JP6568214B2 - Multistage skip fire - Google Patents

Description

This application is related to US Provisional Patent Application No. 62 / 077,439 “Multi Level Dynamic Skip Fire”, filed on Nov. 10, 2014, US Provisional Patent Application filed on February 17, 2015. No. 62 / 117,426, “Multi Level Dynamic Skip Fire”, US Provisional Patent Application No. 62 / 121,374, filed Feb. 26, 2015, “Using Multi-Level Skip Fire”, 2015 October U.S. Patent Application No. 14 / 919,011 "Multi-level Skip Fire" filed on May 21, and US Patent Application No. 14 / 919,018 filed Oct. 21, 2015 "Multi-level" Also claim “Skip Fire” Each of which is incorporated herein in its entirety for all purposes.

  The present invention relates to a method and system for operating an engine in a skip fire manner. In various embodiments, a skipfire engine control system is described that selectively deactivates a combustion chamber or ignites at a plurality of different power levels.

  The vast majority of vehicles (and many other devices) in use today are powered by an internal combustion (IC) engine. Internal combustion engines typically have multiple cylinders or other combustion chambers in which combustion occurs. Under normal driving conditions, the torque generated by the internal combustion engine needs to vary over a wide range to accommodate the operating requirements of the driver. Over the years, many methods for controlling internal combustion engine torque have been proposed and utilized. Some such methods change the effective displacement of the engine. Engine control for changing the engine's effective displacement can be classified into two types of control, namely, a plurality of fixed displacements and a skip fire. In the multiple fixed displacement control, a fixed set of several cylinders is stopped under a low load condition, for example, an 8-cylinder engine capable of operating the same four cylinders under a specific condition. In contrast, skip fire engine control selectively decimates the ignition of a particular cylinder during a selected ignition opportunity. Thus, a particular cylinder can be ignited in one engine cycle, paused in the next engine cycle, and selectively paused or ignited in the next cycle. For example, when a 4-cylinder engine is ignited every other cylinder, an effective displacement of 1/3 of the total engine displacement can be obtained, but this is a ratio of the displacement that cannot be obtained by simply stopping a pair of cylinders. It is. Similarly, an effective displacement of ½ is obtained by igniting every other cylinder in a three-cylinder engine, but this is the proportion of the displacement that cannot be obtained by simply deactivating a set of cylinders. U.S. Pat. No. 8,131,445 (filed by the assignee of the present application and incorporated herein in its entirety for all purposes) discloses various implementations of skip fire engine control. Yes. In general, skip fire engine control is understood to provide a number of possible benefits, including the potential for significant improvements in fuel economy in many applications. Although the concept of skipfire engine control has been known for many years and its benefits have been understood, skipfire engine control has not yet achieved significant commercial success.

  It is well understood that an engine in operation tends to be a significant source of noise and vibration, collectively referred to in the art as NVH (noise, vibration and discomfort). In general, the fixed idea associated with skipfire engine control is that engine skipfire operation results in rougher engine operation, ie, increased NVH, compared to engines operating in a conventional manner. In many applications, such as automotive applications, vibration control is one of the most significant difficulties in skip fire engine control. In fact, the inability to satisfactorily address the concerns of NVH is considered to be one of the main obstacles that have prevented the widespread use of skipfire engine control.

  U.S. Patent Nos. 7,954,474, 7,886,715, 7,849,835, 7,577,511, 8,099,224 , 8,131,445 and 8,131,447, as well as U.S. patent applications 13 / 004,839, 13 / 004,844 and others. Various engine controllers have been described that allow an internal combustion engine to operate realistically in a skip fire mode of operation. Each of these patents and patent applications is cited herein. Although the described controller works well, efforts continue to further improve the performance of these and other skipfire engine controllers to further reduce the NVH problem in engines operating under skipfire control. This application describes additional skipfire control features and enhancements that can improve engine performance in various applications.

  The present invention relates to skip fire control. In one aspect, a method for controlling an engine is described. The selected combustion cycle is decimated to ignite the selected operating combustion cycle to deliver the desired engine power. One or more combustion chambers can generate multiple possible levels of torque output, eg, for the same cam phaser and / or MAP (intake manifold absolute pressure) setting. A particular level of torque output (eg, high or low torque output) is selected for each ignition combustion chamber (ie, the combustion chamber that is to be ignited). This is called multistage skip fire engine control in this specification. In various designs, the charge to the ignition combustion chamber is adjusted based on whether a high or low torque output is selected for the ignition combustion chamber. Various embodiments relate to engine controllers, software, and systems that are useful for implementing the methods described above.

  In another aspect, an engine controller is described. The engine controller includes a plurality of combustion chambers. Each combustion chamber includes at least one cam operated intake valve. The engine controller includes an ignition ratio calculator, an ignition timing determination module, and an ignition control unit. The ignition ratio calculator is configured to determine an ignition ratio suitable for transmitting the desired torque. The ignition timing determination module is configured to determine a skip fire ignition sequence based on the ignition ratio. The skip fire ignition sequence indicates whether the selected combustion chamber is to be paused or ignited during the selected ignition opportunity, and for each ignition, whether the ignition generates a low torque output or a high torque output. Indicates whether or not. The ignition control unit is configured to operate the combustion chamber in a skip fire manner based on the ignition sequence. In various embodiments, the ignition controller may also determine each ignition combustion chamber (ie, each combustion chamber that is scheduled to ignite) based on whether the ignition sequence indicates a low torque output or a high torque output for the ignition combustion chamber. It is configured to adjust the air supply.

  Multi-stage skipfire engine control can be performed in a wide variety of ways. In some embodiments, for example, a determination as to whether to ignite or pause each combustion cycle, and / or whether to select a particular level of torque output for the ignition combustion chamber is made at each ignition opportunity. Such a determination can be made using one or more of look-up tables, circuits, sigma-delta converters, and other techniques.

  Various systems can be used to control the torque output of the ignition combustion chamber. In some schemes, for example, each of the one or more combustion chambers includes one or more intake valves that are independently controlled. The intake valve can be opened and closed at different times and / or in response to different cycles (eg, Atkinson and Otto cycles), thus helping to change the torque output of the combustion chamber. The combustion chamber intake valves can be activated or deactivated independently for each combustion cycle. In various embodiments, the combustion chamber valve control system allows the combustion chamber to provide a few or more torque output levels under the same engine conditions, eg, the same cam phaser, throttle position, and / or engine speed setting. Can do. It should be understood that the methods described herein for implementing multi-stage skipfire engine control can be utilized in conjunction with any suitable combustion chamber design or valve control system.

  In another aspect, an engine system is described. The engine system includes one intake manifold, one or more combustion chambers, and two or more intake passages. In various embodiments, two intake passages are connected to the combustion chamber. The two intake passages are arranged with respect to the combustion chamber so that the central axis of each intake passage substantially intersects the central axis of the combustion chamber.

  The invention and its advantages are best understood by referring to the following description in conjunction with the accompanying drawings.

FIG. 1A is a cross-sectional view of a combustion chamber and corresponding valve control system according to a particular embodiment of the present invention. FIG. 1B is a cross-sectional view of a combustion chamber and corresponding valve control system according to certain embodiments of the invention. FIG. 2 is an explanatory diagram showing a valve control system according to various embodiments of the present invention. FIG. 3 is an explanatory view showing a valve control system according to various embodiments of the present invention. FIG. 4 is an explanatory view showing a valve control system according to various embodiments of the present invention. FIG. 5 is an explanatory diagram showing a valve control system according to various embodiments of the present invention. FIG. 6 is an explanatory diagram showing a valve control system according to various embodiments of the present invention. FIG. 7 is an explanatory view showing a valve control system according to various embodiments of the present invention. FIG. 8 is a graph illustrating valve lift adjustment of a combustion chamber according to a specific embodiment of the present invention. FIG. 9 is a valve control system according to a particular embodiment of the present invention. FIG. 10 is an explanatory diagram illustrating an example of an intake passage. FIG. 11 is an explanatory view showing an intake passage according to a specific embodiment of the present invention. FIG. 12A is an illustration showing the stages of operation of the combustion chamber and intake valve according to various embodiments of the present invention. FIG. 12B is an illustration showing the stages of operation of the combustion chamber and intake valve according to various embodiments of the present invention. FIG. 12C is an illustration showing stages of operation of the combustion chamber and intake valve according to various embodiments of the present invention. FIG. 12D is an illustration showing stages of operation of the combustion chamber and intake valve according to various embodiments of the invention. FIG. 12E is an illustration showing stages of operation of the combustion chamber and intake valve according to various embodiments of the invention. FIG. 12F is an illustration showing stages of operation of the combustion chamber and intake valve according to various embodiments of the invention. FIG. 13A is a table showing how the valve can be operated to produce different levels of torque output from the combustion chamber in accordance with various embodiments of the present invention. FIG. 13B is a table showing how the valve can be operated to produce different levels of torque output from the combustion chamber, in accordance with various embodiments of the present invention. FIG. 14A is a table illustrating different configurations and features of combustion chambers according to various embodiments of the present invention. FIG. 14B is a table showing different configurations and features of combustion chambers according to various embodiments of the present invention. FIG. 14C is a table illustrating different configurations and features of combustion chambers according to various embodiments of the present invention. FIG. 14D is a table showing different configurations and features of combustion chambers according to various embodiments of the present invention. FIG. 14E is a table showing different configurations and features of combustion chambers according to various embodiments of the present invention. FIG. 14F is a table illustrating different configurations and features of combustion chambers according to various embodiments of the present invention. FIG. 14G is a table showing different configurations and features of combustion chambers according to various embodiments of the present invention. FIG. 14H is a table illustrating different configurations and features of combustion chambers according to various embodiments of the present invention. FIG. 15 is an illustration of a bank of cylinders according to a particular embodiment of the invention. FIG. 16 is a block diagram of an engine controller according to a particular embodiment of the present invention. FIG. 17 is a flow diagram of a method for performing multi-stage skipfire engine control according to a particular embodiment of the present invention. FIG. 18 is an example lookup table showing the maximum allowable combustion chamber output as a function of engine speed and effective ignition ratio. FIG. 19 is an example of a look-up table showing the ignition ratio and level ratio as a function of the effective ignition ratio. FIG. 20 is an illustration of an example circuit for generating a multi-stage skip fire ignition sequence according to a particular embodiment of the invention. FIG. 21 is an explanatory diagram of an example of a circuit that generates a multi-stage skip fire ignition sequence according to another embodiment of the present invention. FIG. 22 is an example of a look-up table that provides a multi-stage skipfire ignition sequence as a function of effective ignition ratio. FIG. 23 is a flow diagram illustrating an example of a method that uses multi-stage skipfire engine control during transitions between ignition ratios. FIG. 24 is a flow diagram illustrating an example method for detecting and managing knocking in an engine according to certain embodiments of the invention. FIG. 25 is a flow diagram illustrating an example method for using multi-stage skipfire engine control in response to a specific engine operation. FIG. 26 is a flow diagram illustrating an example method for diagnosing and managing engine failures according to certain embodiments of the invention.

  In the drawings, like reference numerals may be used to indicate like components. It should also be understood that the depictions in the drawings are schematic and not actual.

  The present invention relates to a system for operating an internal combustion engine by a skip fire method. More specifically, various implementations of the present invention include a skip fire engine control system that can selectively ignite the combustion chamber at a plurality of different torque output levels.

  In general, skip fire engine control selectively stops ignition of a particular cylinder during a selected ignition opportunity. Thus, for example, a specific cylinder can be ignited at a certain ignition opportunity, stopped at the next ignition opportunity, and selectively stopped or ignited at the next opportunity. This is in contrast to conventional variable displacement engine operation that deactivates a certain set of cylinders at certain low load operating conditions.

  One of the difficulties associated with skipfire engine control is reducing undesirable noise, vibration, and discomfort (NVH) to an acceptable level. Noise and vibration generated by the engine may be transmitted to passengers in the passenger compartment via various routes. Some of these paths, such as the drive train, can change the amplitude of the various frequency components present in the engine noise and vibration signatures. Specifically, since the transmission increases the wheel torque and torque fluctuation, the lower the gear ratio, the more the vibration tends to be amplified. Noise and vibration can also excite various vehicle resonances, which can then be transmitted into the vehicle compartment.

  Some of the noise and vibration frequencies can be uncomfortable, especially for vehicle occupants. In particular, an unpleasant vibration that is felt by a vehicle occupant is likely to occur due to a low-frequency repetitive pattern (for example, a frequency component in the range of 0.2 to 8 Hz). There is a possibility that noise is generated in the passenger compartment due to the higher-order harmonics of these patterns. In particular, a frequency around 40 Hz, which is a so-called “boom” frequency, may resonate in the passenger compartment. Commercially possible skipfire engine control requires operation at an acceptable NVH level, while at the same time delivering the engine torque output desired or required by the driver to achieve significant fuel economy gains.

  NVH characteristics vary with engine speed, ignition cycle, and transmission gear. For example, consider an engine controller that selects a specific ignition cycle that indicates the percentage of ignition required to transmit the desired torque to a specific engine speed and gear. Based on the ignition cycle, the engine controller generates a repetitive ignition pattern to operate the engine combustion chamber in a skip fire manner. As is known to those skilled in the art, an engine that runs smoothly in some ignition patterns may produce unpleasant acoustic or vibration effects at a given engine speed and other ignition patterns. Similarly, a given ignition pattern at one engine speed may result in acceptable NVH, but the same pattern may result in unacceptable NVH at other engine speeds. Engine-derived noise and vibration can also be affected by cylinder load or combustion chamber power. The less air and fuel sent to the cylinder, the less power generated by cylinder ignition and the less noise and vibration. As a result, if the cylinder output decreases, some ignition cycles and sequences that could not be used due to poor NVH characteristics may become available.

As described in US patent application Ser. No. 14 / 638,908, which is hereby incorporated by reference in its entirety for all purposes, a skipfire engine controller design minimizes fuel consumption and is an acceptable NVH. It is generally desirable to deliver the required engine power while providing performance. This is a difficult task due to the various operating conditions encountered during vehicle operation. The required engine power can be expressed as a torque demand at a certain engine operating speed. It should be understood that the amount of engine torque transmitted can be expressed as the product of the ignition period and the cylinder load. Therefore, if the ignition cycle (FF) is shortened, the cylinder torque load (CTF) can be reduced to produce the same engine torque, and vice versa. In other words,
Engine torque ratio (ETF) = CTF * FF (Formula 1)
Here, ETF is a value indicating the normalized net or rated engine torque. Since all values in the above equation are dimensionless, they can be used with any kind of engine and any kind of vehicle. That is, various different ignition cycle and CTF combinations may be used to transmit the same engine torque. Equation 1 does not include the effect of engine friction. Similar analysis including friction can be performed. In this case, the calculated parameter is the brake torque ratio. Any of engine net torque ratio, engine brake torque ratio, engine rated torque ratio, or some similar metric may be used as the basis for the control algorithm. For clarity, the term “engine torque ratio” may refer to any of these measures of engine power and is used in the discussion of the engine controller and engine control method below.

