JP6084941B2 - Combustion control device for internal combustion engine - Google Patents

Description

  The present invention relates to a combustion control device for an internal combustion engine, and more particularly to a combustion control device for burning a homogeneous and lean air-fuel mixture by spark ignition.

A lean mixture combustion technique for spark-igniting a lean (high air-fuel ratio) mixture is widely known, for example, as shown in Patent Document 1. In this lean mixture combustion technique, NOx contained in the exhaust gas cannot be purified by a three-way catalyst, so it is important to increase the combustion limit air-fuel ratio. Patent Document 1 discloses a technique for stably combusting while increasing the average air-fuel ratio of the air-fuel mixture in the entire combustion chamber by stratifying and arranging a relatively rich air-fuel mixture (small air-fuel ratio) near the spark plug. Has been.
However, in this technique, since the amount of NOx generated by the combustion in the vicinity of the spark plug is large, it is required to further reduce the NOx emission amount.

  Therefore, the inventor of the present invention proposed a spark-ignition combustion of a homogeneous lean mixture that can burn a homogeneous and lean mixture by spark ignition and reduce NOx emissions while ensuring stable combustion. (Non-Patent Document 1). In Non-Patent Document 1, as a method for reliably burning a homogeneous lean air-fuel mixture having an air-fuel ratio of about “30”, the compression ratio is set so as to raise the air-fuel mixture temperature at ignition to a desired temperature, It has been proposed to use a fuel injection valve.

JP 2002-256927 A

Preprint of the Society of Automotive Engineers of Japan, No. 121-12 (2012/10)

  Non-Patent Document 1 shows that the proposed spark ignition combustion of a homogeneous lean mixture can surely burn a homogeneous lean mixture having an air-fuel ratio of about “30” to reduce NOx emissions. There is still room for further study on specific combustion control methods to be applied to mass-produced engines.

  The present invention has been made in view of such a situation. Combustion of an internal combustion engine that reliably realizes spark ignition combustion of a homogeneous lean air-fuel mixture, is highly efficient, and can reduce NOx emissions. An object is to provide a control device.

In order to achieve the above object, the invention described in claim 1 is a spark ignition means for performing spark ignition of an air-fuel mixture in a combustion chamber (1) of an internal combustion engine, an ignition control means for controlling the spark ignition means, An air-fuel ratio control means for controlling the air-fuel ratio of the air-fuel mixture in the combustion chamber, comprising: a homogeneous lean air-fuel mixture forming means for forming a homogeneous and lean air-fuel mixture in the combustion chamber; The spark ignition means includes an ignition plug (8) and a plurality of ignition coil pairs (71, 72) for generating a discharge in the ignition plug, and a discharge start timing (CAIG) and duration in the ignition plug. (TSPK) can be changed, and the homogeneous lean air-fuel mixture forming means uses a fuel injection valve (6) capable of atomizing the fuel and injecting the fuel into the intake passage (2) of the engine. Inject, said air fuel The control means controls the air-fuel ratio to a leaner air-fuel ratio that is leaner than a predetermined lean air-fuel ratio (AFL1) that is leaner than the stoichiometric air-fuel ratio, and the ignition control means includes a discharge start timing (CAIG) in the spark plug and When controlling the discharge duration (TSPK) and igniting the stratified lean air-fuel mixture formed so that the air-fuel ratio of the air-fuel mixture in the vicinity of the spark plug becomes relatively small in the discharge start timing (CAIG) The ignition control means sets the discharge start timing (CAIG) as the air-fuel ratio increases in the air-fuel ratio range leaner than the predetermined lean air-fuel ratio (AFL1). The discharge angle is advanced and the discharge duration (TSPK) is set to be long .

According to this configuration, the air-fuel ratio is controlled to be a lean air-fuel ratio that is further leaner than the predetermined lean air-fuel ratio, and the atomized fuel is injected into the intake passage. A lean air-fuel mixture is formed in the intake passage and further sucked into the combustion chamber, whereby a lean air-fuel mixture with higher homogeneity can be formed. In addition, since the discharge start timing and discharge duration in the spark plug can be changed, by appropriately setting the discharge start timing and discharge duration, that is, the lead angle from the discharge start timing when igniting a stratified lean mixture is set. By setting the discharge start time on the side, it is possible to set a long discharge duration, and to reliably ignite a homogeneous lean air-fuel mixture. In addition, in the air-fuel ratio range leaner than the predetermined lean air-fuel ratio, control is performed to advance the discharge start timing and set the discharge duration longer as the air-fuel ratio increases. Stable combustion can be realized.

