole 15 by the lift of the piston 11.

  Therefore, in the present embodiment, multiple stages of injection are performed, that is, pre-injection for premixing and main injection for normal compression self-ignition combustion (diesel combustion). And in order to perform such multi-stage injection, optimization of the shape of the said cylinder 10, the piston 11, and the fuel-injection nozzle 13 is achieved. Hereinafter, this point will be described in detail.

[Configuration of Piston 11]
2 and 3 are cross-sectional views of a cylinder top portion of a diesel engine to which the present invention is applied. In particular, FIG. 2 is a state where the piston 11 is below the top dead center (a state where pre-injection is performed), FIG. 3 shows a state where the piston 11 is near top dead center (a state where main injection is performed). As shown in FIG. 2, a concave piston cavity 20 is formed on the top surface of the piston 11. A central protrusion 21 that protrudes toward the cylinder head 12 is formed at the center of the bottom of the piston cavity 20, and the central protrusion 21 has a large outer diameter at the base (bottom) and a smaller outer diameter at the top. The outer peripheral surface 22 is an inclined surface that is inclined so as to expand toward the anti-cylinder head 12 side. C is attached to the expansion angle of the outer peripheral surface 22.

  The bottom surface 24 of the piston cavity 20 is formed in an arc surface that is gently connected to the inclined surface 22 of the central projecting portion 21. The outer peripheral surface 25 of the piston cavity 20 is formed as a circular arc surface continuous from the bottom surface 24 and extends to the top surface of the piston 11. The opening 26 of the piston cavity 20 is slightly inward in the radial direction from the outer peripheral surface 25, thereby forming a lip portion 27 that protrudes inward in the radial direction. In FIG. 2, φY is given to the opening diameter of the piston cavity 20, and φZ is given to the maximum outer diameter of the outer peripheral surface 25 of the piston cavity 20.

[Configuration of Cylinder Head 12]
FIG. 5 is an enlarged cross-sectional view of the cylinder head. As shown in FIGS. 2, 3, and 5, the explosive surface 28 of the cylinder head 12 facing the top surface of the piston 11 is provided with a concave head cavity 29. It is formed in a conical shape. That is, the bottom surface 30 of the head cavity 29 is formed as a flat surface substantially parallel to the explosion surface 28, and the peripheral wall surface 31 of the head cavity 29 is expanded toward the piston 11 side (lower side in FIG. 2). It is inclined to. In FIG. 2, φX is attached to the opening diameter of the head cavity 29, and A is attached to the opening angle of the peripheral wall surface 31.

  6 is a cross-sectional view taken along the line VI-VI in FIG. The cylinder head 12 is provided with two intake and exhaust valve seats 33, and each valve seat 33 is provided with intake and exhaust valves 17A and 17B. The head cavity 29 is surrounded by the four valves 17 </ b> A and 17 </ b> B and has a size that circumscribes each valve seat 33. However, even if it does not necessarily circumscribe each valve seat 33, it can be formed as large as possible within the range of the circumscribed circle.

  As shown in FIG. 5, in the present embodiment, a fuel collision portion 34 formed of a material different from the cylinder head 12 is provided on the peripheral wall surface 31 of the head cavity 29. The fuel collision portion 34 has a width over the entire depth of the head cavity 29 and is formed to have a uniform thickness over the width. As a raw material, what has heat resistance and heat insulation is used, for example, SUH heat-resistant steel etc. are used. However, the fuel collision part 34 may be omitted, and the peripheral wall surface 31 may be made of the material of the cylinder head 12 itself.

[Configuration of Fuel Injection Nozzle 13]
As shown in FIGS. 2 and 3, the fuel injection nozzle 13 is provided at the center of the bottom surface 30 of the head cavity 29. The injection nozzle 13 has a small-diameter portion 36 that accommodates a needle valve in the lower portion and a large-diameter portion 37 that is larger in diameter than the small-diameter portion 36 in the upper portion, and between the small-diameter portion 36 and the large-diameter portion 37. A stepped portion 38 is formed in the upper portion. The injection nozzle 13 is attached to a holding hole 39 formed in the cylinder head 12 via a holder 40. The holder 40 is formed in a cylindrical shape, and the inside of the cylinder is formed in a shape that matches the outer shape of the injection nozzle 13. Between the inner surface of the holder 40 and the small-diameter portion 36, a cylindrical first sealing material 41 is provided to keep the airtight therebetween, and an annular second seal is provided between the holder 40 and the step portion 38. A material 42 is interposed. The tip of the holder 40 is exposed on the bottom surface of the head cavity 29 through the holding hole 39. In addition, a bent portion 43 protrudes from the front end portion of the holder 40 in the radially inward direction, and the first seal material 41 is received from the lower side by the bent portion 43.

