JP2016023632A - engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide an engine 1 capable of easily calculating actual torque TQ of the engine 1 while suppressing influence by fluctuation of a load.SOLUTION: An engine 1 including a first supercharger 6 and a second supercharger 10 as superchargers, and a supercharger rotation speed detection sensor 22 as supercharger rotation speed detection means calculates actual torque TQ of the engine 1 at optional actual engine speed N from an actual supercharger rotation speed NT, a no-load supercharger rotation speed NT0, and a torque increase amount ΔNQ of the engine 1 per a unit increase amount ΔNT of the actual supercharger rotation speed NT.SELECTED DRAWING: Figure 1

Description

  The present invention relates to an engine. Specifically, it relates to a turbocharged engine.

  2. Description of the Related Art Conventionally, an engine that calculates an output torque based on an amplitude (variation width) of an angular velocity of a crankshaft is known. The engine is configured to calculate the output torque of the engine based on the angular velocity fluctuation range with respect to a predetermined load that is measured in advance for each rotational speed of the engine. Thereby, the engine can calculate the output torque only by measuring the angular velocity fluctuation range. For example, as described in Patent Document 1.

  The engine described in Patent Document 1 calculates the output torque of the engine by detecting a change in the angular speed of the crankshaft of the engine. However, the factors that cause fluctuations in the angular velocity of the crankshaft of the engine include not only due to the explosion timing of the engine cylinders but also due to fluctuations in the load. Therefore, when the load variation is large, the influence of the load variation included in the calculated engine output torque may increase.

JP 2007-303382 A

  The present invention has been made in view of the situation as described above, and an object of the present invention is to provide an engine that can easily calculate the actual torque of the engine while suppressing the influence of load fluctuations.

  That is, in the present invention, an engine including a supercharger and a supercharger rotation speed detection means, and at an arbitrary engine rotation speed, an actual supercharger rotation speed, a supercharger rotation speed in an unloaded state, The actual engine torque is calculated from the engine torque increase amount per unit increase amount of the turbocharger rotation speed.

  The present invention further includes a cooling water temperature detecting means, an intake air temperature detecting means, an atmospheric pressure detecting means, and a diesel particulate filter differential pressure detecting means, wherein the actual torque is the cooling water temperature, the intake air temperature The correction is made based on the temperature, the atmospheric pressure, and the diesel particulate filter differential pressure.

  The present invention further includes an EGR device that recirculates exhaust gas to intake air, and an EGR valve opening degree detecting means that adjusts an exhaust gas flow rate for recirculation, and the actual torque is corrected based on the opening degree of the EGR valve. It is what is done.

  The present invention further includes a bypass pipe that bypasses a part of the exhaust gas supplied to the supercharger, and a bypass valve that controls the flow rate of the exhaust gas flowing through the bypass pipe, wherein the actual torque is generated when the bypass valve is opened. It is corrected based on the degree.

  Further, in the present invention, the supercharger is composed of a variable capacity supercharger that changes the supercharger rotation speed by changing the vane opening, and the actual torque is corrected based on the vane opening. It is what is done.

  As effects of the present invention, the following effects can be obtained.

  That is, according to the present invention, the actual torque of the engine is calculated based on the exhaust flow rate discharged according to the actual torque output by the engine. Thereby, it is possible to easily calculate the output torque of the engine while suppressing the influence due to the fluctuation of the load.

  According to the present invention, the actual torque of the engine is calculated without being affected by the operating environment of the engine. Thus, it is possible to easily calculate the engine output torque while suppressing the influence due to the state fluctuation on the engine side and the influence due to the load fluctuation.

  According to the present invention, the actual torque of the engine is calculated without being affected by the configuration of the engine. Thus, it is possible to easily calculate the engine output torque while suppressing the influence due to the state fluctuation on the engine side and the influence due to the load fluctuation.

Schematic which showed the structure of the engine which concerns on 1st embodiment of this invention. The figure which shows the map showing the relationship between the real supercharger rotational speed and the actual torque in arbitrary real engine rotational speeds with the engine which concerns on 1st embodiment of this invention. (A) The figure which shows the map showing the correction coefficient with respect to the cooling water temperature in the engine which concerns on 1st embodiment. (B) The figure which shows the map showing the correction coefficient with respect to the intake air temperature in the engine which concerns on 1st embodiment. (A) The figure which shows the map showing the correction coefficient with respect to atmospheric pressure in the engine which concerns on 1st embodiment (b) The figure which shows the map showing the correction coefficient with respect to DPF differential pressure in the engine which concerns on 1st embodiment. The figure which shows the map showing the correction coefficient with respect to the bypass valve opening degree in the engine which concerns on 1st embodiment. The figure which shows the flowchart showing the control aspect for calculating the actual torque of the engine which concerns on 1st embodiment. Schematic which showed the structure of the engine which concerns on 2nd embodiment of this invention. The figure which shows the map showing the correction coefficient with respect to the EGR valve opening degree in the engine which concerns on 2nd embodiment. The figure which shows the flowchart showing the control aspect for calculating the actual torque of the engine which concerns on 3rd embodiment. (A) Schematic which showed the operation | movement aspect in the fully open state of the variable nozzle vane of the variable capacity | capacitance supercharger provided in the engine which concerns on 3rd embodiment of this invention (b) The variable nozzle vane of a variable capacity supercharger is also the same Schematic which showed the operation | movement aspect in a fully closed state. The figure which shows the map showing the correction coefficient with respect to the nozzle vane opening degree in the engine which concerns on 3rd embodiment. The figure which shows the flowchart showing the control aspect for calculating the actual torque of the engine which concerns on 3rd embodiment.