Various embodiments of the present invention relate to a skipfire engine control system capable of igniting selected combustion chambers at a plurality of different power levels. In this specification, this is called a multistage skip fire operation. In some embodiments, multi-stage skipfire operation can be modeled by modifying Equation 1 above to include the possibility of multiple ignition levels as follows.
Engine torque ratio (ETF) = CTF ₁ * FF ₁ + CTF ₂ * FF ₂ +. . + CTF _n * FF _n (Formula 2)
Here, CTF ₁ is the cylinder torque ratio, FF ₁ is the ignition ratio at the first level, CTF ₂ is the cylinder torque ratio, FF ₂ is the ignition ratio at the second level, CTF _n is the cylinder torque ratio, and FF _n is the nth It is the ignition ratio at the level. The sum of the ignition ratios at each level is equal to the total ignition ratio, ie,
FF = FF ₁ + FF ₂ +. . . + FF _n (Formula 3)
In some embodiments described below, n is equal to 2, but not limited to.

It should be understood that there are many equivalent ways of representing the above concepts. For example, instead of modeling based on the engine torque ratio (ETF), modeling based on the net engine torque (ET) may be performed because the quantities are simply proportional. The cylinder torque ratio (CTF) is proportional to the net mean effective pressure (NMEP), and the n-th level ignition ratio (FF _n ) is proportional to the engine displacement ratio (FEDn) of the cylinder operating at the n-th level. Good. Equation 2 can therefore be formulated equivalently as follows:
ET = NMEP ₁ * FED ₁ + NMEP ₂ * FED ₂ +. . . + NMEP _n * FED _n (Formula 4)
Equation 4 above is only an exemplary modification, and many equivalent modifications are possible. All of these have at least two cylinder groups with a common non-zero output, each having a common quantity related to engine output torque expressed as the sum of a plurality of quantities related to the cylinder group output.

  An example of a multi-stage skip fire operation can be described as follows. The combustion chamber is paused during one selected combustion cycle, ignited at a high level output during the next combustion cycle, and at a lower level output during the next combustion cycle (eg, 0 to 0 at a high level output). 80%). In various implementations, the low level output may substantially correspond to a combustion chamber load that achieves optimum fuel economy, ie, the lowest BSFC (brake specific fuel consumption) operation. As is known, the BSFC combustion chamber load varies as a function of RPM. Thus, the ratio of high and low ignition levels can vary as a function of engine RPM and possibly other variables in various embodiments of the invention. Ignition and pause are adjusted to produce the desired engine torque. The availability of multi-stage skipfire operation allows the engine control system to have more options for finding a balance between engine power, fuel economy, noise, and vibration.

  It should be understood that any suitable technique can be used to enable multi-stage skipfire operation. In some embodiments, for example, the combustion chamber torque output is controlled using throttle control, ignition timing, valve opening / closing timing, MAP adjustment and / or exhaust gas recirculation. In this application, various combustion chamber control systems and configurations are described. Such a system is configured such that the combustion chamber can produce multiple levels of torque output. This application also describes various multi-stage skipfire engine control methods (eg, as described with respect to FIGS. 16-26) that can be performed using the system described above. However, these methods are not limited to the systems described herein and may be used with any suitable combustion chamber design, system, or mechanism.

Combustion Chamber Valve Control System Various embodiments of the present invention relate to a combustion chamber valve control system. Referring initially to FIGS. 1A and 1B, two cross-sectional views of an exemplary combustion chamber valve control system 100 are shown. The combustion chamber valve control system 100 includes a combustion chamber 102 including a piston 104, two intake valves 120a / 120b, and two exhaust valves 122a / 122b. Actuators 116a / 116b control the opening and closing of the intake valves. Intake passages 110a / 110b connect intake valves 120a / 120b to intake manifolds (not shown).

  When the intake valve is opened, air is sent from the intake manifold to the combustion chamber 102 through the corresponding intake passage 110a / 110b. As known to those skilled in the art, when the combustion chamber 102 is to be ignited, air is mixed with fuel in the combustion chamber 102 and the fuel / air mixture is ignited. The resulting combustion drives the piston 104 to the bottom of the combustion chamber 102. As the exhaust valve 122a / 122b opens and the piston 104 moves up, exhaust gas is pushed out of the combustion chamber 102 into the exhaust passage 112a / 112b.

  In many conventional designs, the intake valves 120a / 120b of the combustion chamber 102 are opened and closed simultaneously. That is, it is controlled by the same actuator and / or opened and closed according to the same lift curve. The timing of the lift curve can be adjusted using a cam phaser that shifts the valve opening and closing time relative to the crankshaft motion. However, in various conventional designs, cam phaser mechanics generally allow only a slight change in valve opening and closing timing from cycle to cycle and operate all cylinders in the same manner. However, in the illustrated embodiment, the intake valves 120a / 120b operate and operate independently. The opening / closing timing of one intake valve may be different for each combustion cycle, or may be the same as other intake valves. For example, the intake valve 120a may remain idle or closed during the selected combustion cycle, while the intake valve 120b is open to supply the combustion chamber. Alternatively, the intake valve 120a can be opened and closed based on the Otto cycle during the selected combustion cycle, while the other intake valve 120b can be opened and closed based on Atkinson or other cycles. One or both of the intake valves can be paused or closed during any selected combustion cycle. In various embodiments, each intake valve in the combustion chamber 102 can be activated or deactivated independently for each ignition opportunity.

  The ability to independently control intake valves in the same combustion chamber provides various advantages. As an example, the torque output of the combustion chamber can be adjusted dynamically. For example, in various designs, if both intake valves are open during the intake stroke and closed during the subsequent compression stroke, the result is that one of the intake valves is deactivated during the selected combustion cycle. Less air is sent to the combustion chamber. This therefore reduces the torque produced by ignition of the combustion chamber as compared to the state where both intake valves are open. Similarly, closing one or both of the intake valves prior to completion of the intake stroke results in reduced air intake and reduced combustion cycle torque output. Similarly, leaving one or both intake valves open throughout both the intake and compression strokes results in reduced combustion cycle output. In this case, the air taken into the cylinder is discharged from the cylinder prior to the start of the power stroke. By using independent control of each intake valve and different opening / closing timings for each intake valve, it is possible to output combustion chambers in two or more stages as described later in this application. As described above, the combustion chamber torque output can be changed quickly, for example, for each ignition opportunity, so that the control of vibration, noise, and fuel consumption can be improved.

  The actuators 116a / 116b can use a wide range of mechanisms for controlling the opening and closing of the intake valves 120a / 120b in the combustion chamber 102. For example, in various embodiments, each intake valve is cam operated and / or mechanically controlled. For example, in the illustrated embodiment, the actuators 116a, 116b are each separate cams that independently operate the intake valves 120a, 120b. In some designs, idle, folding valve lift rods, folding backlash regulators, folding roller finger followers, or folding concentric buckets may be placed in the valve train to allow valve rest. These devices allow the intake valve to be activated or deactivated at any given combustion cycle. In some implementations, the operation of the valve can be controlled using an axially moving camshaft that can be shifted to engage different cam lobes with the intake valve system. In this case, one of the cam lobes may be a zero head lobe that effectively deactivates the cylinder. In some embodiments, only one intake valve can be used to open and operate the valve based on two or more different lift curves. These different curves can be generated with different cams or with more complex valve trains. However, it should be understood that various other designs are also possible, as will be described later in this application. Actuation of the intake valve can be performed by mechanical, electromechanical, electrohydraulic or any other suitable mechanism.

  A wide range of systems may be used to operate and control the intake and exhaust valves of the combustion chamber 102. Some design examples are shown in FIGS. 2 to 7 are top views of examples of the combustion chamber valve control system (for example, the combustion chamber control system 100 shown in FIGS. 1A and 1B). 2-7 each show a combustion chamber 102, actuators 116a / 116b, intake valves 120a / 120b, exhaust valves 122a, and possibly additional exhaust valves 122b. A line drawn between the actuator and a particular valve indicates that the actuator controls the opening and closing of the valve. In general, if a line is drawn between the actuator and two or more valves, it means that all valves must be activated during the selected combustion cycle if the actuator is activated. Alternatively, if the actuator is not activated during one combustion cycle, all valves must be deactivated during that combustion cycle. If no line is drawn between the actuator and a specific valve, it means that the actuator does not control the specific valve. The operations described above can be performed using any suitable technique or mechanism, such as using a camshaft assembly that includes one or more cams and / or camshafts.

  There can be a variety of different valve control configurations. For example, in FIG. 2, the intake valve 120a and the exhaust valve 122a are on one side of the combustion chamber 102 (ie, one side of the symmetry axis 105). Intake valve 120b and exhaust valve 122b are on the other side of combustion chamber 102 (ie, the other side of line 105). Actuator 116a controls a valve on one side of combustion chamber 102 (ie, intake valve 120a and exhaust valve 122a), and another actuator (actuator 116b) on the other side of the combustion chamber (ie, intake valve 120b and exhaust valve). 122b).

  FIG. 3 shows a slightly different configuration. In this example, actuators 116a / 116b each control an intake valve on one side of the combustion chamber and an exhaust valve on the other side of the combustion chamber. That is, the actuator 116a controls the intake valve 120a and the exhaust valve 122b, whereas the actuator 116b controls the intake valve 120b and the exhaust valve 122a.

  Different flows in the combustion chamber 102 may result from the above configuration. For example, if the actuator controls intake and exhaust valves (eg, as shown in FIG. 2) on the same side of the combustion chamber, air flowing from the intake valve to the exhaust valve may not pass through the center or central axis 106 of the combustion chamber. The tendency not to flow is shown. When the actuator controls intake and exhaust valves (eg, as shown in FIG. 3) on different sides of the combustion chamber, the air flowing between the intake and exhaust valves tends to pass through the center or central axis of the combustion chamber. This can have different effects on the vortex or rotational flow of air and gas in the combustion chamber. Different control schemes and configurations of actuators and valves help generate the desired amount of vortex flow in the combustion chamber. In general, a moderate amount of eddy current is required. Excessive eddy currents can result in excessive thermal convection toward the combustion chamber walls. If the vortex is too small, the combustion speed of the combustion chamber may be excessively reduced.

  Other valve control configurations are possible. For example, in FIG. 4, the actuator 116a controls the intake valve 120a on one side of the combustion chamber 102 and the exhaust valve 122a / 122b on the other side of the combustion chamber. The other actuator 116b controls the remaining intake valves (intake valves 120b). Accordingly, during each selected combustion cycle, the actuator 116b is activated to open the intake valve 120b and the actuator 116a must also be activated each time an exhaust operation is required. In other words, the actuator 116a is activated each time exhaust action is required during the selected combustion cycle, and the intake valve 120a and both exhaust valves 122a, 122b will be open during the combustion cycle. By opening both exhaust valves, the operation of exhausting exhaust gas from the combustion chamber immediately before blowdown, that is, immediately before the piston reaches top dead center (that is, before the start of the intake stroke) is improved.

  FIG. 5 shows another valve control system. In this example, the actuator 116a controls the intake valve 120a and both exhaust valves 122a and 122b on one side of the combustion chamber 102. The other actuator 116b has a similar function. That is, the intake valve 120b on the other side of the combustion chamber and both the exhaust valves 122a and 122b are also controlled. This configuration also activates both exhaust valves 122a / 122b during selected combustion cycles where exhaust operation is desired and / or whenever one of the intake valves 120a / 120b operates during the selected combustion cycle. If any of the actuators 116a and 116b is activated, the exhaust valves 122a and 122b are activated. However, in contrast to FIG. 4, if a combustion phenomenon is desired, intake valve 120b can be opened during a selected combustion cycle without requiring opening of intake valve 120a.

  Although the above example is a combustion chamber having two intake valves and two exhaust valves, this is not a requirement and the combustion chamber may include any suitable number of intake or exhaust valves. For example, FIG. 6 shows a combustion chamber 102 having two intake valves 120a / 120b and one exhaust valve 122a. The actuator 116a controls the intake valve 120a on one side of the combustion chamber and the exhaust valve 122a. The actuator 116b controls the intake valve 120b on the other side of the combustion chamber 102 and the exhaust valve 116b. Thus, if exhaust operation is desired during the selected combustion cycle, the exhaust valve 122a opens regardless of which intake valve is open.

  FIG. 7 shows a different control scheme involving a combustion chamber 102 having two intake valves 120a / 120b and one exhaust valve 122a. In this exemplary scheme, actuator 116a controls intake valve 120a on one side of combustion chamber 102 and exhaust valve 122a. The actuator 116b controls only the intake valve 120b on the other side of the combustion chamber. In contrast to the control system shown in FIG. 6, actuator 116b does not control exhaust valve 122a. Accordingly, if exhaust operation is desired during the selected combustion cycle, the actuator 116a must be actuated to open the intake valve 120a. That is, intake valve 120b is not the only intake valve that operates during the selected combustion cycle in which combustion and exhaust operations occur within combustion chamber 102, but always operates with intake valve 120a. However, during the selected combustion cycle, intake valve 120a and exhaust valve 122a can be opened while intake valve 120b is at rest.

  FIGS. 8 and 9 show another type of control scheme involving an actuator capable of varying the valve opening time and timing of the intake valve. In other words, in some of the above examples, the actuator can only be in two states: a corresponding intake valve deactivation or a corresponding intake valve actuation. If the intake valve is activated, the opening timing and time of the intake valve are fixed during the selected combustion cycle. However, in other embodiments, the actuator can perform additional functions. That is, the actuator can follow a plurality of cam curves or valve lift settings, each having a different valve opening / closing timing feature.

  An example of this method is shown in FIGS. FIGS. 8 and 9 relate to the combustion chamber 102 having one intake valve 120a, exhaust valve 122a, and actuator 116a (FIG. 9). As shown in FIG. 9, the actuator 116 a controls all the valves in the combustion chamber 102. The actuator 116a is configured to selectively adjust the valve lift of the intake valve 120a based on the valve lift adjustment setting or the cam curve in order to change the output of the combustion chamber.

  FIG. 8 is a graph 800 showing valve lift as a function of time. Two valve lift adjustment settings are represented by curves 802 and 804. The actuator 116a is configured to operate the intake valve 120a based on any valve lift adjustment setting. In various embodiments, the actuator 116a can transition between settings for each combustion cycle. The graph 800 shows how the valve opening time and the degree of the intake valve 120a vary for each setting. That is, in the case of the setting represented by curve 804, the maximum amount of valve lift and the time during which intake valve 120a is open during the selected combustion cycle is greater than that due to the setting represented by graph 802. Accordingly, different amounts of air are sent to the combustion chamber 102 with different settings, resulting in different levels of torque output from the combustion chamber 102. Implementation of different valve head adjustment settings can be done using any suitable technique or valve adjustment mechanism.

  As described above, some of the valve control systems described above can be used to control the rotational and / or vortex flow of gases within the combustion chamber. Control of gas flow in the combustion chamber can be further improved by specific intake channel designs. Various examples of such a design are shown in FIGS.

  For comparison purposes, FIG. 10 is a top view of combustion chamber 1002 and corresponding intake passages 1006a / 1006b according to a conventional design. The two intake passages 1006 a / 1006 b each connect two intake valves of the combustion chamber 102 to the intake manifold 1014. In this example, a single intake passage 1004 is divided by a common passage wall 1112 to form separate intake passages 1006a / 1006b. Note that the central axis (axis 1008a, 1008b) of each intake passage does not intersect the central axis 1010 of the combustion chamber (the central axis 1010 should be understood as a line perpendicular to the page).

