engine

Abstract

To make a heat efficiency improvement request in a middle load operation of an engine and a knocking suppression request in a high load operation compatible.SOLUTION: An engine is equipped with a main combustion chamber that is formed by a cylinder block, a cylinder head, and a piston, a sub chamber 15 that has a plurality of injection holes 18 opening to the main combustion chamber, and an ignition plug 16 that ignites mixture in the sub chamber 15. In the engine, a compression ratio ε of the main combustion chamber is 14 to 24, a drawing ratio β=St / V which is a rate of the total cross-sectional area St of a plurality of injection holes 18 with respect to volume V of the sub chamber 15 is 0.0078 mm-1 to 0.0145 mm-1, and the volume V of the sub chamber 15 is 0.0430 cm3 to 0.5631 cm3.SELECTED DRAWING: Figure 2

Description

[Technical field]

[0001] The present invention is N Regarding. [Background technology]

[0002] Prechamber ignition is known as a technology to improve the thermal efficiency of reciprocating engines. This is a system in which a small prechamber (pre-chamber) is provided separate from the main combustion chamber formed by the cylinder block, cylinder head, and pistons, and the mixture is ignited in the prechamber, and the flame is made into a jet stream from the nozzle hole in the prechamber and ejected into the main combustion chamber. This system has attracted attention as it improves thermal efficiency because the mixture burns in the main combustion chamber faster than normal flame propagation. The prechamber is formed by covering the tip of the spark plug with a cap-shaped partition wall that separates it from the main combustion chamber, and multiple nozzle holes are formed in the partition wall.

[0003] There are two types of pre-chamber ignition: active and passive. The active type has an injector in the pre-chamber, and the mixture required for ignition is formed in the pre-chamber. The passive type does not have an injector in the pre-chamber, and the mixture formed in the main combustion chamber during the compression stroke is guided from the nozzle to the area around the spark plug. Patent Document 1 shows an example of active pre-chamber ignition, and Patent Document 2 shows an example of passive pre-chamber ignition. [Prior art documents] [Patent documents]

[0004] [Patent Document 1] JP 2018-66369 A [Patent Document 1] JP 2021-113549 A Summary of the Invention [Problem to be solved by the invention]

[0005] In the case of the passive type, once the pre-chamber volume and the diameter and number of the nozzle holes are determined, the jet potential of the flame ejected from the pre-chamber to the main combustion chamber, i.e., the maximum energy transfer rate RETmax from the pre-chamber to the main combustion chamber, is uniquely determined. Therefore, it is considered difficult to simultaneously achieve the requirement of improving thermal efficiency during medium load operation of the engine and the requirement of suppressing knocking during high load operation.

[0006] For example, during medium load operation, EGR (exhaust gas recirculation) is performed to reduce NOx emissions and improve fuel efficiency. Because the EGR gas flows into the pre-chamber, misfires are likely to occur even if the engine is ignited. In order to meet the requirements for preventing misfires and improving thermal efficiency during medium load operation, it is possible to determine the specifications of the pre-chamber from the perspective of making it easier for the mixture to flow into the pre-chamber and increasing the jet potential. However, in that case, during high load operation, a rich mixture enters the pre-chamber, so the jet potential becomes excessively large and knocking is likely to occur. On the other hand, if the specifications of the pre-chamber are determined from the perspective of suppressing the jet potential in order to suppress knocking at high load, the above requirements during medium load operation cannot be met.

[0007] An object of the present invention is to simultaneously satisfy the requirement for improving thermal efficiency during medium load operation of an engine and the requirement for suppressing knocking during high load operation. [Means for solving the problem]

[0008] After extensive research into pre-chamber ignition, the inventors discovered that this ignition combustion method can also satisfy both the requirements for medium load operation and the requirements for high load operation described above.

[0009] The engine disclosed herein is an engine including a main combustion chamber formed by a cylinder block, a cylinder head, and a piston, an auxiliary combustion chamber having a plurality of nozzle holes opening into the main combustion chamber, and an ignition plug for igniting the air-fuel mixture in the auxiliary combustion chamber, The compression ratio ε of the main combustion chamber is 14 or more and 24 or less, The ratio of the total cross-sectional area St of the multiple nozzle holes to the volume V of the sub-chamber is 0.0078 mm -1More than 0.0145mm -1 is as follows: The volume V of the above sub-chamber is 0.12 cm 3 More than 0.328cm 3 The present invention is characterized in that:

[0010] The compression ratio ε can be determined based on the thermal efficiency and ignition timing requirements. On the other hand, when the compression ratio ε is small, the degree of filling of the air or mixture from the main combustion chamber to the pre-chamber is low. Therefore, from the viewpoint of obtaining a mixture in the pre-chamber that can be reliably ignited by ignition during medium-load EGR operation, the compression ratio ε is set to 14 or more. However, as a large compression ratio ε is more likely to cause knocking at high loads and high revolutions, the upper limit is set to 24.