  Below, the engine 1 which is an inline 6 cylinder diesel engine provided with the 1st supercharger 6 and the 2nd supercharger 10 which concern on 1st embodiment of this invention using FIG. 1 is demonstrated. In the present embodiment, the engine 1 is an in-line six-cylinder engine including the first supercharger 6 and the second supercharger 10, but the type of engine, the number of cylinders, and the number of superchargers are limited. It is not a thing. Further, in the present embodiment, the engine 1 is provided with the bypass valve 16, but can be applied to an engine that does not include the bypass valve 16.

  As shown in FIG. 1, the engine 1 is a diesel engine. In the present embodiment, the engine 1 is an in-line six-cylinder engine 1 having six cylinders 3. The engine 1 mixes and burns the air supplied through the intake device 2 and the fuel supplied from the six fuel injection valves 4 at an appropriate ratio inside each cylinder 3 under the control of the ECU 24. Rotate the output shaft. The engine 1 discharges the intake air generated by the combustion of fuel to the outside through the exhaust device 5.

  The intake device 2 includes intake pipes 2a, 2b, 2c, and 2d, a first compressor unit 8 of the first supercharger 6, a second compressor unit 12 of the second supercharger 10, a first intercooler 14, and a second intercooler 15. Consists of The exhaust device 5 includes exhaust pipes 5 a, 5 b, 5 c, and 5 d, a first turbine unit 7 of the first supercharger 6, and a second turbine unit 11 of the second supercharger 10. The exhaust device 5 communicates the exhaust pipe 5a and the exhaust pipe 5b with a bypass pipe 5d via a bypass valve 16.

  The engine 1 is connected to the first intercooler 14 via the intake pipe 2d of the intake device 2. The engine 1 is connected to the second turbine unit 11 of the second supercharger 10 via the exhaust pipe 5 a of the exhaust device 5.

  The first supercharger 6 compresses and compresses intake air using the exhaust pressure of the intake air as a drive source. The first supercharger 6 includes a first turbine unit 7 and a first compressor unit 8. The first turbine unit 7 is configured to be rotatable by the exhaust pressure of the exhaust gas supplied from the second turbine unit 11 of the second supercharger 10 via the exhaust pipe 5b. The first turbine section 7 is communicated to the outside via the exhaust pipe 5c, and is configured to be able to discharge the intake air to the outside.

  The first compressor unit 8 is connected to the first turbine unit 7 by a connecting shaft 9 and is configured to be rotatable. The first compressor unit 8 is configured to be able to suck in air by rotation and be compressed under pressure. The first compressor unit 8 is configured to be able to suck external air. The first compressor unit 8 is connected to the first intercooler 14 via the intake pipe 2a.

  The second supercharger 10 pressurizes and compresses the supply air compressed and compressed by the first supercharger 6 using the exhaust pressure as a drive source. The second supercharger 10 includes a second turbine unit 11 and a second compressor unit 12. Further, the second turbine unit 11 is configured to be rotatable by the exhaust pressure of the exhaust supplied from the engine 1 through the exhaust pipe 5a. The second turbine unit 11 is connected to the first turbine unit 7 of the first supercharger 6 through the exhaust pipe 5b.

  The second compressor unit 12 is connected to the second turbine unit 11 by a connecting shaft 13 and is configured to be rotatable. The second compressor unit 12 is configured to suck air by rotation and be compressed by pressure. The second compressor unit 12 is connected to the second intercooler 15 through the intake pipe 2c.

  The 1st intercooler 14 and the 2nd intercooler 15 cool supply air. The first intercooler 14 and the second intercooler 15 cool the supply air by exchanging heat between the coolant supplied by a cooling water pump (not shown) and the supply air. The first intercooler 14 is connected to the second compressor unit 12 of the second supercharger 10 via the intake pipe 2b. The second intercooler 15 is connected to the engine 1 via the intake pipe 2d.

  The exhaust device 5 includes an exhaust pipe 5 a on the upstream side of the second turbine unit 11 and a downstream side of the second turbine unit 11 in order to bypass exhaust gas supplied to the second turbine unit 11 of the second supercharger 10. A bypass pipe 5d communicating with the exhaust pipe 5b is provided. A bypass valve 16 (waist gate valve) is provided in the middle of the bypass pipe 5d. The bypass valve 16 limits the flow rate of the exhaust gas that passes through the bypass pipe 5d. The bypass valve 16 is composed of a normally closed electromagnetic flow control valve. The bypass valve 16 can change the opening degree by acquiring a signal from the ECU 24 described later. In the present embodiment, the bypass valve 16 is composed of a normally closed electromagnetic flow control valve. However, any valve that can limit the flow rate of exhaust gas may be used.

