JPH05263671A - Valve timing control device of internal-combustion engine with turbocharger - Google Patents

Valve timing control device of internal-combustion engine with turbocharger

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Publication number
JPH05263671A
JPH05263671A JP6516492A JP6516492A JPH05263671A JP H05263671 A JPH05263671 A JP H05263671A JP 6516492 A JP6516492 A JP 6516492A JP 6516492 A JP6516492 A JP 6516492A JP H05263671 A JPH05263671 A JP H05263671A
Authority
JP
Japan
Prior art keywords
exhaust
valve
turbine
combustion engine
internal combustion
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Pending
Application number
JP6516492A
Other languages
Japanese (ja)
Inventor
Hisashi Oki
久 大木
Original Assignee
Toyota Motor Corp
トヨタ自動車株式会社
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Toyota Motor Corp, トヨタ自動車株式会社 filed Critical Toyota Motor Corp
Priority to JP6516492A priority Critical patent/JPH05263671A/en
Publication of JPH05263671A publication Critical patent/JPH05263671A/en
Pending legal-status Critical Current

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Classifications

    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02BINTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B37/00Engines characterised by provision of pumps driven at least for part of the time by exhaust
    • F02B37/02Gas passages between engine outlet and pump drive, e.g. reservoirs
    • YGENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02TCLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00Road transport of goods or passengers
    • Y02T10/10Internal combustion engine [ICE] based vehicles
    • Y02T10/14Technologies for the improvement of mechanical efficiency of a conventional ICE
    • Y02T10/144Non naturally aspirated engines, e.g. turbocharging, supercharging

Abstract

(57) [Abstract] [Purpose] To reduce the pumping loss of the internal combustion engine while maintaining the required rotation of the turbine even when the exhaust pressure becomes excessive. [Structure] A pair of exhaust ports 8 and 9 are provided in each cylinder # 1 to # 4 of an engine 2, and each exhaust port 8 and 9 is connected to each exhaust passage 1.
3 and 15 are separately communicated with each other, and the exhaust passage 13 is connected to the turbine 1
The exhaust passage 15 is connected to the inlet side of the turbine 7 and to the outlet side of the turbine 17. Exhaust valves 10 and 11 driven by hydraulic actuators are provided in the exhaust ports 8 and 9, respectively, to open the exhaust valve 11 later than the exhaust valve 10 in the exhaust stroke, and to exhaust the exhaust valve 1 during high-speed operation.
The hydraulic actuator is controlled so as to increase the valve opening period of 1. Therefore, only the energy of the exhaust pressure at the time of blowdown effective for the rotation of the turbine 17 is utilized through the exhaust passage 13, and unnecessary exhaust pressure after the blowdown is removed through the exhaust passage 15 without resistance.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an internal combustion engine equipped with a turbocharger, and more particularly to a turbocharger in which the opening / closing timing of an exhaust valve in the internal combustion engine is controlled in accordance with the operation of the turbocharger. The present invention relates to a valve timing control device for an internal combustion engine with a supercharger.

[0002]

2. Description of the Related Art Conventionally, in an internal combustion engine equipped with a turbocharger, when the exhaust port is opened in the low speed operation region, the energy of the pressure of exhaust gas (exhaust pressure) is insufficient and the turbocharger There was a case where the turbine rotation of was insufficient. In such a case, when the internal combustion engine tries to shift to the high speed operation region, there is a problem that the followability of the turbocharger is deteriorated and the acceleration performance is impaired.

A technique for dealing with the above problem is disclosed in, for example, Japanese Patent Laid-Open No. 1-277654. In the technique of this publication, the engine is provided with two independent exhaust ports, and one of the first exhaust ports is connected to the turbine downstream side of the turbocharger via the first exhaust passage. The remaining second exhaust port communicates with the turbine upstream side via the second exhaust passage. In addition, a branch passage that branches into the second exhaust passage is provided in the middle of the first exhaust passage, and a waste gate valve (WGV) for opening and closing is provided in the branch passage. Then, during deceleration operation of the vehicle or when the clutch is disengaged, the first exhaust port is opened a little earlier than the second exhaust port by opening the branch passage with the waste gate valve and controlling the opening / closing of the exhaust valve. Let As a result, a large exhaust pressure is obtained at high temperature and high pressure during blowdown, and the large energy of the exhaust pressure is input to the turbine through the first exhaust port, the first exhaust passage and the branch passage, and the turbocharger is supplied. Is maintained at the required number of revolutions.