[0011] The restriction ratio β is set at 0.0078 mm in order to ensure ignition during high-load, high-speed operation and to improve thermal efficiency during medium-load EGR operation. -1 More than 0.0145mm -1 The following is the case.

[0012] First, one of the issues during high-load, high-speed operation is gas exchange in the pre-chamber. In other words, in pre-chamber ignition, the mixture passes through a narrow nozzle hole that connects the main combustion chamber and the pre-chamber, so gas exchange in the pre-chamber becomes an issue when the piston moves faster. The restriction ratio β is related to whether or not the air or mixture passes smoothly through the nozzle hole. Therefore, from the perspective of ensuring ignition during high-load, high-speed operation, the lower limit of the restriction ratio β was set to 0.0078 mm so that the air or mixture is packed reliably into the pre-chamber. -1 It states that:

[0013] The upper limit of the drawing ratio β is set at 0.0145 mm. -1 Therefore, it is possible to prevent the jet potential from becoming excessively small during the medium load EGR operation. In other words, it becomes easy to obtain a desired jet potential and increase the thermal efficiency.

[0014] Next, the volume V of the sub-chamber will be explained. This volume V affects the magnitude of the jet potential. If this volume V is 0.12 cm 3 As a result, a relatively large jet potential is obtained even under medium loads. This is advantageous for improving thermal efficiency. On the other hand, the upper limit of the volume V is set to 0.328 cm 3 This makes it possible to prevent the jet potential from becoming excessive at high load and high revolution speeds, which is advantageous in suppressing knocking.

[0015] In one embodiment, the compression ratio ε is equal to or greater than 16 and equal to or less than 18. This is more advantageous for filling the air-fuel mixture into the pre-chamber during medium load EGR operation and for suppressing knocking during high load and high revolution operation.

[0016] In one embodiment, the volume V of the antechamber is 0.2 cm 3 More than 0.328cm 3 This makes it possible to obtain the desired jet potential during medium load operation and to suppress knocking during high load and high revolution operation with a high degree of accuracy.

[0017] In one embodiment, the drawing ratio β is 0.0078 mm -1 More than 0.011mm -1 This is more advantageous in ensuring ignition during high-load, high-speed operation and in improving thermal efficiency during medium-load EGR operation.

[0018] The engine disclosed herein is an engine including a main combustion chamber formed by a cylinder block, a cylinder head, and a piston, an auxiliary combustion chamber having a plurality of nozzle holes opening into the main combustion chamber, and an ignition plug for igniting the air-fuel mixture in the auxiliary combustion chamber, The compression ratio ε of the main combustion chamber is 14 or more and 24 or less, The ratio of the total cross-sectional area St of the multiple nozzle holes to the volume V of the sub-chamber is 0.0078 mm -1 More than 0.0145mm -1 is as follows: The volume V of the sub-chamber is expressed by the following equation using the compression ratio ε and the throttle ratio β:

[0019] V = (1 / 2.6080) × ln(A) The above A is expressed as a / (0.3600×(0.9698×ln(ε)-1.5623)×(0.8494×ln(β)+5.1483)), where a is greater than or equal to 0.95 and less than or equal to 1.6.

[0020] That is, in this engine, the volume V of the sub-chamber depends on the compression ratio ε and the restriction ratio β. Specifically, the larger the compression ratio ε or the larger the restriction ratio β, the smaller the volume V. If the compression ratio ε is large, knocking is more likely to occur at high load and high revolutions, so the volume V is made smaller as the compression ratio ε increases so that the jet potential does not become excessively large. Also, since the jet potential increases as the restriction ratio β increases, in order to suppress this, the volume V is made smaller as the restriction ratio β increases.

[0021] This engine is advantageous in ensuring ignition and suppressing knocking during high-load, high-speed operation, as well as ensuring ignition and improving thermal efficiency during medium-load EGR operation. In particular, by making the size of the volume V dependent on the compression ratio ε and the throttling ratio β, a desired jet potential is obtained during medium-load operation, improving thermal efficiency, and the reliability of suppressing knocking during high-load, high-speed operation is increased.

