JPH06229247A - Engine with supercharger - Google Patents

Abstract

PURPOSE:To ensure antiknock property and improve engine torque for optimum conditions, in an engine with a supercharger equipped with two intake valves, from a relationship between the pressure ratio of the supercharger, the cylinder bore, geometrical compressive ratio and the like. CONSTITUTION:In an engine provided with two intake valves and a supercharger 17 at an intake passage 10, cylinder bore B is set to 50mm<<=B<=70mm, and a relationship between supercharging pressure ratio (ratio to the atmospheric pressure in supercharger discharging side pressure), geometrical compressive ratio gamma of the engine, and cylinder bore B(mm) is set as follows: gamma>-0.29.epsilon+6.0-0.022.B.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a supercharged engine in which one cylinder is provided with two or more intake valves and an intake passage is provided with a supercharger.

[0002]

2. Description of the Related Art Conventionally, there has been widely known an engine in which a supercharger is provided in an intake passage to improve charging efficiency and torque. In this way, when the charging efficiency is increased by supercharging, knocking easily occurs in the low rotation and high load region, and the exhaust temperature easily rises in the high rotation and high load region. Therefore, improvement of the filling efficiency is likely to be hindered.

In order to prevent knocking in an engine with a supercharger, the intake valve closing timing is delayed or intake / exhaust is performed as disclosed in Japanese Patent Laid-Open No. 2-119620. A technique for increasing the amount of overlap in the valve opening period is known. That is,
When the intake valve closing timing is delayed more than the bottom dead center, the effective compression ratio is reduced and the temperature rise due to the compression work is suppressed, so that the knock resistance is enhanced. Further, when the overlap amount of the intake / exhaust valve opening period is set to be large, the scavenging property is enhanced and the knock resistance is enhanced by reducing the residual exhaust gas.

Further, as a means for suppressing the exhaust gas temperature rise in an engine with a supercharger, for example, JP-A-3-23 is available.
As disclosed in Japanese Unexamined Patent Publication No. 327, there is one that makes the air-fuel ratio lean in the supercharging region.

[0005]

In order to suppress knocking, in addition to the above, it is conceivable to reduce the cylinder bore diameter to shorten the flame propagation distance, and to retard the ignition timing. There is also a method. Also,
In order to improve the output, it is desirable to provide two or more intake valves in one cylinder, because this will increase the intake opening area.

In order to suppress knocking and improve output, various factors such as the cylinder bore diameter, ignition timing, engine compression ratio, supercharger pressure ratio, etc., are involved. However, in the conventional overpayment engine, the relationship between these various factors is not fully taken into consideration, and there is room for improvement in terms of knock resistance and output improvement.

In view of the above circumstances, the present invention aims to increase the torque of the engine while ensuring the knock resistance in relation to the pressure ratio of the supercharger, the cylinder bore diameter of the engine, the geometric compression ratio and the like. It is an object of the present invention to provide a supercharged engine that can provide optimum conditions.

[0008]

In order to achieve the above object, the present invention provides a supercharger in the intake passage of a reciprocating piston type engine having two or more intake valves in one cylinder. Cylinder bore diameter is 70 mm in engines with a feeder
The pressure ratio (γ = P) of the supercharger discharge side pressure (P) to the atmospheric pressure (P ₀ ) at the time of high load is set as follows.
/ P ₀ ), the geometric compression ratio (ε) of the engine, and the cylinder bore diameter (B) are set so as to satisfy the following equation.

Γ ≧ −0.29 · ε + 6.0−0.022 · B In the present invention, the cylinder bore diameter is preferably 50 mm or more.

Further, the closing timing of the intake valve is set to 60 ° or more after the bottom dead center in the crank angle, or the overlap amount of the opening period of the intake valve and the exhaust valve is set to 20 ° or more in the crank angle. , Pressure ratio (γ = P / P ₀ ) of supercharger discharge side pressure (P) to atmospheric pressure (P ₀ ) at high load
And the geometrical compression ratio (ε) of the engine and the bore diameter (B) of the cylinder are preferably set to satisfy the following equation.

