JP2014234821A - Ultra-high-efficiency internal combustion engine with centralized combustion chamber - Google Patents

Abstract

PROBLEM TO BE SOLVED: To achieve ultra-high-efficiency for thermal efficiency by adopting ultra-rapid combustion using a centralized combustion chamber and a multi-point plug, and a mirror cycle, and further to achieve ultra-low NOx by significantly delaying the ignition timing of a spark gap through an ultra-high-efficiency combustion method.SOLUTION: Around a combustion chamber 1, a squish part S including a fine gap between a piston top surface at a top dead center position and a cylinder head wall surface is formed and defined as the centralized combustion chamber. Two or more spark gaps oppose the combustion chamber 1. The ignition timing of the spark gap is controlled to a compression top dead center or after that in a high-load area. Further, in a low-load area, close timing of suction valves 3 and 3' is changed by a valve opening/closing timing control device 15 which varies opening/closing timing of the suction valves. Therefore, a mirror cycle is implemented while reducing an effective suction step.

Description

The present invention adopts a centralized combustion chamber, multi-point ignition system, high compression ratio, high expansion ratio, Miller cycle, etc. to achieve ultra-high efficiency, greatly reducing CO ₂ and reducing NOx generation to a minimum level The present invention relates to a high-efficiency four-cycle internal combustion engine.

The problems imposed on internal combustion engines, especially automobile internal combustion engines, are great, and there is an urgent need for ultra-high efficiency and ultra-low pollution to cope with global warming (CO ₂ reduction) and air pollution. In the spark ignition engine, since the output is controlled by restricting the intake air, the loss ratio is large as the pump loss, and the compression ratio is a low compression ratio of 12 or less, so the heat efficiency is low.
In addition, since ignition and combustion start by the spark plug is performed before the top dead center, negative work occurs, the maximum combustion pressure and temperature increase, friction loss and cooling loss increase, thermal efficiency deteriorates, A large amount of NOx is generated. On the other hand, although the diesel engine is highly efficient, the friction loss is large and there is room for improvement. The spark ignition engine is the same, but since the compression ratio and the expansion ratio are equal, the exhaust energy is excessive and the exhaust loss is large. As an exhaust purification measure, a common rail fuel injection device, diesel particulate filter (DPF), lean NOx catalyst, etc. reaching 2000 atmospheres are required, and these are very expensive.

Problems to be solved by the invention

The object of the present invention is to achieve high thermal efficiency. For this reason, a concentrated combustion chamber (compact combustion chamber) and a multi-point ignition method are adopted to shorten the combustion period to the limit and to perform ultra-rapid combustion (1/3 of the conventional). In this way, the ignition timing by the spark plug can be greatly delayed (the ignition timing is near the top dead center and the top dead center or higher in the high load range), so there is no negative work before the top dead center, Maximum combustion pressure and temperature are reduced, and friction loss and cooling loss are reduced. Furthermore, a high compression ratio of 15 to 18 is adopted, and a mirror cycle having a larger expansion ratio than compression is adopted. Since the Miller cycle is a cycle that shortens the effective suction stroke, the pump loss is reduced. Second, as described above, the ultra-rapid combustion method is used to achieve a very low NOx emission by greatly delaying the ignition timing (in some cases, the top dead center or later). This eliminates the need for an expensive lean NOx catalyst and premix combustion, so that a DPF such as a diesel engine is also unnecessary.

Means to solve the problem

In order to achieve the above object, the present invention forms a squish portion having a minute gap between the top surface of the piston and the cylinder head wall surface at the top dead center position around the combustion chamber, thereby making the combustion chamber a concentrated combustion chamber, The combustion chamber is configured so that two or more spark gaps face each other, and the ignition timing of the spark gap is controlled to be at or above the compression top dead center in the high load range. A valve opening / closing timing control device that varies the air-fuel ratio, thereby reducing the effective intake stroke by changing the closing timing of the intake valve in a low load range, thus igniting the compressed supply air in the combustion chamber by a spark gap. The mirror cycle was performed after burning.

Effect of the invention

According to the present invention, two or more spark gaps are faced in a compact centralized combustion chamber to achieve ultra-rapid combustion as a multi-point ignition system (the combustion period is 1 / 2.5 of the conventional crank angle). ˜3 (about 15 °), the ignition timing by the spark plug can be greatly delayed to the vicinity of the top dead center (in the high load range or above). Therefore, not only the maximum combustion temperature and pressure are low, the cooling loss and the friction loss are reduced, but also the combustion before the top dead center is hardly or not at all, so the cooling loss and the friction loss corresponding to this period are also small. . In the high load range, the ignition timing is set to the top dead center or later, so there is no worry of knocking (combustion is performed in the expansion stroke), and a high compression ratio can be adopted (if the fuel is gasoline, 14th to 15th) In the case of natural gas, the 17th to 18th positions), in the low load range, the mirror cycle is performed to increase the expansion ratio with respect to the effective compression ratio, so the thermal efficiency becomes very high. Moreover, since this mirror cycle is a cycle for shortening the effective suction stroke, the pump loss can be greatly reduced by reducing the intake throttle accordingly. Thus, the thermal efficiency is dramatically improved. Next, as described above, in the present invention, the ignition timing is greatly delayed by the ultra-rapid combustion method so that it is near the top dead center. For example, in the low load range, the ignition timing is set at the top dead center or slightly before that, but since a lean air-fuel mixture is employed, NOx is hardly generated. In the high load range, the ignition timing is set to the top dead center or later, so combustion is performed in the expansion stroke, and NOx generation is almost zero.
Therefore, an expensive lean NOx catalyst is unnecessary (only an inexpensive oxidation catalyst).
In addition, the exhaust gas temperature is low because the compression ratio and expansion ratio are high as described above, and it is not necessary to supply a rich mixture to protect the catalyst. (In the past, when the exhaust port is insulated, the durability of the catalyst deteriorates in a high load range). In addition, since super-rapid combustion is possible, the ignition timing can be greatly delayed during engine cold to increase the exhaust gas temperature, thereby enabling early activation of the catalyst. In the present invention, since the premixed combustion method is used, an expensive system such as a DPF or a common rail type fuel injection device is not required as an exhaust gas purification measure. In addition, since the maximum combustion temperature is low in the present invention, one of the features is that loss due to thermal dissociation is small.