[0022] In one embodiment of the engine, an injector for injecting fuel to form the air-fuel mixture is provided to inject the fuel into the main combustion chamber. 。 Effect of the Invention

[0023] According to the present invention, it is possible to improve the thermal efficiency during medium load EGR operation and suppress knocking during high load, high speed operation. [Brief description of the drawings]

[0024] [Figure 1] FIG. 1 is a schematic diagram of an engine. [Diagram 2]2(a) is a partially cut-away side view of the pre-chamber plug, and FIG. 2(b) is a view from below. [Diagram 3] 1 is a process diagram of an engine manufacturing method. [Figure 4] 5 is a graph showing pressure changes in the main combustion chamber and the auxiliary combustion chamber during the compression stroke and the expansion stroke of the engine. FIG. [Diagram 5] 5 is a graph showing the change in the energy transfer rate from the auxiliary combustion chamber to the main combustion chamber during the compression stroke and the expansion stroke of the engine. FIG. [Figure 6] FIG. 13 is a graph showing the correlation between the ΔPmax calculation results and the F1(ε)×F1(β) calculation results by simulation. [Figure 7] A graph showing the correlation between the ΔPig calculation results and the F2(ε)×F2(β) calculation results from a simulation. [Figure 8] A graph showing the correlation between the RETmax calculation results and the F3(ε)×F3(β)×F3(V) calculation results from a simulation. [Figure 9] FIG. 4 is a graph showing the relationship between RETmax and a knock index Ki. [Figure 10] 3 is a graph showing changes in the heat release rate of an engine. DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

[0025] Hereinafter, the embodiments of the present invention will be described with reference to the drawings. The following description of the preferred embodiments is merely illustrative in nature and is not intended to limit the present invention, its application, or its uses.

[0026] <Engine configuration> The present embodiment relates to a reciprocating engine for driving a vehicle.

[0027] As shown in Fig. 1, the engine 1 includes a cylinder block 2 and a cylinder head 3. A cylinder 4 is formed in the cylinder block 2. A connecting rod 7 connected to a crankshaft is connected to a piston 5 provided in the cylinder 4, so that the piston 5 reciprocates in the cylinder 4. The cylinder block 2, the cylinder head 3, and the piston 5 form a main combustion chamber 6 of the engine.

[0028] An intake port 8 and an exhaust port 9 are formed in the cylinder head 3. Although not shown, an intake valve and an exhaust valve are provided in the intake port 8 and the exhaust port 9 so as to open and close the openings on the main combustion chamber 6 side.

[0029] Although not shown in the figure, an intake passage is connected to the intake port 8, and an exhaust passage is connected to the exhaust port 9. An EGR passage that recirculates a portion of the exhaust gas that has passed through the three-way catalyst to the intake passage is connected to the exhaust passage. An EGR cooler and an EGR valve that controls the flow rate of the exhaust gas flowing through the EGR passage are provided in this EGR passage.

[0030] An injector 11 that injects fuel into the main combustion chamber 6, a pre-chamber plug 12 equipped with an auxiliary chamber, which will be described later, and a normal ignition plug 13 that does not have an auxiliary chamber are attached to the cylinder head 3. The injector 11 is provided on the cylinder axis, with its tip facing the center of the main combustion chamber 6. The pre-chamber plug 12 and the normal ignition plug 13 are located on either side of the injector 11.

[0031] The pre-chamber plug 12 is provided on the intake port 8 side, and extends diagonally downward from the intake port 8 side, with its tip facing the main combustion chamber 6. The normal ignition plug 13 is provided on the exhaust port 9 side, and as shown in Figure 2, the normal ignition plug 13 is provided on the exhaust port 9 side, and extends diagonally downward from the exhaust port 9 side, with its tip facing the main combustion chamber 6. It is also possible to provide the pre-chamber plug 12 on the exhaust port 9 side, and the normal ignition plug 13 on the intake port 8 side.

[0032] 2, the pre-chamber plug 12 has an auxiliary chamber 15 formed at its tip, and an ignition plug 16 is provided in this auxiliary chamber 15. Similar to the normal ignition plug 13, the ignition plug 16 has a center electrode 16a and a side electrode (earth) 16b.

[0033] The auxiliary chamber 15 is provided in the main combustion chamber 6, but the mixture in the auxiliary chamber 15 can be burned independently of the main combustion chamber 6. More specifically, the mixture in the auxiliary chamber 15 is ignited by a spark plug 16, and functions as an auxiliary combustion chamber that causes flame propagation in the auxiliary chamber 15.

[0034] 2(a) and 2(b), the auxiliary chamber 15 is defined by a semi-spherical auxiliary chamber forming portion 17 having a predetermined diameter and thickness. A plurality of injection holes 18 communicating with the main combustion chamber 6 are formed in the auxiliary chamber forming portion 17.

[0035] These nozzle holes 18 are provided to allow the mixture in the main combustion chamber 6 to flow into the auxiliary chamber 15, and are provided to ignite the mixture that has flowed in, thereby ejecting / radiating the flame generated in the auxiliary chamber 15 into the main combustion chamber 6, thereby accelerating the combustion of the mixture in the main combustion chamber 6.

[0036] Basically, when the EGR valve is closed, the mixture is a mixture of fresh air from the intake port 8 and fuel injected from the injector 11, and when the EGR valve is open, the mixture is a mixture of fresh air from the intake port 8, exhaust gas from the EGR passage, and fuel injected from the injector 11.

