JPS5932646B2 - internal combustion engine - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention relates to an internal combustion engine, and particularly to an internal combustion engine that can reduce harmful components in exhaust gas as much as possible.

Harmful components emitted from internal combustion engines, namely nitrogen oxides (
NOx), carbon monoxide (CO) and hydrocarbons (HC)
Various types of devices have been developed to reduce this, but in any case, the basic objective of these devices is to prevent the generation of harmful components within the combustion chamber of the engine.

As for HC and Co, they can be oxidized relatively easily by providing a re-combustion device such as an oxidation catalyst or a thermal reactor in the engine exhaust system and further maintaining the exhaust temperature at a high temperature.

However, with regard to NOx, post-treatment in the exhaust system is difficult, and although reduction catalysts are being researched, these catalysts have problems in terms of performance, durability, cost, etc.

Therefore, how to suppress the generation of NOx in the combustion chamber of the engine is extremely important.

Exhaust gas recirculation (EGR) is a known method for reducing NOx, which recirculates a portion of the exhaust gas to the intake system to suppress the maximum combustion temperature of the fuel in the combustion chamber to some extent. This suppresses the production of NOx.

This EGR mixes inert exhaust gas that is not involved in fuel combustion into the intake air-fuel mixture to suppress the maximum combustion temperature and suppress the generation of NOx. Carbon dioxide (C
O□), water vapor (H2O) and nitrogen (N2) in the intake air
In the case of a lean mixture, NOx can be reduced by increasing the ratio of gas that does not participate in combustion, such as excess oxygen (02) (hereinafter referred to as inert gas including EGR), to the fuel.

However, if the amount of these inert gases is increased, the amount of NOx produced can be reduced almost proportionally.
On the other hand, the longer the combustion time and the higher the engine speed, the more difficult it becomes to effectively utilize the combustion energy as output, resulting in deterioration of output performance and fuel efficiency.As a result, NOx reduction is also affected. There was a limit.

In addition, as mentioned above, since the maximum combustion temperature is suppressed,
There is a problem in that HC and CO generated within the combustion chamber increase.

Therefore, in order to ensure the stability and output of the engine even if a large amount of inert gas is mixed into the intake mixture, it is necessary to shorten the combustion time of the intake mixture as much as possible, that is, to perform quick combustion. In this way, the combustion pressure can be used effectively.

This reduction in combustion time can be achieved by keeping the flame propagation distance from the ignition point as short as possible.

Furthermore, by maintaining the exhaust temperature at a high temperature, HC and CO generated in the combustion chamber can be effectively oxidized in the reburning device of the exhaust system.

In realizing this technical idea, the present invention realizes shortening of combustion time by optimally determining the mounting position of the spark plug installed in the combustion chamber, the squish area, and the installation location of the intake and exhaust valves. Achieves a significant reduction in NOx by increasing the amount of inert gas mixed into the intake air-fuel mixture.
Furthermore, when adding an inert gas, the overlap of the intake and exhaust valves is made into a dog to improve fuel efficiency, and by creating a loop-type characteristic that takes its maximum value in a specific engine operating range, Especially in city driving
The purpose of the present invention is to provide an internal combustion engine that can achieve a significant reduction in .

Embodiments of the present invention will be described below with reference to the drawings.

In FIG. 1, 1 is a carburetor as a fuel supply device; 2 is a carburetor as a fuel supply device;
3 is the intake manifold, 3 is the engine body, 4 is a thermal reactor as a re-combustion device attached to the exhaust system, 5 is the exhaust pipe, and the thermal reactor 4 is the unburned harmful component in the exhaust, that is, carbon monoxide (CO). , hydrocarbons (HC,) are oxidized under secondary air from the secondary air supply device 6.

A part of the exhaust gas is passed through the exhaust gas recirculation passage 8 and the exhaust gas recirculation control valves 30, 3.
1.40 Exhaust gas is recirculated into the intake manifold 2 by an exhaust gas recirculation device 7 that is properly controlled according to engine operating conditions as described later.

As shown in Figures 3 to 5, the installation position of the spark plug 11 installed in the combustion chamber 10 is designed to reduce the amount of inert gas, including recirculated exhaust gas that does not participate in combustion, for the purpose of significantly reducing NOx. Even if the ratio of fuel to fuel is increased (even if it is set at the optimal position to ensure stable combustion with little combustion cycle fluctuation).

That is, the optimal position is such that the volume of the combustion chamber surrounded by a spherical surface drawn from the ignition point of the spark plug 11 (the midpoint of the spark gap 14) with a radius R expressed by the following equation is the top dead center of the piston 15. It is set so that the combustion chamber volume is 35 inches or more, and the combustion chamber is located near the maximum thickness and close to the combustion chamber center (cylinder center line).