However, in the above technique, during high-speed operation or the like, the exhaust pressure in the first exhaust passage becomes excessive and there is a risk of over-rotation or overheating of the turbine. Therefore, when the exhaust pressure becomes excessively high, the waste passage is closed by a waste gate valve, and the exhaust pressure in the first exhaust passage is directly released to the downstream side of the turbine without being input to the turbine. ing.

[0005]

However, in the prior art disclosed in the above publication, when the exhaust pressure becomes excessive, the waste gate valve is closed from the beginning of the exhaust stroke, so that part of the exhaust pressure is reduced from the beginning. Was dumped directly to the downstream side of. Therefore, in consideration of the exhaust pressure discarded through the exhaust stroke, as shown in FIG. 10A, when the wastegate valve (WGV) is closed, at the time of blowdown that can most effectively affect the rotation of the turbine. Exhaust pressure was just largely discarded, and the amount of exhaust pressure discarded after blowdown, which is unnecessary for turbine rotation, did not become so large. Therefore, FIG.
As shown in (b), the energy at the time of blowdown that should be input to the turbine and the average energy were relatively low, and it was not possible to sufficiently secure the energy effective for the rotation of the turbine. Moreover, as shown in FIG. 10 (a), the average back pressure of the internal combustion engine does not become too low, and it becomes a resistance in the exhaust stroke to increase the pressure loss and cause the pumping loss of the engine to increase. It was

The present invention has been made in view of the above-mentioned circumstances, and an object thereof is to reduce pumping loss of an internal combustion engine while maintaining necessary rotation of a turbine even when exhaust pressure becomes excessive. Another object of the present invention is to provide a valve timing control device for an internal combustion engine with a turbocharger capable of achieving the above.

[0007]

To achieve the above object, in the present invention, as shown in FIG. 1, a compressor M is rotated by rotating a turbine M1 by exhaust gas.
2, an internal combustion engine M4 including a turbocharger M3 that supercharges intake air, a first exhaust port M6 and a first exhaust port M6 that are independently provided in a cylinder M5 of the internal combustion engine M4. Two exhaust ports M7, one end communicates with the first exhaust port M6, the other end communicates with the inlet side of the turbine M1, and a first exhaust passage M8, and one end communicates with the second exhaust port M7. A second exhaust passage M9 having the other end communicated with the outlet side of the turbine M1, a first exhaust valve M10 for opening / closing the first exhaust port M6, and a second exhaust valve M7 for opening / closing the second exhaust port M7. Second exhaust valve M11, first valve driving means M12 for opening and closing the first exhaust valve M10, second valve driving means M13 for opening and closing the second exhaust valve M11, and internal combustion Includes engine M4 speed Based on the operating state detecting means M14 for detecting the operating state, and the detection result of the operating state detecting means M14, the second exhaust valve M11 is opened later than the first exhaust valve M10 in the exhaust stroke of the internal combustion engine M4. At the same time, the valve control means M1 for driving and controlling the first and second valve drive means M12, M13 so as to increase the valve opening period of the second exhaust valve M11 during high speed operation.
5 and.

[0008]

According to the above construction, as shown in FIG. 1, the first valve drive means M is provided in the exhaust stroke of the internal combustion engine M4.
12 is driven and the first exhaust valve M10 is opened, so that the exhaust gas emitted from the cylinder M5 passes through the first exhaust port M6 and the first exhaust passage M8 to the turbine M
1 is supplied to the inlet side of the turbine 1 to rotate the turbine M1. By this action, the compressor M2 is driven and the intake air is supercharged. On the other hand, in the exhaust stroke of the internal combustion engine M4, the second valve drive means M13 is driven to open the second exhaust valve M11, so that the exhaust gas emitted from the cylinder M5 is discharged to the second exhaust port M7 and the exhaust port M7. It is escaped to the outlet side of the turbine M1 through the second exhaust passage M9.

Then, during operation of the internal combustion engine M4, the valve control means M is operated based on the detection result of the operating state detection means M14.
By operating 15 the first and second valve drive means M12, M13 are drive-controlled, and the second valve drive means M12, M13 are controlled in the exhaust stroke.
The exhaust valve M11 is opened later than the first exhaust valve M10, and the opening period of the second exhaust valve M13 is increased during high speed operation.

Therefore, since the second exhaust valve M11 is opened later than the first exhaust valve M10, only the exhaust pressure at the time of blowdown most effective for the rotation of the turbine M1 is the first exhaust passage M8. First through the turbine M1
Is supplied to. Exhaust pressure unnecessary for the rotation of the turbine M1 is delayed from the time of blowdown, and the second exhaust passage M9
Through to the outlet side of the turbine M1. Therefore, only the energy of the exhaust pressure effective for the rotation of the turbine M1 is used, and the unnecessary exhaust pressure is removed without resistance. Moreover, during high-speed operation, the second exhaust valve M13
Since the valve opening period is increased, a portion of the exhaust pressure that becomes excessive during high-speed operation that is unnecessary for the rotation of the turbine M1 sufficiently escapes to the outlet side of the turbine M1 through the second exhaust passage M9. The unnecessary exhaust pressure increased by high speed operation is removed without resistance.