  FIG. 11 illustrates another intake path design according to certain embodiments of the invention. In FIG. 11, two intake passages 1106 a / 1106 b connect the intake manifold 1114 to the combustion chamber 1102, each connected to a separate intake valve on the combustion chamber 1102. The intake passages 1106a / 1106b are connected to the combustion chamber 1102 at a certain angle, that is, do not extend parallel to each other. In the illustrated embodiment, an intake passage 1106b for one combustion chamber 1102 shares an air flow path with an intake passage 1122 for an adjacent combustion chamber 1120, but in other embodiments, an adjacent combustion chamber The intake passage for is completely separate.

  The angle at which each intake passage 1106a / 1106b connects to the combustion chamber 1102 (substantially) intersects the central axis 1108a / 1108b of each intake passage 1106a / 1106b with the central axis 1110 of the combustion chamber 1102. With this design, the air sent using the intake passages 1106a / 1106b is sent directly to the center of the combustion chamber, possibly reducing the degree of vortex or mixing compared to the configuration of FIG. Such a configuration is optionally combined with the valve system shown in FIGS. 1-7 to help improve control of gas movement within the combustion chamber 1102.

  Additional adjustments can be made to the combustion chamber design to further control the charge to the combustion chamber and / or the gas flow within the combustion chamber. In some embodiments, for example, the size and / or diameter of the combustion chamber intake valves (eg, intake valves 120a / 120b in FIGS. 1A and 1B) are different. That is, the speed of airflow through the valve varies depending on their shape, size, or design. The asymmetric charge to the combustion chamber tends to create vortices in the combustion chamber, which is desirable under certain conditions.

  If the combustion chamber intake valves are independently controlled (eg, as shown in FIGS. 1-7), they may also follow different valve lift curves and / or open / close times may be different. These curves and valve opening / closing points can be mixed and adapted in any desired manner consistent with available valve control mechanisms. For example, one intake valve is actuated to perform a lift curve that opens and closes immediately after BDC throughout the entire intake stroke. The lift curve makes it possible to induce the maximum supply air, and is also called a normal timing and a lift curve. The other intake valve is actuated to follow an intake valve early closing (EIVC) or intake valve late closing (LIVC) curve. Both the EIVC and LIVC curves and timing result in reduced air introduction compared to the normal lift curve. With normal timing and lift curves, the engine operates on an Otto cycle, i.e. valve opening and closing timing results in substantially maximum charge. When the EIVC or LIVC valve opening / closing timing is used, the effective compression ratio decreases because the supply air decreases. This is often described as engine operation using an Atkinson or Miller cycle. Using different lift curves and timings helps provide additional control over combustion chamber power, vibration, noise, and fuel consumption.

  A particular scheme that generates a particular level of torque by using a particular lift curve and / or valve opening and closing timing for one or more intake valves is referred to herein as a valve control scheme. Thus, there can be different valve control schemes that each generate different (eg, low, medium and / or high) levels of torque from the ignition combustion chamber. Each valve control scheme independently controls each intake valve in the combustion chamber to operate each intake valve according to a specific lift curve and / or timing cycle (eg, Otto, Atkinson, etc.). Depending on the particular valve control scheme, multiple intake valves of a combustion chamber can be operated according to the same or different lift curves and / or timing cycles.

  Referring now to FIGS. 12A-12E, some differences between the valve control system described above and a conventional valve control system will be described. For comparison, FIG. 12A shows the various stages of combustion chamber operation in an exemplary Otto cycle intake and compression stroke currently used in many automotive engines. The combustion chamber includes two intake valves (intake valves 1202a and 1202b) that operate on the Otto cycle by both operating based on normal timing and lift curves.

  During the intake stroke, both valves 1202a / 1202b operate. The piston 1206 moves from top dead center (TDC) to bottom dead center (BDC). About 40 degrees before the piston 1206 reaches the BDC, the valve lift reaches its highest point. When the piston 1206 reaches the BDC, the compression stroke is started. The piston then moves backward to top dead center (TDC). The intake valve is closed approximately 40 ° after BDC.

  During the Atkinson cycle, the intake valve can be closed earlier or later. The former is called intake valve early closing (EIVC). An example of EIVC valve operation is shown in FIG. 12B. In FIG. 12B, both intake valves 1202a / 1202b are operating according to the EIVC Atkinson cycle. The intake valves 1202a / 1202b are closed before the piston 1206 reaches the BDC at the end of the intake stroke. This is significantly faster than the Otto cycle shown in FIG. 12A where the intake valve closes after 40 °. Therefore, compared with the Otto cycle, the intake valve closes earlier and the opening time is shorter, resulting in less air in the combustion chamber and lower torque output.

  FIG. 12C shows an alternative Atkinson cycle where both intake valves close late compared to a standard Otto cycle. This method is referred to as intake valve late closing (LIVC). FIG. 12C shows an example of the LIVC valve control system. As shown in the figure, the intake valves 1202a / 1202b are closed about 90 ° after BDC during the compression stroke. In contrast, in the Otto cycle example, the intake valve closes approximately 40 ° after BDC. As a result, more of the air sent to the combustion chamber during the intake phase is pushed out of the combustion chamber during the compression stroke, reducing the air sent to the combustion chamber.

  During the Atkinson cycle, the supply of air from the intake manifold to the combustion chamber is reduced as compared with the Otto cycle, so that the torque output generated by ignition of the combustion chamber is reduced. However, the Atkinson cycle generally has better fuel efficiency than the Otto cycle because of the large proportion of combustion energy converted into effective torque. By operating the combustion chamber in an Atkinson cycle, the combustion chamber can be operated at or near the lowest BSFC operating point.

  In the above example shown in FIGS. 12A-12C, both intake valves are activated simultaneously based on the same cycle. 12D-12E, consider an implementation example in which independently controlled intake valves open and close based on different cycles. The intake valves described in these embodiments may be controlled or actuated using any of the techniques described above (eg, those described with reference to FIGS. 1A, 1B and FIGS. 2-11).

  In FIG. 12D, the intake valve 1202b operates according to the EIVC Atkinson cycle. The intake valve 1202a operates according to the Otto cycle. Therefore, as shown in the figure, the intake valve 1202a closes about 40 ° after BDC, while the piston 1206 is already in the compression stroke. However, the intake valve 1202b closes sooner, that is, just before the end of the intake stroke when the piston is at BDC.

  FIG. 12E shows a system in which the intake valve 1202a operates according to the Otto cycle and the intake valve 1202b operates according to the LIVC Atkinson cycle. Intake valve 1202b is therefore closed later than intake valve 1202a, ie, about 90 ° after BDC during the compression stroke, rather than about 40 ° after BDC.

  The operation of the intake valve following different cycles provides various potential benefits. One provides another means of controlling the flow in the combustion chamber. For example, in FIG. 12D, air enters the combustion chamber 1206 asymmetrically. That is, during the intake phase, more air flows through one intake valve (intake valve 1202a) than the other intake valve for a longer time. This can have a desirable effect on the gas movement in the combustion chamber, i.e. more vortices can be produced. In FIG. 12E, during the compression stroke, more air is pushed from one intake valve (eg, intake valve 1202b) for a longer time than the other intake valve. This asymmetric air flow is a more advantageous feature that increases combustion promoting motion, i.e., vortex and rotational flow, to improve combustion characteristics.

  In some schemes, the intake valves are deviated, i.e., out of phase with each other. An example of this method is shown in FIG. 12F. The intake valves 1202a and 1202b operate based on the same Otto cycle, but the opening / closing timing is shifted. That is, the intake valve 1202a opens and closes earlier than the intake valve 1202b. The system functions somewhat similar to the system shown in FIG. 12E. Since air is exhausted asymmetrically from the combustion chamber, vortex flow may occur in the combustion chamber. The amount of deviation can vary greatly depending on the needs of a particular application.

  A further advantage of operating the combustion chamber intake valves independently according to different cycles is that the combustion chamber torque output can be highly controlled depending on how the valves are operated. Various exemplary valve control schemes will now be described with reference to FIGS. 13A and 13B. That is, the tables shown in FIGS. 13A and 13B show how the intake valves can be operated in different ways to generate different levels of torque. In some embodiments, the valve control schemes shown in FIGS. 13A and 13B each use the system shown in FIGS. 12D and 12E.

  FIG. 13A shows a combustion chamber valve control system having two intake valves that are independently controlled, for example by separate actuators or cams. The valve control system may have any of the features of the system described with reference to FIGS. 2-7 and / or 12D. During the selected combustion cycle, the intake valve 1202a can be paused or actuated according to the Otto cycle (hereinafter referred to as "normal valve"). During the selected combustion cycle, the intake valve 1202b can also be deactivated or activated according to an Atkinson (EIVC) cycle (hereinafter referred to as an “EIVC valve”). Thus, four different valve control schemes are possible for normal and EIVC valves, resulting in four different results 1302/1304/1306/1308, which are shown in Table 1300 of FIG. 13A.

  In results 1302, 1304 and 1306, the combustion chamber is ignited during the selected combustion cycle, and the level of torque output produced by the ignition depends on the valve control scheme. The result 1302 in the table indicates that the highest combustion chamber torque output is obtained when both intake valves are activated. This also creates a moderate vortex. If the EIVC valve is deactivated and the normal valve is activated, the next highest level combustion chamber output can be generated (result 1306). The next highest level of combustion chamber output (ie, lower than results 1302, 1306) occurs when the EIVC valve is activated and the normal valve is deactivated (result 1304). This is because the amount of air sent to the combustion chamber is limited by the EIVC operation. In both results 1304, 1306, actuating only one valve facilitates gas flow and mixing in the combustion chamber, which can generate a larger amount of vortex (ie, higher than result 1302). . Also, both intake valves can be deactivated, which means that no combustion occurs during the selected combustion cycle and no torque output occurs, as shown by the result 1308 in the table of FIG. 13A.

  FIG. 13B includes a similarly structured table 1350, where the intake valve 1202b can be paused or operated according to an Atkinson (LIVC) cycle (hereinafter referred to as a LIVC valve). The valve 1202a can be stopped or operated based on the Otto cycle (hereinafter referred to as a normal valve). Therefore, there are also four different valve control schemes for the selected combustion cycle: 1) LIVC valve operation, normal valve operation, combustion phenomenon occurrence, 2) LIVC valve pause, normal valve operation, combustion phenomenon occurrence, 3) LIVC valve stop, normal valve operation, combustion phenomenon occurs, 4) LIVC valve stop, normal valve stop, combustion phenomenon does not occur. The result of each valve control scheme is shown in FIG. 13B. The valve control system used to implement any of the valve control schemes of FIG. 13B has any of the features of the system described with reference to FIGS. 2-7 and / or 12E.

  The results shown in Table 1350 are completely different from Table 1300 in FIG. 13A. In particular, the maximum combustion chamber torque output is obtained when the normal valve is activated and the LIVC valve is deactivated (result 1356). If both valves are activated (result 1352), a lower, medium level combustion chamber output is obtained. This is because if both valves are activated, some of the air sent through the two valves will be pushed out of the combustion chamber due to the LIVC valve closing delay during the compression stroke. If the normal valve is deactivated and the LIVC valve is activated, a low level (ie, lower than result 1352) combustion chamber output is also obtained (result 1354). In result 1358, both intake valves are at rest and no torque output is generated.

  As described above, results 1354 and 1356 show a larger amount of vortex than result 1352 due to the asymmetric charge to the combustion chamber. Also, the LIVC valve and the normal valve can both be deactivated (result 1358), i.e., the combustion cycle can be thinned out.

  The tables shown in FIGS. 13A and 13B indicate that the flexibility of operation of the combustion chamber can be improved by using independently controlled intake valves and different cycles for different valves. That is, the combustion chamber can achieve three or four different level torque outputs. Also, the combustion chamber is better compared to some other technique (eg, reducing the torque output by adjusting spark timing, throttle, etc.) by selectively using the Atkinson cycle for one valve. It is possible to generate a lower level torque output with fuel efficiency.

  It should be understood that not all engine combustion chambers need have the same valve control system. Instead, the combustion chamber can be divided into two or more different sets, each having different capabilities. For example, one or more combustion chambers can operate in only two modes (ie, pause or ignite with all intake valves operating) or only in one mode (ie, ignite without decimation in all engine cycles) . However, other combustion chambers may have intake valves that are independently controlled as described above with reference to FIGS. This combination of combustion chambers improves flexibility and controllability compared to conventional engines, while at the same time hardware cost and complexity compared to engines where all combustion chambers are capable of multistage torque output. Helps to reduce sexuality.

  Examples of various different combustion chamber configurations are shown in FIGS. Each of these figures includes a table having a plurality of cells and indicators corresponding to power levels and cylinder numbers. Each table shows the different output levels (ie, torque output levels) that each cylinder (denoted by numbers 1-4) can generate in an exemplary four cylinder engine. That is, if a cylinder has a cell marked with output level 1, that cylinder can be ignited to produce a high torque output (eg CTF = 1.0, ie 100% of the maximum allowable output). It means that there is. If a cylinder has a cell marked with power level 2, that cylinder can be ignited to produce a low or partial torque output (eg CTF = 0.7, ie 70% of the maximum allowable output) It means that. If a cylinder has a cell marked with power level 3, it means that the cylinder can be deactivated (thus not producing torque output during the selected combustion cycle).

  Although only three output levels are available in the illustrated embodiment, other embodiments can generate more than two output levels in at least some of the cylinders, for example, as shown in FIGS. The tables 14A-14H in the figure show different configurations and combinations of combustion chamber / valve systems with different capabilities. The cylinders described in the table are configured to generate different power levels using any of the valve control systems, operations, and features described in this application (eg, described with reference to FIGS. 1-13). ing.

  Each table is also associated with fuel efficiency figures. Each fuel consumption value is based on a simulation executed by the present inventor. The numerical value indicates the estimated fuel efficiency gain that the configuration provides compared to a conventional four-cylinder engine (eg, having no ability to deactivate cylinders). It should be understood that the fuel economy figures associated with the tables of FIGS. 14A-14H are prepared in advance based on experimental simulations and may vary for different engine designs and applications.

  For comparison, FIG. 14A is a table showing a cylinder configuration in which all cylinders are capable of only two output levels, that is, each cylinder can be paused or ignited to produce one level of torque output. Such a configuration can be used in a skip fire engine control system. In this design, both intake valves operate at any ignition time. The air supply accompanying ignition can be adjusted by a cam phaser that controls the valve opening and closing time and a throttle that controls the MAP of all cylinders. In general, these control systems cannot adjust the output of a single combustion chamber significantly and rapidly. Although the output of the combustion chamber can be reduced by shifting the spark timing backward, it is often desirable to avoid this control method because of poor fuel consumption. In the cylinder configuration shown in FIG. 14A, ignition under such conditions helps to reduce the combustion chamber pumping loss, and in some cases the cylinder can be ignited with nearly optimal fuel consumption, thus providing moderate fuel consumption. Is good.