Γ ≧ −0.29 · ε + 6.2-0.022 · B

[0012]

According to the above construction, the small bore having the small cylinder bore diameter improves the knock resistance, and the pressure ratio γ, the geometric compression ratio ε, and the cylinder bore diameter B satisfy the above relation. Within the range, the torque is almost maximized while the ignition timing is adjusted so as to secure the knock resistance.

In particular, when the cylinder bore diameter is in the range of 50≤B≤70, the demand for the smaller cylinder bore diameter and the requirement for engine performance are met.

Further, the intake valve is closed late, or
The knock resistance is further enhanced by increasing the overlap amount of the exhaust valve.

[0015]

Embodiments of the present invention will be described with reference to the drawings.
FIG. 1 shows the overall structure of an engine with a supercharger according to an embodiment of the present invention. The engine illustrated in this figure is a V-type engine, and the engine main body 1 includes a pair of V-shaped banks 1A and 1B, and a plurality of cylinders 2 are arranged in each of the banks 1A and 1B. Three cylinders 2 each are arranged.

In each of the cylinders 2, first and second intake ports 3a and 3b and two first and second exhaust ports 4a and 4b are formed so as to open into the combustion chamber. , An intake valve (not shown) in each of the intake ports 3a, 3b
And an exhaust valve (not shown) is provided in each exhaust port 4a, 4b. The valve mechanism that drives the intake valve is provided with a valve timing variable mechanism 5 that can change the opening / closing timing of the intake valve by changing the phase of the camshaft 6 with respect to the cam pulley, for example. Further, each cylinder 2 is provided with an ignition plug 7 as shown in FIG. 2, and this ignition plug 7 is connected to an ignition circuit 8 including an ignition coil, a distributor and the like.

Further, as shown in FIG. 1, an independent intake passage 11 on the downstream side of the intake passage 10 is connected to each of the intake ports 3a and 3b, and a fuel is provided in the vicinity of the intake port of each independent intake passage 11. An injector 9 for injecting and supplying

The intake passage 10 is composed of an intake manifold 12 having the independent intake passages 11 and a common intake passage 13 on the upstream side. An air cleaner 14, an air flow meter 15, a throttle valve 16 and a supercharger 17 are provided in the common intake passage 13. The supercharger 17 shown in the figure is a mechanical supercharger driven by an engine output shaft via a transmission means such as a belt, and particularly preferably, a Risholum supercharger or the like so as to obtain a large pressure ratio. The internal compression type supercharger is used. And to reduce drive loss at low load,
A supercharger 17 is arranged downstream of the throttle valve 16. Further, an intercooler 18 is provided downstream of the supercharger 17.
And a bypass passage 19 that bypasses the supercharger 17, and a bypass control valve 20 that opens and closes the passage 19.

Further, the exhaust gas recirculation to the intake system (EG
The EGR passage is formed so as to perform R), and in the illustrated embodiment, a low load EGR passage 21 and a high load EGR passage 22 are provided. The low load EGR passage 21
Has one end connected to the exhaust manifold 23 and the other end branched to be connected to each of the independent intake passages 11. A low load EGR valve 24 is provided in the middle of the low load EGR passage 21. One end of the high load EGR passage 22 is connected to the exhaust passage 26 downstream of the catalytic converter 25 for purifying exhaust gas, and the other end is connected to the common intake passage 13 upstream of the supercharger 17. There is. The high load EGR passage 22 is provided with a carbon trap 27, an EGR cooler 28, and a high load EGR valve 29. When EGR is performed through the high load EGR passage 22, EGR gas is guided from a relatively downstream side of the exhaust passage to a combustion chamber through a relatively long path, and an EGR cooler 28 is provided. As a result, the EGR gas is sufficiently cooled.

Reference numeral 30 denotes a control unit (ECU) for controlling the engine, which is the air flow meter 1 described above.
5 and signals from the engine speed sensor 31, the throttle opening sensor 32, etc., to control the fuel injection amount from the injector 9 according to the intake air amount and the like, and to control the ignition timing according to the operating state. Further, the valve timing varying mechanism 5, the bypass control valve 20, and the EGR valves 24 and 29 are controlled.