However, θ: ignition advance angle, N: engine speed (1200 to 24
0 Orpm), Vc: Burning rate.

Here, θ is the smallest advance angle determined from the relationship between ensuring exhaust gas temperature and stable engine operation.

The radius R is the combustion volume that spreads and burns in a spherical shape around the ignition point in the combustion chamber 10 from when the intake air-fuel mixture is ignited before the piston top dead center until the piston reaches the piston top dead center. is the radius of the volume W in which the intake air-fuel mixture is combusted.

Therefore, since this radius R can be expressed as the product of the combustion speed ■o and the time t from ignition to the piston top dead center, this time t can be expressed as the engine speed N and the ignition advance angle θ.
Therefore, the radius R can be found using the above equation (1).

Here, when the engine speed is in the range of 1200 to 2400 rpm, as the engine speed increases, θ also increases, so R is substantially constant regardless of the engine speed.

Generally said burning rate V.

Is it almost 15? ? Z/S, and engine speed N=20
In order to obtain the best combustion characteristics when driving in the city at 0 Or, p, m, the ignition advance angle θ is generally set to 30°, but when the radius R at this time is determined using the above formula (1), =3.75×1O-2H =37.5M.

On the other hand, in a normal internal combustion engine, 80% of the total air-fuel mixture is burned when the piston descends from the top dead center of the 15th stroke length, that is, when the crank angle is approximately 40 degrees after the top dead center. If this is done, stable engine characteristics can be obtained. In order to burn 80 parts of the total mixture when the piston is at 40 degrees, it is sufficient to burn about 20 parts of the total mixture at piston top dead center, and the combustion volume ratio at this time is about 35 parts. Become.

Therefore, by setting the volume of the combustion chamber surrounded by the radius R to 35 coefficients or more, stable engine characteristics can be obtained.

By installing the ignition point near the maximum thickness of the combustion chamber and close to the cylinder centerline, the flame propagation distance from the ignition point to each part of the combustion chamber becomes equal and shortest, which is compatible with the above combustion effect. As a result, the combustion time can be further shortened.

If this is exemplified by the hemispherical, wedge-shaped, and bathtub-shaped combustion chambers shown in Figures 3 and 5.6, the position of the spark plug will be as shown in each figure.

Here, by arranging four upper surfaces 5a of the piston 15 as shown in the figure within the range included in the radius R, the combustion volume can be further expanded.

By identifying the installation position of the spark plug as described above,
Inert gases that are not involved in fuel combustion, i.e. residual gas that remains inside the cylinder without being exhausted to the outside during the exhaust stroke, carbon dioxide (CO2) and water vapor (N20) contained in recirculated exhaust gas.
), nitrogen (N2) in the intake air, and surplus oxygen (0□) in a mixture leaner than the stoichiometric air-fuel ratio can be relatively large in proportion to the fuel, suppressing the maximum combustion temperature and significantly reducing NOx. can be reduced to

Therefore, for the fuel burned in the combustion chamber 10,
N2 and CO2 that are not involved in fuel combustion.

The weight ratio P of gases such as N20 and surplus 02 is P-13.5
~22.5.

Here, regarding the range of the weight ratio P, its lower limit is 13
．． 5, if the amount of gas that does not participate in combustion is reduced below that, the NOx reduction effect will be weakened, and the upper limit of 2.
2.5 is that if the amount of gas is increased beyond that, the stability of combustion will be impaired even if the spark plug is set at the optimal position as mentioned above, and the NOx reduction effect will not be significant. This is set in consideration of the fact that the fuel efficiency and output performance will be extremely deteriorated.If the weight ratio P is within this range, the piston 15 will move during the expansion stroke of the engine based on the above-mentioned setting of the spark plug position. The combustion rate when the crank angle is approximately 40 degrees below top dead center is over 80% of the total.
In other words, as a result of numerous experiments, it has become clear that the combustion time is short and the combustion energy can be effectively extracted as engine output.

Next, in order to specifically obtain the weight ratio P, in addition to the above-mentioned exhaust gas recirculation, the residual gas in the cylinder is increased by increasing the overlap between the intake valve 18 and the exhaust valve 19 to some extent.

That is, as shown in FIG. 10, for example, by setting the overlap between the exhaust valve 19 and the intake valve 18 to a crank angle of about 30 to 50 degrees, the proportion occupied by the residual gas becomes 15 to 25 degrees.

In particular, in this way, the overlap is larger than before by 30
~50°, increasing the proportion occupied by residual gas and correspondingly decreasing the proportion of inert gas due to exhaust gas recirculation, thereby improving fuel efficiency.