  FIG. 14B shows the configuration of a conventional engine that performs cylinder deactivation. The two cylinders are ignited during all engine cycles, i.e. they cannot be deactivated. During the selected combustion cycle, the other two cylinders can be ignited to produce a one level torque output or deactivated. Since such an engine cannot stop all cylinders, the fuel consumption may be slightly lower than the configuration shown in FIG. 14A. However, less hardware is required to implement such a system than, for example, a one-step skip fire engine design for every cylinder as shown in FIG. 14A.

  FIG. 14C shows a configuration in which all cylinders can be ignited at three output levels, that is, at rest (no torque output), and at two different output levels. Such a configuration can be used with any of the valve control systems described in the present application (eg, the intake valve is controlled independently for each cylinder, and the intake valve based on the Otto and Atkinson cycles is operated). is there. Such a method can greatly improve fuel efficiency. However, it may also be necessary to provide additional hardware and valve control related features in each cylinder.

  FIG. 14D shows a simpler scheme that allows the three output levels shown in FIG. 14C with two cylinders. However, the remaining two cylinders cannot be deactivated and are ignited at one power level during all engine cycles. Thus, cylinders 2 and 3 require little or no additional hardware compared to cylinders of conventional non-skip fire engines.

  In some embodiments, cylinders 1-4 shown in FIG. 14D are arranged to make the most efficient use of space in the engine. An example of such an arrangement is shown in FIG. FIG. 15 is a top view of a bank or row of cylinders 1 to 4 in engine 1500. The cylinders 1 and 4 are arranged at both ends of the bank, and the cylinders 2 and 3 are arranged at the center of the cylinder row.

  FIG. 15 shows an example in which more output levels / cylinders that can be deactivated are arranged at both ends of the cylinder bank, and cylinders that allow lower torque output levels and / or cannot be deactivated are arranged in the center. This makes it easier to install additional hardware on the cylinders at both ends of the bank, and the cylinders with less hardware requirements are smaller than the space and each side of each cylinder is in contact with another cylinder. Located in the center. Although the illustrated embodiment includes four cylinders, it should be understood that similar devices can be used for banks / rows with more or fewer cylinders (eg, 3-5 or more cylinders). . In other words, in various implementations, the outermost cylinder (e.g., the cylinder (group) at or near the end of the row) can have a higher output level, and the inner cylinder (e.g., at the center of the row). Lesser power levels are possible for cylinders (groups) that are close and / or surrounded by other cylinders on both sides. In an engine having two or more cylinder rows / banks, each cylinder bank / row may be arranged in the same manner as shown in FIG.

  FIG. 14E shows a modification of the configuration shown in FIG. 14d and / or 15. Similarly to FIG. 14D, in FIG. 14E, the cylinders 1 and 4 can output three levels. However, cylinders 2 and 3 are capable of two levels of output (ie, they can be paused or ignited at one torque output level). The configuration shown in FIG. 14E can also be arranged as shown in FIG. 15, with the innermost cylinders (cylinders 2, 3) requiring less hardware than the outermost cylinders (cylinders 1, 4). Less power levels may be associated.

  In FIG. 14F, each cylinder can have two output levels, but the types of possible output levels are different. In this configuration example, the cylinders 1, 4 can have two output levels, can be ignited to produce one torque output level, and can be paused at a selected combustion cycle. The cylinders 2 and 3 cannot be stopped, and can be ignited at two different output levels. Compared with a configuration in which all cylinders can generate three or more output levels, the configuration shown in FIG. 14F requires less hardware. Pre-tests have also shown that such a configuration of fuel is much better compared to a one level skipfire engine system (eg, as shown in FIG. 14A).

  FIG. 14G shows a configuration in which two cylinders (cylinders 1 and 4) are capable of three levels of output (ie, rest and ignition at two different torque output levels). The other two cylinders (cylinders 2 and 3) cannot be deactivated, but can be ignited to generate two different torque output levels. The configuration shown in FIG. 14G may also be arranged as shown in FIG. That is, the cylinders 1 and 4 capable of a higher output level are arranged at both ends of the cylinder row / bank, whereas the cylinders (cylinders 2 and 3) capable of a lower output level are arranged in the row / bank. It is arranged in the center or inside. As described above, in various embodiments, cylinders 1 and 4 require additional hardware to achieve additional power levels, and the outer ends of the cylinder row / bank have such hardware installed. Provide as much space as possible.

  FIG. 14H shows a modified example in which not all cylinders can be stopped or thinned out. However, each cylinder can be ignited to produce two different torque output levels. In various implementations, the configuration requires less hardware compared to a cylinder system with a lower NVH compared to a conventional skipfire engine control system and capable of more power levels. .

  The embodiments shown in FIGS. 14A-14H may be implemented using any of the valve control systems described in this application. That is, the various embodiments shown in FIGS. 14A-14H include one or more cylinders that can be paused and / or ignited to generate multiple levels of torque output. Such multi-stage torque output can be realized in various ways. In some implementations, for example, each cylinder includes two intake valves, and each intake valve is controlled by a different actuator (eg, as shown in FIGS. 2-7). To generate high torque output, air is supplied through both intake valves during the selected combustion cycle. To produce a low torque output, only one intake valve is charged during the selected combustion cycle or air is pushed out of the cylinder by the LIVC valve. As shown in FIGS. 2-7, control of one or more exhaust valves may be handled by one or more actuators. In some schemes, a cylinder has a single intake valve with adjustable valve lift so that the cylinder can be ignited to generate different torque output levels (eg, as described in connection with FIGS. 8 and 9). It is configured to have. The configurations shown in FIGS. 14A-14H can also be used in an engine system with any of the valve passage configurations described above (eg, as described in connection with FIGS. 10A, 10B, 11). In some designs, each cylinder capable of multi-stage torque output operates a different intake valve according to a different cycle (eg, as described in connection with FIGS. 12A-12E and 13A-13B). That is, the torque outputs at different levels shown in the tables of FIGS. 14A to 14H are different from those of the techniques shown in the tables of FIGS. 13A and 13B (e.g. For example, by pausing one of the valves to generate a torque output of 2).

Multi-stage Skip Fire Engine Control System Various embodiments of the present invention relate to a multi-stage skip fire engine control system. One or more combustion chambers of the engine can be ignited to generate at least two different levels of non-zero torque output. The combustion chamber output torque can be controlled for each ignition opportunity. The overall engine torque output can be controlled by igniting or deactivating the cylinder at each ignition opportunity. Based on the desired engine torque, the engine control system determines an ignition sequence to operate the engine in a skip fire manner. The sequence shows a series of pauses and ignitions. For each ignition, the sequence indicates the corresponding torque output level. The engine combustion chamber is operated based on an ignition sequence to transmit the desired engine torque. Such a skip fire ignition sequence is referred to herein as a multi-stage skip fire ignition sequence.

  The above-described embodiments of the multi-stage skipfire engine control system may be used with any of the engines, combustion chambers, intake channels, and valve control system designs described in this application. In various embodiments, for example, the system generates an ignition sequence that ignites from one or more combustion chambers at multiple torque output levels. Each of these combustion chambers uses independently controlled intake and / or exhaust valves, operates the same combustion chamber intake valves according to different cycles (eg, Otto and Atkinson), and / or drawings. Any other feature or technique described in connection with can cause such a high or low torque output ignition to occur. However, it should be understood that the described multi-stage skipfire engine control system is not limited to such systems and operations, and can be applied to any engine or combustion chamber design capable of generating multiple levels of combustion chamber output. The present invention is particularly applicable to a control system that determines ignition for each ignition opportunity, but is not limited to such a control system.

  Referring now to FIG. 16, a multi-stage skipfire engine controller 1630 according to a specific embodiment of the present invention will be described. The engine controller 1630 includes an ignition ratio calculator 1602, an ignition timing determination module 1606, an ignition control unit 1610, a power system parameter adjustment module 1608, and an engine diagnosis module 1650. The engine controller 1630 is arranged to operate the engine by a skip fire method.

  The engine controller 1630 receives input signals 1614 representing desired engine power and various vehicle operating parameters, such as engine speed 1632 and transmission gear 1634. The input signal 1614 can be treated as a request for the desired engine power or torque. Signal 1614 may be received or introduced from an accelerator pedal position sensor (APP) or other suitable source, such as a cruise controller, torque calculator, or the like. An optional preprocessor can change the accelerator pedal signal before it is transmitted to the engine controller 1630. However, it should be understood that in other implementations, the accelerator pedal position sensor can communicate directly with the engine controller 1630.

  The ignition ratio calculator 1602 is configured to receive the input signal 1614 (and other suitable source, if present) and engine speed 1632 and determine an ignition ratio suitable for delivering the desired output. In various embodiments, the ignition ratio may be any data that indicates or represents the ratio of ignition to ignition opportunity (ie, pause in addition to ignition).

In some implementations, the ignition ratio calculator 1602 first generates an effective ignition ratio. In various embodiments, the effective ignition ratio (EFF) is the product of the ignition ratio and the weighted average normalized reference cylinder charge during ignition operation. (Thus, in such an embodiment, the effective ignition ratio differs from the ignition ratio and does not necessarily clearly indicate the ratio of ignition to ignition opportunity). In various embodiments, the normalized reference cylinder charge or cylinder torque ratio has at least two potentially different non-zero values, each associated with a group of cylinders. Mathematically, the engine torque ratio (ETF) can be expressed by an effective ignition ratio (EFF).
ETF = EFF * CT ^act _H (Formula 5a)
Here, CTF ^act _H is the actual air supply in the cylinder group at the highest air supply level. In the case of a system having two supply levels, the high level torque supply may be referred to as complete supply, and the low torque level supply may be referred to as partial supply. In the various examples described above in this application, the amount of torque produced by combustion chamber ignition is characterized by a cylinder torque ratio (CTF), which indicates the combustion chamber output relative to a reference value. For example, the CTF value may relate to a reference ambient pressure and temperature, i.e. the maximum possible output torque generated by the combustion chamber when the throttle is fully opened at 100 kPa and 0 C, and appropriate valve opening and closing and ignition timing. Of course, other ranges and reference values may be used. In this application, CTF is generally a value between 0 and 1.0, but may be greater than 1.0 under certain conditions such as low ambient temperature and / or operation at sea level or in a supercharged engine. For some embodiments described in this application, the reference CTF value is 1.0 for full charge and the reference CTF value is 0.7 for partial charge. For convenience of explanation, these values will be used in the following description of the invention, but it should be understood that these values may vary depending on the exact engine design and engine operating conditions. It should be understood that the actual CTF transmitted from the combustion chamber can be adjusted from these reference values.

In some embodiments, the ignition ratio calculator 1602 determines one or more combinations of level ignition ratios and cylinder torque levels (eg, as shown in Equation 2) suitable for delivering the desired output. It is configured accordingly. These combinations can also be expressed as an effective ignition ratio (EFF) 1611. In some designs, the engine torque ratio (ETF) can be expressed as the product of EFF and the adjustment factor α.
ETF = EFF * CT ^act _H = EFF * CTF ^R _H * α (Formula 5b)
Here, CTF ^R _H is a reference cylinder torque ratio associated with the cylinder having the maximum cylinder charge. As described above, CTF ^R _H is assumed to be 1 here, but this is not essential. The adjustment coefficient α varies depending on engine parameter settings such as ignition timing and throttle and cam phaser positions.

  The ignition ratio calculator 1602 can generate the effective ignition ratio in various ways depending on the needs of a particular application. In some implementations, for example, the effective ignition ratio is selected from a library of predetermined effective ignition ratios and / or from a look-up table. Various implementations utilize lookup tables based on one or more engine parameters (eg, gear, engine speed, etc.), fuel consumption, maximum allowable CTF, and / or NVH associated with various effective ignition ratios. The effective ignition ratio is determined. These and other schemes are described in detail below.

  If the effective ignition ratio is determined by the calculator 1602, it is passed to the ignition timing determination module 1606. The ignition timing determination module 1606 is configured to issue a sequence of ignition commands that, based on the received effective ignition ratio, communicates to the engine the percentage of ignition and ignition output torque level necessary to generate the desired engine output. ing. The sequence can be generated in various ways, for example using a sigma-delta converter, using one or more lookup tables, or using a state machine. An ignition command sequence output from the ignition timing determination module 1606 (sometimes referred to as a drive pulse signal 1616) is an ignition control unit 1610 that adjusts actual ignition via an ignition signal 1619 directed to the engine combustion chamber 1612. Passed to.

  The sequence of ignition commands issued by the ignition timing determination module 1606 indicates the combination of pause and ignition and the torque level associated with the ignition. In various embodiments, for each ignition, the sequence indicates a particular torque output level selected from two or more possible torque output levels. The sequence may be in any suitable form. In some embodiments, for example, the sequence consists of values of 0, 0, 0.7, 1, and so on. In this example, during the next four ignition occasions, the corresponding combustion chamber is paused, paused, ignited (with a low level combustion chamber output, such as 70% of the reference cylinder torque output), and 100% of the reference cylinder torque output. Igniting (with high level combustion chamber output, etc.). An ignition sequence that exhibits pause and ignition with multiple levels of combustion chamber output is referred to herein as a multi-stage skipfire ignition sequence.

  The ignition timing determination module 1606 can determine the ignition determination and the ignition sequence in various ways. In various implementations, for example, the ignition timing determination module 1606 searches one or more lookup tables to determine an appropriate multi-stage ignition sequence. A suitable multi-stage ignition sequence may be configured to maximize fuel economy consistent with the realization of acceptable NVH features. Factors affecting NVH may include transmission gear, engine speed, cylinder charge, and / or other engine parameters. Based on the effective ignition ratio, fuel economy, NVH considerations and / or one or more of the factors discussed above, module 1606 selects a multi-stage ignition sequence from a plurality of ignition sequence options. In other implementations, the module 1606 uses a sigma delta converter or algorithm to determine an appropriate ignition sequence. Any suitable algorithm or process can be used to generate an ignition sequence that transmits the desired engine torque. Various techniques for determining the ignition sequence are described below with reference to FIGS.

In the embodiment shown in FIG. 16, a power system parameter adjustment module 1608 that cooperates with the ignition timing determination module 1606 is provided. The power system parameter adjustment module 1608 instructs the engine combustion chamber 1612 to properly set the selected power system parameters to ensure that the actual engine output is substantially equal to the requested engine output. To do. For example, under certain conditions, the output resulting from each ignition of the combustion chamber must be adjusted to transmit the desired engine torque. The power system parameter adjustment module 1608 can set any appropriate engine setting (eg, mass supply, ignition timing, cam timing, valve control, exhaust gas recirculation, throttle, etc.), and the engine output at which the actual engine output is requested. It plays a role in setting to help ensure that it is met. The engine output is therefore not limited to operation at a discrete level and can be adjusted continuously, i.e., analogy by adjusting the engine settings in various implementations. Mathematically, in some schemes this can be expressed by including a multiplicative factor in the output of each cylinder group. Therefore, Equation 2 can be modified and combined with Equation 5 as follows:
ETF = α * CTF ^R _H * EFF = α ₁ * CTF ^R ₁ * FF ₁ + α ₂ * CTF ^R ₂ * FF ₂ +. . . + Α _n * CTF ^R _n * FF _n (Formula 6)
Here, α ₁ , α _2, and α _n represent cylinder load adjustment coefficients associated with the respective cylinder groups, and CTF ^R ₁ , CTF ^R _2, and CTF ^R _n represent reference cylinder torque ratios of the respective cylinder groups. Some engine settings, such as throttle position, will affect the adjustment of all cylinder groups, while some settings, such as ignition timing and / or fuel injection amount, may vary from cylinder group to cylinder, and even from cylinder to cylinder. It should be understood that it can be adjusted to. In various implementation examples, the ignition timing and the amount of injected fuel differ for each different cylinder group. The ignition timing of each cylinder group can be adjusted so as to obtain the optimum fuel consumption for the cylinder group, and the injected fuel amount can obtain a substantially stoichiometric air / fuel ratio in all the cylinder groups. It can be adjusted. In this case, the amount of fuel injected is substantially proportional to the generated cylinder torque.