The variable valve timing mechanism 5 is used.
As shown in FIG. 3, the intake valve opening / closing timing is relatively advanced to the first timing IV1 and the second timing later than the first timing IV1.
It is possible to change the valve opening overlap amount O / L with the exhaust valve (EV) and the intake valve closing timing IC by changing to the timing IV2.
Then, the control unit 30 controls the intake valve opening / closing timing according to the operating state, for example, the first timing IV1 on the low load side and the second timing IV2 on the high load side.

The bypass control valve 20 is controlled so as to be opened on the low load side and closed on the high load side.
In addition, as for EGR control, both EGs are controlled in the idle region.
The R valves 24 and 29 are closed, the low load EGR valve 24 is opened in the low load region excluding the idle region, and the high load EGR valve 29 is opened in the high load region.

In such an engine, the cylinder bore diameter B is 70 mm or less, preferably 50 mm.
≦ B ≦ 70 mm.

Further, the pressure ratio γ (= P / P ₀ ) of the supercharger discharge side pressure P to the atmospheric pressure P ₀ at high load, the geometric compression ratio ε of the engine, and the cylinder bore diameter B (mm). The relation with and is set so as to satisfy the following equation (1).

[0025]

[Formula 1] γ ≧ −0.29 · ε + 6.0−0.022 · B (1) Then, under the setting that satisfies the above formula (1), the ignition timing depends on the above pressure ratio γ. Depending on the filling efficiency and the geometric compression ratio ε, the retard amount is controlled so that knocking is prevented.

According to the apparatus of the present embodiment as described above, the cylinder bore diameter is reduced so as to be advantageous in suppressing knocking and knocking occurs within a range satisfying the requirements of ensuring reliability and preventing a sudden increase in intake resistance. The ignition timing is adjusted so as not to occur, the charging efficiency is increased, and the engine torque is increased to the maximum. Such an operation will be described below based on the data shown in FIGS.

Grounds and Action of Setting Cylinder Bore Diameter In order to improve knock resistance in a supercharged engine, it is desirable to make the cylinder bore diameter as small as possible, and also to reduce the bearing load of the crankshaft. Is desirable. On the other hand, when trying to secure a predetermined displacement by increasing the piston stroke while reducing the cylinder bore diameter, it is necessary to consider the limit on the reliability of the piston speed and the limit due to the sudden increase in intake resistance due to the reduction of the intake valve. There is.

That is, considering the reliability limit of the piston speed, the average piston speed Um
Is calculated from the engine speed N and the piston stroke S by the following equation (2).

[0029]

[Equation 2] Um = (N / 30) · S (2) By the way, the limit value of the average piston speed Um is usually said to be Um = 20 m / s (= 2.0 × 10 ⁴ mm / s). By substituting this into the above equation, the piston stroke at that time can be obtained according to the engine speed. Then, the single chamber volume corresponding to the cylinder bore diameter B is obtained based on this piston stroke. This Um = 2.0 ×
A single chamber volume of 10 ⁴ mm / s is required for an engine speed of 4
1000 rpm from 000 rpm to 8000 rpm
When each of them is shown in correspondence with the cylinder bore diameter, it becomes like a one-dot chain line in FIG.

Further, it is generally known that when the average intake Mach number Mi becomes Mi = 0.5, the intake resistance rapidly increases, which causes the volume efficiency to rapidly decrease. The average intake Mach number Mi can be expressed by the following equation (3).

[0031]

[Equation 3] Mi = {Vh · (η _V / 100)} / {a · Fim · (θic−θio) / 6 · N} (3) where, Vh: Single chamber volume η _V : Volume efficiency a: sound velocity θic: intake valve opening timing θio: intake valve closing timing N: engine rotation speed Fim: average intake opening area. The average intake opening area Fim is Fim = Fia /
It can be expressed as (θic−θio). However, Fia
Is the intake valve effective angle area.

Further, the conditions are set as follows.

(1) Each cylinder has two intake valves and two exhaust valves, the intake valves have the same size and the exhaust valves have the same size, and the area ratio between the intake valves and the exhaust valves is 1.5 at the throat portion. .

(2) The volumetric efficiency is 100%.