  As shown in FIG. 5, a plurality of fuel injection ports 44 are formed at the tip of the injection nozzle 13. Four to ten fuel injection ports 44 are formed radially about the nozzle axis O1. Each injection port 44 opens toward the peripheral wall surface 31 (fuel collision part 34) of the head cavity 29, respectively.

[Relationship Between Head Cavity 29 and Fuel Injection Nozzle 13]
As shown in FIG. 2, the fuel injected from the fuel injection port 44 of the injection nozzle 13 diffuses in a radial direction conically or horizontally as a whole. When the diffusion angle (spray angle) of the spray axis X1 at this time is B, 120 ° ≦ B ≦ 180 ° is set. By setting in this way, the fuel directly collides with the peripheral wall surface 31. By this collision, the direction of the fuel is changed and the fuel is atomized.

  Therefore, when pre-injection is performed (FIGS. 1B and 2), it is possible to prevent the fuel from directly colliding with the cylinder liner 46 and adhering thereto. Further, since the fuel is atomized by the collision with the peripheral wall surface 31, the mixing with the air can be further promoted. In FIG. 2, as a comparison, the spray axis X <b> 2 when it does not collide with the peripheral wall surface 31 of the head cavity 29 is illustrated by a two-point difference line. In this case, the fuel collides with the cylinder liner 46.

  If the spray angle B is set to 120 ° or more, it is necessary to form the head cavity 29 deep in order to cause the fuel to collide with the peripheral wall surface 31 of the head cavity 29. This is because it involves a change in structure. The reason why the spray angle B is set to 180 ° or less is that if it is more than that, the flow coefficient in the injection nozzle 13 decreases. In addition, when performing the main injection, a large amount of fuel diffuses in the upper side (the bottom surface 30 side) in the head cavity 29, and the fuel concentration in the portion becomes high, which causes the generation of smoke.

  On the other hand, the expansion angle A of the peripheral wall surface 31 of the head cavity 29 is set to 60 ° ≦ A ≦ 120 °. The smaller the spread angle A, the lower the possibility that the fuel after the collision will adhere to the cylinder liner 46. However, reducing the expansion angle A hinders fuel diffusion when performing main injection (FIGS. 1C and 3). That is, it becomes difficult to promote the mixing of the fuel after the main injection and collides with the peripheral wall surface 31 and a large amount of air existing outside in the cylinder 10 in the radial direction, which hinders the formation of a uniform air-fuel mixture. Further, a large amount of fuel colliding with the peripheral wall surface 31 of the head cavity 29 is diffused to the upper side in the head cavity 29, and smoke is likely to be generated due to excessive fuel concentration in the head cavity 29. Therefore, in the present embodiment, by setting the expansion angle A as described above, appropriate mixture formation is realized in both the pre-injection and the main injection.

  Further, assuming that the intersection angle between the peripheral wall surface 31 of the head cavity 29 and the fuel spray axis X1 on the piston 11 side is D, 120 ° ≦ D <180 ° is set. When D is 120 ° or less, in the case of performing main injection, the diffusion of fuel to the upper side in the head cavity 29 increases, and smoke is likely to be generated. This is because the fuel hardly collides with the peripheral wall surface 31.

  The head cavity 29 has a bottom surface 30 formed on a flat surface substantially parallel to the explosion surface 28 of the cylinder head 12, and a corner 47 is formed at the boundary between the bottom surface 30 and the peripheral wall surface 31 as shown in FIG. 5. Has been. Due to the presence of the corner portion 47, a space R wider than the conventional technique (dome-shaped or conical head cavity) is secured on the upper side in the head cavity 29. Therefore, even if the fuel that collided with the peripheral wall surface 31 of the head cavity 29 diffuses to the upper side (the bottom surface 30 side) of the head cavity 29, mixing with air is promoted as compared with the prior art.

  The fuel injected from the fuel injection port 44 does not have to collide with the peripheral wall surface 31 as long as it is set so that at least the spray axis X1 collides with the peripheral wall surface 31. In other words, when the distance in the cylinder axial direction between the portion where the spray axis X1 collides with the peripheral wall surface 31 and the explosion surface 28 is E, it is only necessary to satisfy the condition of E> 0. The speed of the fuel droplets is highest on the spray axis X1, and the speed decreases due to the shearing action with the surrounding air as it goes radially outward, so the fuel below the spray axis X1 becomes mist-like. The penetrating force is reduced, so that the cylinder liner 46 is less likely to reach the cylinder liner 46 without colliding with the peripheral wall surface 31, and the cylinder liner 46 is diffused by the fuel colliding with the peripheral wall surface 31 being diffused as a downward component. This is because the penetrating ability to go to is further weakened.