[0011]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment embodying a valve timing control device for an internal combustion engine with a turbocharger according to the present invention will be described in detail below with reference to FIGS.

FIG. 2 is a schematic diagram showing an in-line four-cylinder type engine 2 as an internal combustion engine equipped with a turbocharger 1 and its intake system and exhaust system. An intake port 3 that is bifurcated is formed in each of four cylinders # 1, # 2, # 3, and # 4 of the engine 2. Also, each cylinder # 1
A pair of intake valves 4 and 5 for opening and closing are provided for each of the cylinders # 1 to # 4 at both open ends of each intake port 3 that branches to open to # 4. Further, the inlet side of each intake port 3 is connected to a common intake passage 7 via an intake manifold 6.

On the other hand, each of the cylinders # 1 to # 4 has a pair of first exhaust ports 8 having the same port diameter independently of each other.
And a second exhaust port 9 are formed respectively.
Further, in each of the cylinders # 1 to # 4, a first exhaust valve 10 for opening and closing is provided at the opening end of the first exhaust port 8. Similarly, a second exhaust valve 11 for opening and closing is provided at the opening end of the second exhaust port 9. Further, the outlet side of each first exhaust port 8 is connected to a common first exhaust passage 13 via a first exhaust manifold 12, respectively. Similarly, the outlet side of each second exhaust port 9 is connected to the second exhaust manifold 14 respectively.
Through the common second exhaust passage 15.

The turbocharger 1 is provided with a copresser 16 and a turbine 17 which are integrally rotatable, and its compressor 1
6 is provided in the middle of the intake passage 7. Further, the turbine 17 is provided corresponding to one end of the first exhaust passage 13. That is, one end of the first exhaust passage 13 is connected to the turbine 1
It is connected to the entrance side of 7. Further, the outlet side of the turbine 17 is connected to another third exhaust passage 18. Further, one end of the second exhaust passage 15 communicates with the outlet side of the turbine 17, that is, the third exhaust passage 18.

Therefore, exhaust gas is guided from the cylinders # 1 to # 4 to the first exhaust passage 13 through the first exhaust ports 8 and the first exhaust manifolds 12, so that the exhaust pressure of the turbine 17 is reduced. It is supplied to the entrance side. Then, the turbine 17 is rotated by the energy of the exhaust pressure, so that the compressor 16 is integrally driven to rotate and the intake air in the intake passage 7 is supercharged, so that each cylinder # passes through each intake manifold 6 and each intake port 3. 1
To # 4. Further, the exhaust pressure is supplied to the turbine 17 by guiding the exhaust gas from each of the cylinders # 1 to # 4 to the second exhaust passage 15 through each of the second exhaust ports 9 and each of the second exhaust manifolds 14. Without being exhausted to the third exhaust passage 18.

Next, a valve operating hydraulic drive system for opening and closing the first and second exhaust valves 10 and 11 in each of the cylinders # 1 to # 4 will be described. In this embodiment, the valve actuating hydraulic drive system is designed so that the first and second exhaust valves 10 and 11 are independently opened and closed for each of the cylinders # 1 to # 4. Separate hydraulic circuits corresponding to # 1 to # 4 are provided. In this embodiment, assuming that the hydraulic circuits have a common configuration, one cylinder # 1 will be represented below, and a valve operating hydraulic drive system including the hydraulic circuit will be described.

In this embodiment, each cylinder # 1 to # is
The intake valves 4, 5 of No. 4 are opened and closed by the valve operating hydraulic drive device like the exhaust valves 10, 11, but the description thereof is omitted here.

FIG. 3 is a schematic diagram showing the first and second exhaust valves 10 and 11 in cylinder # 1 and the hydraulic drive system for operating the valves. The cylinder block 19 of the engine 2 is provided with first and second exhaust valves 10 and 11 corresponding to the first and second exhaust ports 8 and 9, respectively. Valve stem 10a, 1 of each exhaust valve 10, 11
1a extends upward through the cylinder block 19 and is attached to the cylinder block 19 so as to be vertically movable. Cylinder block 19 and cylinder head 2
Between 0 and 0, retainers 21 and 22 are fixed to the valve stems 10a and 11a near the tips. Further, between the retainers 21 and 22 and the cylinder block 19, a valve spring 23 that urges the exhaust valves 10 and 11 upward, that is, urges the exhaust valves 10 and 11 in a closing direction. 24 are respectively interposed.