(3) Combustion chamber shape: Pent roof type (4) Space between both intake valve seats: 2.5 mm or more (5) Space between intake / exhaust valve seats: 3.5 mm or more (6) Both exhausts Distance between valve seats: 4.0 mm or more (7) Distance between intake valve seat and spark plug: 2.5
mm or more (8) Distance between exhaust valve seat and spark plug: 3.5
mm or more (9) Valve grip angle: 30 degrees (10) Plug diameter: 14 mm diameter (11) Stem diameter: 6 mm diameter (12) Throat diameter = valve seat diameter-5 mm (13) Valve lift: 8.5 mm (14 ) Valve opening period: 256 degCA If these conditions are specified, the intake valve effective angle area F
The ia and the cylinder bore diameter have a correspondence relationship as shown in FIG. 4, and the average Mach number Mi depends on the single chamber volume, the cylinder bore diameter, and the engine rotation speed. Then, the engine speed is 1 from 4000 rpm to 8000 rpm.
Mi = 0.5 for cylinder bore diameter every 000 rpm
When the volume of the single chamber is calculated, the solid line in FIG. 5 is obtained.

By the way, assuming that the engine rotation speed at which Um = 2.0 × 10 ⁴ mm / s is Na and the engine rotation speed at which Mi = 0.5 is Nb, if Nb> Na, the reliability is improved. The diameter of the cylinder bore was set to be large enough to allow air to reach the speed higher than the limit, which violates the demand for a smaller cylinder bore. On the other hand, Na-Nb> 20
If it becomes 00 rpm, the maximum horsepower generation rotation speed (rotation speed near Nb) becomes too low as compared with the reliability limit, which is not preferable in terms of performance. In terms of performance, the maximum horsepower generation rotational speed is preferably 5000 rpm or more, and the reliability limit may be 8000 rpm or less.

From these points, Na ≧ Nb, Na-Nb
≦ 2000 rpm, Nb ≧ 5000 rpm, Na ≦ 80
The range of 00 rpm, that is, the range shown by hatching in FIG. 4, satisfies the requirements for the cylinder bore diameter reduction and the engine performance requirements. As for the cylinder bore diameter B, a preferable range is approximately 50 ≦ B ≦ 70.
Further, it is preferable that the volume of the single chamber is in the range of 150 cc to 400 cc, and the ratio of the piston stroke to the cylinder bore diameter is larger than 1.

Grounds and settings for setting that satisfy the above equation (1)
In FIG. 6, ε = 6 with the geometric compression ratio as a parameter.
In each case of 7, 8, and 9, the charging efficiency and the ignition timing M
The relationship with the retard amount from BT is shown by lines A1, A2, A3.
It is indicated by A4, and the retard amount is a value adjusted so that knocking does not occur depending on the filling efficiency. In addition,
This data is the data when the cylinder bore diameter B is B = 63 mm.

As shown in this figure, knocking is more likely to occur as the geometric compression ratio ε becomes higher and as the packing efficiency η _c becomes higher. The packing efficiency η _c is reduced as the geometrical compression ratio ε is increased, and the packing efficiency η _c is reduced by about 28% when the geometrical compression ratio ε is increased by 1.0. Further, the ignition as retard amount of the ignition timing as the charging efficiency eta _c is increased in epsilon certain geometric compression ratio is proportionally increased, the geometric compression ratio at a constant charging efficiency eta _c epsilon is high It is necessary to increase the retard amount proportionally in order to prevent knocking. Therefore, these geometric compression ratio ε and packing efficiency η
The relationship between _c and the ignition timing retard amount has the characteristic shown in FIG.

FIG. 7 shows the relationship between the retard amount from the MBT of the ignition timing and the torque reduction rate. As shown in FIG. 7, the torque reduction rate increases as a quadratic function as the retard amount increases.

FIG. 8 shows the relationship between the charging efficiency η _c and the geometric compression ratio ε and the retard amount of the ignition timing on the lines A2, A3.
A4 (these are the same as those shown in FIG. 3), and the substantial charging efficiency improvement allowance considering the torque decrease due to the ignition timing retard is shown by broken lines B2, B3, B4. This substantial charging efficiency improvement amount corresponds to the engine torque, and the torque increase when the charging efficiency is increased while retarding the ignition timing at each compression ratio shown in the figure is determined by the MBT ignition timing. It is converted into the charging efficiency improvement charge when it is assumed that the same amount of torque is obtained by increasing the charging amount.