  The engine 1 is provided with an engine rotation speed detection sensor 17. The engine rotation speed detection sensor 17 detects the actual engine rotation speed N of the engine 1. The engine rotation speed detection sensor 17 includes a sensor and a pulsar and is provided on the output shaft of the engine 1. In this embodiment, the engine rotation speed detection sensor 17 is composed of a sensor and a pulsar. However, any sensor capable of detecting the actual engine rotation speed N may be used.

  Further, the engine 1 is provided with a fuel injection pressure detection sensor 18. The fuel injection pressure detection sensor 18 detects the fuel injection pressure Pf of the fuel injected from the fuel injection valve 4. The fuel injection pressure detection sensor 18 is provided in the middle of the fuel supply pipe 4a. In the present embodiment, the fuel injection pressure detection sensor 18 may be any sensor that can detect the fuel injection pressure Pf.

  Further, the engine 1 is provided with a cooling water temperature detection sensor 19 which is a cooling water temperature detection means. The cooling water temperature detection sensor 19 detects the cooling water temperature Tw of the cooling water that is cooling the cylinder block and the cylinder head that constitute the main body of the engine 1. The cooling water temperature detection sensor 19 is provided in the cooling water flow path of the engine 1. In the present embodiment, the cooling water temperature detection sensor 19 may be any sensor that can detect the cooling water temperature Tw.

  Further, the intake device 2 is provided with an intake air temperature detection sensor 20 as an intake air temperature detection means in the vicinity of the outside air intake port of the first compressor unit 8 of the first supercharger 6. The intake air temperature detection sensor 20 detects the intake air temperature Ti. In the present embodiment, the intake air temperature detection sensor 20 may be any sensor that can detect the intake air temperature Ti.

  Further, the intake device 2 is provided with an atmospheric pressure detection sensor 21 which is an atmospheric pressure detection means. The atmospheric pressure detection sensor 21 detects the atmospheric pressure Pa. In the present embodiment, the atmospheric pressure detection sensor 21 may be any sensor that can detect the atmospheric pressure Pa.

  Further, the intake device 2 is provided with an atmospheric pressure detection sensor 21 which is an atmospheric pressure detection means. The atmospheric pressure detection sensor 21 detects the atmospheric pressure Pa. In the present embodiment, the atmospheric pressure detection sensor 21 may be any sensor that can detect the atmospheric pressure Pa.

  The intake device 2 is provided with a supercharger rotation speed detection sensor 22 which is a supercharger rotation speed detection means. The supercharger rotation speed detection sensor 22 detects the actual supercharger rotation speed NT. In the present embodiment, the supercharger rotation speed detection sensor 22 may be any sensor that can detect the actual supercharger rotation speed NT.

  In the exhaust device 5, DPF differential pressure detection sensors 23, which are diesel particulate filter (hereinafter simply referred to as “DPF”) differential pressure detection means, are provided on an inlet side and an outlet side of a DPF (not shown). The DPF differential pressure detection sensor 23 detects the DPF differential pressure Pdpf. In the present embodiment, the DPF differential pressure detection sensor 23 may be any sensor that can detect the DPF differential pressure Pdpf.

  As described above, in the intake device 2, the first compressor section 8, the intake pipe 2 a, the first intercooler 14, the intake pipe 2 b, the second compressor section 12, the intake pipe 2 c, and the second intercooler 15 are arranged in the engine 1 in order from the upstream side. Connected and configured. The exhaust device 5 is configured by connecting an exhaust pipe 5a, a second turbine unit 11, an exhaust pipe 5b, a first turbine unit 7, and an exhaust pipe 5c in order from the upstream side. Further, the exhaust device 5 communicates the exhaust pipe 5a and the exhaust pipe 5b with a bypass pipe 5d via a bypass valve 16.

  The ECU 24 controls the engine 1. Specifically, the engine 1 main body and the bypass valve 16 are controlled. The ECU 24 stores various programs and data for controlling the engine 1. The ECU 24 may have a configuration in which a CPU, ROM, RAM, HDD, and the like are connected by a bus, or may be configured by a one-chip LSI or the like.

  The ECU 24 is connected to the fuel injection valves 4, 4, 4, 4, 4, and 4, and injects an arbitrary fuel injection amount F at an arbitrary fuel injection pressure Pf at an arbitrary fuel injection timing θ.・ 4 ・ 4 ・ 4 ・ 4 ・ 4 can be controlled.

  The ECU 24 is connected to the bypass valve 16 and can control the bypass valve 16 to an arbitrary bypass valve opening Vb.

  The ECU 24 is connected to the engine rotation speed detection sensor 17 and can acquire the actual engine rotation speed N detected by the engine rotation speed detection sensor 17.

  The ECU 24 is connected to the fuel injection pressure detection sensor 18 and can acquire the fuel injection pressure Pf detected by the fuel injection pressure detection sensor 18.