Detailed Description of the Invention

FIG. 1 shows an ultra-high efficiency internal combustion engine according to the present invention, which includes an intake valve 3 driven by a cam provided on an intake camshaft (not shown in FIG. 1A) and a cam 7 provided on an exhaust camshaft 6. An exhaust valve 2 to be driven is provided. The exhaust valve 2 is driven by a cam 7 via a rocker arm 5, and the intake valve 3 is also driven in the same manner, although not shown. FIG. 1 (a) shows a multi-cylinder engine. FIG. 1 (b) shows one cylinder as viewed from above. As shown in the figure, there are two intake valves and two exhaust valves per cylinder, the exhaust valve 2 and the intake valve 3 are provided in the combustion chamber 1, and the exhaust valve 2 'and the intake valve 3' are provided in the squish portion S. (The exhaust valve 2 'and the intake valve 3' are also driven by cams). 9 is a catalyst for purifying exhaust gas, 11 is an intake throttle valve driven by an actuator 12 (electric motor), 20 is an electronic control unit, and 21 is a hydraulic control valve. The internal combustion engine is a four-cycle engine, and the intake, compression, expansion, and exhaust processes are performed by two rotations of the crankshaft. The exhaust camshaft 6 is provided with a valve opening / closing timing control device 14 'integrally with the sprocket 14 (or pulley), and the intake camshaft (not shown) is similarly provided with a valve opening / closing timing control device 15' (see FIG. The valve opening / closing timing control device 14'15 'has the same structure), and the sprockets 14, 15 are driven by being decelerated to 1/2 through a crankshaft and a chain (or a toothed belt). A known valve opening / closing timing control device 15 'includes a housing 16 and a rotor 17. The housing 16 is connected and integrated with a sprocket 15 and a bolt, and the rotor 17 is connected and integrated with an intake camshaft and a bolt (not shown). Yes. The advance chamber 18 and the retard chamber 19 are formed by the housing 16 and the rotor 17, and hydraulic oil is supplied from the hydraulic control valve 21. The hydraulic control valve 21 is supplied with hydraulic pressure from a hydraulic pump (not shown), and a plunger 22 whose movement in the axial direction is driven and controlled by an electromagnetic solenoid 23 according to an output signal from an electronic control unit 20 (hereinafter referred to as ECU) which will be described later. A spool valve 24 is provided, and the spool valve 24 is positioned at a position where the repulsive force of the spring 25 and the pressing force of the plunger 22 are balanced.
The ECU 20 receives a signal from a crank angle sensor that outputs a crank angle for each predetermined crank angle, and a signal from a cam angle sensor that outputs a cam angle signal for each predetermined cam angle. The ECU 20 outputs the signal from the cam angle sensor. The cam phase angle can be detected based on the output phase difference between the rotation angle pulse generated and the rotation angle pulse output from the crank angle sensor. The ECU 20 outputs a control signal to the hydraulic control valve 21 so that the phase difference of the rotor 17, that is, the intake camshaft with respect to the crankshaft becomes a target value. Then, the ECU 20 takes in the detected cam phase angle as a feedback signal, and feedback-controls the drive duty ratio of the hydraulic control valve 21 in accordance with the deviation from the control target phase angle. The hydraulic control valve 21 drives the electromagnetic solenoid 23 in accordance with the duty ratio instructed from the ECU 20, adjusts the supply / discharge of hydraulic pressure to the advance chamber 18 and the retard chamber 19, and advances the intake camshaft to the target phase angle. Maintain an angle, retard angle, or neutral (fix to an arbitrary cam phase angle). The ECU 20 is mainly composed of a microcomputer including a ROM, a RAM, a CPU, an input / output port and the like, and these are connected to each other by a bidirectional bus. The ECU 20 includes various sensors for parameters necessary for grasping the operating state of the engine, such as a crank angle sensor, a cam angle sensor, an accelerator sensor that detects an accelerator opening, a water temperature sensor that detects an engine cooling water temperature, an atmospheric pressure sensor, an engine Each signal from an air flow sensor, a knock sensor, an O ₂ sensor, and the like sucked in is transmitted to the input port via a corresponding A / D converter.
The engine speed can be known from the output signal from the crank angle sensor. The output port is connected to the fuel injection valve 13, the spark plug 4, the actuator 12, the hydraulic control valve 21 and the like via corresponding drive circuits, and transmits respective control signals. The ROM has a control routine for controlling the engine, such as a control routine for determining the injection amount and injection timing of the fuel injection valve 13, a control routine for controlling the energization of the spark plug 4, and controls used for them. A map containing the values is stored. Various data stored in the RAM are replaced with the latest data every time the engine speed sensor outputs a signal.
The CPU operates in accordance with an application program stored in the ROM, and executes fuel injection control, ignition timing control, and the like.