That is, according to the inventor's experiments, as shown in Fig. 18, when adding the same amount of inert gas, the deterioration of fuel efficiency due to recirculated exhaust gas is more significant than that due to residual gas. If it is possible to reduce NOx to the same degree, it is preferable to suppress the recirculated exhaust gas, which has a high rate of deterioration in fuel efficiency, and to increase the amount of residual gas, which has a low rate of deterioration, as much as possible.

However, as is clear from the engine stability characteristics of the engine of the present invention shown in FIG.
° ) will be limited.

The minimum value of 30° indicates the lowest limit at which fuel efficiency can be improved.

Further, by making the air-fuel ratio of the air-fuel mixture uniform (leaner than the stoichiometric air-fuel ratio), excess oxygen that does not participate in combustion can be increased.

Here, the theoretical formula for determining the weight ratio P will be briefly explained.

Assuming that the gas components in the air are 21% oxygen and 9% nitrogen, the air required to burn 1kq of fuel is 14.7k.
q, then the weight of nitrogen contained in the air that does not participate in combustion is the molecular weight of N2: 28.0□0 molecular weight: 3
2, the following equation is obtained.

Next, the weight of exhaust gas is almost equal to the weight of air (per unit volume), and the excess air in the mixture when the air-fuel mixture is made leaner than the stoichiometric air-fuel ratio is thin (and does not participate in combustion). If the ratio of recirculated exhaust gas, residual gas, and excess air to the air required for combustion is, for example, 10, the weight is 14.7 kg ×
10/100=1.47#y, and if it is 20th section,
Similarly, 14.7kg x 20/100 = 2.
It weighs 94 kg.

From these facts, the proportion of the weight of the gas that does not participate in combustion is the proportion of the weight of the gas that does not participate in combustion, plus the nitrogen of formula (3) above, and is specifically determined as shown in the following table.

Therefore, for example, when the residual gas is 15 parts and the recirculated exhaust gas is 30 parts, and the air-fuel ratio A/F is 14.7, the weight ratio P of gas not involved in combustion to the burned fuel is 45 parts of exhaust gas + nitrogen = Since 6.6+11.3=17.9,
It becomes 1:17.9.

In addition, the air-fuel ratio is A when the residual gas is 10% and the recirculated exhaust gas is 10%.
Similarly, when /F=16.2 (1.1 in surplus air ratio λ), the weight ratio P is 20 units of exhaust gas + 10 units of excess air +
Since nitrogen=2.9+1.5+11.3=15.7, the ratio is 1:15.7.

To obtain a predetermined weight ratio P from Konoko, at the same time as exhaust gas recirculation, the residual gas is controlled by overlapping the intake and exhaust valves or by providing an exhaust pressure regulating valve (not shown) in the exhaust passage. In combination with this, the air-fuel ratio is diluted.

In this way, even if a relatively large amount of inert gas components that are not involved in combustion are present in the combustion chamber 10, combustion can be carried out in an extremely short time by igniting the ignition plug 110 installed at an appropriate position. , it is possible to achieve a significant reduction in NOx while ensuring stable combustion with little cycle fluctuation.

According to experiments conducted by the present inventors, the bore diameter of the cylinder 16 is 95φ or less, and the cylinder volume per cylinder is 60 mm.
By setting the compression ratio ρ to 8.0 to 10.5, the engine stability will be prevented even if a large amount of inert gas is mixed into the intake mixture under the above conditions. It has been confirmed that a significant reduction in NOx can be achieved without any problems.

Furthermore, if the cylinder bore diameter and cylinder volume are made larger than the above values, even if the spark plug mounting position is set to the optimal position as described above, no improvement in the NOx reduction effect will be seen, and the engine output will be unnecessarily reduced. It only led to a decline.

Furthermore, it has been confirmed that when the compression ratio ρ is lowered below the lower limit of 8.0, the engine output deteriorates, and when it is set higher than the upper limit of 10.5, abnormal combustion such as knocking occurs, and on the other hand, thermal efficiency is improved. It has also been confirmed that fuel efficiency improves.

However, even within the above range, as shown in FIG. 22, when the compression ratio is increased, HC increases.

This is thought to be due to the fact that increasing the compression ratio increases the density of the quench layer, which is the source of mainly generating HC. Merely increasing the compression for better operation and at the same time improving fuel efficiency will not improve the performance of the engine as a whole since this will lead to an increase in HC.

In this respect, by associating the above-mentioned dog overlap with this high compression ratio, it is possible to suppress the increase in HC and improve engine performance for the first time.

Next, a specific example of an exhaust gas recirculation device that introduces recirculated exhaust gas that is part of the inert gas when the inert gas is mixed into the intake air-fuel mixture under the above conditions will be described.