The invention according to claim 2 is the combustion control device for an internal combustion engine according to claim 1 , further comprising flow generation means (2a, 4) for generating a flow of the intake air-fuel mixture in the combustion chamber. It is characterized by.

  According to this configuration, since the flow of the sucked air-fuel mixture is generated in the combustion chamber, a powerful initial flame nucleus is formed by setting the discharge duration of the spark plug and the air-fuel mixture flow, and high-efficiency combustion Can be realized.

According to a third aspect of the present invention, in the combustion control device for an internal combustion engine according to the first or second aspect , the homogeneous lean air-fuel mixture forming means is configured to inject the fuel during a closing period of the intake valve (21) of the engine. The first fuel injection (FI1) by the valve (6) is executed, and the second fuel injection (FI2) by the fuel injection valve is executed during the valve opening period of the intake valve.

  According to this configuration, since the first fuel injection in the valve closing period of the intake valve and the second fuel injection in the valve opening period are performed, a lean mixture with high homogeneity is formed in the intake passage by the first fuel injection. By performing the second fuel injection during the subsequent intake valve opening period, it is possible to form a lean air-fuel mixture with higher homogeneity in the combustion chamber.

According to a fourth aspect of the present invention, in the combustion control apparatus for an internal combustion engine according to any one of the first to third aspects, the predetermined lean air-fuel ratio (AFL1) is determined by an amount of NOx discharged from the combustion chamber. It is set so that it may become below an allowable upper limit.

  According to this configuration, the predetermined lean air-fuel ratio is set so that the amount of NOx discharged from the combustion chamber is less than or equal to the allowable upper limit value, so that the air-fuel ratio is controlled to an air-fuel ratio range leaner than the predetermined lean air-fuel ratio. By doing so, the NOx emission amount can be surely reduced below the allowable upper limit value.

It is a figure which shows the structure of the internal combustion engine and its control apparatus concerning one Embodiment of this invention. It is a figure for demonstrating arrangement | positioning of the tumble flow control valve (4) provided in an intake passage. It is a circuit diagram which shows the structure of the ignition circuit unit (7) corresponding to one cylinder. It is a figure for demonstrating the injection state of the fuel injected by a fuel injection valve (6). It is a figure which shows the valve opening time of a fuel injection valve (6), valve opening time (TFI), and the lift amount (LFT) at the time of valve opening. It is a figure which shows the relationship between an air fuel ratio (AF) and NOx density | concentration (CNOx) in exhaust_gas | exhaustion. It is a figure which shows transition of the cylinder pressure (PCYL) in a compression stroke and a combustion stroke. It is a figure which shows the table which sets the discharge start time (CAIG) and discharge continuation time (TSPK) of a spark plug according to the target air fuel ratio (AFCMD) of an ultra lean mixture.

Embodiments of the present invention will be described below with reference to the drawings.
FIG. 1 is a diagram showing the configuration of an internal combustion engine (hereinafter referred to as “engine”) and its control device according to an embodiment of the present invention. For example, a throttle valve 3 is placed in the middle of an intake passage 2 of a four-cylinder engine 1. Is arranged. A fuel injection valve 6 is provided on the downstream side of the throttle valve 3 in the intake passage 2, and its operation is controlled by an electronic control unit (hereinafter referred to as “ECU”) 5.

  As shown in FIG. 2, a partition wall 2a and a tumble flow control valve 4 disposed in one flow path formed by the partition wall 2a are provided immediately upstream of the intake valve 21 in the intake passage 2. The tumble flow control valve 4 is configured to be opened and closed by an actuator 4a. The actuator 4a is connected to the ECU 5, and its operation is controlled by the ECU 5. The tumble flow control valve 4 generates a tumble flow of the air-fuel mixture in the combustion chamber 1a.

  A spark plug 8 is attached to each cylinder of the engine 1, and the spark plug 8 is connected to the ECU 5 via the ignition circuit unit 7. The ECU 5 controls the discharge start timing CAIG and the discharge duration time TSPK in the spark plug 8 as will be described later.