  The engine controller 1630 also includes an engine diagnostic module 1650. The engine diagnosis module 1630 is configured to detect any engine malfunction (eg, knocking, ignition failure, etc.) in the engine. Any known technique, sensor, or detection process may be used to detect the failure. In various embodiments, if a malfunction is detected, engine diagnostic module 1650 instructs ignition controller 1610 to perform a procedure that reduces the likelihood of a malfunction that may occur in the future. In various embodiments, a multi-stage skipfire ignition sequence has been generated to address potential failures. Various examples of operations that can be performed by the engine diagnostic device 1650 will be described later in this specification with reference to FIGS.

  It should be understood that the engine controller 1630 is not limited to the specific configuration shown in FIG. One or more of the illustrated modules can be integrated. Alternatively, the characteristics of a particular module may be distributed across multiple modules. One or more features from one module / component may (alternatively) be performed by another module / component. Engine controllers are also disclosed in U.S. Pat. Nos. 7,954,474, 7,886,715, 7,849,835, 7,577,511, 8,099,224, 131,445, 8,131,447 and 8,616,181, U.S. patent applications 13 / 774,134, 13 / 963,686, 13 / 953,615, 13 / 953,615, 13 / 886,107, 13 / 963,759, 13 / 963,819, 13 / 961,701, 13 / 963,744, 13/843 567, 13 / 794,157, 13 / 842,234, 13 / 654,244, 13 / 654,248, 14 / 638,908, 14 / 799,389 14th / Additions based on other patent applications including 07,109, and 14 / 206,918, US provisional patent applications 61 / 080,192, 61 / 104,222, 61 / 640,646 All of which are incorporated herein by reference for all purposes. Any of the features, modules, and operations described in the aforementioned patent documents can be added to the engine controller 1630 shown. In various alternative implementations, these functional blocks may be implemented using microprocessors, ECUs or other computing devices, using analog or digital components, using programmable logic, using the combinations described above, and / or Or it can be implemented algorithmically in any other suitable manner.

  Referring now to FIG. 17, a method for determining a multi-stage skipfire ignition sequence according to a particular embodiment of the present invention will be described. This method can be executed by the engine controller 1630 shown in FIG.

  Initially, at step 1705, the engine controller 1630 determines a desired engine torque based on the input signal 1614 (FIG. 16), current engine operating speed, transmission gear and / or other engine parameters. Input signal 1614 is derived from any suitable sensor (s) or operating parameter (s) including, for example, an accelerator pedal position sensor.

  In step 1710, the ignition ratio calculator 1602 determines an effective ignition ratio suitable for transmitting the desired torque. In various embodiments, as described above, the effective ignition ratio includes both the ignition ratio of each cylinder group and the torque level corresponding to the cylinder group. The determination of the effective ignition ratio may be based on any suitable engine parameters such as gear, engine speed, etc., and other engine characteristics such as NVH and fuel economy. In some embodiments, the effective ignition ratio is selected from a set of predetermined effective ignition ratios determined to be fuel efficient and / or have acceptable NVH characteristics given the engine parameters. The effective ignition ratio may be generated or selected using any suitable mechanism, such as one or more look-up tables described in connection with FIG. 18 of the present application. One way of determining an appropriate effective ignition ratio is shown in FIG. FIG. 18 shows an exemplary look-up table 1800 that includes indicators of engine speed and effective ignition ratio (EFF). This table is associated with a specific gear, i.e. other gears may correspond to other tables. Alternatively, in other versions of the illustrated table, gear is an additional indicator of the table. For each effective ignition ratio and engine speed, the table shows the maximum allowable level of combustion chamber torque output that still provides acceptable NVH performance. Each effective ignition ratio is based on a combination of the ignition ratio associated with each ignition level and the output at each level. In the case of a multi-stage skipfire engine having two cylinder groups with different torque levels, the effective ignition ratio (EFF) is expressed as the ratio of high-level ignition to total ignition, denoted as ignition ratio (FF) and HLF (high). be able to. The FF and HLF values associated with different effective ignition ratios are shown in FIG.

  The maximum allowable combustion chamber power value reflects the fact that NVH generally tends to increase at higher levels of combustion chamber power. Therefore, for any given engine speed and effective ignition ratio, it is desirable to ensure that the combustion chamber output does not exceed a certain level so that NVH is maintained at an acceptable level. In various embodiments, the ignition ratio calculator 1602 searches the table to find one or more effective ignition ratios that are suitable for transmitting the desired torque and satisfy the combustion chamber output requirements in the table.

  An example will be described to make the usage of the table easier to understand. In this example, the desired engine torque ratio is 0.2 and the engine speed is 1300 RPM. If the reference torque value associated with the high level combustion cylinder group is the maximum torque value, the effective ignition ratio must be greater than or equal to the engine torque ratio in order to generate the desired torque. Therefore, in this example, the required torque output can be generated only at an EFF value of 0.2 or more. Table 1800 of FIG. 18 shows in column 1802 a list of arrays of EFF values greater than 0.2.

  The ignition ratio calculator can search the row of column 1802 for engine speed 1300 RPM to find a suitable effective ignition ratio that provides optimum fuel economy and acceptable NVH simultaneously with the transmission of required engine torque.

For example, when the engine load (engine torque ratio) is 0.2, the effective ignition ratio is considered to be 0.57. By examining table 1800, the torque level associated with high torque ignition (CTF ^{act in} equations 5a, 5b) must be less than 0.14 CTF value for cell 1804 to obtain acceptable NVH performance. Indicates. However, in this case, the ETF value is only 0.57 * 0.14 = 0.08, which is significantly lower than the required torque level. Therefore, in this case, the use of the EFF value of 0.57 is excluded because the NVH and torque requirements are not met simultaneously. In various embodiments, the ignition ratio calculator 1602 searches the rows of the table 1800 until an appropriate effective ignition ratio is found. For example, at an effective ignition ratio of 0.70, the combustion chamber output (CTF) required for transmitting the desired torque is 0.2 / 0.70 = 0.29. By examining the table shown in FIG. 19, it can be seen that an EFF value of 0.7 corresponds to FF = 1 and HLF = 0. Accordingly, all ignitions are low level ignitions corresponding to the low level reference CTF value 0.7, and ignition is performed at all ignition opportunities, and in this case there is no pause.

  The high level combustion chamber output required to transmit the desired torque is 0.29, which is below the high level combustion chamber output threshold (0.58 of cell 1806) shown in Table 1800, so that effective ignition is performed for engine operation. It is conceivable to use a ratio. The ignition ratio calculator 1602 can continue to search for rows and determine that multiple effective ignition ratios meet the maximum combustion chamber output requirement of the table. Such effective ignition ratios are referred to as effective ignition ratio candidates in this specification.

  The ignition ratio calculator 1602 then selects one of the effective ignition ratio candidates. This selection may be made in any suitable manner. In some implementations, for example, the ignition ratio calculator 1602 searches another table or module that indicates relative fuel consumption or fuel consumption for each of a plurality of effective ignition ratios. Based on this fuel consumption information, the calculator selects one of the effective ignition ratio candidates. That is, calculator 1602 selects an ignition ratio candidate with the highest or highest fuel efficiency. The selected effective ignition ratio is used to transmit the desired engine power by adjusting the engine parameters to achieve the desired adjustment factor (described in connection with Equation 5) for each high and low level ignition. Assume the required torque output. In various implementations, the selected effective ignition ratio is generally selected based on maximizing fuel economy while operating with acceptable NVH performance. If an effective ignition ratio is selected or generated, it is passed to the ignition timing determination module 1606.

  Thereafter, in step 1715 of FIG. 17, the ignition timing determination module 1606 determines a multi-stage skip fire ignition sequence. The multistage skip fire ignition sequence indicates an ignition determination sequence (that is, ignition and skip). A combustion chamber torque output level is selected for each ignition in the sequence. In various embodiments, this selection is shown in the sequence.

  The multi-stage skipfire ignition sequence can be generated in various ways depending on the needs of a particular application. In some embodiments, for example, the ignition timing determination module 1606 searches one or more look-up tables indicating appropriate ignition sequences based on one or more selected engine parameters including an effective ignition ratio. Additionally or alternatively, the ignition timing determination module 1606 may include a sigma delta converter or circuit that outputs an ignition determination and / or an ignition sequence. Various different implementations are shown below in FIGS.

  19 to 20 show a specific implementation example. In this implementation, the ignition timing determination module 1606 determines the characteristics of the multi-stage skip fire ignition sequence using one or more lookup tables. An example of the reference table is shown in FIG. FIG. 19 is a table showing the ignition ratio (FF) and the high level ratio (HLF) for each set of effective ignition ratios (EFF). The ignition ratio (FF) indicates the ratio of ignition to ignition opportunity (eg, ignition and skip) over a period of multiple ignition opportunities. The ignition ratio does not necessarily assume a constant torque output level for each ignition. The level ratio (LF) is any value that serves to indicate the ratio of the ignition to the total number of ignitions each generating a particular (eg, high or low) level torque output. In the illustrated embodiment, a high level ratio (HLF) is used that indicates the ratio of high level torque output ignition to the total number of ignitions.

  In this particular example, the ignition of the combustion chamber is performed at two different levels of combustion chamber output: a high level (eg, 100% of the reference cylinder torque output) and a low level (eg, 70 of the reference cylinder torque output). %) Torque output. Since each ignition can generate two torque output levels, if the HLF is 1/3, 1/3 of the ignition over a period of time will generate a high level of torque output and 2/3 of the ignition will be low. Generate torque output. The systems and indicators described above can be modified as appropriate for different implementations, such as when the combustion chamber torque output is greater than two levels.

  Using the reference table shown in FIG. 19, the ignition timing determination module 1606 uses the effective ignition ratio (EFF) determined in step 1710 to characterize the multistage skip fire ignition sequence (eg, high level ratio and ignition ratio). To decide. Accordingly, when EFF is 0.57 in the example shown in FIG. 19, the ignition ratio is 2/3 and the high level ratio is 1/2.

  In various embodiments, the ignition timing determination module 1606 then generates a multi-stage skipfire ignition sequence in response to the determined ignition characteristics. That is, in the above example, if the ignition ratio is 2/3 and the high level ratio is 1/2, the ignition timing determination module 1606 generates an ignition sequence that mixes the results of the ignition opportunity over a selected period. Generate. During this period, 2/3 of the ignition determination is ignition and 1/3 is rest. One half of the ignition is associated with high torque output and the remainder is associated with low torque output. In some embodiments, the ignition sequence is in the form of a series of CTFs, indicating a pause, a high torque output ignition, a low torque output ignition, and another pause, for example, with a numerical sequence of 0, 1, 0.7, 0. be able to. The ignition sequence can be generated using any suitable algorithm, circuit, or mechanism.

  One such circuit is shown in FIG. FIG. 20 shows a sigma delta circuit 2000 that forms part of the ignition timing determination module 1606. In the illustrated example, the ignition timing determination module 1606 provides the ignition ratio (FF) and high level ratio (HLF) obtained from the table of FIG. 19 to the sigma delta circuit 2000 to generate an appropriate multi-stage skip fire ignition sequence. input. The circuit 2000 can be implemented in hardware or software (eg, implemented as part of a software module or in executable computer code). In the figure, the symbol 1 / z indicates a delay.

  The top of circuit 2000 effectively implements a first order sigma delta algorithm. In circuit 2000, the ignition ratio (FF) is provided at input 2002. A subtractor 2004 adds the ignition ratio 2002 and the feedback 2006. The sum 2008 is passed to the accumulator 2010. The accumulator 2010 adds the sum 2008 and the feedback 2014 to obtain the sum 2012. The sum 2012 is fed back to the accumulator 2010 as feedback 2014. The sum 2012 is passed to the quantizer 2018 and converted into a binary stream. That is, the quantizer 2018 generates an ignition value 2020 having a sequence of 0 and 1. Each 0 indicates that the associated combustion chamber needs to be paused. Each 1 indicates that the associated combustion chamber needs to be ignited. The ignition value is converted into a floating point number by the converter 2019 to generate a value 2022, which is input to the subtractor 2004 as feedback 2006.

  The bottom of the circuit shows the level of torque output that should be generated for each ignition indicated by the value 2020 for that ignition to transmit the desired torque. The value 2022 is passed to a multiplier 2023 that also receives the HLF 2001. Multiplier 2023 multiplies these two inputs. Accordingly, when pause is indicated by the value 2022, the output of the multiplier 2023 becomes zero. As a result of the above multiplication, a value 2026 is obtained and passed to the subtractor 2035. The subtractor 2035 subtracts the feedback 2027 from the value 2026. The resulting value 2037 is passed to the accumulator 2028. Accumulator 2028 adds value 2037 to feedback 2030. The resulting value 2032 is fed back as feedback 2030 to the accumulator 2028 and also passed to the quantizer 2040. The quantizer 2040 converts the input to a binary value, ie, 0 or 1. (Example: If the input value 2032 is> = 1, the output of the quantizer is 1. Otherwise, the output is 0.), the resulting high level flag 2042 will have the corresponding ignition ( Indicates whether the ignition is to generate a high level torque output (as indicated by the ignition value 2020). That is, in this example, if the high level flag 2042 is 0, the corresponding ignition should generate a low level output. If the high level flag 2042 is 1, the corresponding ignition should generate a high level output. (If the ignition value 2020 indicates rest, the high level flag 2042 is 0 and is irrelevant.) The high level flag 2042 is passed to the converter 2044 to convert the value to a floating point value. The resulting number 2046 is passed to the subtractor 2035 as feedback 2027.

  The above described circuit thus provides a multi-stage skipfire ignition sequence that can be utilized for engine operation. In this example, an ignition value 2020 is generated based on the ignition ratio (FF) (eg, as determined in step 1710 in FIG. 17 and / or a look-up table in FIG. 19). If the ignition value 2020 is 1, the corresponding combustion chamber is ignited. For each such ignition, the high level flag 2042 may be 0 or 1 depending on the (high) level ratio 2001 (eg, as determined using the lookup table of FIG. 19). If the high level flag is 1, the ignition is an ignition that generates a high level output. If the high level flag is 0, the ignition is an ignition that generates a low level output. If the ignition value 2020 is 0, the corresponding combustion chamber is deactivated. By passing the zero value to the multiplier 2023, the corresponding high level flag is also zero. As the time elapses, the circuit determines the ignition determination and the combustion chamber output level, for example 1 to 0 (that is, the ignition value 2020 is 0 or 1, the high level flag 2042 is 0 or 1), 0 to 0, 1 to 0, Two streams of binary values indicating 0 to 1 and 1 to 1 can be generated.

