JP2020186696A - Internal combustion engine with supercharger - Google Patents

Abstract

To provide an internal combustion engine with a supercharger which is reduced in a pumping loss.SOLUTION: An internal combustion engine with a supercharger comprises an exhaust manifold 60 for merging exhaust gases discharged from an exhaust port through a plurality of exhaust pipes 61 with each other in an aggregation part 62, and a supercharger turbine 70 arranged at an exhaust downstream side of the exhaust manifold 60. A throttle part is formed at exhaust downstream-side end parts of the exhaust pipes which are coagulated in the aggregation part which is most approximate to the supercharger turbine 70 in the aggregation part 62. The throttle part has a cross section area which is smaller than a cross section are of the exhaust pipe 61 except for the exhaust downstream-side end parts, and equal to 80% and smaller than 100% with respect to an effective cross section area St of the supercharger turbine 70 which is defined by a flow rate dm/dt of the exhaust gas passing the supercharger turbine 70, fluid density ρ, and a pressure ratio p1/p2 of the supercharger turbine 70. By this constitution, exhaust interference can be reduced without largely increasing a pumping loss, and can obtain high turbine power.SELECTED DRAWING: Figure 3

Description

The present invention relates to a multi-cylinder turbocharged internal combustion engine installed in a vehicle.

Conventionally, a multi-cylinder structure is configured such that the exhaust gas discharged from the exhaust port of each cylinder is merged by using an exhaust manifold and the exhaust gas is sent out into the turbine housing of the turbocharger turbine immediately downstream of the exhaust manifold. Internal combustion engines with a supercharger are known. One of such an internal combustion engine with a supercharger (hereinafter, referred to as a "conventional engine") is exhaust gas by providing a tapered nozzle-shaped flow path portion immediately upstream of the collecting portion of the exhaust manifold. The dynamic pressure at the outlet part of the is once increased. As a result, the exhaust pulse generated from each cylinder wraps around to other cylinders (so-called exhaust interference) is reduced. Further, the conventional engine is equipped with an exhaust gas recirculation device (external EGR), and by temporarily increasing the dynamic pressure at the outlet portion of the exhaust gas as described above, the exhaust pulse is effectively applied to the external EGR. The EGR rate can be improved (see, for example, Patent Document 1).

JP-A-2007-64099 (Fig. 2) Japanese Unexamined Patent Publication No. 2005-16313 (Fig. 4)

By the way, in a conventional engine, in order to improve the EGR rate and reduce exhaust interference, the minimum flow path cross-sectional area (cross-sectional area of the nozzle at the tip of each exhaust pipe assembly) is the maximum opening area of the exhaust valve per cylinder. It is set to be about 25% of. Further, in the case of the internal combustion engine disclosed in Patent Document 2, the sum of the cross-sectional areas of the two exhaust passages at the end position of the partition portion of the exhaust pipe is the sum of the cross-sectional areas of the exhaust passage at the boundary between the introduction portion and the scroll portion in the turbine housing. It is set to be 50% to 80% of the cross-sectional area of.

As described above, in the conventional engine and the internal combustion engine disclosed in Patent Document 2, the cross-sectional area of the "throttle portion" in each exhaust pipe of the exhaust manifold is relative to the maximum opening area of the exhaust valve or the cross-sectional area of the entrance of the scroll portion. It has been made excessively small. However, with such a configuration, the pressure in the exhaust manifold rises excessively, so that the pumping loss may increase or the exhaust valve may open unintentionally. Further, in the conventional engine and the internal combustion engine disclosed in Patent Document 2, it is necessary to provide a diffuser between the exhaust manifold and the scroll portion of the turbine to reduce the pressure loss.

The present invention has been made to address the above problems. That is, one of the objects of the present invention is to provide an internal combustion engine with a supercharger capable of reducing exhaust interference without significantly increasing pumping loss and thus achieving a high turbine power.

The internal combustion engine with a supercharger of the present invention (hereinafter, also referred to as "the engine of the present invention") includes a plurality of cylinders (22), an exhaust manifold (60), and a turbine for a supercharger (70). I have.

The plurality of cylinders include an exhaust valve (25). The exhaust manifold collects exhaust gas discharged from the exhaust port through a plurality of exhaust pipes (61) connected to the exhaust ports (26) of the plurality of cylinders when the exhaust valve is open. It is designed to merge in (62). The turbocharger turbine is provided on the exhaust downstream side of the exhaust manifold.

However, in the gathering part of a plurality of exhaust pipes, exhaust interference occurs between the exhaust pipes corresponding to the two cylinders in which the opening periods of the exhaust valves overlap each other, and the exhaust gas is used for the turbocharger on the downstream side of the exhaust gas. Since the amount of exhaust gas directed to the turbine is reduced, the energy efficiency of the turbocharger turbine is reduced.

Therefore, in the engine of the present invention, the exhaust downstream end is excluded from the exhaust downstream end (61a) of the exhaust pipe that collects at the exhaust pipe closest to the supercharger turbine (62a). The flow rate (dm / dt) of exhaust gas passing through the turbocharger turbine, the fluid density (ρ), and the pressure ratio of the turbocharger turbine (ρ), which is smaller than the cross-sectional area (Sep) of the exhaust pipe. A throttle portion (63) having a cross-sectional area (Seo) of 80% or more and less than 100% with respect to the effective cross-sectional area (St) of the turbocharger turbine determined by p ₁ / p ₂ ) is formed.

Generally, a turbocharger turbine is designed so that its cross-sectional area is smaller than that of an exhaust pipe because work is taken out by a pressure ratio (expansion ratio). That is, it is designed so that the pressure at the turbocharger turbine inlet is higher. For this reason, the proportion of exhaust gas that goes around to other cylinders tends to increase due to exhaust interference at the gathering portion immediately before the turbocharger turbine. Therefore, according to the engine of the present invention, for example, in an exhaust manifold having a plurality of collecting portions by gradually assembling exhaust pipes, at least a turbocharger turbine among the collecting portions of a plurality of exhaust pipes of the exhaust manifold. A "throttle portion" is formed at the exhaust downstream end of the exhaust pipe that gathers at the gathering portion closest to. For example, in an exhaust manifold having only one collecting portion, a throttle portion is formed at each of the exhaust downstream ends of the exhaust pipe that gathers at the collecting portion.