In order to open and close the exhaust valves 10 and 11, the cylinder head 20 of the engine 2 is formed with a pair of hydraulic chambers 25 and 26 corresponding to the valve stems 10a and 11a, respectively. Each hydraulic chamber 25, 26
Plungers 27 and 28 are respectively assembled to be movable up and down, and the hydraulic chambers 25 and 26 and the plungers 27 and 28 serve as a first valve drive means.
And a second hydraulic actuator 30 as a second valve drive means. The lower end of each plunger 27, 28 is engaged with the upper end of each valve stem 10a, 11a. And
In principle, the exhaust valves 10 and 11 are moved up and down to be opened and closed by controlling the operating oil pressure of the oil pressure chambers 25 and 26 to move the plungers 27 and 28 up and down. ..

A hydraulic circuit for operating the hydraulic actuators 29 and 30 will be described. This hydraulic circuit includes a hydraulic pump 31 driven by the engine 2, and the suction side of the hydraulic pump 31 is connected to a hydraulic tank 33 via a strainer 32. Further, the discharge side of the hydraulic pump 31 is branched into two hydraulic pipes 34 and 35, and in the middle of one hydraulic pipe 34, a high pressure accumulator 36 and a first control valve 37 are arranged in order from the hydraulic pump 31. ,
A second control valve 38 and a low pressure accumulator 39 are provided in series. A return line 40 extending from the outlet side of the low pressure accumulator 39 is connected to the hydraulic tank 33. Then, the first control valve 37 and the second control valve 3
8 is connected to the hydraulic chamber 25 of the first hydraulic actuator 29 at one end thereof. Similarly, in the middle of the other hydraulic line 35, the hydraulic pump 31
High-pressure accumulator 42, third control valve 4
The third and fourth control valves 44 and the low pressure accumulator 45 are provided in series. In addition, the low pressure accumulator 45
A return pipe line 46 extending from the outlet side of the is connected to the hydraulic tank 33 via the return pipe line 40. Then, one end of a hydraulic line 47 extending from between the third control valve 43 and the fourth control valve 44 is connected to the hydraulic chamber 26 of the second hydraulic actuator 30.

Here, the first to fourth control valves 37, 3
8, 43, 44 are respectively provided with actuators 37a, 38a, 43a, 44a for opening and closing them. Each actuator 37a, 38a, 43a, 4
Reference numeral 4a is composed of a high-speed response type laminated piezoelectric element or the like, and is driven by energization. Then, by selectively energizing the actuators 37a, 38a, 43a, 44a, the control valves 37, 38, 43,
44 is selectively opened and closed.

Therefore, by appropriately controlling the opening / closing timing between the first control valve 37 and the second control valve 38, the first hydraulic actuator 29 is driven, and the first exhaust gas corresponding thereto is driven. The opening / closing timing of the valve 10 and its lift amount (opening / closing amount) are controlled. Similarly, by appropriately controlling the opening / closing timing between the third control valve 43 and the fourth control valve 44, the second hydraulic actuator 30 is driven, and the second exhaust valve 11 corresponding thereto is driven. The opening / closing timing and the opening / closing amount thereof are controlled.

Each of the actuators 37a, 38 described above
In this embodiment, a valve controller 48 as a valve control means is provided in order to control the energization timing of a, 43a and 44a. The valve controller 48 is mainly composed of a microcomputer, and the controller 48 is connected with a rotation speed sensor 49 as an operation state detecting means for detecting a rotation speed (engine rotation speed) NE of the engine 2. Then, the valve controller 48, based on the detection signal of the rotation speed sensor 49, discharges the exhaust valves 10, 1 according to the operating state at that time.
The opening / closing timing and the opening / closing amount of No. 1 are calculated, and each actuator 37a, 38a, 43a,
A control signal for energization is output to 44a.

In the valve controller 48, the second exhaust valve 11 is connected to the first exhaust valve 1 in the exhaust stroke of the engine 2 based on the detection result of the rotation speed sensor 49.
The first and second control valves 37, 38, 43, 44 are opened and closed so as to open the valve later than 0 and increase the opening period of the second exhaust valve 11 during high speed operation.
The hydraulic actuators 29 and 30 are controlled to be driven.

That is, when the valve controller 48 determines that the engine 2 is operating at a low speed, as shown in FIG. 4, the first and second exhaust valves 10 are opened and closed at a predetermined timing in the exhaust stroke. Control valve 37,
38 is opened and closed.