  FIG. 21 shows another circuit 2100 configured to generate a multi-stage skipfire ignition sequence based on, for example, the effective ignition ratio (EFF) determined in step 1710 of FIG. Such a circuit is sometimes referred to as a multi-bit or multi-stage sigma delta. The circuit is configured to generate from output 2102 representing an effective ignition ratio an output 2130 indicative of rest, ignition at high level torque output, or ignition at low level torque output.

  In the circuit, the input 2102 which is the EFF determined in step 1710 is passed to the subtractor 2104. Feedback 2132 is subtracted from input 2102. The resulting value 2106 is passed to the accumulator 2107. Accumulator 2107 adds feedback 2108 to value 2106. The resulting sum 2110 is fed back to accumulator 2107 as feedback 2108. Sum 2110 is also passed to subtractor 2126 and subtractor 2112. The value 2124 is defined as 1 indicating a high level of combustion chamber output. The value 2124 is passed to switch 2122 and subtractor 2126. The subtractor 2126 subtracts the value 2128 from the sum 2110 to generate a value 2124 and passes it to the switch 2122.

  In this example, the value 2114 is defined as 0.7, indicating a low level combustion chamber output. The value 2114 is passed to the subtractor 2112 and switch 2118. The subtractor 2112 subtracts the value 2114 from the sum 2110 to generate a value 2140 and passes it to the switch 2118.

  Switch 2118 receives three inputs: value 2114, value 2140, and value 2116. The value 2116 indicates the minimum level of combustion chamber output (eg, pause without generating torque). The switch 2118 passes the value 2114 or 2116 as its output in response to the value 2140. If the value 2140 is less than 0, the output of the switch 2118 is equal to the value 2116. If the value 2140 is greater than or equal to 0, the output of the switch 2118 is the value 2114. The switch output 2120 is passed to the switch 2122.

  Switch 2122 receives three inputs: value 2120, value 2128, and value 2124. The switch passes the value 2120 or the value 2124 as an output depending on the value 2128. If the sum 2128 is less than 0, the output of the switch 2130 is the value 2120. If the value 2128 is greater than or equal to 0, the output of the switch 2130 is the value 2124. The output of the switch 2122 is passed to the subtractor 2104 as feedback 2132.

  An output 2130 of the switch 2122 indicates an ignition determination, and if the ignition determination involves ignition, indicates an output level of torque due to ignition. In the illustrated embodiment, the output 2130 is either 0, 1, or 0.7. Accordingly, output 2130 indicates, based on input 2102, whether to suspend the corresponding combustion chamber during a particular combustion cycle, ignite with a high level output, or ignite with a low level output. Circuit 2100 skips over time (eg, skip, ignition at high level torque, ignition at low level torque, ignition at low level torque, pause, ignition at high level torque, etc.) It is configured to generate a sequence of values (eg, 0, 1, 0.7, 0.7, 0, 1, etc.) that form a fire ignition sequence.

  Note that in the above example the multi-stage skip ignition sequence is a mix of at least three different levels, 0, 0.7 and 1. Three different levels can be used to obtain the same or similar effective firing ratio from many different sequences. Using the ignition ratio calculator 1602 or the ignition timing determination module 1606 (FIG. 16), simultaneously transmit the required output torque level and acceptable NVH characteristics, and which of these multi-stage skip ignition sequences has the best fuel economy. You can decide what to give. Although slightly counterintuitive, the high engine torque output can be used to shift the noise and vibration generated by the engine from resonance and other undesirable frequencies, so that only the low engine torque output is used. May be desirable to insert high torque output ignition.

  FIG. 22 shows another method for determining the multi-stage skip fire ignition sequence based on the effective ignition ratio determined in step 1710 of FIG. In this scheme, the ignition timing determination module 1606 uses one or more lookup tables to select a multi-stage skip fire ignition sequence based on the effective ignition ratio (EFF) determined in step 1710.

  FIG. 22 includes an exemplary lookup table 2200. The lookup table 2200 shows a plurality of different multi-stage skip fire ignition sequences. Each sequence (ie, each row of the table) contains many ignition opportunity results and is associated with a different effective ignition ratio. The result of each ignition opportunity is defined in the table as 0 (indicating rest), 1 (indicating ignition at a high torque output level), or 0.7 (indicating ignition at a low torque output level). Each ignition opportunity is associated with a particular cylinder, as indicated by the columns associated with cylinders 1-4 of the four cylinder engine.

  In this example, the ignition timing determination module 1606 uses the table 2200 to determine a multi-stage skipfire ignition sequence that conveys an amount of engine torque that is substantially equal to the effective ignition ratio determined in step 1710. For example, if the effective ignition ratio is 0.47, the associated ignition sequence is 0.7, 0.7, 0, 0.7, 0.7, 0, 0.7, 0.7, 0, 0.7 , 0.7, 0. This means that in a continuous combustion cycle, the combustion chamber is ignited, ignited, paused, ignited, ignited, paused, ignited, ignited, paused, ignited, ignited, and paused. Using 0.7 for each ignition and the absence of 1 indicates that all ignition combustion chambers are ignited to produce a low torque output rather than a high torque output.

  It should be understood that FIGS. 18-22 show only a few ways of determining an appropriate multi-stage skipfire ignition sequence, and that the above techniques can be modified as appropriate to meet the needs of different applications. In some implementations, for example, there is no need to calculate an effective ignition ratio and / or no sigma-delta converter is required. Various embodiments determine a required torque (eg, as described in connection with step 1705 of FIG. 17) and reference one or more look-up tables to determine a skip fire ignition sequence based on the required torque. It is. In some schemes, table functionality is alternatively provided by software modules, software code, algorithms, or circuits.

  Returning to FIG. 17, in step 1720, the ignition timing determination module 1606 transmits the skip fire sequence to the ignition controller 1610. The ignition controller 1610 then assigns an ignition determination to the corresponding combustion chamber and causes the combustion chamber to operate accordingly. That is, as described in connection with step 1715, in various embodiments, each ignition in the sequence is associated with a selected torque output level (eg, high torque output, low torque output). The ignition control unit 1610 assigns each ignition in the sequence and its corresponding torque output level to a specific combustion chamber. The combustion chamber is ignited and operates to generate a corresponding torque output level.

  For example, if the ignition sequence indicates that the combustion chambers are sequentially paused, ignited with a high torque output, and then ignited with a low torque output, the ignition control unit 1610 instructs the corresponding combustion chamber to operate as such. . In various embodiments, this may be to independently control the corresponding combustion chamber intake valves to generate different torque output levels as indicated in the skipfire ignition sequence. The combustion chamber is described in connection with the valve control techniques described herein to generate different torque output levels (eg, FIGS. 1A, 1B, 2-11, 12A-12F, 13A-13B, 14A-14H, 15). ) May be used to operate. The combustion chamber may also have any design or configuration shown in this specification or the drawings described above. In various embodiments where not all combustion chambers are capable of ignition / pause or control at different torque levels, the control method described in FIGS. It should be understood that the room high level ignition / low level ignition-suggestions may be included to properly indicate ignition / pause.

  In various embodiments, the determination of the effective ignition ratio (step 1710), the determination of the ignition sequence and / or the selection of the selected combustion cycle and high or low level torque output of the combustion chamber (step 1715) is performed at each ignition opportunity. Is done. Therefore, the various operations described above can be quickly executed in response to changes in required torque and other conditions. In other embodiments, the period in which the above operations are performed is slightly less, for example, every second ignition opportunity or every engine cycle.

  The operations of method 1700 of FIG. 17 may be performed using any of the systems described in FIGS. For example, the method 1700 generates an ignition sequence in which each ignition is associated with a specific torque output level. In various embodiments, these torque output levels are the different output levels or torque output levels described in connection with FIGS. 13A-13B and 14A-14H. That is, if an ignition sequence (step 1720 in FIG. 17) is executed on the engine and the selected combustion chamber is ignited to produce a different level of torque output, any valve control mechanism and / or other system shown The torque output of the different level can be generated even when using.

Transition between engine torque ratio and effective ignition ratio One of the difficulties in skipfire engine control is the management of transitions between different engine output torque levels. Consider an example in which the accelerator is stepped in slightly because you want a little more torque. This greater torque requirement can only be realized by increasing the cylinder load beyond a level that indicates an acceptable level of NVH. Accordingly, different ignition ratios and level ratios are selected. However, when a new pattern is suddenly used, a separate NVH problem arises because the resulting change in the resulting torque is too rapid. As a result, it may be desirable to make a more gradual transition between the two effective ignition ratios.

  Such transitions can be managed using various techniques. As an example, the ignition timing can be adjusted so as to reduce the torque output during the transition. However, when the ignition timing is used in this way, the fuel efficiency is generally deteriorated. Another option is to manage transitions using multistage skipfire engine control.

  An exemplary technique is shown in FIG. FIG. 23 illustrates a method 2300 for managing transitions between first and second effective ignition ratios using multi-stage skipfire engine control. First, in step 2305, the engine is operated using a specific effective ignition ratio. Thereafter, the engine is operated using a second, different effective ignition ratio (step 2310). These different effective ignition ratios are generally associated with different engine output torque levels, but in some cases the engine torque can be kept constant throughout the effective ignition ratio transition.

  Each effective ignition ratio may be one that causes the engine to operate in a skip fire manner. In some cases, there may be various ignition patterns, while in other cases there may be a limited number of ignition patterns, for example, a ring cylinder deactivation where the cylinder alternates between ignition and deactivation at every ignition opportunity. . In some cases, the effective ignition ratio may correspond to a variable displacement operation where, for example, a certain set of cylinders are deactivated or all cylinder operations are used. Even if the variable displacement operation by a certain set of cylinders is not a skip fire operation, it is possible to transition between various fixed displacement levels using skip fire control if the engine hardware supports it. In some cases, such as during coasting, the effective ignition ratio may be zero. During each operating state of operating the engine using a specific ignition ratio, the engine can be operated using any of the techniques described in connection with FIGS. 16-22 or using other engine control techniques. .

  In step 2315, the engine is operated according to a multi-stage skip fire ignition sequence during the transition between the two effective ignition ratios. The multi-stage skipfire ignition sequence can be generated in various ways depending on the needs of a particular application. In some embodiments, for example, the effective ignition ratio is stepped up to one or more intermediate ignition ratios during the transition. A multi-stage skip fire ignition sequence is generated based on the intermediate ignition ratio (s) and used for engine operation for transition. The rate of change of the effective ignition ratio during the transition may be based on any suitable engine parameter, such as absolute manifold pressure. Any of the techniques described in connection with the drawings (eg, one or more lookup tables, sigma delta converters, etc.) may be used to generate the multi-stage skipfire ignition sequence. Various techniques using skip fire operation during transitions between modes are described in co-assigned US patent application Ser. No. 13 / 799,389, which is incorporated herein in its entirety for all purposes. Yes. Any of the techniques described in the patent literature can also be used.

  One method is to store a predetermined multistage skip fire ignition sequence in a library (eg, one or more reference tables). In various embodiments, each skipfire ignition sequence is associated with a specific effective ignition ratio. The ignition timing determination module 1606 examines the library and selects one of the predetermined sequences to determine an appropriate multi-stage ignition sequence for use in the transition. The selected sequence is then used for engine operation during the transition.

  Consider an example in which a four-cylinder engine is operated according to an ignition sequence in which four combustion chambers are ignited or stopped based on patterns 0.7, 0, 0.7, 0. That is, the combustion chambers 1 to 4 are repeatedly ignited, paused, ignited, and paused, and each ignition is a low-level output ignition (eg, including CTF = 0.7). Therefore, the effective ignition ratio equivalent to this type of engine operation is 0.35. The engine then transitions to another type of engine operation where the ignition pattern is 0.7, 0.7, 0.7, 0.7. That is, the combustion chamber repeats ignition and does not pause any combustion chamber. Each ignition produces the same low level output (eg, CTF = 0.7). The effective ignition ratio for this type of engine operation is therefore 0.7. That is, the engine output torque is assumed to be constant from the first effective ignition ratio (0.35) to the second effective ignition ratio (0.75), assuming that other engine parameters such as MAP and ignition time are kept constant. ) Is doubled at the transition to.

In this example, the ignition timing determination module 1606 looks up one or more lookup tables. Based on the corresponding effective ignition ratio, the lookup table (s) provide the following transition multi-stage skip fire ignition sequence (underlined below).

  Combustion chambers 1-4 then operate as engine transitions between the two effective ignition ratios based on the transition pattern described above. As a result, the engine torque increases in a stepwise manner, thus helping to smooth the transition and improve passenger comfort.

  It should be understood that the above-described use of the transition multi-stage skipfire ignition sequence is applicable to a wide variety of engines. Thus, it is not necessary for each combustion chamber of the engine to be able to pause and / or ignite at multiple torque output levels. For example, as described in connection with FIGS. 14A-14H above, only one or some of the combustion chambers may have the functions described above. In the above example, for example, only the first or third cylinder can be deactivated. The second and fourth cylinders are ignited in all engine cycles and the combustion chamber output can be adjusted between high and low levels.

  In some situations, it may be desirable to change the level ratio while transitioning between two effective ignition ratios. That is, in an engine control system that allows multiple levels of combustion chamber torque output, it may be useful to change the frequency at which a particular combustion chamber output level is used during transitions between effective ignition ratios.

  Consider an example where the engine is transitioning between two effective ignition ratios. When operating the engine using the first effective ignition ratio, the effective ignition ratio is ½ and the combustion chambers 1-4 of the engine are sequenced 1-0-1-0 (ie, high level combustion chambers). Ignition / pause at torque output and ignition / pause at high level combustion chamber torque output. When operating the engine using the second effective ignition ratio, the effective ignition ratio is 1, and the engine is sequence 1-1-1-1 (ie, all combustion chambers are ignited at a high level output). ). Therefore, the engine torque output doubles during the transition between the two effective ignition ratios, assuming that other engine parameters are kept constant.

  Since all of the above ignitions produce the maximum combustion chamber output, the ignition ratio for each of the above operating conditions is equal to the effective ignition ratio (assuming each ignition is CTF = 1.0) and both conditions The high level ratio (HLF) at 1 is 1 (ie, 100% of the ignition is high level output). In this example, each combustion chamber can also be ignited with a low level combustion chamber torque output (eg, CTF = 0.7). Each effective ignition ratio is characterized by the following values: That is, (X, Y), as shown in FIG. 19, X is an ignition ratio, and Y is HLF. Thus, the two states are characterized by (1/2, 1) and (1, 1).

  While transitioning between two different effective ignition ratios, it may be desirable to operate the engine in a skipfire manner using a different level ratio than that used when the engine is operated in one or both states. In the context of the above example, there is a change from (1/2, 1) to (1, 0) during the transition, ie the ignition sequence 0.7-0.7-0.7-0.7. That is, the combustion chamber is ignited at a low level output (eg, CTF = 0.7) within a subset of ignitions that are transitioning between two states. The effective ignition ratio therefore transitions from 1/2 to 1 via 0.7. The advantage of using low level ignition during the transition is that the NVH generated by such ignition is lower. This is because ignition includes a lower cylinder load and there is no pause in the ignition pattern.