  The ECU 24 is connected to the cooling water temperature detection sensor 19 and can acquire the cooling water temperature Tw detected by the cooling water temperature detection sensor 19.

  The ECU 24 is connected to the intake air temperature detection sensor 20 and can acquire the intake air temperature Ti detected by the intake air temperature detection sensor 20.

  The ECU 24 is connected to the atmospheric pressure detection sensor 21 and can acquire the atmospheric pressure Pa detected by the atmospheric pressure detection sensor 21.

  The ECU 24 is connected to the DPF differential pressure detection sensor 23, and can acquire the DPF differential pressure Pdpf detected by the DPF differential pressure detection sensor 23.

  The ECU 24 is connected to the supercharger rotation speed detection sensor 22 and can acquire the actual supercharger rotation speed NT detected by the supercharger rotation speed detection sensor 22.

  The ECU 24 stores, for each actual engine speed N, an actual torque map M0 for calculating the actual torque TQ output from the engine 1 in the case of the actual turbocharger speed NT under predetermined operating conditions. Yes. The ECU 24 also calculates a coolant temperature correction map M1 for calculating the correction coefficient Cf1 of the actual torque TQ at an arbitrary coolant temperature Tw, and a correction coefficient Cf2 of the actual torque TQ at an arbitrary intake air temperature Ti. Intake temperature correction map M2, atmospheric pressure correction map M3 for calculating correction coefficient Cf3 of actual torque TQ at an arbitrary atmospheric pressure Pa, DPF for calculating correction coefficient Cf4 of actual torque TQ at an arbitrary DPF differential pressure Pdpf A differential pressure correction map M4 and a bypass valve correction map M5 for calculating a correction coefficient Cf5 of the actual torque TQ at an arbitrary bypass valve opening Vb are stored. That is, the ECU 24 is configured to correct the correction coefficient Cf1 to the correction coefficient Cf5 when the operation condition of the engine 1 is different from the operation condition of the engine 1 when the actual torque map M0 is set.

  Next, the actual torque map M0 will be described in detail with reference to FIG.

実過給機回転速度ＮＴを算出する数１は、以下の数２に示すように、実過給機回転速度ＮＴからエンジン１の実トルクＴＱ（補正前実トルクＴＱ０）を算出する式に変形される。実トルクＴＱ（補正前実トルクＴＱ０）は、数２に示すように、実過給機回転速度ＮＴと無負荷過給機回転速度ＮＴ０との差に回転速度増加量ΔＮＴに対するトルク増加量ΔＴＱの比を乗じることで算出される。そして、実トルクマップＭ０は、数２に基づいて、予め試験等で取得した所定の運転条件での第２過給機１０の無負荷過給機回転速度ＮＴ０および回転速度増加量ΔＮＴに対するトルク増加量ΔＴＱの比が取得した実エンジン回転速度Ｎ毎に定められている。従って、ＥＣＵ２４は、取得した任意の実エンジン回転速度Ｎと、実過給機回転速度ＮＴとから実トルクマップＭ０に基づいてエンジン１の実トルクＴＱ（補正前実トルクＴＱ０）を算出することが可能である。

As shown in FIG. 2, the actual turbocharger rotational speed NT increases in proportion to the actual torque TQ of the engine 1. That is, the actual turbocharger rotational speed NT is a supercharger rotational speed when the engine 1 is in an unloaded state at an arbitrary actual engine rotational speed N under a predetermined operating condition, as shown in the following formula 1. A value obtained by multiplying the ratio of the load supercharger rotational speed NT0 (engine rotational speed component) and the turbocharger rotational speed increase amount ΔNT to the torque increase amount ΔNT of the engine 1 by the actual torque TQ (actual torque TQ0 before correction). Torque component).

Formula 1 for calculating the actual turbocharger rotational speed NT is transformed into an equation for calculating the actual torque TQ (the actual torque TQ0 before correction) of the engine 1 from the actual turbocharger rotational speed NT, as shown in the following formula 2. Is done. As shown in Equation 2, the actual torque TQ (actual torque TQ0 before correction) is the difference between the actual turbocharger rotational speed NT and the no-load supercharger rotational speed NT0, and the torque increase amount ΔTQ relative to the rotational speed increase amount ΔNT. Calculated by multiplying by the ratio. The actual torque map M0 is based on the equation (2), and the torque increase with respect to the no-load supercharger rotational speed NT0 and the rotational speed increase amount ΔNT of the second supercharger 10 under the predetermined operating condition acquired in advance by a test or the like. The ratio of the amount ΔTQ is determined for each acquired actual engine speed N. Therefore, the ECU 24 can calculate the actual torque TQ (the actual torque TQ0 before correction) of the engine 1 based on the actual torque map M0 from the acquired arbitrary actual engine rotation speed N and the actual turbocharger rotation speed NT. Is possible.

  Next, the cooling water temperature correction map M1, the intake air temperature correction map M2, the atmospheric pressure correction map M3, the DPF differential pressure correction map M4, and the bypass valve correction map M5 will be described in detail with reference to FIGS. In each of the correction maps M1, M2, M3, M4, and M5, correction coefficients Cf1, Cf2, Cf3, Cf4, and Cf5 under predetermined operating conditions are set to 1. In the present embodiment, the correction coefficient is determined from each preset correction map. However, the present invention is not limited to this, and the correction formula may be determined for each correction coefficient and calculated from the correction formula.