As shown in this figure, when the charging efficiency η _c is increased while retarding the ignition timing, the increase (line A
2, A3, A4), the amount obtained by subtracting the amount of torque reduction caused by ignition timing retard from the amount of torque increase caused by ignition timing retard (broken lines B2, B3, B4), and when the ignition timing retard amount increases to some extent It becomes the maximum, and if the retard amount further increases, the packing efficiency η
_{Even if c} increases, the torque decreases. Therefore, the charging efficiency η _c
In order to increase the torque by increasing the charging efficiency η _c while suppressing the knocking which tends to occur with the increase of the ignition timing by retarding the ignition timing, the torque near the maximum value in the characteristics shown by the broken lines B2, B3, B4 can be obtained. It is desirable that the maximum value be within 2%. Then, the retard amount and the charging efficiency η _{c at} which the torque reaches this level are plotted (points c ₂ , c ₃ and c ₄ ) and connected to each other to form a line C.

Based on the data in FIG. 8, the compression ratio ε and the charging efficiency η _c when the charging efficiency η _c is increased to the point where a torque lower than the maximum torque by 2% is obtained while retarding the ignition timing. When the relationship is calculated, the line E in FIG. 9 is obtained. That is, this line E represents the characteristics of the line C in FIG. 8 with the horizontal axis being the compression ratio ε. Further, the line D in FIG. 9 shows the case where the charging efficiency η _c is adjusted so that the knock limit of the ignition timing matches the MBT (point a ₁ in FIG. 6,
(corresponding to a ₂ , a ₃ and a ₄ ), the relationship between the compression ratio ε and the packing efficiency η _c is shown.

As is clear from the comparison between the lines D and E, the torque can be increased by increasing the charging efficiency η _c while retarding the ignition timing. When the characteristic of the line E is mathematically expressed,

[0045]

[Equation 4] η _c = −24 · ε + 375 (4) However, this data shows the cylinder bore diameter B as B =
It is when it is set to 63 mm. As for the relationship with the cylinder bore diameter B, the filling efficiency η at which a constant knock resistance is obtained can be obtained.
_c increases as the cylinder bore diameter B decreases,
According to the experiments conducted by the present inventor, the filling efficiency η _c could be increased by 7% when the cylinder bore diameter B was reduced by 4 mm. From this, the following formula (5) is obtained by adding the cylinder bore diameter B to the formula (4).

[0046]

[Equation 5] η _c = −24 · ε + 375-7 / 4 (B−63) (5) Further, the atmospheric temperature To is To = 300 ° K, and the supercharging air temperature Tb (almost due to the action of the intercooler). Tb = 3)
33 ° K (60 ° C), volume efficiency η _v is η _v = 0.9 × 1
When 00% as the relationship between the pressure ratio of the charging efficiency eta _c and supercharger gamma (the ratio P / P ₀ of the atmospheric pressure P ₀ of the supercharger discharge side pressure P),

[0047]

## EQU6 ## η _c = η _v × (To / Tb) × γ = 90 × (300/333) × γ ≈81 · γ (6) This is illustrated in FIG. 10.

Substituting equation (6) into equation (5),

[0049]

(7) γ = −0.296 · ε + 6.0−0.022B (7), and when B = 63 mm, the line G in FIG. 11 is obtained. Then, approximately the maximum torque is obtained above the line given by the equation (7).

Therefore, the relation of the above equation (1) is approximately given, and by the setting satisfying this relation, the torque is increased to almost the maximum while the knock resistance is secured by the ignition timing retard. .

Knocking by late closing of intake valve and scavenging
When the suppression action is applied In the valve operating characteristics as shown in FIG. 3, when the intake valve closing timing IC is set to be late at 60 ° or more after the bottom dead center at the crank angle, the effective compression ratio decreases and the temperature rise due to compression work. Knock resistance is improved by suppressing this. Also,
When the overlap amount O / L during the opening period of the intake / exhaust valve is set to a large value of 20 ° or more in crank angle, the scavenging property is enhanced and the residual exhaust gas is reduced to enhance the knock resistance.