[0037] In this embodiment, in a plan view seen from below as shown in FIG. 2(b), four of these nozzle holes 18 are provided at 90° intervals around an axis passing through the apex A of the pre-chamber forming part 64. All of the nozzle holes 18 are formed so as to extend in a direction at 45° to a position 45° from the apex A of the hemispherical pre-chamber forming part 17, as shown in FIG. 2(a), so that a flame is ejected from the nozzle hole 18 at an angle of 45° to the axis.

[0038] The number and positions of the injection holes 18 are not limited to these values, and for example, in a plan view seen from below, five or six injection holes 18 may be provided at equal intervals around an axis passing through the apex A, and the number of injection holes may be less than four or seven or more. In addition, the injection holes 18 may be straight holes extending in the above-mentioned 45° direction, or may be swirl holes extending at a lateral incline with respect to the 45° direction.

[0039] <Engine and auxiliary compartment specifications> In this embodiment, the bore-stroke ratio (S / B), which is the ratio between the cylinder bore (inner diameter B) and the piston stroke (stroke S), is preferably 1 to 1.5. The stroke volume is preferably 500cc to 700cc. The compression ratio ε is preferably 14 to 24, and more preferably 16 to 18.

[0040] The volume V of the antechamber 15 is 0.12 cm 3 More than 0.328cm 3 It should be less than 0.2cm. 3 More than 0.328cm 3 The ratio of the total cross-sectional area St of the nozzle holes 15 (the sum of the cross-sectional areas of the nozzle holes 15) to the volume V of the sub-chamber 15 is preferably 0.0078 mm or less. -1 More than 0.0145mm -1 It should be less than 0.0078mm. -1 More than 0.011mm -1 It is preferable that:

[0041] <Engine manufacturing> A manufacturing method of the engine will now be described.

[0042] 3, in the first step S1, the compression ratio ε is set within the range of 14 to 24. From the viewpoint of the desired thermal efficiency, the ignition timing, etc., the compression ratio ε is determined to be, for example, 15 or 17.

[0043] In the second step S2, the pressure difference between the main combustion chamber and the auxiliary combustion chamber depends on the compression ratio ε and the throttle ratio β, and the upper limit βmax and lower limit βmin of the throttle ratio β are determined based on the compression ratio ε determined in the first step S1. The throttle ratio β is set within the range between the upper limit βmax and the lower limit βmin.

[0044] In the third step S3, by utilizing the fact that the jet potential depends on the compression ratio ε, the throttle ratio β, and the pre-chamber volume V, the pre-chamber volume V is set based on the compression ratio ε set in the first step S1 and the throttle ratio β set in the second step S2.

[0045] In the fourth step S4, the specifications of other engine components (engine stroke volume Vst, clearance volume, piston stroke S, bore diameter B, number and arrangement of nozzle holes connecting the main combustion chamber and the pre-chamber, etc.) are set. Since the pre-chamber volume V and the aperture ratio β are determined, when the nozzle holes are circular, the diameter of each nozzle hole is determined once the number of nozzle holes is determined.

[0046] In the fifth step S5, each part constituting the engine is designed, manufactured and assembled based on the specifications set in the second step S2 to the fourth step S4.

[0047] The specifications of the engine components that are not related to the compression ratio ε, the throttle ratio β, and the sub-chamber volume V can be set in parallel with the first step S1 to the third step S3 or in advance.

[0048] (Regarding the second step S2 (determination of βmax and βmin)) The upper limit restriction ratio βmax and the lower limit restriction ratio βmin are determined on the premise that the pressure difference ΔP between the main combustion chamber 6 and the auxiliary combustion chamber 15 depends on the compression ratio ε and the restriction ratio β. A specific description will be given below.

[0049] The upper limit restriction ratio βmax is determined based on the compression ratio ε so that the pressure difference ΔPmax between the main combustion chamber 6 and the auxiliary combustion chamber 15 is equal to or greater than a predetermined value when the pressure in the auxiliary combustion chamber 15 is at its maximum under the conditions of medium load, medium speed, EGR operation of the engine. In pre-chamber ignition, the pressures in the main combustion chamber 6 and the auxiliary combustion chamber 15 basically change as shown in Figure 4.

[0050] To explain this pressure change, during the compression stroke, the nozzle hole 18 of the auxiliary combustion chamber 15 creates air flow resistance, so the pressure in the auxiliary combustion chamber 15 rises more slowly than in the main combustion chamber 6. In other words, the pressure in the auxiliary combustion chamber 15 is lower than that in the main combustion chamber 6. Then, the air-fuel mixture in the auxiliary combustion chamber 15 is ignited by ignition before the top dead center of the compression stroke, and the pressure in the auxiliary combustion chamber 15 rises rapidly, and after the top dead center of the compression stroke, the pressure in the auxiliary combustion chamber 15 becomes higher than the pressure in the main combustion chamber 6. When the pressure in the auxiliary combustion chamber 15 reaches its maximum, the pressure difference between the main combustion chamber 6 and the auxiliary combustion chamber 15 is ΔPmax.