In FIG. 11, the exhaust gas recirculation device 7 is a two-valve type in which a first control valve 30 and a second control valve 31 are installed in an exhaust gas recirculation passage 8 that communicates the exhaust pipe 5 and the intake manifold 2. .

The first control valve 30 is fixed to a negative chamber 30a into which the negative pressure of the carburetor venturi section 1a amplified by the negative pressure amplifier 32 is introduced, and a diaphragm 30e that separates the negative pressure chamber 30a from the atmospheric chamber 30b. and a valve body 30d arranged in the exhaust gas recirculation passage 8, and as shown in FIG. The exhaust gas recirculation passage 8 begins to open gradually from then on, and is configured to fully open when the negative pressure reaches a predetermined negative pressure value 22 or more.

On the other hand, the second control valve 31 separates a negative pressure chamber 31a that communicates with the exhaust gas recirculation passage 8 near the communication part with the intake manifold 2 and into which suction negative pressure is introduced, and an atmospheric chamber 31b from the negative pressure chamber 31a. The first control valve 3 is fixed to a diaphragm 31e that
As shown in FIG. 15, the valve body 31d changes from the fully open state to a predetermined negative pressure value as the suction negative pressure increases. It is configured to gradually start narrowing the passage 8 from P3 and close the passage 8 at a predetermined negative pressure value P4.

By configuring the first and second control valves 30 and 31 to control the EGR rate (the ratio of the amount of exhaust gas recirculation to the amount of intake air), the EGR rate pattern can be changed to a loop type as shown in FIG. By setting the negative pressure IJ (exhaust gas recirculation amount) characteristics of the first and second control valves 30 and 31 arbitrarily, the maximum value of the EGR rate can be adjusted to the engine normal operating range, that is, the engine speed. is 1400~2000r
, p, m in the city driving driving range and at a position where acceleration frequency is high, that is, in the area where acceleration is most frequently performed in the regular driving range, the suction negative pressure is -150 tran HV ~ -2
It can be set in the region of 0.005 nHf.

That is, since the first control valve 30 is operated by the venturi negative pressure amplified by the negative pressure amplifier 32, here, for example, the amplification factor is set to 10, and the operating pressure point P1. P2 at -50 mHf and -150, respectively.
mmHs', the EGR amount starts to gradually increase from an equal venturi negative pressure of 15 mmHf, reaches the maximum amount at -15 mmHf, and thereafter maintains this maximum amount.

However, the amount of intake air that determines the EGR rate increases as the engine speed increases or the intake negative pressure decreases.
Hf? As the degree of negative pressure increases, the EGR rate decreases in a region where the EGR amount is constant, and as a result, the EGR rate characteristic of the first control valve 30 becomes a claw-shaped EGR rate as shown in FIG. It becomes a pattern.

Therefore, if an attempt is made to perform EGR control using only this first control valve 30, the EGR rate will increase in the low speed and low load region of the flat road running load curve R/L, which will impair engine stability. Although there is a need to increase the EGR rate in areas with high acceleration frequency where the intake air volume is large and the combustion chamber temperature is high and a large amount of NOx is likely to be generated, this
It becomes impossible to set the maximum value of the GR rate high.

In addition, there is a region where the EGR rate is low even in the slight downhill traveling region or in the engine deceleration region, which is the lower side of the flat road traveling load curve R/L, so the combustion chamber temperature decreases when the engine decelerates. TeHC.

The generation of CO becomes significant, and the reaction temperature of the thermal reactor installed in the exhaust system cannot be maintained (as a result, large amounts of HC and CO are released into the atmosphere without being able to be processed).

Therefore, by disposing a second control valve 31 that throttles the passage 8 in response to the suction negative pressure in the exhaust gas recirculation passage 8 downstream of the first control valve 30, this problem can be solved.

That is, the operating pressure point P3 of this second control valve 31.

P4 for example -200rIvrLH1, -400wIL
When the temperature is high, the second control valve 31 alone reduces the EGR amount as the suction negative pressure increases, so the EGR rate characteristic of the first control valve 30 shown in FIG. 16a is affected by the throttling action. As shown in b, it shows the opposite characteristics, and the suction negative pressure is -2
The EGR rate is reduced in the range of 00WrrrLHf to -400rrvnHf.

Therefore, by adding this second control valve 31, E
The GR rate characteristic can be expressed as C, which is a combination of the characteristics a and b (
As a result, as shown in Fig. 12, the center of the maximum value of the EGR rate is at the suction negative pressure -1501rrrILH1 to -200
1rrrrLH, engine speed 1400~200Or,
This results in a loop-type EGR rate characteristic defined in the range of p and m.