Therefore, at least in the low rotation and high load region,
After increasing the knock resistance by delaying the intake valve closing timing or increasing the overlap amount as described above,
When the charging efficiency η _c is increased to the point where a torque lower than the maximum torque is obtained by 2% while retarding the ignition timing, the correspondence relationship between the compression ratio ε and the charging efficiency η _c at B = 63 mm is shown in FIG. It becomes like the line F in the middle, and the filling efficiency η _c is about 20% higher than the characteristic of the line E. Therefore, when the cylinder bore diameter B is added to this and the correspondence is made into a mathematical expression, the following expression (8) is obtained.

[0053]

[Equation 8] η _c = −24 · ε + 395-7 / 4 (B−63) (8) When the charging efficiency η _c is converted into the supercharging pressure ratio γ,

[0054]

[Formula 9] γ = −0.296 · ε + 6.2−0.022B (9), and when B = 63 mm, the line H in FIG. 11 is obtained.

Therefore, when the intake valve is late closed or the scavenging action is added in this way, approximately the torque is almost maximized while the knock resistance is secured by the setting that satisfies the following expression (10). Will be increased.

Γ ≧ −0.29 · ε + 6.2-0.022 · B (10) Considering EGR and octane number of fuel In an engine as shown in FIG. 1, EGR for high load at high load
When the EGR gas sufficiently cooled by the passage 22 is supplied to the combustion chamber, the knock resistance is enhanced. Further, as the octane number of the fuel increases, the knock resistance is enhanced.

Therefore, when the EGR rate when supplying the cooled EGR gas to the combustion chamber is Re (%) and the octane number of the fuel is Ro, and these effects are added to the equation (1), γ · (1 + Re / 100) ≧ −0.29 · ε + (a + b · Ro) −0.022 · B + c · Re / 100 (11) In this formula, a, b, and c are constants, and it was found by the inventor of the present invention that a = 0.09, b.
= 0.059, c = 0.007. In addition, while substituting the values of a, b, and c, the octane number Ro
Is about 100 and the EGR rate Re is 0, the above equation (1) is obtained.

Effect on Exhaust Temperature When the effect of the compression ratio ε on the exhaust temperature is considered when setting is made so as to satisfy the above formula (1), (10) or (11), When the ratio ε is increased, the expansion ratio is increased, so that the exhaust temperature is lowered, and as the compression ratio ε is decreased, the exhaust temperature is increased.

However, since the ignition timing is adjusted according to the compression ratio ε and the charging efficiency η _c in order to secure the knock resistance, the ignition timing is advanced when the compression ratio ε becomes low, so that the exhaust gas temperature Rise is avoided. This action will be described based on FIG. 6 and FIG. 12 described above.

FIG. 12 shows data experimentally investigated with respect to reduction of exhaust temperature, showing effects of increase in compression ratio ε and advance of ignition timing on exhaust temperature.
For the purpose of comparison, the effects of supplying cooled EGR gas, making the air-fuel ratio rich, reducing the intake air temperature, and reducing the exhaust pressure are also shown for comparison. As is clear from this figure, changing the compression ratio ε by 1 changes the exhaust temperature by about 20 ° C, while changing the ignition timing by 5 ° changes the exhaust temperature by about 40 ° C. Therefore, when the compression ratio ε is decreased by 1, the ignition timing may be advanced by about 2.5 deg in order to cancel the increase in exhaust gas temperature.

Further, as shown in FIG. 6, in order to keep the knock resistance the same, it is necessary to decrease the charging efficiency η _c by about 28% when the compression ratio ε is increased by 1, while the ignition timing is changed. It is necessary to reduce the filling efficiency η _c by about 4% if the distance is advanced by 1 deg. Therefore, when the compression ratio ε is decreased by 1, the ignition timing may be advanced by about 7 deg in order to keep the charging efficiency η _c and the knock resistance the same. When the angle of advance is advanced to this extent, the exhaust gas temperature is lowered more than the amount of increase due to the reduction of the compression ratio ε.

By adjusting the ignition timing in this way, the exhaust temperature does not rise even when the compression ratio ε is low.

In the engine with a supercharger of the present invention, the supercharger is not limited to the Rishorum type supercharger shown in the embodiment, but may be another mechanical supercharger or a turbocharger. .