The valve controller 48 is the engine 2
When it is determined that the vehicle is operating at medium speed, as shown in FIG. 5, the first exhaust valve 10 is opened and closed at a predetermined timing in the exhaust stroke and at a slightly shorter opening period than during low speed operation. And the second control valves 37, 38 are opened and closed. In addition, the second exhaust valve 11 is replaced with the first exhaust valve 1
The third and fourth control valves 43, 44 are driven to open and close so that the valve opening time is slower than 0, the valve opening period is shorter, and the lift amount is smaller than that of the first exhaust valve 10. There is.

Further, when the valve controller 48 determines that the engine 2 is operating at high speed, as shown in FIG.
In the exhaust stroke, the first and second control valves 37 and 38 are driven to open and close in order to open and close the first exhaust valve 10 at a predetermined timing and in a valve opening period shorter than that during medium speed operation. In addition, the second exhaust valve 11 is replaced by the first exhaust valve 10
The third and fourth control valves 43 and 44 are opened and closed to open the valve so that the valve opening period is slower and longer than that at the time of medium speed operation, and is equal to the lift amount of the first exhaust valve 10. It is designed to let you.

In each of FIGS. 4 to 6 described above, the broken line shows the lift amount and the opening / closing timing of the intake valves 4 and 5, and the opening / closing timing and the opening / closing amount are always constant.

Next, the operation of the valve timing control system for the internal combustion engine with a turbocharger including the valve drive hydraulic drive system constructed as described above will be described with reference to FIGS. Figure 7
Represents the change in the lift amount of each exhaust valve 10 and 11 and the change in exhaust pressure in the first and second exhaust passages 13 and 15 in the present embodiment at low speed operation, medium speed operation and high speed operation respectively. It is a time chart shown in comparison with technology. Further, FIG. 8 also shows P in the present embodiment at low speed operation, medium speed operation, and high speed operation, respectively.
7 is a graph showing -V line and pumping loss PL in comparison with the related art.

As shown in FIG. 7, during low speed operation of the engine 2, the two exhaust valves 10 and 11 are used in the exhaust stroke.
Among them, only the first exhaust valve 10 is opened at a predetermined timing. Then, all of the exhaust gas emitted from each of the cylinders # 1 to # 4 is supplied to the turbine 17 through the first exhaust passage 13.

Therefore, all the energy of the exhaust pressure acts on the turbine 17 even during the blowdown, the turbine 17 is driven as much as possible even in the low speed range, and the turbocharger 1 transfers it to the engine 2. Will be supercharged. At this time, the pumping loss PL is also relatively small as shown in FIG. 8 (a).

After that, as shown in FIG. 7, when the engine 2 operates at a medium speed and the supercharging pressure by the turbocharger 1 exceeds a predetermined level, the first exhaust valve 10 is opened in the exhaust stroke. In addition to the valve, the second exhaust valve 11 is opened with a small opening degree for a relatively short period of time after that. Therefore, when the first exhaust valve 10 is opened first, a higher exhaust pressure is taken out at the time of blowdown, as compared with the case where both the exhaust valves 10 and 11 are opened simultaneously. Then, out of all the exhaust gas emitted from each of the cylinders # 1 to # 4, the exhaust gas in the first half of the exhaust stroke including the blowdown is supplied to the turbine 17 through the first exhaust passage 13. Further, the exhaust gas in the latter half of the exhaust stroke is allowed to escape through the second exhaust passage 15 to the third exhaust passage 18 as it is.

Therefore, the energy of the exhaust pressure including the exhaust pressure at the time of effective blowdown acts on the turbine 17, so that the turbine 17 is efficiently driven in the medium speed range, so that the engine by the turbocharger 1 is driven. 2 is effectively supercharged. In addition, the relatively low exhaust pressure in the latter half of the exhaust stroke close to the intake stroke is discarded as it is without being supplied to the turbine 17. Therefore, more energy than necessary does not act on the turbine 17. Further, since the exhaust pressure in the latter half of the exhaust stroke is released and the exhaust pressure becomes low in the portion close to the intake stroke, as shown in FIG.
As shown in (b), the pressure loss of the engine 2 is reduced and the pumping loss PL can be reduced.