  The ECU 5 includes an intake air flow sensor 11 that detects the intake air flow rate GAIR of the engine 1, an intake air temperature sensor 12 that detects the intake air temperature TA, a throttle valve opening sensor 13 that detects the throttle valve opening TH, and an intake pressure PBA. An intake pressure sensor 14 to detect, a coolant temperature sensor 15 to detect the engine coolant temperature TW, and other sensors (not shown) are connected, and detection signals from these sensors are supplied to the ECU 5.

  A crank angle position sensor 16 that detects a rotation angle of a crankshaft (not shown) of the engine 1 is connected to the ECU 5, and a pulse signal corresponding to the rotation angle of the crankshaft is supplied to the ECU 5. The crank angle position sensor 16 outputs a plurality of pulse signals indicating the crank angle position. These pulse signals are used for various timing controls such as fuel injection timing, ignition timing (discharge start timing of the spark plug 8), and the like. Used to detect the engine speed NE.

  A three-way catalyst 10 for exhaust purification is provided in the exhaust passage 9. A proportional oxygen concentration sensor 17 (hereinafter referred to as “LAF sensor 17”) is mounted on the upstream side of the three-way catalyst 10 and on the downstream side of the collection portion of the exhaust manifold communicating with each cylinder. 17 outputs a detection signal substantially proportional to the oxygen concentration (air-fuel ratio AF) in the exhaust gas and supplies it to the ECU 5.

  The ECU 5 shapes input signal waveforms from various sensors, corrects the voltage level to a predetermined level, converts an analog signal value into a digital signal value, a central processing unit (CPU), It comprises a storage circuit for storing various calculation programs executed by the CPU and calculation results, an output circuit for supplying a drive signal to the fuel injection valve 6, the ignition circuit unit 7, the actuator 4a, and the like.

  The fuel injection amount by the fuel injection valve 6 is controlled by correcting the basic fuel amount calculated according to the intake air flow rate GAIR with the air-fuel ratio correction coefficient KAF corresponding to the air-fuel ratio AF detected by the LAF sensor 17. The The air-fuel ratio correction coefficient KAF is calculated so that the detected air-fuel ratio AF coincides with the target air-fuel ratio AFCMD.

  FIG. 3 is a circuit diagram showing the configuration of the ignition circuit unit 7 corresponding to one cylinder. The ignition circuit unit 7 includes a first coil pair 71 including a primary coil 71a and a secondary coil 71b, a primary coil 72a, A second coil pair 72 including a secondary coil 72b, a booster circuit 73 that boosts the output voltage VBAT of the battery 30 and outputs a boosted voltage VUP, transistors Q1 and Q2 that control energization of the primary coils 71a and 72a, Diodes D1 and D2 connected between the secondary coils 71b and 72b and the spark plug 8 are provided.

  The bases of the transistors Q1 and Q2 are connected to the ECU 5, and on / off control (primary coil energization control) is performed by the ECU 5. It is possible to change the discharge duration (discharge duration) TSPK in the spark plug 8 according to the operating state of the engine 1 by alternately conducting energization while overlapping a part of the energization periods of the two primary coils. it can. Further, the first energization end timing of the primary coil corresponds to the discharge start timing CAIG, and the discharge start timing CAIG can also be changed according to the operating state of the engine 1.

  The fuel injection valve 6 is capable of atomizing and injecting fuel, and has an SMD (Sauter Mean Diameter) characteristic of about 35 μm (injection at a fuel pressure of 350 kPa and SMD 50 mm below the injection port). And a valve that can change the opening degree (lift amount) of the valve in at least two stages. FIG. 4A schematically shows a fuel injection state (a state in which the injected fuel is diffused) by the fuel injection valve 6, and FIG. 4B shows a fuel by a normal fuel injection valve shown for comparison. The injection state of is shown. In the normal fuel injection valve, the fuel concentration in the peripheral portion of the fuel distributed in a conical shape is high, whereas in the fuel injection valve 6, the reach of the atomized fuel is shortened and the concentration distribution in the diffusion region is reduced. High homogeneity (small density difference).