[0051] As described above, because the injection hole 18 creates airflow resistance, the pressure difference between the main combustion chamber 6 and the auxiliary combustion chamber 15 before ignition increases as the compression ratio ε increases (the pressure in the main combustion chamber 6 increases), so ΔPmax after ignition decreases as the compression ratio ε increases. On the other hand, looking at the throttle ratio β, in medium-load, medium-speed operation, if the throttle ratio β increases, pressure is more likely to escape through the injection hole 18 during the pressure rise in the auxiliary combustion chamber 15 after ignition, so ΔPmax decreases as the throttle ratio β increases.

[0052] That is, ΔPmax depends on the compression ratio ε and the throttle ratio β. Therefore, ΔPmax can be expressed as follows using a function F1(ε) of the compression ratio ε and a function F1(β) of the throttle ratio β.

[0053] ΔPmax=F1(ε)×F1(β) ……(1) F1(ε)=-0.0062×ε+0.1949 …… (2) F1(β) = 9.51 × 10 -5 ×β -1.754 ... (3) Here, F1(ε) and F1(β) are derived from a parametric study using a zero-dimensional simulation tool. This tool prepares a small chamber simulating a pre-chamber in the main combustion chamber of an engine, and calculates the pressure, temperature, and density of the main combustion chamber and the pre-chamber by taking into account the energy exchange between the pre-chamber and the main combustion chamber, and the heat radiation from each of the main combustion chamber and the pre-chamber.

[0054] In the calculation, the following conditions are set: bore diameter B, stroke S, compression ratio ε, pre-chamber volume V, throttling ratio β, nozzle hole diameter D, engine speed, fuel injection amount, excess air ratio, EGR rate, intake valve closing timing, exhaust valve opening timing, temperature and pressure at intake valve closing timing, wall temperature and heat release rate of the main combustion chamber and pre-chamber. The fuel injection amount, EGR rate, heat release rate, etc. are input by simulating the results of a real single-cylinder engine.

[0055] Figures 4 and 5 show an example of the results of calculations under a given medium-load EGR operating condition. As explained above, Figure 4 shows the pressure changes in the main combustion chamber and the auxiliary combustion chamber, with ΔPmax and ΔPig serving as the evaluation indexes. ΔPig is the pressure difference between the main combustion chamber and the auxiliary combustion chamber when the mixture in the auxiliary combustion chamber is ignited before the top dead center of the engine compression stroke, specifically, at 10° before the top dead center of the compression stroke in this embodiment. Figure 5 shows the change in the energy transfer rate from the auxiliary combustion chamber to the main combustion chamber, with the maximum value of this energy transfer rate, RETmax, serving as the evaluation index. Hereinafter, RETmax may be referred to as the jet potential.

[0056] The above calculations are performed for several hundred conditions, and the degree of influence of the compression ratio ε and the throttle ratio β on ΔPmax is regression-analyzed from the calculation results to derive the formulas for F1(ε) and F1(β) above.

[0057] Figure 6 shows the correlation between the calculation results of ΔPmax using a simulation tool and the calculation results of F1(ε) × F1(β) using the above formulas (1) to (3). It can be seen that the derived formulas (2) and (3) are appropriate.

[0058] It has been confirmed through experiments that in medium load EGR operation, when ΔPmax is 0.02 MPa or more, the thermal efficiency of the engine becomes higher than the standard. Therefore, the compression ratio ε set in the first step S1 is substituted into equation (2) to obtain F1(ε), and this value of F1(ε) is substituted into equation (1) and ΔPmax is set to the predetermined value of 0.02 MPa to obtain F1(β), and the upper limit restriction ratio βmax is obtained from equation (3). Incidentally, when the compression ratio ε=17, the upper limit restriction ratio βmax is 0.011.

[0059] The lower limit restriction ratio βmin is determined based on the compression ratio ε set in the first step S1 so that the pressure difference ΔPig between the main combustion chamber and the auxiliary combustion chamber is equal to or smaller than a predetermined value under high-load and high-speed engine operating conditions.