  The fuel is in the form of a liquid column immediately after exiting from the fuel injection port 44 and then reaches the peripheral wall surface 31 while diffusing in a mist form. In the present embodiment, the distance (length) L of the spray axis X1 from the fuel injection port 44 to the peripheral wall surface 31 and the diameter φd of the fuel injection port 44 are set to have a relationship of L / φd ≧ 50. By setting in this way, the liquid column portion S1 (the length is indicated by L1) of the fuel S hardly collides and adheres directly to the peripheral wall surface 31.

  If the liquid columnar fuel adheres to the peripheral wall surface 31 of the head cavity 29 during low load operation such as when starting the engine, white smoke is discharged without being involved in combustion. Such a problem can be solved by setting the diameter φd. The above relationship between the distance L and the diameter φd is particularly effective when the fuel injection pressure is about 50 MPa or more.

  As described above, the fuel collision portion 34 having heat insulation and heat resistance is provided on the peripheral wall 31 of the head cavity 29. Although the cylinder head 12 is cooled by the cooling water, the temperature of the fuel collision part 34 is difficult to be lowered due to the heat insulating property. Therefore, the fuel that has collided with the fuel collision part 34 is easily evaporated by the heat of the fuel collision part 34, and the attached fuel is prevented from being discharged as white smoke.

  As shown in FIG. 6, the head cavity is provided so as to circumscribe the four valve seats 33 for intake and exhaust. That is, the valve seat 33 is formed to the maximum extent that does not erode. As a result, the distance from the fuel injection nozzle 13 to the peripheral wall surface 31 of the head cavity 29 can be made as long as possible, and the liquid column portion S1 of the fuel is prevented from adhering to the peripheral wall surface 31.

[Relationship between head cavity and piston cavity]
The opening diameter φX of the head cavity 29 is formed smaller than the diameter φY of the opening 26 of the piston cavity 20. Specifically, there is a relationship of φY ≧ 1.2 × φX. Therefore, in the case of performing main injection (FIGS. 1C and 3), most of the fuel injected from the injection nozzle 13 and colliding with the peripheral wall surface 31 of the head cavity 29 is introduced into the piston cavity 20. The squish flow into the piston cavity 20 during the compression stroke and the reverse squish flow out of the piston cavity 20 during the expansion stroke cause a strong turbulence in the flow of fuel, and the air flows outside the piston 11 in the radial direction. Mixing is promoted more.

  The diameter φY of the opening 26 of the piston cavity 20 is formed smaller than the outer diameter (maximum outer shape) φZ of the outer peripheral surface 25 of the piston cavity 20, and the opening 26 has a lip portion protruding radially inward. Since 27 is formed, a stronger squish flow and a reverse squish flow can be obtained.

[Relationship between injected fuel and piston cavity 20]
The protrusion 21 formed in the piston cavity 20 is formed in a mountain shape as described above, and the expansion angle C of the outer peripheral surface 22 is set to C ≦ 140 °. In the case of performing the main injection (FIGS. 1C and 3), the fuel injected from the injection nozzle 13 and colliding with the head cavity 29 spreads further in a conical shape, and the outer peripheral surface 22 of the protrusion 21 in the piston cavity 20. Collide with the slope. Therefore, even after colliding with the projecting portion slope 22, the fuel is further guided radially outward to promote mixing with air.

  As shown in FIG. 3, the diameter φV of the portion where the fuel collides with the piston cavity 20 and the maximum outer diameter φZ of the piston cavity 20 are in a relationship of φV ≧ 0.5 × φZ. Therefore, when the main injection is performed, the fuel spreads radially outward in the piston cavity, and mixing with air is promoted. Therefore, the fuel is injected near the center of the piston cavity 20 to prevent the occurrence of “thermal pinch” in which the flame contracts without using radially outward air.

  The main injection may be delayed not only when the piston 11 is substantially at the top dead center but also after passing the top dead center mainly for NOx reduction. In such a case, if the injected fuel is injected out of the piston cavity 20, the generation of a uniform air-fuel mixture is hindered and smoke is increased. Therefore, in the present embodiment, when the piston 11 is within the range of the crank angle ± 40 ° from the top dead center, the fuel injection angle B and the peripheral wall surface are ensured so that the injected fuel surely enters the piston cavity 20. An expansion angle A of 31, an opening diameter φY of the piston cavity 20 and the like are set. As an example, FIG. 4 shows the arrangement of the pistons 11 when the crank angle is top dead center ± 30 °.

  In the present embodiment, when the fuel spray axis X1 that collides with the peripheral wall surface 31 spreads out to the maximum in the radial direction, that is, after the fuel spray axis X1 collides with the peripheral wall surface 31, The spray axis X 1 is set so as to enter the piston cavity 20, assuming the case of diffusion along the inclination of the wall surface 31.