The cross-sectional characteristic obtained by cutting this loop type characteristic along the straight line Y-Y, which obliquely intersects the equal venturi negative pressure curve X-X passing through the central maximum value and the equal suction negative pressure straight line and passing through the central maximum value, is the 12th cross-sectional characteristic.
It will look like the one shown in the figure.

According to such an exhaust gas recirculation device, the flat road running load curve R/
Above L, in the engine normal operating range and in the high acceleration range, that is, the engine speed is 1400 to 2000 r, p, m, and the suction negative pressure is -150rrrmHff to -200mmH.
It has a loop-type EGR rate characteristic in which the center of the maximum value of the EGR rate is in the region of f, and the EGR rate gradually decreases from the center. and a large amount of HC.

High control accuracy without generating CO and E
Since the maximum value of the GR rate can be increased, therefore, when EGR is mixed into the intake mixture as a part of the inert gas in the intake mixture, the gas weight ratio can be maintained under the above conditions to be effective. This makes it possible to achieve significant NOx reduction.

Here, the maximum value of the exhaust recirculation rate characteristic is set to 1400 to 2000 rpm at engine speed and -150 to 2000 rpm at suction negative pressure, as described above.
The reason for setting -200 mm Hf will be briefly explained.

First, we examined NOx emissions based on 10 modes that represent typical urban driving conditions for cars, and found that NOx emissions during acceleration were 70% of the total NOx emissions.
%, and furthermore, the amount of this 70th factor is the emission ratio by operating condition (represented by a hook, it becomes a distribution as shown in Fig. 20).

From now on, when the engine speed is 1400 to 200 rpm and the shaft torque is 5 kg, ? ? N in the operating region near Z
It has been found that the amount of Ox emissions is large, and therefore it is preferable to maximize the exhaust gas recirculation rate in this operating region.

In this case, in the low speed, low load region, high speed region, and deceleration region, in consideration of engine stability and a decrease in NOx emissions, exhaust recirculation must be omitted or minimized, and the exhaust recirculation rate characteristics must be set to a loop type. It is preferable that the characteristics be as follows.

Next, when performing exhaust gas recirculation control based on the above characteristics, it is preferable to express the characteristics in terms of the relationship between suction negative pressure and engine rotational speed, and therefore it is necessary to convert shaft torque to suction negative pressure.

However, it is difficult to unambiguously determine the relationship between the suction negative pressure and the equiaxed torque, which changes depending on the amount of exhaust gas recirculation.

Therefore, convert as follows.

In other words, a graph (
Determine the exhaust gas recirculation rate necessary to achieve the NOx reduction target value from Figure 21A), and then use the graph (Figure 21B) showing the relationship between suction negative pressure and exhaust gas recirculation rate to determine the exhaust gas recirculation rate. This is to find the suction negative pressure.

By performing such conversion individually for each of a plurality of representative conditions, the frequency distribution corresponding to FIG. 20 can be expressed on a characteristic diagram of suction negative pressure and engine speed (for example, FIG. 12). , As a result, the above-mentioned region (engine speed 14
00-2000rpm.

NOx at negative suction pressure -150 to -200 Hf)
The emission rate will be extremely large.

By setting the maximum value of the exhaust recirculation rate in this range, NOx during city driving, represented by the 10 modes, can be reduced.
can be effectively reduced.

The embodiment shown in FIG. 17 is designed to obtain the above-mentioned loop-type EGR rate characteristic using a single control valve.

This control valve 40 is constructed by introducing the negative pressure of the carburetor venturi portion 1a amplified by the negative pressure amplifier 32 into the negative pressure chamber 40a, in the same way as the first control valve 30 in the previous embodiment. The EGR amount is controlled by adjusting the opening degree of the exhaust gas recirculation passage 8 by changing the lift amount of It is attached.

This negative pressure amplifier 32 has a negative pressure chamber 32a into which venturi negative pressure is introduced, an atmospheric chamber 32c separated from the negative pressure chamber 32a by a relatively large diaphragm 32b, and a comparison with the atmospheric chamber 32C. The relay chamber 32e is separated by a diaphragm 32d with a small area and communicates with the negative pressure tank 32i and the negative pressure chamber 40a of the control valve 40.

The diaphragm 32d has a through hole 3 in the peripheral wall that communicates with the atmospheric chamber.
2g, and is connected to the diaphragm 32b by a cylindrical connecting member 32f equipped with a check valve 32h inside, and when the diaphragm 32b operates, the check valve 32h connects the end of the negative pressure outlet nozzle 32j communicating with the negative pressure tank 32i. It is designed to open and close.

Therefore, when the engine speed is low and the venturi negative pressure is small, the diaphragm 32b is located below the position shown in the figure, and the diaphragm 32d is displaced downward via the connecting member 32f, thereby causing the check valve 32h to close to the nozzle 32j.
The open end is closed to cut off the negative pressure from the negative pressure tank 32i.