Thereafter, as shown in FIG. 7, when the engine 2 is operating at high speed and the supercharging pressure by the turbocharger 1 reaches a sufficient level, the first exhaust valve 10 is opened in the exhaust stroke. The period is further shortened, and a little later than that, the second exhaust valve 11 is opened with a large opening for a relatively long period. For this reason, a higher exhaust pressure is taken out at the time of blowdown, as compared with the case where both exhaust valves 10 and 11 are opened simultaneously. Then, of all the exhaust gas emitted from each of the cylinders # 1 to # 4, only the high exhaust pressure at the time of blowdown is supplied to the turbine 17 through the first exhaust passage 13. Further, the exhaust gas after the blowdown is released to the third exhaust passage 18 as it is through the second exhaust passage 15. That is, as shown in FIG.
As shown in (b), of the total exhaust pressure, the exhaust pressure during blowdown is applied to the first exhaust passage 13, and the exhaust pressure after blowdown is applied to the second exhaust passage 15.
Then, only the energy of the exhaust pressure at the time of blowdown is input to the turbine 17.

Therefore, since only the energy of the high exhaust pressure at the time of blowdown acts on the turbine 17, more energy than necessary is not applied to the turbine 17 in the high speed range, and the turbine 17 is driven without excess or deficiency. The turbocharger 1 effectively supercharges the engine 2. In addition, the relatively low exhaust pressure after the blowdown is discarded without being supplied to the turbine 17. Therefore, the exhaust pressure is not applied to the turbine 17 more than necessary. Further, since the exhaust pressure after the blowdown is discarded without resistance, the pressure loss of the engine 2 is reduced and the pumping loss PL can be significantly reduced, as shown in FIG. 8 (c).

As described above, according to this embodiment, when the exhaust pressure becomes excessively high such as during the medium / high speed operation of the engine 2, the second exhaust valve 11 has the first exhaust valve.
Of the exhaust valve 10 is opened later than the exhaust valve 10 of FIG.
First, the gas is supplied to the turbine 17 through the exhaust passage 13. Further, the exhaust pressure unnecessary for the rotation of the turbine 17 after the blowdown is released to the outlet side of the turbine 17 through the second exhaust passage 15.

Therefore, only the energy of the exhaust pressure at the time of blowdown effective for the rotation of the turbine 17 is utilized,
Unnecessary exhaust pressure is removed without resistance. Moreover, the faster the engine 2 operates, the second exhaust valve 1
Since the valve opening period of No. 1 is increased, the portion of the exhaust pressure that becomes excessive during high-speed operation, which is unnecessary for the rotation of the turbine 17, is sufficiently passed through the second exhaust passage 15 to the outlet side of the turbine 17. Unnecessary exhaust pressure, which is released and increased by high speed operation, is removed without resistance.

Therefore, as shown in FIG. 9B, the average back pressure of the engine 2 can be made relatively small, and the average energy to be supplied to the turbine 17 can be made relatively high at the same time. .. As a result, even when the exhaust pressure becomes excessively high, it is possible to maintain the rotation required for the turbine 17, and it is possible to effectively reduce the pumping loss PL of the engine 2 while maintaining the rotation.

The present invention is not limited to the above-described embodiment, but can be implemented as follows with a part of the structure appropriately changed without departing from the spirit of the invention. (1) In the above embodiment, the first and second exhaust valves 1
Although the hydraulic actuators 29 and 30 are provided as the first and second valve driving means for driving the opening and closing of the valves 0 and 11, the valve operating device centering on the cam shaft is provided as the first and second valve driving means. You can also In that case, by providing a variable valve timing mechanism for each exhaust valve, it becomes possible to individually change the opening / closing timing of each exhaust valve.

(2) In the above embodiment, each exhaust valve 1
Hydraulic actuator 2 for opening and closing 0, 11
Although the intake valves 4 and 5 are driven to open and close by a hydraulic actuator equivalent to that of the exhaust valves 1 and 9,
It is also possible to drive only the valves 0 and 11 by the hydraulic actuators 29 and 30, and to drive the intake valves 4 and 5 only by the valve operating device centering on the camshaft.

(3) In the above embodiment, the inline 4-cylinder type engine 2 is embodied, but any internal combustion engine having a turbocharger may be used regardless of the in-line multi-cylinder type or the V-type multi-cylinder type. Can also be embodied in this type.

(4) In the above embodiment, each exhaust port 8,
Although the port diameters of 9 are the same, those diameters may be different.