Further, in the engine of the present invention, the cross-sectional area of the "throttle portion" formed at the downstream end of the exhaust gas is set to 80% or more and less than 100% of the effective cross-sectional area of the turbocharger turbine. There is no possibility that the pumping loss will increase depending on the part. In addition, since the flow velocity of the exhaust gas input to the turbocharger turbine does not become excessively high, a diffuser is not required between the exhaust manifold and the turbocharger turbine. Therefore, according to the engine of the present invention, exhaust interference can be reduced without significantly increasing pumping loss, and as a result, high turbine power can be achieved.

In the above description, in order to help the understanding of the invention, the reference numerals used in the embodiments are attached in parentheses to the configurations of the invention corresponding to the embodiments, but each constituent requirement of the invention is based on the reference numerals. It is not limited to the specified embodiment.

FIG. 1 is a schematic configuration diagram of an internal combustion engine with a supercharger according to the first embodiment of the present invention. 2A and 2B are external views of the exhaust manifold shown in FIG. 1, FIG. 2A is a perspective view, and FIG. 2B is a top view. FIG. 3 is a schematic cross-sectional view of the exhaust manifold and the turbine shown in FIG. FIG. 4 is a diagram schematically showing a gathering portion and a throttle portion of the exhaust manifold shown in FIG. FIG. 5 is a diagram showing a change in the blowdown work amount of the turbine shown in FIG. 1 with respect to the area ratio. FIG. 6 is a diagram showing a change in the gas flow rate immediately before the gathering portion of the exhaust manifold shown in FIG. 1 with respect to the crank angle. FIG. 7 is a diagram showing a change in the area ratio of the gas flow rate at the time of blowdown input to the turbine shown in FIG. FIG. 8 is a diagram showing a change in the exhaust pressure with respect to the crank angle in one of the exhaust pipes of the exhaust manifold shown in FIG. FIG. 9 is a diagram showing a change in the area ratio of the maximum pressure (peak value) at the time of blowdown in one of the exhaust pipes of the exhaust manifold shown in FIG. FIG. 10 shows the change in the area ratio of the maximum pressure while the valve opening period of the exhaust valve in one of the exhaust pipes of the exhaust manifold shown in FIG. 1 overlaps with the valve opening period of the exhaust valve of the other cylinder. It is a figure shown. 11A and 11B are views showing a trade-off curve between the maximum output and low-speed torque of a general internal combustion engine, FIG. 11A is a view when the exhaust manifold is not provided with a throttle portion, and FIG. The figure in the case where the diaphragm portion with a small degree of aperture is provided, and (C) is the figure in the case where the diaphragm portion with a relatively large degree of aperture is provided. FIG. 12 is a diagram schematically showing a gathering portion and a throttle portion of the exhaust manifold according to the second embodiment of the present invention. FIG. 13 is a diagram schematically showing one gathering portion and a throttle portion of the exhaust manifold according to the third embodiment of the present invention and a diagram showing the relationship between the exhaust timing of each cylinder. FIG. 14 is a diagram schematically showing one gathering portion and a throttle portion of the exhaust manifold according to the third embodiment of the present invention and a diagram showing the relationship between the exhaust timing of each cylinder. FIG. 15 is a diagram schematically showing a gathering portion and a throttle portion of the exhaust manifold according to the fourth embodiment of the present invention and a diagram showing the relationship between intake / exhaust timing of each cylinder. FIG. 16 is a diagram schematically showing a gathering portion and a throttle portion of an exhaust manifold according to a first modification of the present invention, and FIG. 16A is a diagram showing an application example of a 5-cylinder internal combustion engine. B) is a diagram showing an example of application to a 3-cylinder internal combustion engine. FIG. 17 is a diagram schematically showing a gathering portion and a throttle portion of the exhaust manifold according to the second modification of the present invention, and is a diagram showing an application example to a twin scroll type turbine. FIG. 18 is a diagram schematically showing the collecting portion and the throttle portion of the exhaust manifold according to the third modification of the present invention, and FIG. 18A shows a bypass pipe connected to the exhaust manifold having only one collecting portion. The figure which showed the example, (B) is the figure which showed the example which connected the bypass pipe to the exhaust manifold which has three gathering parts.

<First Embodiment>
(Constitution)
FIG. 1 is a schematic configuration diagram of an internal combustion engine with a supercharger (hereinafter, also referred to as “first engine”) 10 according to the first embodiment of the present invention. The first engine 10 is mounted on a vehicle (not shown). The first engine 10 is a multi-cylinder (in-line 4-cylinder in this example), 4-cycle, piston reciprocating type, diesel engine. The first engine 10 includes an engine main body 20, a fuel supply system 30, an intake system 40, and an exhaust system 50.

The engine body 20 includes a body 21 including a cylinder block, a cylinder head, a crankcase, and the like. Four cylinders (combustion chambers) 22 are formed in the main body 21. These four cylinders 22 are also hereinafter also referred to as first cylinder # 1, second cylinder # 2, third cylinder # 3, and fourth cylinder # 4. A fuel injection valve (injector) 23 is arranged above each cylinder 22.

The fuel injection valve 23 opens in response to an instruction from an electronic control device (ECU) (not shown) to directly inject fuel into the cylinder 22. Two intake valves 24 are provided in each cylinder 22. The intake valve 24 includes a first intake valve 241 corresponding to the first cylinder # 1, a second intake valve 242 corresponding to the second cylinder # 2, a third intake valve 243 corresponding to the third cylinder # 3, and a fourth cylinder. It includes a fourth intake valve 244 corresponding to # 4. The intake valve 24 opens at a predetermined timing in synchronization with the rotation angle (hereinafter, referred to as “crank angle”) CA of a crankshaft (not shown), and closes at a predetermined timing.

Two exhaust valves 25 are provided in each cylinder 22. The exhaust valve 25 includes a first exhaust valve 251 corresponding to the first cylinder # 1, a second exhaust valve 252 corresponding to the second cylinder # 2, a third exhaust valve 253 corresponding to the third cylinder # 3, and a fourth cylinder. It includes a fourth exhaust valve 254 corresponding to # 4. The exhaust valve 25 opens at a predetermined timing in synchronization with the crank angle CA and closes at a predetermined timing. The period from the opening of the exhaust valve 25 to the closing of the valve (hereinafter referred to as "valve opening period") depends on the angle at which the crankshaft rotates while the exhaust valve 25 is separated from the valve seat. It is defined and hereinafter referred to as "exhaust valve working angle θex". The exhaust valve working angle θex is configured to be adjustable by a well-known variable valve mechanism (not shown).