[Application example of this embodiment]
The basic configuration of the diesel engine of the present embodiment is as described above, and an application example in which this diesel engine is systemized together with various peripheral devices will be described below. FIG. 7 is a schematic view showing the engine system. The engine system 65 includes an engine controller 66, a supercharger 67, an intercooler 68, an EGR device 69, an intake valve operating device 70, an exhaust variable valve operating device 71, and a variable swirl device 72.

  The intake hole 14 provided in the cylinder head 12 communicates with the outside air via the intake passage 73, the intercooler 68, the first intake pipe 74, the compressor portion 75 of the supercharger 67, and the second intake pipe 76. Is communicated with the outside air via the exhaust passage 77, the turbine section 78 of the supercharger 67 and the exhaust pipe 79. An oxidation catalyst 80 is disposed in the exhaust pipe 79. An EGR pipe 81 of the EGR device 69 communicates between the exhaust passage 77 and the intake passage 73, and the EGR pipe 81 is provided with an EGR valve 82 whose opening degree can be adjusted.

  The engine controller 66 includes an ECU that controls various devices, together with an arithmetic device and various storage devices. The fuel injection nozzle 13, the intake valve operating device 70, the exhaust variable valve operating device 71, the variable swirl device 72, and the EGR device 69 are connected to the output portion of the engine controller 66. The operation is controlled by the ECU.

  The input of the engine controller 66 includes an engine speed sensor (not shown), a load sensor (accelerator opening detection sensor) (not shown), an in-cylinder (in-cylinder) pressure sensor (not shown), an intake flow sensor 85, An intake (supply) temperature sensor 86, an intake (supply) pressure sensor 87, a cooling water temperature sensor 88, and an abnormal combustion detection sensor (such as a knock sensor) 89 are connected.

  The engine speed sensor is arranged on a crankshaft or a gear or wheel fixed to the crankshaft, and detects the crankshaft speed (engine speed). The load sensor is provided in a fuel increasing / decreasing mechanism or the like of an accelerator device or a fuel injection pump, and detects a fuel injection amount as a load. The intake air flow sensor 85 is provided in the second intake pipe 76 on the upstream side of the supercharger 67, and detects the flow rate of air flowing through the second intake pipe 76. The intake air temperature sensor 86 and the intake pressure sensor 87 are provided in the first intake pipe 74 on the downstream side of the intercooler 68, and detect the temperature and pressure of the air flowing through the first intake pipe 74. The cooling water temperature sensor 88 is disposed, for example, in a cooling water jacket around the cylinder 10 and detects the temperature of the cooling water around the cylinder 10. The abnormal combustion detection sensor 89 is disposed in the side wall of the cylinder 10 or in the cylinder 10 and detects abnormal combustion such as knocking by detecting abnormal vibration of the cylinder 10 or abnormal pressure fluctuation in the combustion chamber 18. ing.

[About fuel injection timing]
FIG. 8 is a time chart showing the timing of fuel injection from the fuel injection nozzle 13. In the diesel engine of the present embodiment, as already described, main injection and pre-injection (premature injection) are performed. The main injection is started near the top dead center of the compression stroke, for example, 10 ° to 0 ° before the top dead center, or after that. The pre-injection is performed in the initial stage or middle stage of the compression stroke, for example, between 120 ° and 40 ° before top dead center.

  In this way, not only pre-injection but also main injection improves output during high loads and burns carbon monoxide (CO) and unburned hydrocarbons (HC) generated by combustion by pre-injection. Can be made. In addition, NOx reduction associated with main injection can be expected due to the combustion gas of the premixed gas having an internal EGR effect.

[Setting of injection stage number for pre-injection]
8A shows an example in which the pre-injection and the main injection are performed only once, while FIG. 8B shows an example in which the pre-injection is performed in two stages. In the present embodiment, as shown in FIG. 2, the fuel injected from the fuel injection port 44 collides with the peripheral wall surface 31 of the head cavity 29, thereby preventing the fuel after the collision from directly contacting the cylinder liner 46. However, depending on the operating conditions, it cannot be said that there is no possibility. Therefore, as shown in FIG. 8B, the pre-injection is divided into a plurality of stages, and the injection amount (injection period) of each stage is reduced, so that the fuel can be reliably prevented from reaching the cylinder liner 46. .

Whether or not to divide the pre-injection and how many times it is divided are determined as follows.
FIG. 10 is a flowchart for determining fuel injection conditions. According to this flowchart, first, in step S1, the basic injection amount is determined from the detected engine speed and load in consideration of the accelerator opening and the injection pressure. In step S2-1, the pre-injection amount is calculated from the basic injection amount, and in step S2-2, the pre-injection timing is calculated. In step S3, the injection period is estimated from the calculated pre-injection amount and pre-injection timing.