  As shown in FIG. 3A, the cooling water temperature correction map M1 determines a correction coefficient Cf1 for an arbitrary cooling water temperature Tw. When the cooling water temperature Tw is higher than the cooling water temperature Tws under a predetermined operating condition, the actual torque TQ increases due to a reduction in torque loss due to the viscosity or friction of the lubricating oil. Therefore, when the coolant temperature Tw is higher than the coolant temperature Tws, the coolant temperature correction map M1 is set to a value where the correction coefficient Cf1 is larger than 1. On the other hand, when the cooling water temperature Tw is lower than the cooling water temperature Tws, the actual torque TQ decreases due to an increase in torque loss caused by the viscosity or friction of the lubricating oil. Therefore, when the cooling water temperature Tw is lower than the cooling water temperature Tws, the correction coefficient Cf1 is set to a value smaller than 1 in the cooling water temperature correction map M1.

  As shown in FIG. 3B, the intake air temperature correction map M2 determines a correction coefficient Cf2 for an arbitrary intake air temperature Ti. When the intake air temperature Ti is higher than the intake air temperature Tis under a predetermined operating condition, the actual torque TQ decreases due to a decrease in oxygen concentration per unit volume during intake air. Therefore, when the intake air temperature Ti is higher than the intake air temperature Tis, the correction coefficient Cf2 is set to a value smaller than 1 in the intake air temperature correction map M2. On the other hand, when the intake air temperature Ti is lower than the intake air temperature Tis, the actual torque TQ increases due to an increase in the oxygen concentration per unit volume in the intake air. Therefore, when the intake air temperature Ti is lower than the intake air temperature Tis, the intake air temperature correction map M2 is set to a value where the correction coefficient Cf2 is larger than 1.

  As shown in FIG. 4A, the atmospheric pressure correction map M3 determines a correction coefficient Cf3 for an arbitrary atmospheric pressure Pa. When the atmospheric pressure Pa is lower than the atmospheric pressure Pas (1 atm) under a predetermined operating condition, the actual torque TQ decreases due to a decrease in oxygen concentration per unit volume during intake. Therefore, when the atmospheric pressure Pa is lower than the atmospheric pressure Pas (1 atmospheric pressure), the atmospheric pressure correction map M3 is set to a value where the correction coefficient Cf3 is smaller than 1.

  As shown in FIG. 4B, the DPF differential pressure correction map M4 determines a correction coefficient Cf4 for an arbitrary DPF differential pressure Pdpf. When the DPF differential pressure Pdpf is higher than the DPF differential pressure Pdpfs under a predetermined operating condition, the actual torque TQ is caused by a decrease in the actual supercharger rotational speed NT caused by an increase in the outlet back pressure of the second supercharger 10. Decrease. Therefore, when the DPF differential pressure Pdpf is higher than the DPF differential pressure Pdpfs, the correction coefficient Cf4 is set to a value smaller than 1 in the DPF differential pressure correction map M4. On the other hand, when the DPF differential pressure Pdpf is lower than the DPF differential pressure Pdpfs, the actual torque TQ is increased due to the increase in the actual turbocharger rotational speed NT caused by the decrease in the outlet back pressure of the second supercharger 10. Will increase. Therefore, when the DPF differential pressure Pdpf is lower than the DPF differential pressure Pdpfs, the correction coefficient Cf4 is set to a value larger than 1 in the DPF differential pressure correction map M4.

  As shown in FIG. 5, the bypass valve correction map M5 determines a correction coefficient Cf5 for an arbitrary bypass valve opening Vb. When the bypass valve opening degree Vb is larger than the bypass valve opening degree Vbs (closed state) under a predetermined operating condition, the actual torque TQ is caused by a decrease in the actual turbocharger rotational speed NT caused by an increase in the exhaust bypass amount. Decrease. Therefore, when the bypass valve opening degree Vb is larger than the bypass valve opening degree Vbs (closed state), the bypass valve correction map M5 has the correction coefficient Cf5 set to a value smaller than 1.

  Next, the flow of intake air (supply air) and exhaust gas will be described with reference to FIG.

  As shown in FIG. 1, the exhaust from the engine 1 is supplied to the second turbine unit 11 of the second supercharger 10 via the exhaust pipe 5a. The second turbine unit 11 is rotated by the exhaust pressure of the exhaust. The rotational power of the second turbine unit 11 is transmitted to the second compressor unit 12 via the connecting shaft 13. The second compressor unit 12 is rotated by the rotational power transmitted from the second turbine unit 11. The exhaust gas supplied to the second turbine unit 11 is discharged from the second supercharger 10 through the exhaust pipe 5b.