The fuel supply system 30 includes a fuel pressurizing pump 31, a fuel delivery pipe 32, and a fuel delivery pipe 33. The discharge port of the fuel pressurizing pump 31 is connected to the fuel delivery pipe 32. The fuel delivery pipe 32 is connected to the fuel delivery pipe 33. The fuel delivery pipe 33 is connected to the fuel injection valve 23. The fuel pressurizing pump 31 pumps up fuel stored in a fuel tank (not shown), pressurizes the fuel, and supplies the pressurized high-pressure fuel to the fuel delivery pipe 33 through the fuel delivery pipe 32.

The intake system 40 includes an intake manifold 41, an intake pipe 42, a compressor 43 of a supercharger TC, and an intercooler 44.

The exhaust system 50 includes an exhaust manifold 60, a turbocharger TC turbine (supercharger turbine) 70, and the like.

The exhaust manifold 60 includes a branch portion 61 and a gathering portion 62. The branch portion 61 includes a first exhaust pipe 611, a second exhaust pipe 612, a third exhaust pipe 613, and a fourth exhaust pipe 614. One end of the first exhaust pipe 611 communicates with the exhaust port 261 of the first cylinder # 1, and the other end communicates with the collecting portion 62. One end of the second exhaust pipe 612 communicates with the exhaust port 262 of the second cylinder # 2, and the other end communicates with the collecting portion 62. One end of the third exhaust pipe 613 communicates with the exhaust port 263 of the third cylinder # 3, and the other end communicates with the collecting portion 62. One end of the fourth exhaust pipe 614 communicates with the exhaust port 264 of the fourth cylinder # 4, and the other end communicates with the collecting portion 62.

A throttle portion 63 is formed on the exhaust upstream side of the gathering portion 62. The diaphragm portion 63 includes a first diaphragm portion 631, a second diaphragm portion 632, a third diaphragm portion 633, and a fourth diaphragm portion 634.

More specifically, as shown in FIG. 2A, the exhaust manifold 60 further includes a mounting flange 64 that is coupled to the cylinder head of the main body 21. Each exhaust upstream end of the first exhaust pipe 611, the second exhaust pipe 612, the third exhaust pipe 613, and the fourth exhaust pipe 614 is welded to the mounting flange 64, respectively.

The first throttle portion 631 is formed at the exhaust downstream end portion 611a of the first exhaust pipe 611. The second throttle portion 632 is formed at the exhaust downstream end portion 612a of the second exhaust pipe 612. The third throttle portion 633 is formed at the exhaust downstream end portion 613a of the third exhaust pipe 613. The fourth throttle portion 634 is formed at the exhaust downstream end portion 614a of the fourth exhaust pipe 614. The cross-sectional area Seo1 of the first throttle portion 63 is smaller than the cross-sectional area Sep1 of the first exhaust pipe 611 excluding the exhaust downstream end portion 611a. Similarly, the cross-sectional area Seo2 of the second throttle portion, the cross-sectional area Seo3 of the third throttle portion, and the cross-sectional area Seo4 of the fourth throttle portion are the cross-sectional areas Seo2 of the second exhaust pipe 612 and the cross-sectional area of the third exhaust pipe 613, respectively. It is smaller than the cross-sectional area of Seo3 and the fourth exhaust pipe 614. Hereinafter, when it is not necessary to distinguish between the exhaust downstream end portions 611a, 612a, 613a and the exhaust downstream end portion 614a, they are simply referred to as the exhaust downstream end portion 61a.

As shown in FIG. 2B, the collecting portion (collecting member) 62 has an inlet opening 621 and an outlet opening 622. A flange 623 protruding in the outer peripheral direction is formed around the outlet opening 622. In the first to fourth exhaust pipes 611 to 614, the exhaust downstream end portions 61a are welded to the inlet opening 621 of the collecting portion (exhaust collecting pipe) 62, respectively.

As shown in FIG. 3, the turbine 70 includes a turbine housing 71, a turbine wheel 72, a rotating shaft 73, and the like. The turbine housing 71 includes an exhaust introduction section 711 and a scroll section 712. The turbine housing 71 is formed so that the cross-sectional area of the flow path is constant from the exhaust introduction portion 711 to the scroll portion 712. That is, the opening area of the inlet of the turbine housing 71 (hereinafter, referred to as “turbine inlet”) 711a is equal to the opening area of the inlet 712a of the scroll portion 712. A flange 74 protruding in the outer peripheral direction is formed around the turbine inlet 711a.

ここで、κは比熱比、Ｒは気体定数（Ｊ／(mol・Ｋ)）、Ｔは絶対温度（Ｋ）である。このタービン有効断面積Ｓｔは、タービン７０の圧力比ｐ₁／ｐ₂及び質量流量ｄｍ／ｄｔ等が変動した場合であっても、大きく変動することはなく、主にタービン７０の形状及び構造によって決定される。 By the way, the effective cross-sectional area St (m ² ) of the turbine 70 is the mass flow rate dm / dt (kg / s) of the exhaust gas (fluid) flowing in the turbine 70, and the fluid density ρ (kg / kg / kg / s) of the exhaust gas at the turbine inlet 711a. m ³ ), calculated based on the ratio of the exhaust gas pressure p ₁ at the turbine inlet 711a to the fluid pressure p _{2 at} the turbine outlet (hereinafter referred to as the "pressure ratio") p ₁ / p ₂ (hereinafter referred to as "pressure ratio"). (1).).

Here, κ is the specific heat ratio, R is the gas constant (J / (mol · K)), and T is the absolute temperature (K). This turbine effective cross-sectional area St does not fluctuate significantly even when the pressure ratio p ₁ / p ₂ and the mass flow rate dm / dt of the turbine 70 fluctuate, and mainly depends on the shape and structure of the turbine 70. It is determined.

The cross-sectional area Seo of the throttle portion 63 of each exhaust pipe 61 is set to 90% of the effective turbine cross-sectional area St. That is, the cross-sectional area Seo of the throttle portion 63 is smaller than the cross-sectional area Sep of each exhaust pipe 61 excluding the exhaust downstream end 61a of each exhaust pipe 61, and is 90% of the effective turbine cross-sectional area St. The exhaust downstream end of each exhaust pipe is formed. In this example, the cross-sectional areas Seo1, Seo2, Seo3, and Seo4 of the exhaust downstream end portion 61a (throttle portion 63) are equal to each other.

By the way, if the length of the throttle portion 63 in the flow path direction is increased, the pressure loss increases. On the other hand, if the length of the throttle portion 63 in the flow path direction is shortened, contraction may occur in the throttle portion 63 and the loss may increase. Therefore, in the first engine 10, the length Ln of the throttle portion 63 in the flow path direction is set to be substantially equal to the length Ls of the short side of the throttle portion 63.