  Next, in step S <b> 4, a spray tip reaching distance model (described later) when fuel is injected for the estimated injection period is calculated, and it is verified whether or not the spray adheres to the cylinder liner 46. If it is determined in step S5 that there is no adhesion to the cylinder liner 46 as a result of verification, in step S6, a condition for performing pre-injection only once is established.

  As a result of the verification, if it is determined that there is adhesion to the cylinder liner 46, the injection amount and the injection timing of each stage are calculated assuming that the injection is divided into two stages (steps S7 and S8). . Then, in the same manner as described above, the injection period is estimated (step S3), the spray tip reach distance model of each stage is calculated, and whether or not the spray adheres to the cylinder liner 46 is verified (steps S4 and S5). If it is determined that the fuel does not adhere, the condition for dividing the pre-injection into two stages is determined (step S6). If it is determined that the fuel is attached, the number of pre-injection divisions is further increased to perform the same operation. repeat.

The fuel tip reach distance model is calculated as follows.
In this embodiment, the fuel injected from the fuel injection nozzle 13 once sprays into the combustion chamber 18 after colliding with the peripheral wall surface 31 of the head cavity 29. Therefore, if the spray tip reaching distance when performing normal spraying (free spraying) with no collision with the head cavity 29 is Lp, the spray tip reaching distance Lp ′ when colliding with the head cavity 29 is set to Lp. Can be calculated by Lp ′ = σ · Lp.

  The correction function σ is a function that takes into account the injection pressure and spray angle of the fuel injection nozzle 13, the expansion angle of the head cavity peripheral wall surface 31, the atmospheric pressure and temperature in the combustion chamber 18, and the like. The value of σ can also be obtained by quoting from a map created based on data obtained in advance by a rig test.

The spray tip reach distance Lp of free spray can be calculated by the following equation.
Lp = σ · A · (Δp / ρa) ^0.25 · (dn · t) ^0.5
Here, σ is an experimental constant, and is obtained by a test bench or the like. A is the area of the nozzle hole and is shown as A = π / 4 · dn ² . Δp is the difference between the injection pressure and the cylinder internal pressure. ρa is the density of the gas in the cylinder. dn is the diameter of the nozzle. t is an injection period.

  Further, when pre-injection is performed in a plurality of stages, the first-stage spray tip arrival distance Lp ′ and the second-stage spray tip reach distance Lp ″ are further different values. Fuel injection is performed in an atmosphere where the flow of air in the spraying direction is small, whereas the second stage fuel injection is caused by the flow of ambient air in the injection direction caused by the first stage fuel injection. Therefore, the second stage spray tip arrival distance Lp ″ is obtained by multiplying the first stage spray tip arrival distance Lp ′ by a predetermined correction coefficient (acceleration coefficient) α, Lp ″ = α Lp ′ (or Lp ″ = α · σ · Lp). The correction coefficient α is obtained empirically from the atmospheric conditions in the combustion chamber 18 and the injection conditions.

  FIG. 9 is a graph illustrating the relationship between the elapsed crank angle (time) and the spray tip reaching distance, and the diagram A shows the spray tip reaching distance when the pre-injection is not divided. For example, when the wall surface of the cylinder liner 46 is at a distance La, the fuel spray collides with the cylinder liner 46 when the elapsed crank angle (elapsed time) is θa. In such a case, the pre-injection is divided as described in the above flowchart. FIG. 9 is a diagram B showing a state in which the pre-injection is divided into three stages. In this case, the injection period is set so that the pre-injection at each stage stops before reaching the cylinder liner 46. It can be seen that in each pre-injection, the second stage distance is longer than the first stage, and the third stage distance is longer than the second stage.

  As described above, in the present embodiment, in addition to the configuration in which the fuel collides with the peripheral wall surface 31 of the head cavity 29, the pre-injection is divided, and the spray tip arrival distance of each injection is shortened. Fuel adhesion is reliably prevented.

  FIG. 8C is an example in which spraying is performed not only in the pre-injection and the main injection but also in the overlap TDC (TDC (OL)) during the subsequent period from the exhaust stroke to the intake stroke. In this example, since the fuel is injected when the piston 11 is near the top dead center, it does not adhere directly to the cylinder liner 46. Further, when the piston temperature after exhaust is high, the injected fuel can be evaporated and the piston 11 can be cooled.