As a result, the output negative pressure to the control valve 40 becomes &bJ'i.

At this time, depending on the position of the diaphragm 32d, the atmospheric chamber 32c and the relay chamber 32e are communicated through the through hole 32g of the connecting member 32f, thereby reducing the negative pressure acting on the negative pressure chamber 40a of the control valve 40. do.

Further, when the engine speed increases and the venturi negative pressure increases, the diaphragm 32b is displaced upward, and the connecting member 32
Move the check valve 32h upward via f.

As a result, the end of the nozzle 32j is opened, and the negative pressure tank 32
Negative pressure from i is introduced into the negative pressure chamber 40a of the control valve 40 via the relay chamber 32e.

That is, this negative pressure amplifier 32 amplifies the venturi negative pressure according to the area ratio of the diaphragm 32b and the diaphragm 32d, and changes the operating characteristics of the control valve 40 to the aforementioned first control valve 32.
Do the same thing.

In this embodiment, the negative pressure amplifier 32 is provided with an auxiliary mechanism 41.

This auxiliary mechanism 41 includes a negative pressure chamber 41 into which negative pressure generated in a venturi portion 43 formed in a conduit 42 in which a driving member synchronized with engine rotation, such as an oil pump 44, is inserted is introduced.
a and a rotational fulcrum in the middle part, one end abuts approximately the center of the diaphragm 32b, and the other end contacts the negative royal 41
a and a lever 41e connected to a stem 41d fixed to a diaphragm 41b separating the atmospheric chamber 41c.

In other words, when the negative pressure value of the venturi section 43 in the conduit 42 becomes negative as the engine speed increases, this auxiliary mechanism 41 operates as follows:
The diaphragm 41b is displaced upward and acts to push down the diaphragm 32b via the lever 41e.

Conversely, when the negative pressure value of the venturi section 43 becomes small as the engine speed decreases, the diaphragm 41b is displaced downward, and the force acting on the diaphragm 32b is released via the lever 41e.

That is, this auxiliary mechanism 41 functions to reduce the operating negative pressure value of the control valve 40 as the engine speed increases, and has the same operating characteristics as the second control valve 31 in the embodiment described above. This makes the EGR rate characteristics loop-type.

Therefore, the negative pressure-lift amount characteristics of the control valve 40 and the auxiliary mechanism 4
By arbitrarily setting the negative pressure-lift amount characteristic of 1.
Loop type in which the center of the maximum value of the EGR rate is above the flat road running load curve due to engine characteristics, is in the engine normal operating range and is at a position where acceleration frequency is high, and the EGR rate gradually decreases from this center. It can be configured to have EGR rate characteristics.

By specifying the recirculation rate characteristics of the exhaust recirculation device in this way, we can suppress fluctuations in the weight ratio of inert gas mixed into the intake air-fuel mixture, maintain engine stability with fast combustion, and significantly reduce NOx. It is possible to achieve a reduction.

Furthermore, in order to achieve this rapid combustion, it is more effective to make the ignition plug 11 protrude from the mounting wall surface as much as possible within a range that can withstand the heat load.

In other words, flame propagation spreads spherically around the ignition point, so combustion occurs fastest if the ignition point of the spark plug is located at the center of the combustion chamber space.

However, in reality, this is impossible due to the durability of the ignition plug, but as a result of repeated experiments based on this idea, the ignition plug 11 is connected from the mounting wall surface to the ignition point as shown in Figures 7 and 8. By protruding the distance within a range of 2 to 7wn, combustion efficiency can be improved without impairing durability.

As a means for causing the ignition plug 11 to protrude in this way, the seventh
In addition to extending the center electrode 12 and side electrodes 13 as shown in the figure,
As shown in FIG. 8, the ignition plug mounting portion may be made to protrude beyond the general wall surface of the combustion chamber 10.

In this case, the spark gap 14 of the spark plug 11 is set to 1.
1~2.0mm, ignition energy 100? ? It was confirmed that it is more effective when the amount is about Zj (millijoules).

Furthermore, as shown in FIG. 9, the center electrode 12 of the ignition plug 11 is shaped like a truncated cone, and several grooves are provided on the side electrodes 13 facing the center electrode 12 to reduce the discharge area compared to the current one. It was also confirmed that by setting the diameter of the center electrode to a certain degree (currently used center electrode diameter is 2 to 2.5φ), the electric field can be increased and the ignitability can be further improved without the electrode being melted and damaged.

Next, in order to further promote the shortening of the combustion time described above, a squish region S is formed.