[0043]

As described above in detail, according to the present invention, in an internal combustion engine including a turbocharger in which a turbine is rotated by exhaust gas to drive a compressor to supercharge intake air. , The first and second exhaust ports, which are independent of each other in the cylinder, communicate with the first and second exhaust passages, respectively, and one end of the first exhaust passage is connected to an inlet side of the turbine and one end of the second exhaust passage is connected to the first exhaust passage. Are connected to the outlet side of the turbine, and first and second exhaust valves for opening and closing each exhaust port are provided, and the second exhaust valve is opened later than the first exhaust valve in the exhaust stroke of the internal combustion engine. In addition, the opening period of the second exhaust valve is increased during high speed operation. Therefore, only the energy of the exhaust pressure at the time of blowdown effective for the rotation of the turbine is utilized through the first exhaust passage, and the unnecessary exhaust pressure after the blowdown is removed without resistance through the second exhaust passage. Therefore, even if the exhaust pressure becomes excessive, the pumping loss of the internal combustion engine can be effectively reduced while maintaining the rotation required for the turbine.

[Brief description of drawings]

FIG. 1 is a conceptual configuration diagram illustrating a basic conceptual configuration of the present invention.

FIG. 2 is a schematic diagram showing an in-line four-cylinder type engine as an internal combustion engine including a turbocharger and an intake system and an exhaust system thereof in one embodiment embodying the present invention.

FIG. 3 illustrates a first and second cylinder in one embodiment.
FIG. 3 is a schematic configuration diagram showing the exhaust valve of FIG.

FIG. 4 is a time chart illustrating changes in the lift amounts of the first exhaust valve and the intake valve during low speed operation in one embodiment.

FIG. 5 is a time chart for explaining changes in lift amounts of the first exhaust valve, the second exhaust valve, and the intake valve during medium speed operation in one embodiment.

FIG. 6 is a time chart for explaining changes in lift amounts of the first exhaust valve, the second exhaust valve, and the intake valve during high-speed operation in one embodiment.

FIG. 7 is a lift amount of each exhaust valve at a low speed operation, a medium speed operation, and a high speed operation in one embodiment, first and second
7 is a time chart showing changes in exhaust pressure in the exhaust passage of FIG.

FIG. 8 is a graph showing the P-V line and pumping loss in low speed operation, medium speed operation and high speed operation in one example in comparison with the related art.

FIG. 9A is a time chart showing changes in exhaust pressure in a first exhaust passage and a second exhaust passage in comparison with a conventional technique in one embodiment, and FIG. 9B is a turbine input energy; 4 is a time chart showing changes in

FIG. 10A is a time chart showing a comparison of exhaust pressures when a waste gate valve is opened and when the waste gate valve is closed, and FIG. 10B is a time chart showing a change in input energy of a turbine. Is.

[Explanation of symbols]

1 ... Turbocharger, 2 ... Engine as internal combustion engine, 8
... first exhaust port, 9 ... second exhaust port, 10 ... first exhaust valve, 11 ... second exhaust valve, 13 ... first
Exhaust passage, 15 ... Second exhaust passage, 29 ... First hydraulic actuator constituting first valve driving means, 30
... A second hydraulic actuator that constitutes a second valve drive means, 48 ... A valve controller as valve control means.

Claims (1)

[Claims]
1. An internal combustion engine including a turbocharger configured to drive a compressor by rotating a turbine with exhaust gas to supercharge intake air, and a cylinder of the internal combustion engine, which is provided independently of each other. A first exhaust port, a second exhaust port, a first exhaust passage having one end communicating with the first exhaust port and the other end communicating with an inlet side of the turbine; A second exhaust passage that communicates with the second exhaust port and the other end communicates with the turbine outlet side; a first exhaust valve that opens and closes the first exhaust port; and a second exhaust port A second exhaust valve that opens and closes the first exhaust valve, a first valve drive unit that opens and closes the first exhaust valve, and a second valve drive unit that opens and closes the second exhaust valve, The internal combustion engine Operating condition detecting means for detecting operating conditions including the number, based on a detection result of said operating condition detecting means, the internal combustion engine the first and second exhaust valve in the exhaust stroke of the
Valve control means for driving and controlling the first and second valve driving means so as to open the valve later than the exhaust valve and increase the valve opening period of the second exhaust valve during high speed operation. A valve timing control device for an internal combustion engine with a turbocharger, characterized in that:
JP6516492A 1992-03-23 1992-03-23 Valve timing control device of internal-combustion engine with turbocharger Pending JPH05263671A (en)

Priority Applications (1)

Application Number Priority Date Filing Date Title
JP6516492A JPH05263671A (en) 1992-03-23 1992-03-23 Valve timing control device of internal-combustion engine with turbocharger

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
JP6516492A JPH05263671A (en) 1992-03-23 1992-03-23 Valve timing control device of internal-combustion engine with turbocharger

Publications (1)

Publication Number Publication Date
JPH05263671A true JPH05263671A (en) 1993-10-12

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Cited By (13)