[About oxidation catalyst 80]
In the engine system diagram of FIG. 7, an oxidation catalyst 80 is provided in the exhaust pipe 79. As in the present embodiment, CO and HC emissions are unavoidable in the method in which the premixed gas by pre-injection is compressed and ignited. Therefore, by providing the oxidation catalyst 80 in the exhaust pipe 79, highly oxidative CO and HC are removed, and the exhaust to the outside air is reduced to realize clean exhaust.

[EGR device 69]
As shown in FIG. 7, the EGR device 69 of this embodiment is a so-called external EGR device 69, and an EGR pipe 81 that communicates an exhaust passage 77 and an intake passage 73, and an EGR valve 82 provided in the EGR pipe 81. And. Further, a target EGR rate map as shown in FIG. 11 is stored in the storage device of the engine controller 66. In the target EGR rate map, an optimal EGR rate corresponding to a change in the engine operating state, that is, a change in the engine speed and the basic fuel injection amount is recorded. Specifically, the EGR rate decreases as the engine speed increases, and the EGR rate decreases as the basic injection amount increases.

  The EGR valve 82 (FIG. 7) is controlled by the engine controller 66 and adjusts the amount of exhaust gas circulating in the combustion chamber 18 based on the EGR rate obtained from the target EGR rate map according to the engine operating state. FIG. 12 is a flowchart for determining the reflux gas amount. First, in step S1, the basic injection amount is determined by taking into account the accelerator opening from the detected engine speed and load, and in step S2, the target injection amount is determined. An appropriate EGR rate is obtained from the EGR rate map. In step S3, the opening degree of the EGR valve 82 is determined in consideration of detected values of the intake flow rate sensor 85, the intake pressure sensor 87, and the like. In step S4, the actuator of the EGR valve 82 is controlled.

  When performing multi-stage fuel injection of pre-injection and main injection as in this embodiment, increase the pre-injection ratio as much as possible, increase the combustion of the uniform premixed gas, and reduce NOx etc. Is preferred. However, if the pre-injection amount is increased, knocking is likely to occur, so there is a limit to increasing the pre-injection amount.

  Therefore, in the present embodiment, exhaust gas recirculation by the EGR device 69 increases the amount of exhaust gas having a large heat capacity and suppresses the rise in the in-cylinder temperature, thereby extending the ignition delay and uniformly diluting the mixture. As well as suppressing knocking, the pre-injection amount can be increased as much as possible. Moreover, since the temperature rise at the time of combustion accompanying main injection can be suppressed, it is further advantageous for reduction of NOx.

  The EGR device 69 is not limited to an external EGR, but can be an internal EGR or a combination of an external EGR and an internal EGR. As the internal EGR device, a method of adjusting the amount of exhaust gas remaining in the combustion chamber 18 by changing the opening / closing timing of the exhaust valve or the intake valve according to the engine load can be adopted. In this case, it is only necessary to determine the opening / closing timing of the intake valve or the exhaust valve and control the valve operating device of each valve as shown in parentheses in step S3 of FIG.

[Variable valve gear and effective compression ratio setting]
As in the present embodiment, in the method in which a mixture of fuel and air is formed in advance in the combustion chamber 18 and this mixture is self-ignited by adiabatic compression by the piston 11, the load (fuel injection amount) increases. In-cylinder temperature and pressure rise rapidly, and the self-ignition timing may be accelerated, causing knocking. Conversely, when the load is small, the in-cylinder pressure and temperature do not increase, which may cause misfire. . In order to prevent such knocking and misfire, it is necessary to appropriately control the in-cylinder temperature and pressure so that self-ignition occurs near the top dead center of the compression stroke. To that end, the effective compression ratio is changed appropriately. It is effective to do.

  In the present embodiment, exhaust valve reopening that controls the operation of the exhaust valve 17B is employed as a method for obtaining an optimal effective compression ratio in accordance with various engine operating states (operating regions).

  As shown in FIG. 7, the intake valve 17A that opens and closes the intake hole 14 of the cylinder head 12 is connected to an intake valve operating device 70, and the exhaust valve 17B that opens and closes the exhaust hole 15 includes an exhaust variable type movement. A valve device 71 is connected. The exhaust variable valve operating device 71 is configured to temporarily re-open the exhaust valve 17B during the compression stroke of the piston 11, and can change the valve closing timing of the re-opening. It has become.

  The intake valve operating apparatus 70 and the exhaust variable valve operating apparatus 71 are configured such that the valve opening / closing timing is controlled by an ECU included in the engine controller 66. In particular, the exhaust variable valve operating apparatus 71 is The valve closing timing for reopening the exhaust valve 17B is variably controlled. The exhaust variable valve operating device 71 has a configuration in which the valve closing timing is variable depending on the shape of the valve operating cam, a configuration in which the valve is opened and closed using an electric actuator, and the valve closing timing is variable. Any conventionally known configuration can be employed.