  FIG. 5 is a diagram showing the valve opening timing, the valve opening time TFI, and the lift amount LFT when the fuel injection valve 6 is opened, and the horizontal axis is the crank angle CRA. In the present embodiment, the first fuel injection FI1 is executed in the compression stroke within one combustion cycle, and the second fuel injection FI2 is executed in the intake stroke. Moreover, the first fuel injection FI1 is a fuel injection having a relatively large lift amount LFT1 and a relatively short valve opening time TFI1, and the second fuel injection FI2 has a lift amount LFT2 smaller than the lift amount LFT1 and a valve opening time. Let TFI2 be the fuel injection set longer than the valve opening time TFI1. The lift amounts LFT1, LFT2 and the valve opening times (fuel injection times) TFI1, TFI2 are set so that the total fuel injection amount in the first and second fuel injections FI1, FI2 becomes the fuel injection amount corresponding to the target air-fuel ratio AFCMD. Is set.

  A fuel injection valve 6 capable of atomizing and injecting fuel is opened as shown in FIG. 5 to form a relatively homogeneous air-fuel mixture in the intake passage 2 by the first fuel injection FI1, By executing the second fuel injection FI2, an amount of fuel required to make the detected air-fuel ratio AF coincide with the target air-fuel ratio AFCMD is supplied to the combustion chamber, and the air-fuel ratio distribution in the combustion chamber is substantially uniform. It is possible to form a homogeneous (high homogeneity) homogeneous lean mixture.

  It is desirable that the first fuel injection FI1 is executed in the compression stroke of the target cylinder, and the end timing of the second fuel injection FI2 is set immediately before the end timing of the intake stroke.

  In the present embodiment, the target air-fuel ratio AFCMD after completion of warm-up is set to a range of, for example, about “24” to “35” (hereinafter referred to as “ultra-lean air-fuel ratio range”). The air-fuel ratio “24” (predetermined air-fuel ratio AFL1) is set so that the NOx emission amount (NOx concentration in the exhaust gas) from the engine 1 is equal to or less than the allowable upper limit value CNOxHL (for example, 120 ppm). The air / fuel ratio “35” is an air / fuel ratio set as a limit value for obtaining a necessary engine output.

  FIG. 6 is a diagram showing the relationship between the air-fuel ratio AF and the NOx concentration CNOx in the exhaust gas. When the air-fuel ratio AF is “16” or more, the NOx increases as the air-fuel ratio AF increases (lean). The concentration CNOx decreases. Therefore, the predetermined air-fuel ratio AFL1 needs to be set to increase as the allowable upper limit value CNOxHL decreases.

  The discharge start timing CAIG in the spark plug 8 is set to a range of 50 degrees to 15 degrees before top dead center corresponding to the target air-fuel ratio AFCMD in the ultra-lean air-fuel ratio range, and the discharge duration TSPK is a homogeneous lean mixture. In order to ensure ignition, it is set to 1.8 to 3 msec. The boost voltage VUP is set so that the discharge energy when the discharge duration time TSPK is set in this way is 150 to 600 mJ. Conventional lean mixture combustion by spark ignition is stratified mixture combustion realized by generating a flow in the combustion chamber so that the air-fuel ratio in the vicinity of the spark plug becomes relatively small. In the homogeneous lean mixture combustion, the discharge start time CAIG is set from the ignition timing (for example, 8.0 degrees) of the stratified mixture combustion so that the discharge duration TSPK is set relatively long and the discharge duration TSPK can be secured. Set to the advance side.

  Furthermore, the geometrical compression ratio of the engine 1 (ratio of the combustion chamber volume when the piston is located at the bottom dead center and the combustion chamber volume when the piston is located at the top dead center) has a minimum effective compression ratio of 9.0. In order to achieve this, it is set slightly larger than the geometric compression ratio of a normal spark ignition engine.

Further, by changing the opening degree of the tumble flow control valve 4, tumble flow generation control for generating tumble flow with a flow rate of about 5 to 15 m / sec (flow rate when the engine speed NE is 1500 rpm) is performed.
By setting the discharge duration TSPK to be relatively long and generating a tumble flow in the combustion chamber, a strong initial flame kernel is formed, and the flame kernel is grown, so that an unburned mixture at the compression top dead center. By raising the temperature to 1000 ° K or higher, the elementary reaction that governs the combustion laminar flow rate is changed to a reaction in which hydrogen peroxide decomposes to generate OH radicals, ensuring combustion after compression top dead center Can be completed.