  When the bypass valve 16 is in an open state, a part of the exhaust from the engine 1 flows into the bypass pipe 5d that communicates the exhaust pipe 5a and the exhaust pipe 5b. The exhaust gas flowing into the bypass pipe 5d is not supplied to the second turbine section 11 of the second supercharger 10 (bypassing the second turbine section 11), and is discharged from the first supercharger 6 via the exhaust pipe 5b. It is supplied to the first turbine unit 7. That is, the flow rate of the exhaust gas supplied to the second turbine unit 11 is changed depending on the opening degree of the bypass valve 16.

  The exhaust discharged from the second supercharger 10 is supplied to the first turbine section 7 of the first supercharger 6 through the exhaust pipe 5b. The first turbine unit 7 is rotated by the exhaust pressure of the exhaust. The rotational power of the first turbine unit 7 is transmitted to the first compressor unit 8 via the connecting shaft 9. The first compressor unit 8 is rotated by the rotational power transmitted from the first turbine unit 7. Exhaust gas supplied to the first turbine section 7 is discharged to the outside through an exhaust pipe 5c, a purification device (not shown), and the like.

  External air is sucked and pressurized and compressed by the first compressor unit 8 rotated by the rotational power from the first turbine unit 7 of the first supercharger 6. At this time, the intake air is pressurized and compressed, so that compression heat is generated and the temperature rises. The supply air compressed and compressed by the first compressor unit 8 is discharged from the first supercharger 6 through the intake pipe 2a.

  The supply air discharged from the first supercharger 6 is supplied to the first intercooler 14 through the intake pipe 2a to be cooled. The supply air supplied to the first intercooler 14 is discharged from the first intercooler 14 through the intake pipe 2b.

  The supply air discharged from the first intercooler 14 is supplied to the second compressor unit 12 of the second supercharger 10 through the intake pipe 2b. The supplied air is sucked and pressurized and compressed by the second compressor section 12 rotated by the rotational power from the second turbine section 11 of the second supercharger 10. At this time, the supply air is pressurized and compressed to generate compression heat and the temperature rises. The supply air compressed and compressed by the second compressor unit 12 is discharged from the second supercharger 10 through the intake pipe 2c.

  The supply air discharged from the second supercharger 10 is supplied to the second intercooler 15 through the intake pipe 2c and cooled. The air supplied to the second intercooler 15 is discharged from the second intercooler 15 through the intake pipe 2d. The supply air discharged from the second intercooler 15 is supplied to the engine 1 through the intake pipe 2d.

  Hereinafter, a control mode for calculating the actual torque TQ in the engine 1 according to the first embodiment of the present invention will be described with reference to FIG.

  As shown in FIG. 6, in step S110, the ECU 24 acquires the actual engine speed N detected by the engine speed sensor 17 and the actual turbocharger speed NT detected by the supercharger speed sensor 22. , The process proceeds to step S120.

  In step S120, the ECU 24 detects the cooling water temperature Tw detected by the cooling water temperature detection sensor 19, the intake air temperature Ti detected by the intake air temperature detection sensor 20, the atmospheric pressure Pa detected by the atmospheric pressure detection sensor 21, and the DPF differential pressure detection sensor. The DPF differential pressure Pdpf detected by 23 is acquired, and the process proceeds to step S130.

  In step S130, the ECU 24 calculates a correction coefficient Cf1 from the acquired cooling water temperature Tw based on the cooling water temperature correction map M1, and calculates a correction coefficient Cf2 from the acquired intake air temperature Ti based on the intake air temperature correction map M2. The correction coefficient Cf3 is calculated from the acquired atmospheric pressure Pa based on the atmospheric pressure correction map M3, the correction coefficient Cf4 is calculated from the acquired DPF differential pressure Pdpf based on the DPF differential pressure correction map M4, and the bypass valve opening Vb To calculate a correction coefficient Cf5 based on the bypass valve correction map M5, and the process proceeds to step S140.

  In step S140, the ECU 24 calculates the pre-correction actual torque TQ0 based on the actual torque map M0 from the acquired actual engine speed N and the actual turbocharger speed NT, and the process proceeds to step S150.

  In step S150, the ECU 24 calculates the actual torque TQ from the calculated actual torque TQ0 before correction and the calculated correction coefficients Cf1, Cf2, Cf3, Cf4, and Cf5, and the process proceeds to step S110.

  With this configuration, the actual torque TQ of the engine 1 is calculated based on the exhaust gas flow rate discharged in accordance with the actual torque TQ output from the engine 1. Further, the actual torque TQ of the engine 1 is calculated without being influenced by the operating environment of the engine 1 or the configuration of the engine 1. Thus, the actual torque TQ of the engine 1 can be easily calculated while suppressing the influence due to the state fluctuation on the engine 1 side and the influence due to the load fluctuation.

  In the present embodiment, the actual torque TQ may be calculated when the actual turbocharger rotational speed NT continues to be stable for a certain period of time. For example, when it is determined that the operating state of the engine 1 is steady, the ECU 24 can be configured to calculate the actual torque TQ. Further, the actual torque TQ may be calculated while correcting with a correction coefficient under a predetermined condition in consideration of the response delay of the second supercharger 10. For example, when the acceleration or deceleration becomes greater than or equal to a predetermined value, the ECU 24 can be configured to perform calculation using the correction coefficient K1 only for a predetermined time. Here, the correction coefficient K1 is obtained experimentally.