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WO2001036797A1 (en) * 1999-11-15 2001-05-25 Fev Motorentechnik Gmbh Method for operating a piston-type combustion engine, with a controllable turbocharger and a piston-type internal combustion engine for carrying out said method
WO2001081745A1 (en) * 2000-04-20 2001-11-01 Fev Motorentechnik Gmbh Method for influencing the mixture formation and charging movement in a cylinder of a piston internal combustion engine with externally applied ignition
US6438956B1 (en) * 1998-10-05 2002-08-27 Saab Automobile Ab Method of operating an internal-combustion engine, and internal-combustion engine
US6883319B2 (en) * 2001-03-30 2005-04-26 Saab Automobile Ab Method for controlling the charging pressure at a turbocharged combustion engine, and a corresponding combustion engine
WO2005068803A1 (en) * 2004-01-14 2005-07-28 Lotus Cars Limited A turbocharged internal combustion engine
WO2008016052A1 (en) * 2006-08-04 2008-02-07 Toyota Jidosha Kabushiki Kaisha Vehicle control device
JP2008267316A (en) * 2007-04-23 2008-11-06 Toyota Motor Corp Start control device for internal combustion engine with supercharger
JP2009085022A (en) * 2007-09-27 2009-04-23 Toyota Motor Corp Control device of vehicle
JP2009293537A (en) * 2008-06-06 2009-12-17 Toyota Motor Corp Control device for internal combustion engine
DE112008002126T5 (en) 2007-08-13 2010-07-15 Toyota Jidosha Kabushiki Kaisha, Toyota-shi Control device for an internal combustion engine equipped with a turbocharger
US8857177B2 (en) 2007-09-05 2014-10-14 Mahle International Gmbh Piston engine
US9086011B2 (en) 2010-01-22 2015-07-21 Borgwarner Inc. Directly communicated turbocharger
JP2017180360A (en) * 2016-03-31 2017-10-05 マツダ株式会社 Control device of engine

Cited By (18)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US6438956B1 (en) * 1998-10-05 2002-08-27 Saab Automobile Ab Method of operating an internal-combustion engine, and internal-combustion engine
WO2001036797A1 (en) * 1999-11-15 2001-05-25 Fev Motorentechnik Gmbh Method for operating a piston-type combustion engine, with a controllable turbocharger and a piston-type internal combustion engine for carrying out said method
WO2001081745A1 (en) * 2000-04-20 2001-11-01 Fev Motorentechnik Gmbh Method for influencing the mixture formation and charging movement in a cylinder of a piston internal combustion engine with externally applied ignition
US6701887B2 (en) 2000-04-20 2004-03-09 Fev Motorentechnik Gmbh Method for influencing the mixture formation and charging movement in a cylinder of a piston internal combustion engine with externally applied ignition
US6883319B2 (en) * 2001-03-30 2005-04-26 Saab Automobile Ab Method for controlling the charging pressure at a turbocharged combustion engine, and a corresponding combustion engine
DE10297129B4 (en) * 2001-03-30 2010-10-21 Saab Automobile Ab Method for controlling the boost pressure of a turbocharged internal combustion engine and associated internal combustion engine
WO2005068803A1 (en) * 2004-01-14 2005-07-28 Lotus Cars Limited A turbocharged internal combustion engine
WO2008016052A1 (en) * 2006-08-04 2008-02-07 Toyota Jidosha Kabushiki Kaisha Vehicle control device
JP2008039030A (en) * 2006-08-04 2008-02-21 Toyota Motor Corp Vehicle controller
JP2008267316A (en) * 2007-04-23 2008-11-06 Toyota Motor Corp Start control device for internal combustion engine with supercharger
DE112008002126T5 (en) 2007-08-13 2010-07-15 Toyota Jidosha Kabushiki Kaisha, Toyota-shi Control device for an internal combustion engine equipped with a turbocharger
US20110219767A1 (en) * 2007-08-13 2011-09-15 Toyota Jidosha Kabushiki Kaisha Control device for internal combustion engine equipped with turbocharger
US8857177B2 (en) 2007-09-05 2014-10-14 Mahle International Gmbh Piston engine
JP2009085022A (en) * 2007-09-27 2009-04-23 Toyota Motor Corp Control device of vehicle
JP2009293537A (en) * 2008-06-06 2009-12-17 Toyota Motor Corp Control device for internal combustion engine
US9086011B2 (en) 2010-01-22 2015-07-21 Borgwarner Inc. Directly communicated turbocharger
US10215084B2 (en) 2010-01-22 2019-02-26 Borgwarner Inc. Directly communicated turbocharger
JP2017180360A (en) * 2016-03-31 2017-10-05 マツダ株式会社 Control device of engine

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