[0060] As described above, because the injection hole 18 creates air flow resistance, the pressure difference ΔPig between the main combustion chamber and the auxiliary combustion chamber increases as the compression ratio ε increases, and the pressure difference ΔPig decreases as the throttle ratio β increases. In other words, ΔPig depends on the compression ratio ε and the throttle ratio β. Therefore, ΔPig can be expressed as follows using the function F2(ε) of the compression ratio ε and the function F2(β) of the throttle ratio β:

[0061] ΔPig=F2(ε)×F2(β) ……(4) F2(ε)=0.1727×exp(0.1309×ε)……(5) F2(β)=-1.0408×10 -5 ×β -2.0994 ……(6) Like F1(ε) and F1(β), F2(ε) and F2(β) are derived from a parametric study using the above-mentioned 0-dimensional simulation tool. In other words, the formulas for F2(ε) and F2(β) are derived by performing a regression analysis on the calculation results of the degree of influence of the compression ratio ε and the throttle ratio β on ΔPig.

[0062] Figure 7 shows the correlation between the calculation results of ΔPig using a simulation tool and the calculation results of F2(ε) × F2(β) using the above equations (4) to (6). It can be seen that the derived equations (5) and (6) are appropriate. Note that in Figure 7, ΔPig is calculated by subtracting the main combustion chamber pressure from the auxiliary combustion chamber pressure, so it is a negative value.

[0063] In high-load, high-speed operation, gas transfer from the main combustion chamber to the pre-chamber is important (ideally ΔPig = 0). Experiments conducted under several conditions with different throttling ratios β showed that if ΔPig is -0.3 MPa or more (the absolute value of the pressure difference is small), the gas exchange in the pre-chamber is more reliable and comparable thermal efficiency is obtained. Therefore, the compression ratio ε set in the first step S1 is substituted into equation (5) to obtain F2(ε), and this value of F2(ε) is substituted into equation (4) and ΔPig is set to the predetermined value of -0.3 MPa to obtain F2(β), and the lower limit throttling ratio βmin is obtained from equation (6). Incidentally, the lower limit throttling ratio βmax is 0.007 when the compression ratio ε = 17.

[0064] (Regarding the third step S3 (determination of the sub-chamber volume V)) As described above, the pre-chamber volume V is set based on the compression ratio ε set in the first step S1 and the throttle ratio β set in the second step S2, taking advantage of the fact that the jet potential depends on the compression ratio ε, the throttle ratio β, and the pre-chamber volume V.

[0065] Under high-load, high-speed engine operating conditions, the larger the compression ratio ε, the higher the density of the mixture in the pre-chamber, and therefore the larger the jet potential, i.e., RETmax, which is the maximum rate of energy transfer from the pre-chamber to the main combustion chamber. Also, the larger the restriction ratio β, the better the gas exchange in the pre-chamber, and therefore the larger RETmax. Also, the larger the pre-chamber volume V, the more heat is generated in the pre-chamber, and therefore the larger RETmax. In other words, RETmax depends on the compression ratio ε, the restriction ratio β, and the pre-chamber volume V. Therefore, RETmax can be expressed as follows using a function F3(ε) of the compression ratio ε, a function F3(β) of the restriction ratio β, and a function F3(V) of the pre-chamber volume V:

[0066] RETmax=F3(ε)×F3(β)×F3(V) ……(7) F3(ε) =0.9698×ln(ε)-1.5623 ……(8) F3(β) =0.8494×ln(β)+5.1483 ……(9) F3(V) =0.3600×exp(2.6080×V) ……(10) Like F1(ε) and F1(β), F3(ε), F3(β) and F3(V) are derived from a parametric study using the above-mentioned 0-dimensional simulation tool. In other words, the formulas for F3(ε), F3(β) and F3(V) are derived by performing a regression analysis on the calculation results of the degree of influence of the compression ratio ε, the throttle ratio β and the sub-chamber volume V on RETmax.

[0067] Figure 8 shows the correlation between the calculation results of RETmax using a simulation tool and the calculation results of F3(ε) × F3(β) × F3(V) using the above formulas (7) to (10). It can be seen that the derived formulas (8) to (10) are appropriate.

[0068] Therefore, once the RETmax requirement as engine performance is determined, the pre-chamber volume V required for that can be found from the above equations (7) to (10). If RETmax is treated as a constant and RETmax=a, the pre-chamber volume V can be expressed as follows from the above equations (7) to (10).

[0069] V=(1 / 2.6080)×ln(A) ……(11) A=a / (0.3600×(0.9698×ln(ε)-1.5623)×(0.8494×ln(β)+5.1483)) ……(12) That is, by substituting the compression ratio ε set in the first step S1, the throttle ratio β set in the second step S2, and the value a as RETmax into equation (12), A is obtained, and the pre-chamber volume V is obtained from equation (11). According to equations (11) and (12), the pre-chamber volume V becomes smaller as the compression ratio ε increases, the pre-chamber volume V becomes smaller as the throttle ratio β increases, and the pre-chamber volume V becomes larger as the constant a increases.