  FIG. 13 is a graph showing the valve closing timing of the reopening of the exhaust valve 17B and the corresponding change in the in-cylinder temperature. The vertical axis represents the lift amount and in-cylinder temperature of the exhaust valve 17B, and the horizontal axis represents the crank angle. In the same figure, B0 and B1 indicate two re-openings B0 and B1 with the valve closing timings θ0 and θ1, respectively, and the in-cylinder temperature change curves corresponding to the reopenings B0 and B1 are as follows: They are indicated by A0 and A1, respectively. The auto-ignition temperature is indicated by T.

  In the re-opening B0 having the earlier valve closing timing, the effective compression ratio becomes higher because the actual compression stroke becomes longer, thereby the temperature rises faster and the maximum in-cylinder temperature also becomes higher. On the other hand, in the re-envelopment B1, the effective compression ratio is lower than that of B0, the temperature rise is delayed, and the in-cylinder maximum temperature is lowered.

  The self-ignition timing is intersections E0 and E1 between the self-ignition temperature T and the in-cylinder temperatures A0 and A1. In the re-enlightenment B0, it is ignited before reaching the TDC and tends to knock. In the re-enlightenment B1, it can be said that the self-ignition is almost performed at TDC, and the optimum combustion is performed.

  Therefore, in this embodiment, in order to obtain an optimal effective compression ratio according to the engine operating state, the exhaust variable valve gear 71 is controlled so as to perform exhaust re-opening as shown by B1. Yes.

  Specifically, the storage device of the engine controller 66 stores a re-opening valve closing timing map as shown in FIG. 14, and this valve closing timing map indicates changes in the engine operating state, specifically, The optimum valve closing timing corresponding to changes in engine speed and load is recorded.

  This valve closing timing map is used as follows. First, values detected by the engine speed sensor and the load sensor are respectively input to the engine controller 66, and an appropriate valve closing time corresponding to each detected value is obtained from the valve closing time map. Then, the exhaust variable valve gear 71 is controlled by the engine controller 66 so as to close the exhaust valve 17B at the obtained valve closing timing.

  In order to obtain an appropriate effective compression ratio, the exhaust variable valve gear 71 can be controlled using various parameters as the engine operating state instead of or in addition to the engine speed and load described above. For example, the detection values of the in-cylinder pressure sensor, the intake air temperature sensor 86, the intake air pressure sensor 87, the coolant temperature sensor 88, and the abnormal combustion detection sensor 89 can be used.

  When using the detection value of the cooling water temperature sensor 88, the lower the temperature of the cooling water, the later the self-ignition timing will be delayed, and the ignitability will be reduced. When the temperature of the cooling water is low, the effective compression ratio is increased. When the temperature is high, an appropriate valve closing timing may be obtained so as to lower the effective compression ratio.

  When the detection value of the abnormal combustion detection sensor (knock sensor) 89 is used, when the occurrence of knocking is detected, it can be determined that the effective compression ratio is high. Therefore, a valve closing time that delays the valve closing time is obtained. You can do it.

  When the detection value of the intake air temperature sensor 86 or the intake air pressure sensor 87 is used, in consideration of the fact that the lower the intake air temperature or the intake air pressure, the slower the ignition timing and the lower the ignitability. An appropriate valve closing timing may be obtained so as to increase the effective compression ratio and decrease the effective compression ratio when the effective compression ratio is high.

  When using the detection value of the in-cylinder pressure sensor, consider that the lower the in-cylinder pressure, the later the ignition timing and the lower the ignitability, so that the effective compression ratio is increased when the in-cylinder pressure is low and the effective compression ratio is high. What is necessary is just to obtain | require suitable valve closing time so that may be reduced.

  In the above description, the exhaust valve 17B is reopened in order to obtain an appropriate effective compression ratio, but other means may be used. For example, the intake valve operating device 70 may be configured to be variable, and the optimum effective compression ratio may be acquired by early closing or late closing of the intake valve 17A at the end of the intake stroke.

[About variable swirl generator]
FIG. 15 is a schematic diagram illustrating the operation of the variable swirl device 72. The variable swirl device 72 is constituted by a throttle valve provided in the intake passage 73. When the throttle valve 72 is operated in the closing direction, the flow rate of the intake air sucked into the combustion chamber 18 is increased, and the combustion chamber 18 When the throttle valve 72 is operated in the opening direction, the swirl flow is weakened.

  FIG. 16 is a graph showing the appropriate combustion region (operable range) of the premixed gas accompanying pre-injection in correspondence with the change between the load and the equivalence ratio. In the figure, generally, the premixed gas is more likely to knock when the equivalence ratio is higher, and more likely to misfire as the equivalence ratio is lower, and proper combustion can be obtained only within a very narrow range indicated by hatching. I can't.