[0064]

The present invention can be understood from the data shown in FIG. 5 by setting the cylinder bore diameter to 70 mm or less in a supercharged engine having two or more intake valves in one cylinder. It is possible to effectively reduce the cylinder bore diameter and improve knock resistance. In addition, the relationship between the pressure ratio γ of the supercharger, the geometric compression ratio ε of the engine, and the cylinder bore diameter B (mm) is γ ≧ −0.29 · ε +
6 to FIG.
As can be understood from the data of No. 1, the filling efficiency can be increased to the extent that almost the maximum torque can be obtained while ensuring the knock resistance.

Further, the cylinder bore diameter B is 50 ≦ B ≦ 7.
Within the range of 0, the cylinder bore diameter can be reduced while satisfying the engine performance requirements.

Further, the closing timing of the intake valve is set to 60 ° or more after the bottom dead center in the crank angle, or the overlap amount of the opening period of the intake valve and the exhaust valve is set to 20 ° or more in the crank angle. , Γ ≧ −0.29 · ε + 6.2-0.0
22 · B, the intake valve is late closed or the amount of overlap is large, so that the knock resistance is further enhanced, and in connection therewith, a relationship that the engine torque can be further enhanced can be obtained. it can.

[Brief description of drawings]

FIG. 1 is an overall schematic diagram of an engine with a supercharger according to an embodiment of the present invention.

FIG. 2 is a schematic view of an engine body and a portion in the vicinity thereof.

FIG. 3 is a diagram showing valve timing.

FIG. 4 is a diagram showing a relationship between a cylinder bore diameter and an intake valve effective angle area.

FIG. 5 is a diagram showing a relationship between a cylinder bore diameter and a single chamber volume.

FIG. 6 is a diagram showing the relationship between the retard amount from the MBT, the charging efficiency, and the compression ratio.

FIG. 7 is a diagram showing the relationship between the retard amount from the MBT and the torque reduction rate.

FIG. 8 is a diagram showing the relationship among the retard amount from the MBT, the charging efficiency, the compression ratio, and the torque.

FIG. 9 is a diagram showing a relationship between compression ratio and filling efficiency.

FIG. 10 is a diagram showing a relationship between a supercharging pressure ratio and a charging efficiency.

FIG. 11 is a diagram showing a relationship between a compression ratio and a supercharging pressure ratio.

FIG. 12 is a diagram showing an exhaust gas temperature reduction effect by various elements.

[Explanation of symbols]

 1 engine body 3a, 3b intake port 10 intake passage 17 supercharger

Claims (4)

[Claims] 1. A supercharged engine having a supercharger in the intake passage of a reciprocating piston type engine having two or more intake valves in one cylinder. The cylinder bore diameter is 70 mm.
The pressure ratio (γ = P) of the supercharger discharge side pressure (P) to the atmospheric pressure (P ₀ ) at the time of high load is set as follows.
/ P ₀ ), the geometrical compression ratio (ε) of the engine, and the cylinder bore diameter (Bmm) are set so as to satisfy the following equation: supercharged engine. γ ≧ -0.29 ・ ε + 6.0-0.022 ・ B 2. The supercharged engine according to claim 1, wherein the cylinder bore diameter is 50 mm or more. 3. A mechanical supercharger is provided in the intake passage, the closing timing of the intake valve is set to 60 ° or more after bottom dead center in crank angle, and the supercharger discharge side pressure (P )
Pressure ratio (γ = P / P ₀ ) to the atmospheric pressure (P ₀ ), the geometrical compression ratio (ε) of the engine, and the bore diameter (Bm of the cylinder
The supercharged engine according to claim 1, wherein the relationship with m) is set so as to satisfy the following equation. γ ≧ -0.29 ・ ε + 6.2-0.022 ・ B 4. A mechanical supercharger is provided in the intake passage, the overlap amount of the opening period of the intake valve and the exhaust valve is set to 20 ° or more in crank angle, and the supercharger discharges at high load. The relationship between the pressure ratio (γ = P / P ₀ ) of the side pressure (P) to the atmospheric pressure (P ₀ ), the geometrical compression ratio (ε) of the engine and the bore diameter (Bmm) of the cylinder is expressed by the following equation. The supercharged engine according to claim 1, wherein the engine is set to satisfy the above conditions. γ ≧ -0.29 ・ ε + 6.2-0.022 ・ B