The squish area S is formed by the difference in area between the lower surface of the combustion chamber and the cylinder bore, but if the squish area A is the cylinder bore area ~, then A/AO = 0.1 to 0.45, squish thickness 5t-1,05 to 2. .51rmL.

When the squish region S is provided, intense turbulence is generated in the air-fuel mixture during the compression stroke due to the upward movement of the piston 15, stirring is promoted, the fuel distribution becomes more uniform, and this causes turbulence in flame propagation during combustion. Burning time becomes even shorter.

If the squish area and squish thickness are smaller than this, effective squish cannot be expected, and if it exceeds this range, an increase in HC due to flame cooling will be observed, so the above values were set as the optimal ranges.

In addition, if a swirl is generated along with this squish effect, that is, by offsetting the arrangement positions of the intake valve 18 and exhaust valve 19 in each cylinder with respect to the cylinder row line direction, as shown in FIG. The air-fuel mixture drawn into each combustion chamber from the intake manifold 2 is drawn in in the tangential direction of the combustion chamber and forms a swirl within the combustion chamber, promoting mixing of fuel and air and making combustion more stable. It is possible.

In this way, it was possible to significantly reduce NOx without impairing engine operability, but in order to further reduce HC and CO, which are other harmful exhaust components, as shown in Figure 1. , connect the engine exhaust port 20 to the cylinder head 1
A so-called Siamese exhaust port is formed by gathering the cylinders adjacent to each other in the thermal reactor 7 to prevent a drop in exhaust temperature and promote the oxidation reaction of HC and CO in the thermal reactor 4.

That is, in this internal combustion engine, since a relatively large amount of EGR and residual gas is produced, it is undeniable that combustion is suppressed to some extent, and it is difficult to suppress the generation of HC and CO accordingly.

Therefore, as shown in the figure, a reburning device such as a thermal reactor 4, an exhaust manifold, or an oxidation catalyst (not shown) is installed in the exhaust system to oxidize HC and CO.
In either case, it is desirable to maintain the exhaust gas temperature at a high temperature in order to increase the oxidation efficiency.

Generally, in order to increase the exhaust temperature, the ignition timing is delayed considerably from the optimum ignition timing (M, B, T) to prolong the combustion time and increase the exhaust temperature. This contradicts the basic idea, and the output performance and fuel efficiency deteriorate.

Therefore, by making the exhaust port 20 a Siamese port structure and reducing the passage surface area, escape of exhaust heat to the cylinder head 17 is suppressed, and the exhaust temperature can be maintained at a high temperature state.

At the same time, as shown in Figure 3.5.6, a heat-resistant port liner 21 is installed in the exhaust port 20 at a constant distance from the port wall 20a to further prevent the exhaust temperature from decreasing. can be thoroughly implemented.

According to experiments conducted by the inventors, in addition to making the exhaust port 20 a Siamese port, the exhaust valve 19
d, = (0,
40 to 0.50) By setting the DK stage, the exhaust temperature can be maintained at a high temperature even more effectively, and the valve diameter d1 of the intake valve 18 is also set to d□=(0.45 to 0.55)D.
It has been confirmed that by setting this to , pumping loss can be reduced and fuel efficiency can be further improved.

The secondary air supply device 6 is required for the oxidation of HC and CO in the thermal reactor 4 when the air-fuel ratio of the air-fuel mixture supplied to the engine is at the stoichiometric air-fuel ratio or richer than the stoichiometric air-fuel ratio. In the case of air, there is no need for secondary air, and excess oxygen in the exhaust gas assists in the oxidation of HC and Co.

The secondary air supply device 6 may include an air pump 9 as shown in the figure, or a means for sucking secondary air during exhaust through a check valve using intermittent negative pressure generated based on exhaust pulsation. You can also use

In addition to further improving NOx countermeasures, HC.

In order to reduce CO, a three-way catalyst may be installed in the exhaust system.

A three-way catalyst is a catalyst that has the functions of reducing NOx and oxidizing HC and CO. However, in this case, excellent conversion is achieved only when the overall air-fuel ratio of the exhaust gas flowing into the catalyst is at or near the stoichiometric air-fuel ratio. In order to achieve efficiency, air-fuel ratio control must be strictly managed.

Therefore, it is preferable to use an electronically controlled fuel injection device or a combination of an electronically controlled carburetor and a feedback control circuit as the fuel supply device.

A feedback control circuit is one in which, for example, an oxygen sensor is installed in the exhaust system, the state of the air-fuel ratio is determined based on the sensor output, and the fuel supply amount or air supply amount is thereby feedback-controlled. Many people including the applicant (proposed).