[0070] (Regarding RETmax requirements) Suppressing knocking (strong knocking) is important when operating at high load and high revolution speeds in an engine with a high compression ratio ε, and experiments were conducted under several conditions in which RETmax was changed, and it was found that the occurrence of knocking can be suppressed by setting RETmax within a specified range. Figure 9 shows the relationship between RETmax and the knocking index (knock intensity) Ki at 6000 rpm WOT (full-throttle load) obtained in the experiment. The compression ratio ε of the engine is 17. In the experiment, RETmax was changed by changing the pre-chamber volume V and the diameter of the injection hole.

[0071] The Ki value is calculated based on vibration data of pressure waves generated inside the cylinder. The vibration data is detected using a knock sensor, an in-cylinder pressure sensor, etc. The Ki value here indicates the average value of the intensity of knocks that occurred over 300 combustion cycles. Therefore, if strong knocks occur during the sampling period, the Ki value will be high according to their intensity and frequency.

[0072] As shown in Figure 9, an inflection point was observed in the Ki value as RETmax changed from 0 to 2. Specifically, the Ki value was minimal at around RETmax = 1.2 J / deg. It is considered that as RETmax increases, abrupt heat generation occurs accompanied by strong air column oscillation, resulting in a high Ki value. On the other hand, as RETmax decreases, flame propagation in the main combustion chamber slows down, inducing autoignition of the end gas, resulting in a high Ki value.

[0073] Figure 10 shows an example of vibration data D1 and D2 at points P1 and P2 on the graph shown in Figure 9. The solid line graph is vibration data D1 at point P1 where RETmax is 0 (zero) J / deg. The dashed line graph is vibration data D2 at point P2 where RETmax is 1.2 J / deg. The vertical axis represents the heat release rate, and the horizontal axis represents the crank angle.

[0074] Vibration data D1 is the result of a test on an engine that ignites the mixture with a normal spark plug without a pre-chamber. Vibration data D2 is the result of a test on an engine that ignites the mixture with a pre-chamber that has four 1.0 mm diameter nozzle holes and a pre-chamber volume V of 0.31 cc. The ignition timing is just before top dead center (-10° ATDC).

[0075] In the vibration data D1, a large pressure fluctuation, i.e., the occurrence of a strong knock, is observed after a certain period of time has elapsed after ignition. In contrast, in the vibration data D2, such a large pressure fluctuation as in the vibration data D1 is not observed. It can be seen that the strong knock is suppressed in the vibration data D2.

[0076] Based on the data accumulated so far, including measurement error, a practically preferable range of Ki value is 1 or less, and more preferably 0.5 or less. Figure 9 shows the range where Ki value is 1 or less and the range where Ki value is 0.5 or less.

[0077] The range where Ki is 1 or less corresponds to the range where RETmax is 0.95 J / deg to 1.6 J / deg (a = 0.95 to 1.65 in Eq. (12)). The range where Ki is 0.5 or less corresponds to the range where RETmax is 1.05 to 1.5 (a = 1.05 to 1.5 in Eq. (12)).

[0078] According to the above manufacturing method, the upper limit βmax of the restriction ratio β is set so that ΔPmax is equal to or greater than a predetermined value (0.02 MPa at which the thermal efficiency of the engine is higher than the standard) under medium-load EGR operating conditions, and the lower limit βmin of the restriction ratio β is set so that ΔPig is equal to or greater than a predetermined value (−0.3 MPa at which gas exchange in the pre-chamber is facilitated) under high-load, high-speed operating conditions, and the restriction ratio β is set within the range between the upper limit βmax and the lower limit βmin, making it possible to ensure the desired thermal efficiency during medium-load EGR operation and high-load, high-speed operation.

[0079] In addition, the pre-chamber volume V is set so that RETmax under high-load, high-speed operating conditions is in a predetermined range of 0.95 J / deg or more and 1.6 J / deg or less (a = 0.95 or more and 1.65 or less in equation (12)), so that knocking during high-speed, high-load operation can be suppressed.

[0080] That is, according to the above manufacturing method, it is possible to improve the thermal efficiency during medium load EGR operation and suppress knocking during high load, high revolution operation, regardless of the size of the engine's stroke volume.

[0081] Here, regarding the drawing ratio β, when ΔPmax = 0.02 MPa and ΔPig = -0.3 MPa, the upper limit of the drawing ratio βmax when the compression ratio ε = 14 is 0.0133 mm -1 The lower limit of the drawing ratio βmin is 0.0078 mm -1 It becomes.

[0082] When ΔPmax = 0.02 MPa and ΔPig = -0.3 MPa, the upper limit of the drawing ratio βmax when the compression ratio ε = 24 is 0.0083 mm -1 The lower limit of the drawing ratio βmin is 0.0145 mm -1 It becomes.

[0083] That is, when the compression ratio ε is 14 or more and 24 or less, the drawing ratio β is 0.0078 mm -1 More than 0.0145mm -1 The result is as follows.