とする）。即ち、シリンダー中心に１個の点火プラグを配置した従来の場合に比し、図２（イ）では火炎伝播距離が１／４に短縮されるが、実際は点火プラグ４が同一直線上になく対向して備えられている為、目減りして１／３位には短縮されると考えて良い（図２（ロ）の様に点火プラグ４を４個配置すれば確実に１／３に短縮される）。この様に本発明では集中型燃焼室１と多点プラグとの相乗効果により、従来の中心１点プラグの場合に比し必要とされる火炎伝播距離が１／３に短縮され、、言い換えると従来クランク角で４０°〜５０°要した燃焼期間が１／３の１５°位に短縮され、超急速燃焼が可能となるのである。この結果、火花ギャップ（点火プラグ）の着火時期を大幅に遅らせ、上死点付近としても（高負荷域では上死点又はそれ以降とし、低負荷域では同様に上死点又はそれ以降とするか、希薄混合気やＥＧＲを採用する場合は上死点より僅か前とする）超急速燃焼が可能であるから、熱効率の悪化はなく、むしろ最高燃焼温度・圧力の低下によって冷却損失、摩擦損失、更には熱解離損失が減って大幅に熱効率が向上する。極く小排気量のエンジンの場合は火炎伝播距離の絶対値が小さい為、２点の火花ギャップでも良く、本発明では燃焼室に２点以上の火花ギャップを臨ませる事を特徴としている。
尚、本発明では火花ギャップとして点火プラグを使用しているが、図２（ハ）の如く火花ギャップ２６を多数有するセラミックなどの不導体板２７を図２（ニ）の如く燃焼室１に備えても良い。The combustion chamber 1 is provided with an exhaust valve 2 and an intake valve 3, and a squish portion having a minute gap (about 0.8 mm) between the piston top surface and the cylinder head wall surface at the top dead center position around the combustion chamber 1. To form a concentrated combustion chamber (the squish portion S is provided with an exhaust valve 2 'and an intake valve 3'). Combustion chamber 1 has two or more spark gaps (three spark gaps consisting of three spark plugs 4 in the figure), and the ignition timing of this spark gap is top dead center in the high load range. Or it is controlled so that it becomes after that (Because EGR or lean air-fuel mixture is adopted in the low load region, it is desirable to be near the top dead center near the top dead center). FIG. 1 (b) is a view of the combustion chamber 1 of FIG. 1 (a) as viewed from above, and this is shown in a simplified manner in FIG. 2 (a). If the cylinder radius is R, the flame propagation distance when one spark plug is placed at the center of the cylinder is R, and three spark plugs are placed in the centralized combustion chamber 1 as shown in FIG.

And). That is, compared with the conventional case where one spark plug is arranged at the center of the cylinder, the flame propagation distance is shortened to 1/4 in FIG. Therefore, it is possible to think that it will be reduced to 1 / 3rd place by reducing the number of eyes (if 4 spark plugs 4 are arranged as shown in FIG. ) Thus, in the present invention, the synergistic effect of the centralized combustion chamber 1 and the multipoint plug reduces the required flame propagation distance to 1/3 compared to the case of the conventional central single point plug, in other words, The combustion period which conventionally required 40 ° to 50 ° in crank angle is shortened to about 1/3 of 15 °, and ultra-rapid combustion becomes possible. As a result, the ignition timing of the spark gap (ignition plug) is greatly delayed, and even near the top dead center (top dead center or higher in the high load range, and top dead center or lower in the low load range as well) (If lean air-fuel mixture or EGR is used, it should be just before the top dead center.) Super-rapid combustion is possible, so there is no deterioration in thermal efficiency. Rather, cooling loss and friction loss are caused by lowering the maximum combustion temperature and pressure. Furthermore, the thermal dissociation loss is reduced and the thermal efficiency is greatly improved. In the case of an engine with a very small displacement, the absolute value of the flame propagation distance is small, so two spark gaps may be used, and the present invention is characterized by having two or more spark gaps in the combustion chamber.
In the present invention, a spark plug is used as the spark gap. However, as shown in FIG. 2 (c), a non-conductive plate 27 such as ceramic having many spark gaps 26 is provided in the combustion chamber 1 as shown in FIG. 2 (d). May be.