The exhaust manifold 60 and the turbine housing 71 are connected by abutting and fixing the flange 623 of the exhaust manifold 60 and the flange 74 of the turbine 70. The exhaust gas discharged from the turbine 70 is sucked into the catalyst 52 through the downstream exhaust pipe 51.

With such a configuration, the first engine 10 forms an air-fuel mixture in the cylinder 22 by opening the intake valve 24 and supplying fuel into the cylinder 22 by using the fuel injection valve 23. Then, after closing the intake valve, the crankshaft is rotated by compressing and burning the air-fuel mixture. Further, by opening the exhaust valve 25, the exhaust gas after combustion is supplied to the turbine 70 via the exhaust manifold 60. The exhaust gas supplied to the turbine 70 rotates the turbine wheel 72, whereby the air sucked into the intake pipe 42 is compressed by the compressor 43.

As described above, the configuration in which the four exhaust pipes 611 to 614 are assembled in one collecting portion 62 and connected to the turbine 70 is represented by a simple schematic diagram as shown in FIG. In FIG. 4, the collecting portion 62 of each exhaust pipe is represented by the symbol “●”, and the throttle portion 63 is represented by the symbol “) (”. Such a configuration is represented by “4” described later. It may also be referred to as "4-1 configuration" as opposed to "2-1 configuration".

<Action and effect of the first institution>
The reason why the cross-sectional area Seo of the throttle portion 63 of the first engine 10 is set to 90% of the effective turbine cross-sectional area St will be described in detail below. First, in setting the cross-sectional area SEO of the drawing portion 63, the inventor paid attention to the following characteristics, changed the cross-sectional area Seo of the drawing portion 63, and evaluated how these characteristics change.
(1) Blow-down work of the turbine 70 (2) Flow rate in the exhaust pipe 61 (3) Pressure in the exhaust pipe 61

(1) Turbine blowdown work amount The horizontal axis of the graph shown in FIG. 5 is the ratio α of the cross-sectional area Seo of the throttle portion 63 to the effective cross-sectional area St of the turbine α (= Seo / St; hereinafter, simply “area ratio α”. It is called.). On the other hand, the vertical axis represents the turbine work during blowdown (hereinafter referred to as "blowdown work" (unit: a.u.). The blowdown work is that of the nth cylinder #n. It is calculated by time-integrating the turbine power generated during the period from the start of valve opening of the exhaust valve 25n until a certain period of time elapses (while blowdown is occurring) (“n” is the cylinder). Any value from 1 to 4).

As can be understood from FIG. 5, when the area ratio α is larger than 1 (when the cross-sectional area Seo of the throttle portion 63 is larger than the effective turbine cross-sectional area St), the blowdown work amount is as small as the area ratio α (1). (The closer to) it increases. On the other hand, when the area ratio α is smaller than 1 (when the cross-sectional area Seo of the throttle portion 63 is smaller than the turbine effective cross-sectional area St), the blowdown work amount hardly increases even if the area ratio α is reduced. In other words, when the area ratio α is set to be smaller than 1, the blowdown work amount becomes a relatively high value.

(2) Flow rate in the exhaust pipe The horizontal axis of the graph shown in FIG. 6 represents the crank angle CA of the nth cylinder, and the vertical axis is the flow rate (mass flow rate; unit) in the portion of the exhaust manifold 60 immediately before the gathering portion 62. Represents a.u.). The graph shown in FIG. 6 plots the flow rates for the five area ratios α (α = 0.64, 0.87, 1.13, 1.77 and 2.55).

As shown in FIG. 6, when the crank angle CA is around 180 °, the maximum peak flow rate is generated in the forward direction (that is, the direction toward the downstream). This indicates that the exhaust valve 25n of the nth cylinder #n has started to open and a large flow rate is generated in the nth exhaust pipe 61n corresponding to the nth cylinder #n. Further, when the crank angle CA is around 360 °, the maximum peak flow rate is generated in the opposite direction. This is because the exhaust valve of the cylinder (hereinafter, also referred to as "next cylinder") that reaches the exhaust stroke next to the nth cylinder #n when the exhaust valve 25n of the nth cylinder #n is still open. It indicates that the valve is opened and the exhaust gas discharged from the next cylinder wraps around (that is, backflows) to the nth exhaust pipe 61n.

When the area ratio α is large (for example, the area ratio α is 2.55), a peak occurs again in the forward direction when the crank angle CA is around 400 °. This is because the exhaust valve 25n of the nth cylinder #n is closed, so that the exhaust gas flowing back through the nth exhaust pipe 61n cannot flow into the nth cylinder #n, and the nth exhaust pipe 61n It shows that it is folded inside. Further, when the crank angle CA is around 210 °, the flow rate is greatly reduced. This is because the exhaust gas returned from the cylinder (hereinafter, also referred to as "front cylinder") whose exhaust stroke is completed immediately before the nth cylinder #n flows back, and the exhaust gas from the nth cylinder #n is discharged. It indicates that the discharge is hindered.

Therefore, the following can be seen from FIG.
-If the area ratio α is reduced, the maximum peak flow rate in the forward direction decreases.
-If the area ratio α is reduced, the maximum peak flow rate (flow rate around) in the reverse direction decreases.
-If the area ratio α is reduced, the amount of exhaust gas that wraps around from the next cylinder is reduced.
-When the area ratio α is reduced, the amount of exhaust gas that wraps around from the nth cylinder #n to the front cylinder is reduced.

Here, the total mass flow rate (integral) when the crank angle CA is around 180 ° (a certain period after the nth exhaust valve 25n starts opening; the crank angle CA is between 150 ° and 200 °). Value) is defined as "blowdown gas flow rate". As shown in FIG. 7, the blowdown gas flow rate (unit: a.u.) has a small fluctuation amount with respect to the change in the area ratio α in the range where the area ratio α is larger than 1.

This is because, in the range where the area ratio α is larger than 1, the maximum peak flow rate decreases as the area ratio α decreases, but the flow rate of the exhaust gas that wraps around from the front cylinder decreases, so that from the nth cylinder #n. It is considered that the cause is that the exhaust gas is discharged smoothly. On the other hand, the blowdown gas flow rate decreases monotonically as the area ratio α decreases in the range where the area ratio α is smaller than 1. As described above, the smaller the area ratio α, the smaller the influence of the wraparound of exhaust gas from other cylinders. Therefore, it is considered that the monotonous decrease in the area ratio α is due to the increase in pressure loss. As described above, from the viewpoint of the blowdown gas flow rate, it can be seen that the larger the area ratio α is, the better. On the other hand, in the range where the area ratio α is slightly smaller than 1, the flow rate of the exhaust gas that wraps around decreases as the flow rate of the blowdown gas decreases, so that the amount of blowdown work decreases significantly as shown in FIG. do not do.