  Below, the engine 1a in 2nd embodiment of the engine which concerns on this invention is demonstrated using FIG. 7 and FIG. The engine further includes an EGR device 25 in addition to the engine 1 according to the first embodiment of the present invention. In the following embodiments, the same points as those of the above-described embodiments will not be specifically described, and different portions will be mainly described.

  The EGR device 25 returns a part of the exhaust gas to the intake air. The EGR device 25 includes an EGR pipe 26 and an EGR valve 27.

  The EGR pipe 26 is a pipe for guiding the exhaust gas to the intake pipe 2. The EGR pipe 26 is provided so as to communicate the intake pipe 2d and the exhaust pipe 5a. Thereby, part of the exhaust gas passing through the exhaust pipe 5a is guided to the intake pipe 2d through the EGR pipe 26. That is, a part of the exhaust gas is configured to be recirculated to the intake air as EGR gas (hereinafter simply referred to as “EGR gas”).

  The EGR valve 27 limits the flow rate of EGR gas that passes through the EGR pipe 26. The EGR valve 27 is composed of a normally closed type electromagnetic flow control valve. The EGR valve 27 is provided in the middle of the EGR pipe 26. The EGR valve 27 can acquire a signal from the ECU 24 and change the opening degree of the EGR valve 27. In the present embodiment, the EGR valve 27 is composed of a normally closed electromagnetic flow control valve, but any EGR gas flow rate can be used as long as it can be limited.

  The ECU 24 is connected to the EGR valve 27, and can control the EGR valve 27 to an arbitrary EGR valve opening degree Vegr.

  The ECU 24 stores an EGR valve correction map M6 for calculating a correction coefficient Cf6 of the actual torque TQ at an arbitrary EGR valve opening degree Vegr. That is, the ECU 24 is configured to correct the correction coefficient Cf1 to the correction coefficient Cf6 when the operation condition of the engine 1a is different from the operation condition of the engine 1a when the actual torque map M0 is set.

  As shown in FIG. 8, the EGR valve correction map M6 determines a correction coefficient Cf6 for an arbitrary EGR valve opening degree Vegr. When the EGR valve opening degree Vegr is larger than the EGR valve opening degree Vegrs under a predetermined operating condition, the actual torque TQ decreases due to a decrease in the actual supercharger rotational speed NT caused by an increase in EGR gas. Therefore, the EGR valve correction map M6 is set such that the correction coefficient Cf6 is smaller than 1 when the EGR valve opening degree Vegr is larger than the EGR valve opening degree Vegrs. On the other hand, when the EGR valve opening degree Vegr is smaller than the EGR valve opening degree Vegrs, the actual torque TQ increases due to an increase in the actual supercharger rotational speed NT caused by a decrease in EGR gas. Therefore, the EGR valve correction map M6 is set such that the correction coefficient Cf6 is larger than 1 when the EGR valve opening degree Vegr is smaller than the EGR valve opening degree Vegrs.

  Hereinafter, a control mode for calculating the actual torque TQ in the engine 1a according to the second embodiment of the present invention will be described with reference to FIG.

  As shown in FIG. 9, in step S131, the ECU 24 calculates the correction coefficient Cf6 based on the EGR valve correction map M6 from the EGR valve opening degree Vegr, and moves the step to step S140.

  In step S151, the ECU 24 calculates the actual torque TQ from the calculated actual torque TQ0 before correction and the calculated correction coefficients Cf1, Cf2, Cf3, Cf4, Cf5, and Cf6, and the process proceeds to step S110.

  By configuring in this way, the actual torque of the engine is calculated based on the exhaust flow rate discharged according to the actual torque TQ output from the engine 1a. Further, the actual torque TQ of the engine 1a is calculated without being affected by the operating environment of the engine 1a and the configuration of the engine 1a. Thereby, it is possible to easily calculate the actual torque TQ of the engine 1a while suppressing the influence due to the state fluctuation on the engine 1a side and the influence due to the load fluctuation.

  Below, the engine 1c in 3rd embodiment of the engine which concerns on this invention is demonstrated using FIG. 10 and FIG. The engine 1c is configured by a variable displacement supercharger 28 as the second supercharger of the engine 1 according to the first embodiment of the present invention or the engine 1b according to the second embodiment. In the following embodiments, the same points as those of the above-described embodiments will not be specifically described, and different portions will be mainly described.

  As shown in FIG. 10, the variable displacement supercharger 28 changes the characteristics of the supercharger by controlling the flow path configuration of the turbine. The variable displacement supercharger 28 further includes a variable nozzle vane 29 and an actuator 30 that drives the variable nozzle vane 29.