Next, each operation state of the internal combustion engine according to the present invention will be described.
In the present invention, since the valve opening / closing timing control devices 14 'and 15' are provided, the opening / closing timing of the exhaust valve 2 (2 ') and the intake valve 3 (3') can be freely controlled (however, FIG. 1). The valve opening / closing timing control devices 14 'and 15' employed in the above change the valve opening / closing timing while keeping the valve opening / closing period constant). For example, in the low load range, the intake valves 3 'and 3 are opened at 50 ° CA and 20 ° CA after top dead center, respectively, and closed at 90 ° CA after bottom dead center as shown in FIG. On the other hand, the exhaust valves 2 'and 2 are controlled so that both are opened at 10 ° CA after bottom dead center and closed at 70 ° CA after top dead center as shown in FIG. As a result, a large amount of exhaust gas is re-inhaled (EGR) from the exhaust passage 8 and intake air (a mixture of air and exhaust gas) once sucked into the cylinder is returned to the intake passage 10 to reduce the effective intake stroke. Then, a mirror cycle having a lower compression ratio (effective compression ratio) than the expansion ratio is performed. In the present invention, since ignition / combustion by a spark gap is performed at the top dead center or afterwards in a high load range, knocking does not occur even at a high compression ratio. Therefore, when the fuel is gasoline, the compression ratio is 14-15. In the case of natural gas, it is raised to 17-18. Accordingly, the effective intake stroke is shortened by the mirror cycle, and accordingly, the degree of intake throttling is small and the pump loss is reduced. Further, since a large amount of EGR is performed and the air-fuel mixture is also diluted, the pump loss is greatly reduced. Even if the mirror cycle is performed, the effective compression ratio is sufficiently high and the compression end temperature is also high due to EGR, so that combustion is good. On the other hand, since the expansion ratio is 14 to 15 (in the case of gasoline), which is a high expansion ratio, the improvement in thermal efficiency is remarkable in combination with a significant reduction in pump loss.
In addition, the generation of NOx is almost zero due to the use of a large amount of EGR, a lean air-fuel mixture, and a significant delay in the ignition timing of the spark gap (made possible by ultra-rapid combustion). In the high load range, the intake valves 3 'and 3 open at top dead center and 30 ° CA before top dead center as shown in Fig. 3 (a), and close at 40 ° CA after bottom dead center, while the exhaust valve 2' 2 are both opened at 40 ° CA before the bottom dead center as shown in FIG. 3 (b), and closed at 50 ° CA before the top dead center and 20 ° CA after the top dead center, respectively. Although the expansion ratio is the same, the thermal efficiency is improved because it is higher than the conventional 12 or less. Since the ignition timing by the spark gap is at the top dead center or after the top dead center, combustion is performed in the expansion stroke, and NOx is hardly generated. For the same reason, since the maximum combustion temperature and pressure are low, there is little cooling loss and friction loss, and there is no combustion before top dead center, so cooling loss and friction loss are further reduced. By the way, regarding the valve opening / closing timing control, since the intake valve 3 and the exhaust valve 2 are in the combustion chamber 1, they do not collide with the piston no matter how they are controlled, but the intake valve 3 ′ and the exhaust valve 2 ′ are squish parts. In order to avoid collision with the piston, it is necessary to control the former so that it does not open before top dead center, and the latter is closed before top dead center (to form a valve recess). A slight angle of about 10 ° CA is allowed. Note that the intake valves 3 'and 3 open after top dead center in the low load range, but the exhaust valve 2 opens up to 70 ° CA after top dead center, so no pump loss occurs, and the exhaust valve 2' in the high load range. Is closed at 50 ° before the top dead center, but the exhaust valve 2 is open to 20 ° after the top dead center, so that the exhaust is smoothly performed. In the middle load range, the valve opening / closing timing of the intake / exhaust valves is set at an intermediate position in FIGS. When starting the engine, the mirror cycle is not used, and the engine is started at a high compression ratio using the valve opening / closing timing in a high load range (the compression end temperature rises and good combustion is obtained). The present invention employs a multi-point ignition system with two or more spark gaps in the centralized combustion chamber, so that ultra-rapid combustion is possible (as described above, the combustion period is 15% of the conventional 1/3 at the crank angle). Therefore, when the engine is warming up, the ignition timing is greatly delayed to increase the exhaust temperature, and the catalyst can be activated early (when starting, the ignition timing is before the top dead center). In addition, since the compression ratio and expansion ratio are high, the exhaust gas temperature is low (low even in the high load range), and not only does not need to supply a rich mixture to protect the catalyst, but a port liner is used to insulate the exhaust port. The catalyst can be activated at an early stage (conventionally, if the exhaust port is insulated, the exhaust temperature rises and the durability of the catalyst is impaired).

FIG. 4 (a) shows the combustion chamber 1 with one exhaust valve 2 (the valve opening and closing timing is the same as that of the exhaust valve 2 in FIG. 3 (b)), and the squish portion S is provided with intake valves 3a and 3b. is there.
The valve opening / closing timing of the intake valves 3a and 3b is, for example, open at the top dead center in the low load range and closed at 90 ° CA after the bottom dead center to perform the mirror cycle, and open at the top dead center in the high load range, after the bottom dead center Close at 40 ° CA (in the middle load range). However, the valve opening / closing timing control device is the electromagnetic type shown in FIG. 4 (c) or the two-stage cam type shown in FIG. 4 (b) is provided with an intake / exhaust valve only in the combustion chamber 1. The valve opening / closing timing of the exhaust valve 2 is the same as that of the exhaust valve 2 of FIG. 3 (b), and that of the intake valve 3 is shown in FIG. This is the same as the intake valve 3 in (a). 4 (a) and 4 (b), the combustion chamber 1 is exposed to two or more (three in the figure) spark gaps. In the present invention, in addition to the valve opening / closing timing control device shown in FIG. 1, for example, a device that controls the valve opening / closing timing by moving the helical spline in the cam shaft direction, a device that moves the solid cam in the axial direction, and the like can be considered. The valve opening / closing timing control device 28 shown in FIG. 4C is an electromagnetic type, and the electromagnets 29 and 30 are arranged in the housing in a state of facing each other with a plunger 33 integrated with the valve shaft. . The springs 31 and 32 are attached to the hollow portions of the electromagnets 29 and 30 so as to sandwich the plunger 33. Therefore, the valve V (suction / exhaust valve) is located at the middle position of the valve lift by the balance of the forces of the springs 31 and 32. Quiesce. When the electromagnet 29 is magnetized, the plunger 33 is pulled upward and the valve is closed. When the electromagnet 30 is magnetized, the plunger 33 is attracted downward and the valve V is opened. The driver 34 alternately supplies an exciting current to the electromagnets 29 and 30 in accordance with a control signal from the ECU 35 that designates the opening / closing timing of the valve V (which may include valve lift information). According to this, the opening and closing timing of the valve V can be arbitrarily set, and the degree of freedom increases. In the present invention, the valve opening / closing timing of the exhaust valves 2 and 2 'is controlled by the number of re-inhaling exhaust gas by opening the exhaust valves 2 and 2' even in the intake stroke. The valve opening / closing timing may be fixed (for example, opened at 40 ° CA before bottom dead center and closed at top dead center). In this case, since exhaust gas is re-inhaled and EGR is not performed, a further lean air-fuel mixture is used accordingly.