(3) Pressure in the exhaust pipe In FIG. 8, the horizontal axis represents the crank angle CA based on the compression top dead center of the nth cylinder #n, and the vertical axis represents the exhaust pressure in the nth exhaust pipe 61n. ing. The graph shown in FIG. 8 shows the exhaust pressures (α = 0.84, 0.93, 1.00, 1.09, 1.18, 1.40, 1.96) for the seven area ratios α. The unit is a.u.). The exhaust valve 25n of the nth cylinder #n is adapted to start opening when the crank angle CA is 170 ° and to completely close when the crank angle CA is 370 °. That is, when the crank angle CA is 170 °, the exhaust valve 25n is separated from the valve seat, and when the crank angle CA is 370 °, the exhaust valve 25n comes into contact with the valve seat. The exhaust valve working angle θex at this time is 200 °.

As shown in FIG. 8, when the exhaust valve 25n of the nth cylinder #n is opened, the pressure of the nth exhaust pipe 61n rises due to blowdown. The pressure of the nth exhaust pipe 61n reaches a peak when the crank angle CA is about 210 °, then gradually decreases, and when the crank angle CA is about 250 °, it drops to almost the same pressure as before the valve opening starts. .. When the exhaust valve of the next cylinder is opened when the crank angle CA is 350 °, the pressure of the nth exhaust pipe 61n rises again due to the influence of the blowdown of the next cylinder, and peaks when the crank angle CA is about 400 °. After that, it gradually decreases.

In FIG. 9, the area ratio of the exhaust pressure when the crank angle CA is about 210 ° in FIG. 8 (that is, the maximum value of the exhaust pressure at the time of blowdown (hereinafter, referred to as “maximum blowdown pressure”)). The change with respect to α is shown. FIG. 9 plots the results when the exhaust valve working angle θex is 200 °, and when the exhaust valve working angle θex is 220 ° and 240 °, respectively. As can be seen from FIG. 9, the maximum blowdown pressure increases as the area ratio α decreases. Further, when the area ratio α is smaller than 1, the degree of change of the blowdown maximum pressure with respect to the area ratio α is. , The degree of change is greater than when the area ratio α is greater than 1.

In FIG. 10, the maximum value of the exhaust pressure generated by the wraparound of the exhaust gas from the next cylinder when the crank angle CA is about 400 ° in FIG. 8 (hereinafter, referred to as “maximum wraparound pressure”). The change with respect to the area ratio α is shown. In FIG. 10, the results are plotted for the case where the exhaust valve working angle θex is 200 ° and the case where the exhaust valve working angle θex is 220 ° and 240 °, respectively. As can be seen from FIG. 10, when the area ratio α is larger than 1, the change in the maximum wraparound pressure generated by the wraparound with respect to the area ratio α is small. On the other hand, when the area ratio α is smaller than 1, the integrated value of the pressure generated by the wraparound tends to be smaller as the area ratio α is smaller. Further, the smaller the exhaust valve working angle θex, the lower the exhaust pressure at the time of wraparound is the amount (angle) in which the exhaust valve working angle θex of the nth cylinder #n and the exhaust valve working angle of the next cylinder overlap. ) Becomes smaller.

By setting the area ratio α to a value smaller than 1 in this way, the following effects can be obtained.
-Blowdown work is relatively high (ie, does not decrease).
-The maximum peak flow rate in the reverse direction (flow rate from the next cylinder) decreases.
-The amount of return (return amount) of gas that wraps around from the own cylinder to the front cylinder decreases.
However, on the other hand, the smaller the area ratio α, the larger the pressure loss and the lower the flow rate during blowdown. From the above, the area ratio α is preferably less than 1 and close to 1, and preferably set to 0.8 or more and less than 1 (80% or more and less than 100%). The area ratio α is preferably 0.9 (90%).

Next, the effect of the first engine 10 will be described from the viewpoint of the output (driving force) and the torque generated by the first engine 10 with reference to FIG.

FIG. 11 shows the relationship between the maximum output of the first engine 10, the low-speed torque, and the exhaust valve working angle θex. For example, when the area ratio α is set to 1 or more, as shown in FIG. 11A, the larger (wider) the exhaust valve working angle θex, the higher the maximum output of the first engine 10. This is because increasing the exhaust valve working angle θex reduces the pressure loss at the time of a large flow rate. On the contrary, the smaller (narrower) the exhaust valve working angle θex, the higher the low-speed torque of the engine. This is because the amount of overlap between the valve opening period of the exhaust valve 25n of the nth cylinder #n (exhaust valve working angle θex (n)) and the valve opening period of the exhaust valve of the next cylinder becomes small, and the amount of exhaust gas wraparound. Is small. In this way, the locus through which the operating point passes when the exhaust valve working angle θex is changed is called a “trade-off curve”. On the trade-off curve, the operating points at some exhaust valve working angles θex are shown.

When the area ratio α is set to a value slightly smaller than 1 (for example, 0.9), as shown in FIG. 11 (B), the positions of the operating points at some exhaust valve working angles θex are circled in white. It becomes the position represented by. For example, when the exhaust valve working angle θex is large (θex = 240), if the area ratio α is set to 0.9, exhaust interference (wraparound) is reduced and low-speed torque is increased. On the other hand, the maximum output decreases due to the increase in pressure loss. As a result, the position of the white circle almost overlaps with the trade offline. Therefore, in this case, the same effect as reducing the exhaust valve working angle θex can be obtained. On the other hand, when the exhaust valve working angle θex is small (θex = 180), the overlap amount of the exhaust valve is small and the exhaust interference is small, so that the low-speed torque does not increase significantly and the maximum output decreases slightly. ..

However, when the exhaust valve working angle θex is medium (θex = 225), the low speed torque can be increased, but the maximum output is only slightly reduced. Therefore, as shown in FIG. 11B, the position of the operating point is outside the trade-off curve (the region where the maximum output and the low-speed torque are larger than the trade-off curve). That is, at this time, there is a great merit that cannot be obtained simply by reducing the exhaust valve working angle θex.