  The variable nozzle vane 29 can change the relative angle with respect to the turbine body. The variable nozzle vane 29 is provided in the turbine portion of the variable capacity supercharger 28. The variable displacement supercharger 28 is changed from the fully open state (maximum distribution area state) shown in FIG. 10A to the state shown in FIG. 10B by changing the positions of the variable nozzle vanes 29 by the actuator 30. The nozzle opening area can be changed up to the fully closed state (the flow area minimum state). As a result, the variable capacity supercharger 28 is configured to have the characteristics of a supercharger having a large capacity when the opening degree of the variable nozzle vane 29 is fully opened, and the capacity is small when the variable nozzle vane 29 is fully closed. It is comprised so that the characteristic of a supercharger may be provided. That is, the variable capacity supercharger 28 can change the characteristics of the variable capacity supercharger 28 like a supercharger having a different capacity by changing the position of the variable nozzle vane 29. Therefore, the variable displacement supercharger 28 can change the actual supercharger rotational speed NT in a state where the exhaust flow rate does not change by changing the opening of the variable nozzle vane 29.

  The ECU 24 is connected to the actuator 30 and can control the variable nozzle vane 29 to an arbitrary nozzle vane opening degree NV.

  The ECU 24 stores a nozzle vane correction map M7 for calculating a correction coefficient Cf7 of the actual torque TQ at an arbitrary nozzle vane opening degree NV. That is, the ECU 24 is configured to correct the correction coefficient Cf1 to the correction coefficient Cf7 when the operation condition of the engine 1c is different from the operation condition of the engine 1c when the actual torque map M0 is set.

  As shown in FIG. 11, the nozzle vane correction map M7 determines a correction coefficient Cf7 for an arbitrary nozzle vane opening degree NV. When the nozzle vane opening degree NV is larger than the predetermined nozzle vane opening degree NVs under a predetermined operating condition, the actual torque TQ is caused by the fact that the variable displacement supercharger 28 has the characteristics of a supercharger having a large capacity. Decreases due to a decrease in the feeder rotational speed NT. Accordingly, in the nozzle vane correction map M7, when the nozzle vane opening degree NV is larger than the predetermined nozzle vane opening degree NVs, the correction coefficient Cf7 is set to a value smaller than 1. On the other hand, when the nozzle vane opening degree NV is smaller than the predetermined nozzle vane opening degree NVs, the actual torque TQ is caused by the fact that the variable displacement supercharger 28 has the characteristics of a supercharger having a small capacity. Increased with increasing speed NT. Therefore, the nozzle vane correction map M7 is set such that the correction coefficient Cf7 is larger than 1 when the nozzle vane opening NV is smaller than the predetermined nozzle vane opening NVs.

  Hereinafter, a control mode for calculating the actual torque TQ in the engine 1c according to the third embodiment of the present invention will be described with reference to FIG.

  As shown in FIG. 12, in step S132, the ECU 24 calculates a correction coefficient Cf7 from the nozzle vane opening degree NV based on the nozzle vane correction map M7, and moves the step to step S140.

  In step S152, the ECU 24 calculates the actual torque TQ from the calculated actual torque TQ0 before correction and the calculated correction coefficients Cf1, Cf2, Cf3, Cf4, Cf5, Cf6, and Cf7, and the process proceeds to step S110.

  With this configuration, the actual torque of the engine 1c is calculated based on the exhaust flow rate discharged in accordance with the actual torque TQ output from the engine 1c. Further, the actual torque TQ of the engine 1c is calculated without being affected by the operating environment of the engine 1c and the configuration of the engine. Thereby, it is possible to easily calculate the actual torque TQ of the engine 1c while suppressing the influence due to the state fluctuation on the engine 1c side and the influence due to the load fluctuation.

1 Engine 6 1st turbocharger 10 2nd turbocharger 22 Supercharger rotation speed detection sensor N Actual engine rotation speed NT Actual turbocharger rotation speed NT0 No-load turbocharger rotation speed ΔNT Unit of turbocharger rotation speed Increase ΔTQ Unit increment of actual torque TQ Actual torque

Claims (5)

An engine comprising a supercharger and a supercharger rotation speed detection means,
At any engine speed
An engine in which the actual torque of the engine is calculated from the actual turbocharger rotation speed, the turbocharger rotation speed in the no-load state, and the engine torque increase amount per unit increase amount of the turbocharger rotation speed. A cooling water temperature detecting means, an intake air temperature detecting means, an atmospheric pressure detecting means, and a diesel particulate filter differential pressure detecting means;
The engine according to claim 1, wherein the actual torque is corrected based on a coolant temperature, an intake air temperature, an atmospheric pressure, and a diesel particulate filter differential pressure. An EGR device that recirculates exhaust gas to intake air, and an EGR valve opening degree detecting means that adjusts the exhaust gas flow rate for recirculation;
The engine according to claim 1 or 2, wherein the actual torque is corrected based on an opening degree of the EGR valve. A bypass pipe for bypassing a part of the exhaust gas supplied to the supercharger, and a bypass valve for controlling an exhaust flow rate flowing through the bypass pipe;
The engine according to any one of claims 1 to 3, wherein the actual torque is corrected based on an opening degree of the bypass valve. The supercharger is composed of a variable capacity supercharger that changes the supercharger rotation speed by changing the vane opening,
The engine according to any one of claims 1 to 4, wherein the actual torque is corrected based on a vane opening.

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