  In the above example, the engine operates at a high level ratio of 1 when operating at a constant effective ignition ratio and operates at 0 during transitions between constant ignition ratios. The reverse is also possible. In other words, consider an example where each combustion chamber can be ignited again at one of two power levels, ie, a high power level (eg CTF = 1.0) or a low power level (eg CTF = 0.7). At the first effective ignition ratio, the engine operates at (1/2, 0). The engine operates at (1,0) at the target effective ignition ratio. That is, while operating at a constant effective ignition ratio, the engine operates at a high level ratio of 0 (ie, all ignition produces a lower level torque output). However, transitions have different high level ratios. In this example, the engine operates at a high level ratio of 1, that is, (1/2, 1) by the skip fire method. Accordingly, the effective ignition ratio varies from 0.35 to 0.7 via 0.5.

  In other embodiments, the effective ignition ratio can be filtered to slow down the transition between the initial and final ignition ratios. This can be accomplished by filtering the ignition ratio, filtering the level ratio, or filtering both quantities. The firing ratio and level ratio filtering techniques and time constants may be equal or different depending on the characteristics of the transition. Methods for filtering and managing transitions are described in US patent application Ser. Nos. 13 / 654,244 and 14 / 857,371, which are incorporated herein in their entirety for all purposes. Any of these methods may be used during the transition. For example, in some embodiments, the EFF transitions at a constant rate by transitioning the FF at a constant rate and monotonically transitioning the LF at an appropriately calculated rate. Alternatively, the LF or FF may not change monotonically by transitioning to the midpoint first and then to the final ratio (eg, from 1/2 to 0.7 via 1). . The intermediate value can be determined from a lookup table. For example, a two-dimensional table in which one dimension is the starting ratio and the second dimension is the target ratio is extremely useful. A third dimension, such as the rate of change of engine parameters or accelerator pedal position, may be added. In some cases, a constant effective ignition ratio is maintained, but it may be desirable to change the ignition ratio and level ratio. In this case, FF and LF can transition at a constant reverse speed so that their product, EFF, is kept constant.

Detection and management of knocking Knocking can be easily managed using multistage skip fire engine control. Knocking tends to occur more frequently under higher pressures or temperatures, for example, the combustion chamber is ignited with the maximum amount of air and fuel to produce the maximum possible torque output. Therefore, it is desirable to ignite the combustion chamber at a lower torque output level when knocking is detected under selected conditions.

  With reference now to FIG. 24, an exemplary method 2400 for reducing the likelihood of knocking in a multi-stage skipfire engine control system is described. First, in step 2405, the engine is operated according to a multi-stage skip fire ignition sequence. That is, the multistage skip fire engine controller 1630 receives the torque request and generates a multistage skip fire ignition sequence to transmit the desired torque. The engine operates based on an ignition sequence. In various embodiments, the engine operates using any of the multistage skipfire operations, mechanisms, and / or systems described in this application (eg, as shown in FIG. 16 or 17).

  At step 2410, engine diagnostic module 1650 (FIG. 16) detects a (potential) knock in one or more combustion chambers of engine 1612. Any suitable technique or sensor can be used to detect knocking that may occur in the engine. In some implementations, for example, the engine diagnostic module 1650 receives sensor data from one or more knocking sensors that detect a vibration pattern caused by the combustion chamber of the engine 1612. Engine diagnostic module 1650 analyzes the vibration pattern to determine whether knock has occurred.

  In response to detecting (potential) knocking in the combustion chamber of engine 1612, engine diagnostic module 1650 causes one or more selected combustion chambers to output a lower power level during one or more selected combustion cycles. Request to ignite only in (group) (step 2415). Consider an exemplary multi-stage skipfire engine control system where specific combustion chambers can be ignited at low (eg, CTF = 0.5), medium (CTF = 0.7), and high (CTF = 1.0) levels. . In response to detecting a (potential) knock in a particular combustion chamber, engine diagnostic module 1650 may prevent the combustion chamber from being ignited at one or more selected levels (eg, medium and / or high levels). To do. In other words, the (high) level ratio may be reduced / changed (eg from 1 to 0). This restriction may apply to one combustion chamber, some combustion chambers, or all combustion chambers. The constraints may also be applied over a predetermined time for a selected number of combustion cycles or all combustion cycles.

  In various embodiments, the engine diagnostic module 1650 communicates the above request to the ignition timing determination module 1606 so that future skipfire sequences take into account such constraints when determining the sequence for transmitting the required torque. Put in. In step 2420, the engine is operated by the skip fire method based on the request. That is, the engine operates as described in step 2405 except when the required torque is transmitted using only the allowed combustion chamber output level.

  Knocking tends to occur more frequently when the combustion chamber generates a high torque output, i.e., is ignited with a higher CTF. The reason is that the pressure and temperature in the combustion chamber tend to be significantly higher under such conditions. For example, there are means for lowering the pressure and temperature in the combustion chamber, such as adjusting the ignition timing. However, such techniques generally tend to have poor fuel economy. By reducing the supply air and limiting ignition to a lower torque output level, the possibility of knocking can be reduced while improving fuel efficiency.

  Optionally, engine diagnostic module 1650 includes features that re-enable high torque output ignition in response to a high torque request. In step 2425, the engine controller 1630 receives a high torque request, for example, based on data received from an accelerator pedal position sensor. In various embodiments, the high torque demand must exceed a predetermined threshold for the method to proceed to step 2430.

  In step 2430, the engine diagnostic module 1650 causes the engine control system to resume high power ignition in response to the high torque request. In other words, part or all of the restrictions imposed on the high power ignition imposed at step 2415 are released. In step 2435, engine diagnostic module 1650, ignition controller 1610, and / or power system parameter adjustment module 1608 performs one or more appropriate actions to reduce the risk of further knocking. Any known technique, such as ignition timing adjustment, can be used to reduce the risk of knocking.

Reduced cylinder deactivation and start / stop features Multi-stage skipfire engine control can also be used in certain situations where there is no combustion chamber being ignited and the manifold absolute pressure rises to atmospheric pressure level. For example, when the vehicle is coasting and / or on the verge of stopping, the driver can take his / her foot off the accelerator pedal. Under such circumstances, various engine systems can transition to a mode called decelerating cylinder (DCCO). In this mode, to save fuel, the engine cylinders are deactivated while the engine does not require torque. During this period, the intake and exhaust valves are closed and no air is supplied from the intake manifold into the combustion chamber of the engine.

  Another situation is when performing a start / pause feature. That is, some engine systems save fuel by turning off the engine without idling when the vehicle is stationary. In both of the above situations, the manifold absolute pressure (MAP) is equal to the atmospheric pressure because no air is supplied from the intake manifold into the combustion chamber. One problem with this is that the high MAP causes the engine to transmit more torque than necessary when the accelerator pedal is depressed again or other engine control requires torque. If no action is taken to reduce such torque surges, the vehicle and / or engine may accelerate rapidly.

  Multi-stage skipfire engine control can be used to address the above problems. One exemplary method 2500 is shown in FIG. First, in step 2505, the engine is operated according to a multi-stage skip fire ignition sequence. That is, the multistage skip fire engine controller 1630 receives the torque request and generates a multistage skip fire ignition sequence to transmit the desired torque. The engine operates based on an ignition sequence. In various embodiments, the engine operates using any of the multistage skipfire operations, mechanisms, or systems described in this application (eg, as shown in FIG. 16 or 17).

  In step 2510, engine controller 1630 (or any suitable module in the controller) detects that one or more conditions exist. In some embodiments, for example, the controller 1630 detects that the engine was coasting / decelerating, entered a DCCO condition, and / or that current torque is being requested. In other embodiments, the controller 1630 detects that the engine has stopped using the start / pause feature and that torque is being requested again.

  In response to detecting the condition (s), the controller 1630 requests that one or more selected combustion chambers be ignited at only the low torque output level (s) during the one or more selected combustion cycles. (Step 2515). The request can take a variety of forms. In some embodiments, for example, the controller 1630 prevents one or more higher combustion chamber power levels (eg, CTF = 1.0) from being used at all. In other words, the high level ratio is lowered or maintained at a lower level (eg, set to 0, 1/2, etc.). The request may include any of the operations and features described in connection with step 2415 of FIG. 24, for example, any number of combustion chambers or combustion cycles may be constrained in this manner.

  In step 2515, the engine is operated in a multi-stage skip fire system based on the request. That is, the engine operates as described in step 2505 except when the required torque is transmitted using only the allowed combustion chamber output level. In some embodiments, the request is valid until certain conditions are met or for a predetermined period of time, after which normal multi-level skipfire engine operation is resumed. Alternatively or additionally, the high level ratio may increase in steps over time until normal multi-stage skipfire engine operation is resumed. This gradual increase can be dynamically adjusted based on one or more engine parameters, such as manifold absolute pressure. The use of a lower high level ratio and / or a lower combustion chamber torque output level helps mitigate the effects of high MAP.

  As an option, the engine controller 1630 may have features that enable high power ignition again in response to a high torque demand. In step 2525, the engine controller 1630 receives a high torque request based on data received from, for example, an accelerator pedal position sensor. In various embodiments, the high torque demand must exceed a predetermined threshold for the method to proceed to step 2530.

  In step 2530, the engine controller 1630 causes the ignition control unit 1610 to resume high output ignition in response to the high torque request. That is, a part or all of the restrictions imposed on the high torque output ignition imposed in step 2515 are released.

  Any of the steps of method 2500 can be modified as appropriate for different applications. For example, US patent application Ser. No. 14 / 743,581, hereinafter referred to as the '581 application and cited in its entirety for all purposes, describes various techniques for performing start / pause features with skipfire engine control. Has been. Any of the features or operations described in the '581 application may be included in the method 2500.

Engine diagnostic applications The use of multistage skipfire engine control can also affect the design of engine diagnostic systems. In various engine diagnostic systems, engine failures are detected based on measurements of specific engine parameters (eg, crankshaft acceleration). In various embodiments, such a system takes into account the effects of ignition producing different levels of torque output.

  Referring to FIG. 26, an exemplary method 2600 for diagnosing engine failure will be described. First, at step 2605, the engine diagnosis module 1650 obtains ignition information from, for example, the ignition timing determination module 1606 and / or the ignition control unit 1610. The ignition information includes, but is not limited to, ignition determination (eg, skip or ignition), ignition sequence, and corresponding combustion chamber identification information. The ignition information also includes information indicating the level of combustion chamber output associated with each determination to ignite the combustion chamber.

  In step 2610, engine diagnostic module 1650 assigns a window to each ignition opportunity. The window may be any suitable time or interval corresponding to the target ignition opportunity of the target combustion chamber. Certain engine parameters are later measured over the window period to help determine if an engine failure has occurred in the target combustion chamber during the window period. The characteristics of the window can vary depending on the type of engine parameter measurement.

  Consider an example involving a 4-stroke 8-cylinder engine. In this example, the assigned window is the window angle section covering 90 ° rotation of the crankshaft. During the window period, the target combustion chamber is ignited. That is, in this example, the window corresponds to the first half of the power stroke of the target combustion chamber. It should be understood that the window may have any suitable length depending on the needs of a particular application.

  At step 2615, engine diagnostic module 1650 determines a combustion chamber torque output associated with one or more combustion chambers during the assigned window period. In other words, in various embodiments, the ignition timing determination module 1606 and / or the ignition control unit 1610 assigns an ignition determination to each combustion chamber. During the particular window assigned in step 2610, the target combustion chamber has been ignited. During the same window, the other combustion chambers are at different stages of the operating cycle. To use the above example, some combustion chambers have already completed a power stroke, others are in the process of completion, or later move on to a power stroke. Each combustion chamber is configured to pause or ignite for the power stroke associated with it. For each ignition, a specific combustion chamber output level is assigned, for example, ignition at low torque output, ignition at high torque output, and the like. Engine diagnostic module 1650 determines the combustion chamber torque output associated with one, some or all of the combustion chambers during the assigned window period.

  In step 2620, engine diagnostic module 1650 sets an engine parameter threshold or model. In some embodiments, for example, the engine diagnostic module 1650 determines an engine parameter threshold (eg, crankshaft acceleration threshold) that is later used to help determine if an engine failure exists. That is, the threshold value is useful for indicating the expected value of the subsequent engine parameter measurement, assuming ignition information (step 2605) and torque output level determination (step 2615). In other embodiments, the engine diagnostic module 1650 determines a model (eg, torque model) that also helps identify engine failures. For example, a torque model can be used to help show more torque generated by the combustion chamber during the window period. The model is made for one or more combustion chambers during the window period (eg, as indicated by the ignition information obtained at step 2605) and for each ignition (eg, at step 2615). Take into account the corresponding torque output level (as determined by

  In step 2625, engine diagnostic module 1650 measures engine parameters during the window period. Various engine parameters may be used depending on the needs in a particular application and the engine fault to be diagnosed. Some designs measure, for example, crankshaft acceleration, MAP, and / or oxygen sensor output during the window, but any suitable parameter can be measured. It should be understood that different measurements may use different windows.

  Based on the measurement (step 2625) and the threshold / model (step 2620), the engine diagnostic module 1650 then determines whether an engine failure exists. This determination can be made in various ways. In some embodiments, for example, crankshaft acceleration is measured (step 2625). This measurement is used to estimate the actual torque generated during the window period. The torque is compared with the expected torque calculated using the torque model (eg, step 2620). If the actual torque is smaller than the expected torque, the engine diagnosis module 1650 determines that an engine failure (eg, ignition failure) exists. In another implementation, the measured crankshaft acceleration is compared to a threshold (eg, step 2620) and no torque estimation is required. If the actual measurement exceeds the threshold, it is assumed that an engine failure exists or is likely to exist.

  The following example is given to better illustrate how some embodiments of the method are implemented. In this example, the engine has four cylinders and eight cylinders, and the cylinders are ignited in the order of 1-8-7-2-6-5-4-3. Each cylinder has an independently controlled intake valve and / or can operate the valve according to a different cycle as described in connection with FIGS. As a result, each cylinder can be ignited at one of two torque output levels, eg, low torque output (eg, CTF = 0.7) or high output (CTF = 1.0) when ignited.

  The engine diagnostic module 1650 is arranged to determine whether or not an ignition failure has occurred in the combustion chamber 8. The module ignites the combustion chambers 1, 8, 7, 2, 6, 5, 4, and 3, respectively, during a continuous ignition opportunity, indicating that they pause, ignite, pause, ignite, pause, ignite, pause Information is acquired (step 2605). The module assigns a window to the above ignition opportunity of the combustion chamber 8 (step 2610). The assigned window occurs when the cylinder 8 is in the first half of its power stroke and covers the 90 ° rotation of the crankshaft.

  In this example, engine diagnostic module 1650 also determines that each of the ignitions described above, including ignition of combustion chamber 8, is a low torque output (step 2615). In this example, module 1650 determines a crankshaft acceleration threshold that takes into account the cylinder torque output level. That is, unlike the previous case, if the engine diagnostic module 1650 determines that one, some, or all of the ignitions described above have a high torque output, the threshold will be different.