By the way, if the area ratio α is made excessively small, exhaust interference (wraparound) can be greatly reduced, and as a result, low-speed torque can be further increased, but on the other hand, pressure loss increases. As a result, the maximum output is greatly reduced. Therefore, in this case, the operating point cannot be located outside the trade-off curve, as indicated by the symbol “◆” in FIG. 11 (C). In other words, it means that no merit can be obtained by making the area ratio α excessively small.

As described above, the first engine 10 is connected to a plurality of cylinders 22 provided with an intake valve 24 and an exhaust valve 25, and to exhaust ports 26 of the plurality of cylinders when the exhaust valve 25 is open. The exhaust manifold 60 that merges the exhaust gas discharged from the exhaust port 26 through the plurality of exhaust pipes 611, 612, 613 and 614 at the collecting portion 62, and the turbocharger turbine 70 provided on the exhaust downstream side of the exhaust manifold 60. And.

In the first engine 10, throttle portions 63 (631, 632, 633 and 634) are formed at the exhaust downstream end portion 61a of the exhaust pipe 61 that gathers at the gathering portion 62a. The throttle portion 63 is smaller than the cross-sectional area Sep of the exhaust pipe 61 excluding the exhaust downstream end portion 61a, and has an exhaust gas flow rate dm / dt, a fluid density ρ, and a supercharger passing through the turbocharger turbine 70. It has a cross-sectional area Seo of 80% or more and less than 100% with respect to the effective cross-sectional area St of the turbocharger turbine 70 determined by the pressure ratio p ₁ / p ₂ of the turbine 70.

Therefore, according to the first engine 10, exhaust interference can be reduced in the exhaust manifold without significantly increasing the pumping loss. As a result, it is possible to realize a high turbine work rate.

<Second Embodiment>
The internal combustion engine with a supercharger (hereinafter, referred to as “second engine”) 10A according to the second embodiment of the present invention is configured such that the exhaust pipes 61 of the exhaust manifold 60 merge stepwise. It differs from the first embodiment in that it has a total of three collecting portions and a throttle portion is formed at the exhaust downstream end of the exhaust pipe closest to the turbine 70. Therefore, this difference will be mainly described below.

As shown in FIG. 12, the second engine 10A includes a second assembly unit 62b, a third assembly unit 62c, and a fourth assembly unit 62d. The first exhaust pipe 611 and the fourth exhaust pipe 614 are assembled in the second collecting portion 62b. The second exhaust pipe 612 and the third exhaust pipe 613 are assembled in the third collecting portion 62c. The exhaust pipe 61A from the first gathering portion 62b and the exhaust pipe 61B from the second gathering portion 62c gather in the fourth gathering portion 62d. In this way, the configuration in which two exhaust pipes are assembled and the number of exhaust pipes is reduced from four to two and then from two to one is hereinafter referred to as "4-2-1 configuration". Called. It can be said that the configuration of 4-2-1 is a configuration in which two exhaust pipes 61 are assembled in stages.

In the second engine 10A, the throttle portion 63A is located immediately upstream of the fourth gathering portion 62d, which is the gathering portion of the exhaust pipe closest to the turbine 70, that is, at the downstream end of the exhaust pipe 61A from the first gathering portion 62b. The throttle portion 63B is formed at the downstream end of the exhaust pipe 61B from the second collecting portion 62c.

As described above, the effective turbine cross-sectional area St is designed to be smaller than the exhaust pipe cross-sectional area Sep. Therefore, assuming that there is no "throttle" in the exhaust pipe, when the exhaust gas passing through the exhaust pipe 61A branches into the exhaust introduction portion 711 and the exhaust pipe 61B at the fourth collecting portion 62d immediately before the turbine 70, A lot of exhaust gas wraps around the exhaust pipe 61B. On the other hand, when the exhaust gas that has passed through the exhaust pipe 61B branches into the exhaust introduction portion 711 and the exhaust pipe 61A at the fourth collecting portion 62d, a large amount of exhaust gas wraps around the exhaust pipe 61A.

Therefore, the second engine 10A forms the throttle portion 63A at the exhaust downstream end of the exhaust pipe 61A, and forms the throttle portion 63B at the exhaust downstream end of the exhaust pipe 61B. The cross-sectional area Seo of each of the throttle portions 63A and 63B is set in the range of 80% or more and less than 100% (for example, 90%) of the effective cross-sectional area St of the turbine.

According to the second embodiment, at least, a throttle portion is formed at the exhaust downstream end portion of the exhaust pipe that gathers at the gathering portion immediately before the turbine where the exhaust gas wraparound is most likely to occur, and the area ratio α is 80% or more and 100. By setting it to less than%, it is possible to reduce the sneaking of exhaust gas to the next cylinder. Therefore, it is possible to realize a high turbine work rate.

<Third Embodiment>
The internal combustion engine with a supercharger (hereinafter, referred to as “third engine”) 10B according to the third embodiment of the present invention has two exhaust gases gathered at each collecting portion in the configuration of 4-2-1. Of the pipes, when the latter half of the exhaust valve opening period of the own cylinder and the first half of the exhaust valve opening period of the other cylinder overlap in the two cylinders corresponding to the exhaust pipe, the exhaust corresponding to the other cylinder It differs from the first embodiment and the second embodiment in that a throttle portion is formed at the downstream end of the exhaust gas of the pipe. Therefore, this difference will be mainly described below.

As shown in FIG. 13, each exhaust valve 25 is opened in the order of the second exhaust valve 252 → the first exhaust valve 251 → the third exhaust valve 253 → the fourth exhaust valve 254 → the second exhaust valve 252 → ... It is designed to do.

The "engine 10B1", which is one of the third engine 10B, includes three gathering portions (fifth gathering section 62e, sixth gathering section 62f, and seventh gathering section 62g). The first exhaust pipe 611 and the fourth exhaust pipe 614 are assembled in the fifth collecting portion 62e. The second exhaust pipe 612 and the third exhaust pipe 613 are assembled in the sixth collecting portion 62f. The exhaust pipe 61A from the fifth gathering portion 62e and the exhaust pipe 61B from the sixth gathering portion 62f gather in the seventh gathering portion 62g.

The gas discharged from the first cylinder # 1 and the gas discharged from the second cylinder # 2 merge at the seventh collecting portion 62 g. Regarding the relationship between the opening time of the first exhaust valve 251 and the opening time of the second exhaust valve 252, since the opening time of the first exhaust valve 251 is late, the gas discharged from the first cylinder # 1 passes through. A throttle portion 63A is formed at the exhaust downstream end of the exhaust pipe 61A.