おける集中型燃焼室の火炎伝播距離、Ｒはシリンダー中心に１点火花ギャップの従来の火炎伝播距離であるから１／２．４７に短縮される。しかし本発明では圧縮比が１７〜１８と高圧縮比であるから（ガソリンの場合でも１４〜１５と高圧縮比）、圧縮上死点付近では燃料の低温酸化反応が進んでおり、これは火炎伝播速度を増大させるから、火炎伝播距離は実際は１／３位まで短縮されると考えて良い。換言すればクランク角で従来４０°〜５０°であった燃焼期間が１／３の１５°位に短縮され、図１と同様に超急速燃焼が可能である。吸気弁が２個、排気弁が１個の時の燃焼室３８及び点火プラグ４の配置を図５（ニ）に示す。又、吸気弁、排気弁が各々２個づつの時の燃焼室３８、及び点火プラグ４の配置を図５（ホ）に示し、点火プラグ４は３個備えてある（大排気量エンジンに適する）。
再び図５（イ）、（ロ）に戻って吸気弁３７、排気弁３６の弁開閉時期を図６（ホ）、（ヘ）に示す。即ち、低負荷域では吸気弁３７を上死点で開、下死点後９０°ＣＡで閉として有効吸入行程を減少させてミラーサイクルを行なうと共に、排気弁３６を下死点前４０°ＣＡで開、上死点後８０°ＣＡで閉（上死点付近では殆ど閉とした後に）として多量のＥＧＲを行ない、高温の排ガスを再吸入して圧縮端温度を高め、燃焼を良好にしている（排気弁３６の弁開閉時期制御装置については、これは必要不可欠なものではなく、下死点前４０°ＣＡで開、上死点で閉として固定し、ＥＧＲを行なわない様にしても良いが、ここでは可変としている）。この時、混合気を希薄化し、より一層ポンプ損失を減少させる事が望ましい。尚、低負荷でミラーサイクルを行なうに当っては、吸気弁３７を下死点後９０°ＣＡで閉としていたが、下死点前９０°ＣＡで閉としても同様に有効行程を短縮化してミラーサイクルを行なう事ができる。
高負荷域では吸気弁３７は上死点前１０°ＣＡで開、下死点後４０°ＣＡで閉、排気弁３６は下死点前４０°ＣＡで開、上死点で閉としてミラーサイクル、ＥＧＲは行なわない。この様に低負荷域と高負荷域とでは弁開閉時期が異なるが、これはエンジン負荷、回転速度などの運転条件により弁開閉用のカムを切り換えて行なう（２段カム切り換え式）。尚、吸気弁３７は高負荷域では上死点前１０°ＣＡで開くが、弁リフトは僅かなので、適正な弁リセス部によりピストンとの衝突は全く心配ない（排気弁３６も上死点後１０°ＣＡで閉としても良い）。
エンジン暖機時（冷態時）では吸気弁３７は圧縮比の高い高負荷域での弁開閉時期を使用するが、超急速燃焼が可能であるから、着火時期を大幅に遅らせて触媒の早期活性化を図る。この場合、燃焼速度を更に高める為、図５（ロ）のスワール制御弁３９によりシリンダー内スワールを強化するのが良い。次に、弁開閉時期を切り換える機構（弁開閉時期制御装置）について説明する。この機構は公知のものであるが、図６（イ）の如く弁Ｖを駆動するロッカーアーム４５に隣接してロッカーアーム４４がロッカー軸４１（回転しない）に備えられ、Ｘ−Ｘ′軸にカム４２、４３を有するカム軸が配置されてロッカアーム４４、４５を各々駆動する（押し下げる）。ロッカーアーム４４は図示しないロストモーションスプリングによりカム４２に常時接触している。弁Ｖが図５（イ）の吸気弁３７の場合を説明すると、図示しない油圧制御弁により油圧ポンプからの油圧がロッカー軸４１内の油圧に伝達されると、油圧ピストン４８ａがバネ４６に抗してロッカーアーム４４の穴に嵌り込み、ロッカーアーム４４と４５は結合・一体化され、ミラーサイクル用のカム４２のプロフィール通りに弁Ｖが駆動され、上死点で開、下死点後９０°ＣＡで閉となる。エンジン負荷・回転速度などの条件を検出してＥＣＵからの出力信号によりロッカー軸４１内の油圧が低下するとバネ４６により油圧ピストン４８ａがロッカーアーム４４の穴から抜け出し、ミラーサイクルは解除され、カム４３のプロフィール通りに弁Ｖが駆動され、上死点前１０°ＣＡで開、下死点後４０°ＣＡで閉となる。尚、油圧ピストン４８ａとバネ座４７とは各々の端面が常に必らずどこかで接触しており、一瞬でも離れる事がない様になっている。弁Ｖが図５（イ）の排気弁３６の場合も同様の弁開閉時期制御装置により弁開閉時期が制御される（説明略）。
図６（イ）の油圧ピストン４８ａはＸ−Ｘ′軸方向に移動するが、これと直角方向に移動させてカム４２、４３の切り換えを行なう事もできる。即ち、図６（ロ）の如くロッカーアーム４５′に突出部４９を形成し、ここに形成された穴にロッカーアーム４４′に内蔵された油圧ピストン４８ｂがＸ−Ｘ′軸と直角方向に移動して嵌り込むのである。ロッカーアーム４４′、４５′が油圧ピストン４８ｂにより結合されると、ミラーサイクル用カム４２のプロフィール通りに弁Ｖが駆動される。図６（ハ）はロッカーアーム５０に切り換え体５１を備えたもので、ロッカーアーム５０はカム４３により駆動されるが、切り換え体５１はカム４２により駆動されるものである。
Ｃ方向から見た図を図６（ニ）に示すが、常にロストモーションスプリング５２によりカム４２に押し付けられており、このロストモーションスプリング５２は弁Ｖの開閉には何ら関与しない弱いバネである。ロッカー軸４１内の油圧が高まると油圧ピストン４８Ｃ飛び出して切り換え体５１をロックし、ロッカーアーム５０はカム４２のプロフィール通りに駆動され、弁Ｖが吸気弁３７の場合は上死点で開、下死点後９０°ＣＡで閉となって、ミラーサイクルを行なう。ロッカー軸４１内の油圧が低下すると油圧ピストン４８Ｃはバネにより引っ込み（ロック解除）。切り換え体５１はカム４２のプロフィール通りに駆動されるがロッカーアーム５０を駆動せず（空振り）、これはカム４３によって駆動され、上死点前１０°ＣＡで開、下死点後４０°ＣＡで閉となる。Although the combustion chamber 1 is provided in the cylinder head in FIG. 1, the internal combustion engine according to the present invention provided on the piston side is shown in FIG. That is, in FIG. 5 (a), the combustion chamber 38 is provided in the piston, and is configured such that, for example, two spark plugs 4 (spark gaps) face as shown in FIG. 5 (b). 36 is an exhaust valve, and 37 is an intake valve, each provided one by one.
As shown in FIG. 5C, the piston and the combustion chamber 1 are taken out, and the piston diameter is 2R, and the combustion chamber 1 is formed into an oval shape when a circle with a radius r is moved by 2r. Then, the volume v of the combustion chamber 1 is approximately 2πr ³ (the hatched portions A and B are approximately equal in volume). When the fuel is natural gas (CH4), the pressure