[0084] Here, when the compression ratio ε is 24, the lower limit of the drawing ratio βmin (0.0145 mm -1 )>Maximum drawing ratio βmax(0.0083mm -1) In this way, depending on the compression ratio, there is a possibility that the upper limit βmax and the lower limit βmin of the restriction ratio may be reversed (on the high compression ratio side). In this case, it is preferable to prioritize the lower limit βmin of the restriction ratio, that is, to prioritize suppressing excessive throttling in the high engine speed range. This is because, in a high compression ratio engine, even if the flame jet from the pre-combustion chamber is weak (even if ΔPmax<0.02), the temperature in the main combustion chamber is high, so that the weak jet can be compensated for by other means (for example, auto-ignition combustion).

[0085] In order to ensure the desired thermal efficiency during medium load EGR operation and high load high revolution operation, the restriction ratio β is set to 0.0078 mm when the compression ratio ε is 14 or more and 24 or less. -1 More than 0.0145mm -1 It should be less than 0.0078mm. -1 More than 0.011mm -1 It is preferable that:

[0086] In addition, the sub-chamber volume V is 0.0430 cm when RETmax = 0.95 J / deg, the compression ratio ε = 24, and the throttle ratio β = 0.0145. 3 When the compression ratio ε = 14 and the drawing ratio β = 0.0145, V = 0.2045 cm 3 When the compression ratio ε = 24 and the drawing ratio β = 0.0078, V = 0.2018 cm 3 When the compression ratio ε = 14 and the drawing ratio β = 0.0078, V = 0.3635 cm 3 It becomes.

[0087] When RETmax = 1.6 J / deg, the compression ratio ε = 24, the contraction ratio β = 0.0145, and V = 0.2428 cm 3 When the compression ratio ε = 14 and the drawing ratio β = 0.0145, V = 0.4045 cm 3 When the compression ratio ε = 24 and the drawing ratio β = 0.0078, V = 0.4017 cm 3 When the compression ratio ε = 14 and the contraction ratio β = 0.0078, V = 0.5631 cm 3 It becomes.

[0088] Based on the above, the volume of the auxiliary chamber V is 0.0430 cm 3 More than 0.5631cm 3 In order to suppress the occurrence of knocking during high speed and high load operation, the pre-chamber volume V is 0.12 cm 3 More than 0.328cm 3 It should be less than 0.2cm. 3 More than 0.328cm 3 It is preferable that: [Explanation of symbols]

[0089] 1 Engine 2 Cylinder block 3. Cylinder head 4-cylinder 5 Piston 6 Main combustion chamber 8 Intake port 9 Exhaust port 11 Injector 12 Pre-chamber plug 13 Normal spark plug 15 Antechamber 16 Spark plug 16a Center electrode 16b Lateral electrode 17 Subchamber formation part 18 Spray hole

Claims

1. a main combustion chamber formed by a cylinder block, a cylinder head, and a piston; an auxiliary chamber having a plurality of nozzle holes opening into the main combustion chamber; An engine having an ignition plug for igniting the mixture in the sub-chamber, The compression ratio ε of the main combustion chamber is 14 or more and 24 or less, The ratio of the total cross-sectional area St of the plurality of nozzle holes to the volume V of the auxiliary chamber is 0.0078 mm －１ More than 0.0145 mm －１ is as follows: The volume V of the sub-chamber is 0.12 cm ３ 0.328cm or more ３ An engine characterized in that:

2. In claim 1, The engine is characterized in that the compression ratio ε is 16 or more and 18 or less.

3. In claim 1 or 2, The volume V of the sub-chamber is 0.2 cm ３ 0.328cm or more ３ An engine characterized in that:

4. In any one of claims 1 to 3, The drawing ratio β is 0.0078 mm －１ More than 0.011 mm －１ An engine characterized in that:

5. a main combustion chamber formed by a cylinder block, a cylinder head, and a piston; an auxiliary chamber having a plurality of nozzle holes opening into the main combustion chamber; An engine having an ignition plug for igniting the mixture in the sub-chamber, The compression ratio ε of the main combustion chamber is 14 or more and 24 or less, The ratio of the total cross-sectional area St of the plurality of nozzle holes to the volume V of the auxiliary chamber is 0.0078 mm －１ More than 0.0145 mm －１ is as follows: an engine in which the volume V of the auxiliary chamber is expressed by the following equation using the compression ratio ε and the throttle ratio β: V=(1 / 2.6080)×ln(A) The above A is expressed as a / (0.3600×(0.9698×ln(ε)−1.5623)×(0.8494×ln(β)+5.1483)), where a is equal to or greater than 0.95 and equal to or less than 1.

6.

6. In any one of claims 1 to 5, An engine characterized in that an injector for injecting fuel to form the mixture is provided so as to inject the fuel into the main combustion chamber.

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