FIG. 7 is a diagram showing the transition of the in-cylinder pressure PCYL in the compression stroke and the combustion stroke. The solid line L1 corresponds to this embodiment, the thick broken line L2 corresponds to HCCI (homogeneous mixture compression ignition) combustion, and the thin broken line Corresponds to stoichiometric combustion (combustion when the air-fuel ratio is set to the stoichiometric air-fuel ratio). In this figure, the crank angle CRA of 0 degrees corresponds to the compression top dead center. It should be noted that the ignition timing in the case of stoichiometric combustion is set slightly ahead of the compression top dead center (for example, a range within a crank angle of 10 degrees).
It can be confirmed that highly efficient and stable combustion can be obtained by the combustion of the homogeneous lean air-fuel mixture of the present embodiment by spark ignition.

  In this embodiment, after the warm-up of the engine 1 is completed, the target air-fuel ratio AFCMD is set within the ultra-lean air-fuel ratio range according to the required torque of the engine 1, and the spark plug 8 is discharged according to the target air-fuel ratio AFCMD. The start time CAIG and the discharge duration TSPK are set.

  Specifically, the discharge start timing CAIG (defined by the advance amount from the compression top dead center) is set to advance as the target air-fuel ratio AFCMD increases, as shown in FIG. The discharge duration time TSPK is set so as to increase as the target air-fuel ratio AFCMD increases as shown in FIG. 8B. AFL1 and AFL2 shown in FIG. 8 are predetermined air-fuel ratios set to, for example, “24” and “35”, respectively. CAIG1 and CAIG2 shown in FIG. 8A are, for example, “15 degrees” and “50 degrees, respectively. , And TSPK1 and TSPK2 shown in FIG. 8B are predetermined discharge times set to, for example, “1.8 msec” and “3 msec”, respectively.

  It is desirable that the discharge start timing CAIG is set according to the target air-fuel ratio AFCMD as described above, and advanced as the engine speed NE increases. As the engine speed NE increases, the crank angle CRA corresponding to the discharge duration TSPK increases, so that the required discharge duration TSPK is secured even in a high rotation state by advancing the discharge start timing CAIG and stable. Combustion can be obtained.

  As described above, according to the present embodiment, a homogeneous mixture with an ultra lean air / fuel ratio can be formed in the combustion chamber and reliably ignited to complete combustion, so that the homogeneous lean mixture has good controllability. Spark ignition flame propagation combustion can be performed, and combustion with high efficiency and low NOx emission can be realized.

  More specifically, since the fuel atomized by the fuel injection valve 6 is injected into the intake passage 2, a relatively homogeneous lean air-fuel mixture is first formed in the intake passage 2 and further into the combustion chamber 1a. By being inhaled, a lean mixture with higher homogeneity can be formed. Further, since the discharge start timing CAIG and the discharge duration time TSPK in the spark plug 8 can be controlled, the discharge when the stratified lean air-fuel mixture is ignited by appropriately setting the discharge start timing CAIG and the discharge duration time TSPK. By setting the discharge start time CAIG on the advance side from the start time, the discharge duration time TSPK can be set longer and the homogeneous lean air-fuel mixture can be reliably ignited.

  Further, in the ultra-lean air-fuel ratio range from the predetermined air-fuel ratio AFL1 to AFL2 shown in FIG. 8, as the target air-fuel ratio AFCMD increases, the discharge start timing CAIG of the spark plug 8 is advanced and the discharge duration time TSPK is set longer. Since the control is performed, stable combustion can be realized corresponding to the change of the target air-fuel ratio AFCMD.

  Further, since the flow of the air-fuel mixture is generated in the combustion chamber using the tumble flow control valve 4, a powerful initial flame kernel, that is, compression top dead center, is set by setting the discharge duration time TPSK of the spark plug 8 and the air-fuel mixture flow. It is possible to form an initial flame nucleus that can raise the temperature of the combustion mixture at 1000 ° K or higher and realize high-efficiency combustion. That is, in the present embodiment, the air-fuel mixture flow in the combustion chamber is not generated to relatively reduce the air-fuel ratio in the vicinity of the spark plug, but forms a powerful initial flame kernel that completes combustion. Is generated.