The flame propagation distance of the centralized combustion chamber in this case, R, is a conventional flame propagation distance of one spark flower gap at the center of the cylinder, so it is shortened to 1 / 2.47. However, in the present invention, since the compression ratio is a high compression ratio of 17 to 18 (high compression ratio of 14 to 15 even in the case of gasoline), the low-temperature oxidation reaction of fuel proceeds near the compression top dead center. Since the propagation speed is increased, it can be considered that the flame propagation distance is actually shortened to 1/3. In other words, the combustion period of the conventional crank angle of 40 ° to 50 ° is shortened to about 1/3 of 15 °, and ultra-rapid combustion is possible as in FIG. FIG. 5D shows the arrangement of the combustion chamber 38 and the spark plug 4 when there are two intake valves and one exhaust valve. FIG. 5E shows the arrangement of the combustion chamber 38 and the spark plug 4 when there are two intake valves and two exhaust valves, respectively. Three spark plugs 4 are provided (suitable for large displacement engines). ).
Returning again to FIGS. 5 (a) and 5 (b), the valve opening / closing timings of the intake valve 37 and the exhaust valve 36 are shown in FIGS. 6 (e) and 6 (f). That is, in the low load region, the intake valve 37 is opened at the top dead center and closed at 90 ° CA after the bottom dead center to perform the mirror cycle by reducing the effective intake stroke, and the exhaust valve 36 is set to 40 ° CA before the bottom dead center. Open and close at 80 ° C after top dead center (after almost closed near top dead center), perform a large amount of EGR, re-suction hot exhaust gas to increase compression end temperature, and improve combustion (The valve opening / closing timing control device for the exhaust valve 36 is not indispensable. It is fixed at 40 ° CA before bottom dead center and closed at top dead center so that EGR is not performed. Good but variable here) At this time, it is desirable to dilute the air-fuel mixture and further reduce the pump loss. In performing the mirror cycle at a low load, the intake valve 37 was closed at 90 ° CA after bottom dead center. However, if the intake valve 37 is closed at 90 ° CA before bottom dead center, the effective stroke is shortened in the same way. A mirror cycle can be performed.
In the high load range, the intake valve 37 opens at 10 ° CA before top dead center, closes at 40 ° CA after bottom dead center, and the exhaust valve 36 opens at 40 ° CA before bottom dead center, and closes at top dead center. , EGR is not performed. In this way, the valve opening / closing timing differs between the low load region and the high load region, but this is performed by switching the valve opening / closing cam according to operating conditions such as engine load and rotation speed (two-stage cam switching type). The intake valve 37 opens at 10 ° CA before top dead center in a high load range, but the valve lift is slight, so there is no concern about collision with the piston due to the appropriate valve recess (the exhaust valve 36 is also after top dead center). It may be closed at 10 ° CA).
When the engine is warm (cold), the intake valve 37 uses a valve opening / closing timing in a high load region with a high compression ratio. However, since the super-rapid combustion is possible, the ignition timing is greatly delayed to shorten the catalyst early. Try to activate. In this case, in order to further increase the combustion speed, it is preferable to strengthen the swirl in the cylinder by the swirl control valve 39 in FIG. Next, a mechanism for switching the valve opening / closing timing (valve opening / closing timing control device) will be described. Although this mechanism is known, a rocker arm 44 is provided on a rocker shaft 41 (not rotating) adjacent to a rocker arm 45 for driving the valve V as shown in FIG. Cam shafts having cams 42 and 43 are arranged to drive (depress) the rocker arms 44 and 45, respectively. The rocker arm 44 is always in contact with the cam 42 by a lost motion spring (not shown). The case where the valve V is the intake valve 37 of FIG. 5A will be described. When the hydraulic pressure from the hydraulic pump is transmitted to the hydraulic pressure in the rocker shaft 41 by a hydraulic control valve (not shown), the hydraulic piston 48a resists the spring 46. The rocker arms 44 and 45 are combined and integrated, and the valve V is driven according to the profile of the cam 42 for the mirror cycle, opened at the top dead center, and 90 after the bottom dead center. ° Closes at CA. When conditions such as engine load and rotational speed are detected and the hydraulic pressure in the rocker shaft 41 is lowered by an output signal from the ECU, the hydraulic piston 48a is pulled out of the hole of the rocker arm 44 by the spring 46, the mirror cycle is released, and the cam 43 The valve V is driven in accordance with the profile of FIG. 5 and is opened at 10 ° CA before top dead center and closed at 40 ° CA after bottom dead center. Note that the end surfaces of the hydraulic piston 48a and the spring seat 47 are always in contact with each other so that they are not separated even for a moment. When the valve V is the exhaust valve 36 shown in FIG. 5A, the valve opening / closing timing is controlled by a similar valve opening / closing timing control device (not shown).
The hydraulic piston 48a in FIG. 6 (a) moves in the XX 'axis direction, but the cams 42 and 43 can be switched by moving in the direction perpendicular thereto. That is, as shown in FIG. 6 (b), the protrusion 49 is formed on the rocker arm 45 ', and the hydraulic piston 48b built in the rocker arm 44' moves in the hole formed here in the direction perpendicular to the XX 'axis. It fits in. When the rocker arms 44 ′ and 45 ′ are coupled by the hydraulic piston 48 b, the valve V is driven according to the profile of the mirror cycle cam 42. In FIG. 6C, the rocker arm 50 is provided with a switching body 51, and the rocker arm 50 is driven by a cam 43, but the switching body 51 is driven by a cam 42.
FIG. 6 (d) shows a view as seen from the C direction. The lost motion spring 52 is always pressed against the cam 42 by the lost motion spring 52, and the lost motion spring 52 is a weak spring that is not involved in the opening and closing of the valve V. When the hydraulic pressure in the rocker shaft 41 increases, the hydraulic piston 48C pops out to lock the switching body 51, the rocker arm 50 is driven according to the profile of the cam 42, and when the valve V is the intake valve 37, it opens at the top dead center. Closed at 90 ° CA after the dead point, a mirror cycle is performed. When the hydraulic pressure in the rocker shaft 41 decreases, the hydraulic piston 48C is retracted (unlocked) by the spring. The switching body 51 is driven according to the profile of the cam 42 but does not drive the rocker arm 50 (skipping), which is driven by the cam 43 and opens at 10 ° CA before top dead center and 40 ° CA after bottom dead center. It closes at.