  Therefore, in the present embodiment, when the load is high, as shown in FIG. 15B, the throttle valve 72 is controlled so as to increase the swirl flow W, thereby promoting mixing with air and the combustion chamber 18. A homogeneous and lean premixed gas is formed inside to prevent knocking. Conversely, when the load is low, as shown in FIG. 15A, the throttle valve 72 is controlled so as to weaken the swirl flow W, thereby forming the fuel lean region X1 and the fuel rich region X2 in layers. In the fuel rich region X2, misfire is prevented by surely self-igniting.

  The present invention can be appropriately changed in design without being limited to the above embodiment. For example, in the above embodiment, the engine system including the supercharger 67, the EGR device 69, and the variable swirl device 72 is illustrated, but an engine system in which one or more of these are omitted can be used.

  The present invention can be suitably employed in a premixed compression auto-ignition diesel engine.

It is explanatory drawing which shows the outline | summary of the operation process of the diesel engine concerning embodiment of this invention. It is sectional drawing of the cylinder top part of the diesel engine concerning embodiment of this invention. It is sectional drawing of the cylinder top part. It is sectional drawing of the cylinder top part. It is sectional drawing which expands and shows a head cavity and a fuel injection nozzle. FIG. 6 is a cross-sectional view taken along the line VI-VI in FIG. 2. It is the schematic which shows an engine system. It is a time chart which shows the timing of the fuel injection from a fuel injection nozzle. It is a graph which illustrates the relationship between elapsed crank angle (time) and spray tip arrival distance. It is a flowchart for determining fuel injection conditions. It is a figure which shows a target EGR rate map. It is a flowchart for determination of the amount of reflux gas. It is a figure which shows the valve closing time of the re-opening of an exhaust valve, and the change of the cylinder temperature corresponding to it. It is a re-enlightenment valve closing time map. It is the schematic which shows the effect | action of a variable swirl apparatus. It is the figure which showed the appropriate combustion area | region (operable range) of the premixed gas accompanying pre injection corresponding to the change of load and equivalence ratio.

Explanation of symbols

DESCRIPTION OF SYMBOLS 10 Cylinder 11 Piston 12 Cylinder head 13 Fuel injection nozzle 29 Head cavity 30 Bottom surface 31 Peripheral wall surface 44 Fuel injection port 69 EGR device 71 Exhaust variable valve gear 72 Variable swirl device 80 Oxidation catalyst

Claims (7)

A concave head cavity is formed on the explosion surface of the cylinder head facing the top surface of the piston, and a fuel injection nozzle having a plurality of fuel injection ports for injecting fuel toward the peripheral wall surface of the head cavity is formed at the bottom of the head cavity. In the premixed compression self-ignition internal combustion engine provided,
The bottom surface of the head cavity is formed as a flat surface substantially parallel to the explosion surface of the cylinder head, and the peripheral wall surface is inclined so as to expand toward the piston side,
The fuel injection from the injection nozzle is performed by a combination of main injection that starts near the top dead center of the compression stroke and pre-injection that performs fuel injection in the initial or middle of the compression stroke. Ignition internal combustion engine.   The premixed compression self-ignition internal combustion engine according to claim 1, wherein the pre-injection is further performed in a plurality of times.   The premixed compression self-ignition internal combustion engine according to claim 1 or 2, wherein fuel injection from the injection nozzle is further performed in the vicinity of an overlap TDC in a transition period from an exhaust stroke to an intake stroke.   The premixed compression auto-ignition according to claim 1, wherein an oxidation catalyst for removing carbon monoxide or unburned hydrocarbon is provided in an exhaust passage communicating with an exhaust hole formed in the cylinder head. Internal combustion engine.   An EGR device that circulates the EGR gas into the cylinder, and a target EGR rate map in which a target EGR rate corresponding to the engine speed and the fuel injection amount is recorded, and according to the engine speed and the fuel injection amount 2. The premixed compression self-ignition internal combustion engine according to claim 1, wherein the EGR device is controlled so that an EGR gas amount obtained from a target EGR rate map is recirculated into the cylinder. 3.   Equipped with an intake valve and an exhaust valve that open and close the intake and exhaust holes of the cylinder head, respectively, and with a variable valve system that variably sets the opening and closing timing of the intake valve and / or the exhaust valve. The premixed compression self-ignition internal combustion engine according to claim 1, wherein the variable valve gear is controlled so as to obtain a compression ratio. The premixed compression self-ignition internal combustion engine according to claim 1, further comprising a variable swirl generator that generates swirl strength in a variable manner according to engine operating conditions.

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