As described above, according to the present invention, by specifying the mounting position of the spark plug in the combustion chamber and configuring it so that swirl and squish occur optimally, the combustibility of the air-fuel mixture in the combustion chamber is improved and the air-fuel mixture is On the other hand, it is possible to increase the amount of inert gas contained in the exhaust gas and reduce NOx.-1 The amount of inert gas mixed in is maximized by overlapping the intake and exhaust valves, and the rest is carried out by recirculating exhaust. By doing so, it is possible to improve fuel efficiency, and by specifying the maximum value of the exhaust recirculation rate in the engine operating region with high frequency of operation, it is possible to extremely reduce NOx, especially in the urban driving region as represented by the 10 mode. Therefore, it can be effectively reduced.

In addition, in the present invention, in particular, by organically combining a high compression ratio and a high overlap, combustion stability obtained by a high compression ratio (therefore, NOx
In addition to the original effects such as improving fuel efficiency and solving problems caused by high compression ratios, high overlap increases the internal residual exhaust gas, combusts the increased HC next time, and reduces the increase in HC. This also has the effect of suppressing emissions as shown in Figure 23, and the overall performance of the engine can be improved.

[Brief explanation of the drawing]

FIG. 1 is a schematic plan view of the internal combustion engine of the present invention, FIG. 2 is an explanatory diagram showing the relationship between combustion volume and combustion rate in a normal internal combustion engine, FIG. 3 is a cross-sectional explanatory diagram of the combustion chamber, and FIG. The figure is a plan view of the same, Figures 5 and 6 are cross-sectional explanatory diagrams showing six different examples of combustion chamber shapes, Figures 7 and 8 are cross-sectional explanatory diagrams showing examples of six different installations of the spark plug, and Figure 9 is a cross-sectional explanatory diagram showing examples of six different installations of the spark plug. FIG. 10 is an explanatory diagram showing the overlap of intake and exhaust valves. FIG. 11 is an explanatory diagram of an embodiment of the exhaust gas recirculation device. FIG. 12 is a recirculation rate characteristic diagram. is a reflux rate characteristic diagram of the first control valve in the device shown in FIG. 11, and FIG. 14 is a negative pressure characteristic diagram of the first control valve.
A lift amount diagram, FIG. 15 is a negative pressure-lift amount diagram of the second control valve, and FIG. 16 is an explanation showing the reflux rate characteristics on the equal venturi negative pressure curve with the first and second control valves. Figure, 1st
FIG. 7 is an explanatory diagram of different embodiments of the exhaust gas recirculation device, and FIG. 18 is a graph showing the relationship between the proportion of inert gas and the fuel efficiency deterioration rate.
Fig. 19 is a graph showing the upper limit of overlap, Fig. 20 is a NOx emission sharing ratio graph during 10 mode acceleration,
Figures 21A and B are graphs showing the relationship between NOx concentration, suction negative pressure and exhaust recirculation rate, Figure 22 is a graph showing compression ratio and HC amount, and Figure 23 is a graph showing the relationship between NOx emissions and various characteristics. This is a graph showing. 1... Carburetor, 2... Intake manifold, 3... Engine body, 4... Thermal reactor, 5... Exhaust pipe,
6... Secondary air supply device, 7... Exhaust recirculation device, 1
0... Combustion chamber, 11... Spark plug, 15... Piston, 16... Cylinder, 17... Cylinder head,
18... Intake valve, 19... Exhaust valve, 20... Exhaust port, 21... Port liner, 30, 31, 40.
...Exhaust recirculation control valve, 32...Negative pressure amplifier, ...
Cylinder bore area, A... squish area, S...
Squish area, St... Squish thickness.

Claims (1)

[Scope of Claims] 1. The volume surrounded by a spherical surface drawn from the ignition point of a single ignition plug with the following radius R is 35 times or more of the combustion chamber volume at the top dead center of the piston, and, It is installed near the maximum thickness of the combustion chamber and close to the center of the combustion chamber, while the intake and exhaust valves are offset in the cylinder row direction, and a squish area with a relatively large area ratio to the cylinder pore area is provided. A combustion chamber is formed, and the weight ratio of gases such as nitrogen, carbon dioxide, water vapor, and excess oxygen that do not participate in the combustion of fuel to the fuel burned in the combustion chamber is 13.5 to 22.5. At the same time, the compression ratio is set to 8.0 to 10.5, the overlap of the intake and exhaust valves is set to 30 to 50 degrees, and the exhaust recirculation device is set to an engine speed of 1400 to 2000 rpm, 1 suction negative pressure - 150 degrees. An internal combustion engine characterized by having loop-type characteristics such that the exhaust gas recirculation rate is maximized in the region of ~-200rrrmHg. However, θ: ignition advance angle, ゛N: engine speed (1200 to 24
00 r, p, m), Vc: burning rate.