By the way, in FIG. 5A, if a part of the multi-cylinder engine is in a low load region (for example, two cylinders in a four-cylinder engine) are deactivated, the pump loss is reduced and the thermal efficiency is further improved ( The same applies to FIG. 1). Next, a mechanism for operating and stopping the intake / exhaust valves will be described. That is, in FIG. 7 (a), the rocker shaft 41 is provided with rocker arms 54 and 55 adjacent to the rocker arm 53 for driving the valve V, and the interior of the rocker shaft 41 is separated by the partition wall 56. The control valves 57 and 58 are connected. The hydraulic control valve 57 receives an output signal from an ECU (not shown) according to conditions such as engine load, rotation speed, and coolant temperature, and supplies hydraulic pressure from the hydraulic pump into the rocker shaft 41. The valve V is an intake valve or an exhaust valve, and the description will be given for the case of the intake valve, but the same applies to the case of the exhaust valve. In the engine high load range or at the time of start-up, the hydraulic pressure from the hydraulic pump is shut off by the hydraulic control valve 57 (maintained at a low pressure), and therefore the hydraulic piston 48b (moves perpendicularly to the XX ′ axis as in FIG. 6 (b)). Is pushed into the hole of the rocker arm 54 by a spring, and the rocker arms 53 and 54 are combined and integrated as shown in the figure, the valve V is driven by the cam 43, and the valve opening / closing timing in a high load region as shown in FIG. It becomes. In the partial load region, the hydraulic control valve 57 supplies the high pressure from the hydraulic pump and pulls out the hydraulic piston 48b (the loker arms 53 and 54 are separated), and the hydraulic control valve 58 supplies the high pressure from the hydraulic pump to the hydraulic piston. 48a is pushed into the hole of the rocker arm 55, and the rocker arms 53 and 55 are combined and integrated. As a result, the valve V is driven by the cam 42, and the valve opening / closing timing in the low load region is reached as shown in FIG. .
In the low load region (a constant region in the low load region where the intake / exhaust valves should be stopped), the hydraulic control valves 57 and 58 are controlled to release the coupling between the rocker arms 53, 54 and 55, and the rocker arm Since 53 is not driven by the cam, the operation of the valve V is stopped. The hydraulic piston 48b in FIG. 6 (b) is a type that couples the rocker arms against the spring 46 when a high pressure acts, whereas the hydraulic piston 48b in FIG. 7 (a) resists the spring when a high pressure acts. This is a type in which the connection between the rocker arms is released, and the action is simply reversed. When the inside of the rocker shaft 41 becomes high pressure, the rocker arms 53 and 54 are uncoupled and the rocker arms 53 and 55 are joined. When the oil pressure inside the rocker shaft 41 becomes low, the rocker arms 53 and 54 Since they are coupled and the rocker arms 53 and 55 are uncoupled, the hydraulic control valves 57 and 58 can be shared. The same applies to the exhaust valve (valve V is the exhaust valve). In the case where the valve opening / closing timing of the exhaust valve 36 is fixed in FIG. 5 (a), it is only necessary to add a mechanism for stopping the exhaust valve 36 in the low load region, and the inside of the rocker shaft 41 'as shown in FIG. 7 (b). When the hydraulic pressure is low, the hydraulic piston 48C is fitted into the mating hole by a spring to fix the switching body 51 to the rocker arm, and the exhaust valve 36 is driven by a cam. Pushing out, the switching body 51 is driven according to the profile of the cam, but does not drive the rocker arm.
That is, the exhaust valve 36 is in a resting state. The hydraulic piston 48C has a reverse action to that of FIG. As described above, in FIG. 7 (a), adjacent rocker arms are coupled or released by the hydraulic piston, and in FIG. 7 (b), the switching body 51 is locked or unlocked to the rocker arm by the hydraulic piston. Although the valve opening / closing timing of the valve V has been changed or stopped by this, more embodiments can be considered as shown in FIGS. 7C and 7D by appropriately combining them. First, when the rocker arms 60 and 59 are coupled and integrated by a hydraulic piston (not shown) in FIG. 7C (the switching body 51 is unlocked), the valve V is driven according to the profile of the cam 42, When the coupling between the rocker arms 59 and 60 is released and the driving body 51 is coupled and integrated (locked) with the rocker arm 59 by a hydraulic piston (not shown), the valve V is driven according to the profile of the cam 43. Then, if the coupling between the rocker arms 59 and 60 is released and the lock of the switching body 51 is released, the valve V is deactivated. Next, in FIG. 7D, the valve V is deactivated when both the left and right switching bodies 51 are unlocked, and the right side switching body 51 is locked (the left switching 51 is unlocked). When the right switching body 51 is unlocked and the left switching body 51 is locked, the valve V is driven according to the profile of the cam 43. 6 and 7 have a rocker shaft at one end thereof, the present invention is similarly applied to a rocker arm having a rocker shaft at an intermediate position of the rocker arm, an example of which is shown in FIG. Shown in (e).