The gas discharged from the first cylinder # 1 and the gas discharged from the third cylinder # 3 merge at the seventh gathering portion 62 g. Regarding the relationship between the opening time of the first exhaust valve 251 and the opening time of the third exhaust valve 253, since the opening time of the third exhaust valve 253 is late, the gas discharged from the third cylinder # 3 passes through. A throttle portion 63B is formed at the exhaust downstream end of the exhaust pipe 61B.

The gas discharged from the third cylinder # 3 and the gas discharged from the fourth cylinder # 4 merge at the seventh collecting portion 62 g. Regarding the relationship between the opening time of the third exhaust valve 253 and the opening time of the fourth exhaust valve 254, the opening time of the fourth exhaust valve 254 is late, so the gas discharged from the fourth cylinder # 4 passes through. The throttle portion 63A may be formed at the exhaust downstream end of the exhaust pipe 61A.

The gas discharged from the second cylinder # 2 and the gas discharged from the fourth cylinder # 4 merge at the seventh collecting portion 62 g. Regarding the relationship between the opening time of the second exhaust valve 252 and the opening time of the fourth exhaust valve 254, the opening time of the second exhaust valve 252 is late, so the gas discharged from the second cylinder # 2 passes through. The throttle portion 63B may be formed at the exhaust downstream end of the exhaust pipe 61B.

As shown in FIG. 14, the "engine 10B2", which is one of the third engine 10B, includes three gathering portions (8th gathering section 62h, 9th gathering section 62i, and 10th gathering section 62j). .. The first exhaust pipe 611 and the second exhaust pipe 612 are assembled in the eighth collecting portion 62h. The third exhaust pipe 613 and the fourth exhaust pipe 614 are assembled in the ninth collecting portion 62i. The exhaust pipe 61C from the 8th gathering section 62h and the exhaust pipe 61D from the 9th gathering section 62i gather in the 10th gathering section 62j.

It is the eighth gathering portion 62h that the gas discharged from the first cylinder # 1 and the gas discharged from the second cylinder # 2 merge. Regarding the relationship between the opening time of the first exhaust valve 251 and the opening time of the second exhaust valve 252, since the opening time of the first exhaust valve 251 is late, the gas discharged from the first cylinder # 1 passes through. A throttle portion 631 is formed at the exhaust downstream end of the first exhaust pipe 611.

It is at the tenth gathering section 62j that the gas discharged from the first cylinder # 1 and the gas discharged from the third cylinder # 3 merge. Regarding the relationship between the opening time of the first exhaust valve 251 and the opening time of the third exhaust valve 253, since the opening time of the third exhaust valve 253 is late, the gas discharged from the third cylinder # 3 passes through. A throttle portion 63D is formed at the exhaust downstream end of the exhaust pipe 61D.

It is at the ninth gathering section 62i that the gas discharged from the third cylinder # 3 and the gas discharged from the fourth cylinder # 4 merge. Regarding the relationship between the opening time of the third exhaust valve 253 and the opening time of the fourth exhaust valve 254, the opening time of the fourth exhaust valve 254 is late, so the gas discharged from the fourth cylinder # 4 passes through. A throttle portion 634 is formed at the exhaust downstream end of the fourth exhaust pipe 614.

It is at the tenth gathering section 62j that the gas discharged from the second cylinder # 2 and the gas discharged from the fourth cylinder # 4 merge. Regarding the relationship between the opening time of the second exhaust valve 252 and the opening time of the fourth exhaust valve 254, the opening time of the second exhaust valve 252 is late, so the gas discharged from the second cylinder # 2 passes through. A throttle portion 63C is formed at the exhaust downstream end of the exhaust pipe 61C.

According to the third embodiment, when an exhaust manifold having a plurality of gathering portions is adopted by merging in stages, exhaust interference can be effectively reduced.

<Fourth Embodiment>
The supercharged internal combustion engine (hereinafter referred to as “fourth engine”) 10C according to the fourth embodiment of the present invention joins the assembly portion closest to the turbine 70 in the configuration of 4-2-1. Not only is a throttle formed at the downstream end of the exhaust of the exhaust pipe, but the intake valve of the own cylinder is opened in the two cylinders corresponding to the exhaust pipe among the two exhaust pipes that gather at each gathering part. When the latter half of the valve opening period (intake valve opening period) and the first half of the exhaust valve opening period of the other cylinder overlap, the throttle part is at the exhaust downstream end of the exhaust pipe corresponding to the other cylinder. It differs from the first embodiment, the second embodiment, and the third embodiment in that it is formed. Therefore, this difference will be mainly described below.

In the case of the configuration of 4-2-1 shown in FIG. 15, for example, the latter half of the intake valve opening period of the first cylinder # 1 and the first half of the exhaust valve opening period of the fourth cylinder # 4 overlap. There is. When the exhaust gas from the 4th cylinder # 4 wraps around to the 1st cylinder # 1 in the latter half of the valve opening period of the 1st intake valve 241, the exhaust pressure of the 4th cylinder # 4 reaches its peak. On the other hand, since the pressure in the cylinder of the first cylinder # 1 is low, the first exhaust valve 251 may unintentionally open. Therefore, in this case, by forming the throttle portion 634 at the downstream end of the fourth exhaust pipe 614, it is possible to reduce the wraparound of the exhaust gas to the first exhaust pipe 611.

Considering the exhaust gas that wraps around the third cylinder # 3, it is conceivable that the wraparound from the second cylinder # 2 is in addition to the wraparound from the fourth cylinder # 4. The exhaust valve 252 of the second cylinder # 2 starts opening before the intake stroke of the third cylinder # 3 is completed. Therefore, the throttle portion 632 is formed at the exhaust downstream end of the second exhaust pipe 612 through which the exhaust gas of the second cylinder # 2 passes.

Similarly, the exhaust valve 251 of the first cylinder # 1 starts opening before the intake stroke of the fourth cylinder # 4 is completed. Therefore, the throttle portion 631 is formed at the exhaust downstream end of the first exhaust pipe 611 through which the exhaust gas of the first cylinder # 1 passes.

The exhaust valve 253 of the third cylinder # 3 starts opening before the intake stroke of the second cylinder # 2 is completed. Therefore, the throttle portion 633 is formed at the exhaust downstream end of the third exhaust pipe 613 through which the exhaust gas of the third cylinder # 3 passes.