  Further, a lean mixture with high homogeneity is formed in the intake passage 2 by the first fuel injection FI1 during the closing period of the intake valve 21, and combustion is performed by performing the second fuel injection FI2 in the subsequent intake valve opening period. In the chamber 1a, it is possible to form a lean air-fuel mixture with a higher degree of homogeneity.

  The predetermined air-fuel ratio AFL1, which is the lower limit value of the ultra-lean air-fuel ratio range, is set so that the amount of NOx discharged from the combustion chamber is less than or equal to the allowable upper limit value, so that the ultra-lean air on the lean side of the predetermined air-fuel ratio AFL1 is set. By controlling the air-fuel ratio within the fuel ratio range, the NOx emission amount can be reliably reduced below the allowable upper limit value.

  In the present embodiment, the ignition circuit unit 7 and the spark plug 8 correspond to spark ignition means, and the partition wall 2a and the tumble flow control valve 4 correspond to flow generation means. The ECU 5 constitutes an ignition control means and an air-fuel ratio control means, and the fuel injection valve 6 and the ECU 5 constitute a homogeneous lean mixture forming means.

  The present invention is not limited to the embodiment described above, and various modifications can be made. For example, in the above-described embodiment, a mechanism that generates a tumble flow is used as the flow generation means, but a mechanism that generates a swirl flow may be employed. Moreover, you may comprise the shape of the combustion chamber 1a and the piston top part so that a squish flow may be produced | generated. Further, although two coil pairs of the ignition circuit unit 7 are provided corresponding to one ignition plug, three or more coil pairs may be provided.

  In the above-described embodiment, an example of a four-cylinder engine is shown, but the present invention can be applied regardless of the number of cylinders. The present invention can also be applied to a combustion control device such as an engine for a marine propulsion device such as an outboard motor having a vertical crankshaft.

DESCRIPTION OF SYMBOLS 1 Internal combustion engine 2 Intake passage 2b Partition (flow production | generation means)
4 Tumble flow control valve (flow generation means)
5 Electronic control unit (ignition control means, air-fuel ratio control means, homogeneous lean mixture formation means)
6 Fuel injection valve (homogeneous lean mixture formation means)
7 Ignition circuit unit (spark ignition means)
8 Spark plug (spark ignition means)
CAIG Discharge start time TSPK Discharge duration AFCMD Target air-fuel ratio

Claims (4)

Spark ignition means for performing spark ignition of the air-fuel mixture in the combustion chamber of the internal combustion engine, ignition control means for controlling the spark ignition means, and air-fuel ratio control means for controlling the air-fuel ratio of the air-fuel mixture in the combustion chamber, In a combustion control device for an internal combustion engine,
A homogeneous lean mixture forming means for forming a homogeneous and lean mixture in the combustion chamber;
The spark ignition means includes an ignition plug and a plurality of ignition coil pairs for generating a discharge in the ignition plug, and is capable of changing a discharge start timing and duration in the ignition plug,
The homogeneous lean air-fuel mixture forming means injects the fuel into the intake passage of the engine using a fuel injection valve capable of atomizing and injecting the fuel,
The air-fuel ratio control means controls the air-fuel ratio to a leaner air-fuel ratio than a predetermined lean air-fuel ratio leaner than the stoichiometric air-fuel ratio,
The ignition control means controls the discharge start timing and the discharge duration in the spark plug, and the discharge start timing is formed so that the air-fuel ratio of the air-fuel mixture in the vicinity of the spark plug becomes relatively small. Set to the advance side from the discharge start timing when igniting the lean mixture ,
In the air-fuel ratio range leaner than the predetermined lean air-fuel ratio, the ignition control means advances the discharge start timing and sets the discharge duration time longer as the air-fuel ratio increases. Combustion control device for an internal combustion engine. The combustion control device for an internal combustion engine according to claim 1, further comprising a flow generation means for generating a flow of the air-fuel mixture sucked into the combustion chamber. The homogeneous lean air-fuel mixture forming means executes the first fuel injection by the fuel injection valve during the closing period of the intake valve of the engine, and the second fuel injection by the fuel injection valve during the valve opening period of the intake valve. The combustion control apparatus for an internal combustion engine according to claim 1 or 2 , wherein The combustion of the internal combustion engine according to any one of claims 1 to 3 , wherein the predetermined lean air-fuel ratio is set so that an amount of NOx discharged from the combustion chamber is not more than an allowable upper limit value. Control device.

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