In the present invention of FIG. 5, if the valve opening / closing timing of the intake valve 37 is used in the high load region even in the low load region, the compression ratio becomes very high and a large amount of EGR is performed. , The HCCI combustion (premixed compression ignition combustion) can be performed, and the thermal efficiency can be greatly improved.

1 is a diagram showing an ultra-high efficiency internal combustion engine according to the present invention. The figure which shows arrangement | positioning of a concentrated combustion chamber and a spark gap. The figure which shows valve opening / closing timing control of an intake / exhaust valve. The figure which shows the various Example of this invention. 1 is a diagram showing an ultra-high efficiency internal combustion engine according to the present invention. The figure which shows a valve opening / closing timing control apparatus. The figure which shows the valve opening / closing timing control apparatus which also enables valve operation stop.

1 and 38 are combustion chambers, 2 and 2 'and 36 are exhaust valves, 3 and 3' and 37 are intake valves, 3a and 3b are intake valves, 4 is a spark plug, 5 · 44 · 45 · 44 '· 45' 50, 53, 54, 55, 59, 60 and 61 are rocker arms, 6 is a camshaft, 7 is 42, 43 is a cam, 8 is an exhaust passage, 9 is a catalyst, 10 is an intake passage, and 11 is an intake throttle valve 12 and 40 are actuators, 13 is a fuel injection valve, 14 and 15 are sprockets, 14 'and 15' are valve timing control devices, 16 is a housing, 17 is a rotor, 18 is an advance chamber, 19 is a retard chamber , 20, 35 are ECUs, 21, 57, 58 are hydraulic control valves, 22 are plungers, 23 are electromagnetic solenoids, 24 are spool valves, 25, 31, 32, 46 are springs, 26 are spark gaps, 27 are non- Conductor plate, 28 is a valve opening / closing timing control device, 29/3 Is an electromagnet, 33 is a plunger, 34 is a driver, 39 is a swirl control valve, 41 and 41 'are rocker shafts, 47 is a spring seat, 48a, 48b and 48C are hydraulic pistons, 49 is a protrusion, 52 is a lost motion spring , 56 are partition walls, V is a valve, and 51 is a switching body.

Claims (2)

In a four-cycle internal combustion engine having an intake valve and an exhaust valve, a squish portion having a minute gap is formed around the combustion chamber between the top surface of the piston at the top dead center position and the wall surface of the cylinder head, and the combustion chamber is intensively combusted. The combustion chamber is configured so that two or more spark gaps face the combustion chamber, and the ignition timing of the spark gap is controlled to be at or above the compression top dead center in the high engine load range. A valve opening / closing timing control device for varying the valve opening / closing timing of the valve is provided, thereby changing the closing timing of the intake valve to reduce the effective intake stroke in the engine low load range, thus reducing the compressed supply in the combustion chamber. An ultra-high-efficiency internal combustion engine configured to perform a mirror cycle by igniting and burning a gas through the spark gap. 2. An ultra-high efficiency internal combustion engine according to claim 1, wherein a recess is formed in the piston to form a combustion chamber, and two or more spark gaps face the combustion chamber.

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