Further, similarly to the second engine 10A, a throttle portion 63A is formed at the exhaust downstream end of the exhaust pipe 61A that joins the 13th gathering portion 62m closest to the turbine 70, and a throttle portion 63A is formed at the exhaust downstream end of the exhaust pipe 61B. 63B is formed. Therefore, in the case of this example, a throttle portion is formed at the exhaust downstream end portion of all the exhaust pipes.

According to the fourth embodiment, not only the wraparound at the 13th gathering portion 62m but also the wraparound at the 11th gathering portion 62k and the 12th gathering portion 62l can be reduced.

<Modification example>
The present invention is not limited to the above embodiment, and various modifications can be adopted within the scope of the present invention.

<First modification>
In the above embodiment, the first engine 10, the second engine 10A, the third engine 10B, and the fourth engine 10C are in-line 4-cylinder internal combustion engines, but the internal combustion engine may be an in-line 5-cylinder internal combustion engine. It may be an in-line 3-cylinder internal combustion engine.

As shown in FIG. 16A, the in-line 5-cylinder internal combustion engine (hereinafter referred to as “fifth engine”) 10D according to the first modification of the present invention has the 14th gathering unit 62o and the first It has 15 gathering parts 62p. The fourth exhaust pipe 614 and the fifth exhaust pipe 615 are assembled in the 14th collecting portion 62o. The first exhaust pipe 611, the second exhaust pipe 612, the third exhaust pipe 613, and the exhaust pipe 61D in which the fourth exhaust pipe 614 and the fifth exhaust pipe 615 are assembled are assembled in the fifteenth collecting portion 62p.

In the case of this configuration, the throttle portion 631 is formed at the exhaust downstream end of the first exhaust pipe 611, the throttle portion 632 is formed at the exhaust downstream end of the second exhaust pipe 612, and the exhaust downstream end of the third exhaust pipe 613 is formed. A throttle portion 633 is formed in the portion, a throttle portion 634 is formed at the exhaust downstream end portion of the fourth exhaust pipe 614, and a throttle portion 635 is formed at the exhaust downstream end portion of the fifth exhaust pipe 615.

The in-line 3-cylinder internal combustion engine (hereinafter, referred to as “sixth engine”) 10E according to the first modification of the present invention has a 16th assembly portion 62q as shown in FIG. 16B. doing.
The first exhaust pipe 611, the second exhaust pipe 612, and the third exhaust pipe 613 are assembled in the 16th gathering portion 62q.

In the case of this configuration, the throttle portion 631 is formed at the exhaust downstream end of the first exhaust pipe 611, the throttle portion 632 is formed at the exhaust downstream end of the second exhaust pipe 612, and the exhaust downstream end of the third exhaust pipe 613 is formed. A throttle portion 633 is formed in the portion.

<Second modification>
In the above embodiment, the turbocharger turbine 70 is a single scroll type having one scroll unit, but the turbocharger turbine may include a twin scroll type turbocharger turbine.

As shown in FIG. 17, the twin-scroll type internal combustion engine with a supercharger (hereinafter, “7th engine”) 10F according to the second modification of the present invention has the 17th assembly section 62r and the 18th assembly section 62s. have.

The first exhaust pipe 611 and the fourth exhaust pipe 614 are assembled in the 17th collecting portion 62r. The second exhaust pipe 612 and the third exhaust pipe 613 are assembled in the 18th collecting portion 62s. In the case of this configuration, the throttle portion 631 is formed at the exhaust downstream end of the first exhaust pipe 611, the throttle portion 632 is formed at the exhaust downstream end of the second exhaust pipe 612, and the exhaust downstream end of the third exhaust pipe 613 is formed. A throttle portion 633 is formed in the portion, and a throttle portion 634 is formed in the exhaust downstream end portion of the fourth exhaust pipe 614.

<Third modification example>
The internal combustion engine according to the third modification includes a bypass pipe 53 and a waist gate valve 54 that bypass the turbine 70 in the exhaust system. In this case, the internal combustion engine according to the third modification is the exhaust upstream of the bypass pipe 53 as shown in (A) “4-1 configuration” (B) “4-2-1 configuration” in FIG. The end may communicate with the exhaust pipe 61 on the exhaust upstream side of the throttle portion 63, and the exhaust downstream end of the bypass pipe 53 may communicate with the downstream exhaust pipe 51.

Assuming that the exhaust upstream end of the bypass pipe 53 communicates with the exhaust downstream side of the throttle portion 63, the exhaust gas whose pressure has been increased by the throttle portion 63 is discharged to the downstream side exhaust pipe 51. That is, according to the configuration of the internal combustion engine according to the third modification, it is possible to prevent the exhaust gas whose pressure has been increased by the throttle portion 63 from being discharged to the downstream exhaust pipe 51.

In the above embodiment, the internal combustion engine is a so-called diesel internal combustion engine, but it may be a gasoline internal combustion engine having an ignition device.

10 ... Internal combustion engine with supercharger (first engine), 20 ... Engine body, 21 ... Main body, 22 ... Cylinder, 23 ... Fuel injection valve, 24 ... Intake valve, 25 ... Exhaust valve, 26 ... Exhaust port, 30 ... Fuel supply system, 40 ... Intake system, 50 ... Exhaust system, 60 ... Exhaust manifold, 61 ... Exhaust pipe, 611 ... First exhaust pipe, 612 ... Second exhaust pipe, 613 ... Third exhaust pipe, 614 ... Fourth Exhaust pipe, 62 ... Meeting part, 63 ... Squeezing part, 70 ... Turbine, 71 ... Turbine housing, 72 ... Turbine wheel, Seo ... Squeezing part cross section, St ... Turbine effective cross section.

Claims (1)

Multiple cylinders with exhaust valves and
An exhaust manifold that merges exhaust gas discharged from the exhaust port at a collecting portion through a plurality of exhaust pipes connected to the exhaust ports of the plurality of cylinders when the exhaust valve is open.
A turbocharger turbine provided on the downstream side of the exhaust of the exhaust manifold and
In an internal combustion engine with a supercharger equipped with
Of the collecting portions, the exhaust downstream end of the exhaust pipe that gathers at the collecting portion closest to the turbocharger turbine is smaller than the cross-sectional area of the exhaust pipe excluding the exhaust downstream end, and the excess The cross-sectional area of 80% or more and less than 100% of the effective cross-sectional area of the supercharger turbine determined by the flow rate of exhaust gas passing through the supercharger turbine, the fluid density, and the pressure ratio of the supercharger turbine. The squeezed part to have was formed,
Internal combustion engine with supercharger.

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