  In various embodiments, the crankshaft acceleration threshold is particularly strongly influenced by the operation of the combustion chamber 8, i.e., whether the cylinder 8 has been ignited with a low or high torque output. However, torque output levels associated with other cylinders can also have an effect. For example, if the cylinder 8 is in the first half of the power stroke during the assigned window period, the cylinder 1 is in the second half of its power stroke. Whether the cylinder 1 has been ignited with a low torque output instead of a high torque output may also significantly affect the threshold.

  Engine diagnostic module 1650 then measures the actual crankshaft acceleration during the window (step 2625). Module 1650 compares the measured value to a threshold value. When the measured value falls below (substantially) the threshold value, it is determined that an ignition failure has occurred in the combustion chamber 8 (or an ignition failure may have occurred).

  The example and method 2600 described above may be modified in various ways for different applications. For example, co-assigned US patent applications Nos. 14 / 207,109, 14 / 582,008, 14 / 700,494, and 14/14, which are incorporated herein in their entirety for all purposes. Nos. 206 and 918 describe various engine diagnostic systems and operations. Any of the features or operations described in these applications can be included in method 2600.

  Any and all of the above components may be configured to update their decisions / calculations very quickly. In some preferred embodiments, these determinations / calculations are updated at every ignition opportunity, but this is not required. In some embodiments, for example, determination of (effective) ignition ratio (step 1710 in FIG. 17), determination of multi-stage skipfire ignition sequence (step 1715), and / or engine operation based on the sequence (step 1720) may include: , Executed at every ignition opportunity. One advantage of controlling the various components at each ignition opportunity is that the engine's responsiveness to altered inputs and / or conditions is greatly improved. Although the operation at each ignition opportunity is very effective, good control can still be achieved even if the various components are updated more slowly (e.g., the ignition ratio / sequence is determined by two at each crankshaft rotation). It should be understood that this may be done at each ignition opportunity).

  The present invention has been described on the premise of operating a naturally aspirated four-stroke internal combustion piston engine that is primarily suitable for use in an automobile. However, it should be understood that the above application is very well suited for use with a wide range of internal combustion engines. These include engines for virtually any type of vehicle, including automobiles, trucks, boats, aircraft, motorcycles, scooters, etc., and virtually any other application that uses an internal combustion engine to ignite the combustion chamber. It is out. The various schemes described include virtually any type of two-stroke piston engine, diesel engine, Otto cycle engine, dual cycle engine, Miller cycle engine, Atkinson cycle that works with engines operating under a wide range of different thermodynamic cycles. Includes engines, Wankel engines and other types of rotary engines, combined cycle engines (such as combined Otto and diesel engines), hybrid engines, and star engines. Also, the method described above will work well with newly developed internal combustion engines, whether or not they operate using currently known or future developed thermodynamic cycles. . A supercharged engine such as one using a supercharger or turbocharger may be used. In this case, the maximum cylinder load may correspond to the maximum cylinder air supply obtained by supercharging the intake air.

  It should be understood that any of the methods or operations described herein can be stored in the form of computer code executable on a suitable computer readable medium. These operations are performed when the processor executes computer code. Such operations may include an ignition ratio calculator 1602, an ignition timing determination module 1606, an ignition controller 1610, a power system parameter adjustment module 1608, an engine controller 1630, an engine diagnostic module 1650, or any other module described in this application. , Components, or all operations performed by a controller.

  Some of the embodiments described above refer to combustion chamber deactivation. In various implementations, the combustion chamber pause prevents air from passing through the pause combustion chamber during one or more selected pause combustion cycles by a pumping action. The combustion chamber can be paused or stopped in various ways. In various systems, after exhaust gas is exhausted from the combustion chamber in the preceding combustion cycle, a low pressure spring is formed in the combustion chamber, and neither the intake valve nor the exhaust valve is opened during the subsequent combustion cycle. Is formed. In yet another embodiment, a high pressure spring is formed in the quiescent combustion chamber to prevent air and / or exhaust gas from escaping from the combustion chamber. The combustion chamber may be paused in any suitable manner that generates little or no power during the power stroke of the combustion chamber.

  The present application also refers to the concept of a combustion chamber that is used to generate different levels of torque or to have different charge or cylinder load levels. For example, these levels of torque output can be indicated in a multi-stage skipfire ignition sequence and / or stored in a lookup table or library. As mentioned above, in some embodiments, each level of such torque output is different for the different sets of operations described in this application (eg, opening only one intake valve, opening both intake valves, different For example, using a different cycle for the intake valve). In some schemes, the level of torque generated by the combustion chamber may vary from ignition opportunity to ignition opportunity. For example, the cylinder is deactivated during the combustion cycle, ignited with a high torque output during the next combustion cycle, ignited with a low torque output during the next combustion cycle, and then deactivated or ignited at any torque output level Can do.

  Various embodiments of the present invention primarily pause the cylinder during the resting combustion cycle by pausing both the intake and exhaust valves to prevent air from passing through the cylinder during the resting combustion cycle. It has been described on the assumption of a skip fire control device. However, some skip-fire valve operation schemes take into account effectively deactivating the cylinder by deactivating only the exhaust valve or only the intake valve and preventing air from passing through the cylinder by the pumping operation. I want you to understand. Some of the described schemes work equally well in such applications. Further, it is generally preferred to prevent air from passing through the idle cylinder during the combustion cycle that is culled by deactivating the cylinder, but air may pass through the cylinder during the selected idle combustion cycle. There are some specific cases that are desirable. This may be desirable, for example, when engine braking is desired and / or in certain exhaust device related diagnostic or operational requirements. This may also be useful in getting out of a DCCO (Decelerated Cylinder) condition. The described valve control scheme works equally well in such applications.

  This application refers to various systems and techniques that selectively generate a plurality of different (eg, high or low) torque output levels from an ignition combustion chamber. It should be understood that in various embodiments, various engine conditions remain substantially the same (although not essential) during the selected combustion cycle in which the combustion chamber ignites. Such engine conditions include, but are not limited to, manifold absolute pressure, cam phasor settings, engine speed, and / or throttle position. In other words, the present application generates different level torque outputs for each ignition combustion chamber without the need to vary, for example, throttle position, MAP, engine speed, and / or cam phasor settings to generate different levels of torque output. Various exemplary valve control systems and techniques are described (eg, as described in connection with FIGS. 1A, 1B, 2-11, 12A-12F, 13A, 13B, 14A-14H, and 15). doing.

  Various implementations of the present invention have accumulators and other mechanisms to determine the portion of ignition that was requested but not transmitted or transmitted but not required so that an ignition decision is made at each ignition opportunity. Very well suited for use in conjunction with a dynamic skipfire action to track. However, the described techniques are applicable to virtually any skipfire application (including skipfire operation using a fixed firing pattern or sequence) that can occur when rotating cylinder deactivation and / or various other skipfire techniques are used. It is equally well suited for operation modes in which the individual cylinders are sometimes ignited and occasionally paused when operating in a particular operating mode. Similar techniques can be used in variable stroke engine control in which the number of strokes in each combustion cycle is changed in order to effectively change the engine displacement.

  Although only a few embodiments of the present invention have been described in detail, it should be understood that the present invention can be implemented in many other forms without departing from the spirit or scope of the invention. We often refer to the term ignition ratio. It should be understood that the ignition ratio can be interpreted or expressed in various ways. For example, the ignition ratio may take the form of an ignition pattern, sequence, or any other ignition feature that includes or essentially represents the percentage of ignition described above. We often refer to the term “cylinder”. It should be understood that in various embodiments, the term cylinder should be construed broadly to include any suitable type of combustion chamber. The engine can also use skip-fire technology that operates with either low torque or high torque output ignition rather than cylinders that operate with pauses and ignition. In this control scheme, called dynamic ignition level modulation, the cylinder does not pause. With dynamic ignition level modulation, the output of the ignited cylinder dynamically varies according to the rest / ignition type pattern. For example, a particular cylinder can sometimes be ignited with a "high" or "higher" torque output level, sometimes ignited with a "low" or "lower" torque output level, and a "low" output level in a skip fire pattern Corresponds to “pause” and “high” power level corresponds to ignition. Accordingly, the embodiments are to be considered as illustrative and not restrictive, and the invention is not limited to the details set forth herein.

Claims (30)

A method of controlling the operation of an internal combustion engine having a plurality of combustion chambers to transmit a desired output, each combustion chamber having at least one intake valve and at least one exhaust valve, the method comprising:
The selected low combustion cycle is ignited at a low torque output, the selected high combustion cycle is ignited at a high torque output, and the engine is operated in a multi-stage skipfire system that skips the selected combustion cycle. Dynamically determining at each ignition opportunity during operation of the engine whether to ignite each combustion cycle with a high or low torque output or skip a selected combustion cycle;
Adjusting the charge to each ignition combustion cycle based on whether the high or low torque output has been selected for the ignition combustion cycle.   The method of claim 1, wherein the low torque output combustion cycle includes the use of an intake valve early closing (EIVC) cycle for the high torque output combustion cycle.   The method of claim 1, wherein the low torque output combustion cycle includes the use of an intake valve late closure (LIVC) cycle for the high torque output combustion cycle. The method according to any one of claims 1 to 3,
Determining a desired effective ignition ratio that represents an ignition opportunity ratio that needs to be ignited at a reference output level to deliver the desired output;
Determining a combustion cycle for igniting at a high output and a combustion cycle for igniting at a low output based on the effective ignition ratio. The method of claim 4, wherein
The effective ignition ratio is determined by an ignition ratio calculator;
A method wherein the determination of a combustion cycle for igniting at high power and a combustion cycle for igniting at low power is made by an ignition level determination module.   6. The method of claim 5, wherein the ignition level determination module uses a sigma delta converter to determine a combustion cycle that ignites at high power and a combustion cycle that ignites at low power.   6. The method of claim 5, wherein the ignition level determination module uses a look-up table to determine a combustion cycle that ignites at high power and a combustion cycle that ignites at low power.   8. The method according to claim 1, wherein the selection of the high or low torque output in each combustion cycle is determined using a sigma delta converter. 9. The method according to claim 6 or 8, wherein the sigma delta converter is
Analog parts,
Digital parts,
A method that is implemented using at least one of programmable logic.   10. A method according to claim 6, 8 or 9, wherein the sigma delta converter is implemented using program instructions executed by a processor. 11. A method as claimed in any preceding claim, wherein the selection of the high or low torque output is based on a look-up table .   12. A method according to any one of the preceding claims, wherein during each low torque combustion cycle, the corresponding combustion chamber is not stopped and fuel is supplied to the corresponding combustion chamber, Combustion occurs and net torque is transmitted by the low torque combustion cycle.   13. A method according to any one of the preceding claims, wherein each of the combustion chambers has at least two corresponding intake valves and the at least two of the combustion chambers associated with the combustion cycle. A method of adjusting supply air in each combustion cycle so as to generate a high torque output or a low torque output by independently controlling an intake valve.   14. A method according to any one of the preceding claims, characterized in that all of the valves are actuated by one or more cam lobes connected to one or more camshafts. 15. The method according to any one of claims 1 to 14, wherein each combustion chamber includes a first intake valve and a second intake valve, the method further comprising:
A method comprising operating the engine in a multi-stage skipfire manner during a selected combustion cycle and simultaneously opening and closing the first and second intake valves based on different timing cycles. The method of claim 15, wherein
The first intake valve operates based on one of an intake valve early close (EIVC) cycle and an intake valve late close (LIVC) cycle;
The method wherein the second intake valve operates based on an Otto cycle. A method according to any one of claims 1 to 16,
Each combustion chamber includes a first intake valve and a second intake valve;
When a combustion cycle is ignited with a high torque output, the first and second intake valves of the corresponding combustion chamber are independently controlled based on a high torque valve control scheme;
When a combustion cycle is ignited with a low torque output, the first and second intake valves of the ignition combustion chamber are controlled independently based on a low torque valve control scheme different from the high torque valve control scheme. Feature method. The method of claim 17, wherein
The high torque valve control scheme allows air to be supplied through the first and second intake valves during a selected combustion cycle;
The method wherein the low torque valve control scheme does not charge through the first intake valve during a selected combustion cycle. The method of claim 17, wherein
The high torque valve control scheme supplies air through the first intake valve instead of the second intake valve during the selected combustion cycle;
The high torque valve control scheme further operates the first intake valve based on an Otto cycle during the selected combustion cycle;
The low torque valve control scheme supplies air through the first and second intake valves during a selected combustion cycle;
The low torque valve control scheme further operates the first intake valve based on an Otto cycle during the selected combustion cycle, and based on an intake valve late close (LIVC) cycle during the selected combustion cycle. And operating the second intake valve. The method of claim 17, wherein
The high torque valve control scheme supplies air through the first intake valve instead of the second intake valve during a selected combustion cycle;
The high torque valve control scheme further operates the first intake valve based on an Otto cycle during the selected combustion cycle;
The low torque valve control scheme supplies air through the first and second intake valves during a selected combustion cycle;
The low torque valve control scheme further operates the first intake valve based on an Otto cycle during the selected combustion cycle and is based on an intake valve early closing (EIVC) cycle during the selected combustion cycle. And operating the second intake valve. A method of controlling the operation of an internal combustion engine having a plurality of combustion chambers to transmit a desired output, each combustion chamber having at least one intake valve and at least one exhaust valve, the method comprising:
The selected low combustion cycle is ignited at a low torque output, the selected high combustion cycle is ignited at a high torque output, and the engine is operated in a multi-stage skipfire system that skips the selected combustion cycle. Dynamically determining at each ignition opportunity during operation of the engine whether to ignite each combustion cycle with a high or low torque output or skip a selected combustion cycle;
Adjusting the charge to each ignition combustion cycle based on whether the high or low torque output has been selected for the ignition combustion cycle; and
A method wherein the combustion cycle ignited at the low torque output ignites with a minimum brake specific fuel consumption condition for the combustion cycle. An ignition controller used to control operation of an internal combustion engine having a plurality of combustion chambers to transmit a desired output, each combustion chamber having at least one intake valve and at least one exhaust valve, Ignition controller
To the determination of output for the combustion cycle, low power combustion cycle that is selected so as to ignite at low torque output, selected high power combustion cycle so as to ignite at high torque output, and the ignition is allowed so as not selected skipped ignition level determination unit identifies the combustion cycle, i.e. an ignition level determination unit configured to make the determination before Kide force to each ignition opportunities during operation of the engine,
Ignition controller in a low output combustion cycle, characterized in that it comprises a configured ignition control unit in order to direct the operation of the intake valves so as to supply less air than a high output combustion cycle.   23. The ignition controller according to claim 22, wherein the ignition control unit is configured to indicate that a low torque output combustion cycle uses an intake valve early closing (EIVC) cycle for the high torque output combustion cycle. Features an ignition controller.   23. The ignition controller according to claim 22, wherein the ignition control unit is configured to indicate to use an intake valve late closing (LIVC) cycle for the high torque output combustion cycle for a low torque output combustion cycle. Features an ignition controller. The ignition controller according to any one of claims 22 to 24,
Comprising an ignition ratio calculator configured to determine a desired effective ignition ratio that represents an ignition opportunity ratio that needs to be ignited at a reference output level to transmit the desired output;
The ignition controller, wherein the ignition level determination unit determines a combustion cycle to be ignited at a high output and a combustion cycle to be ignited at a low output based on the effective ignition ratio. 26. The ignition controller according to any one of claims 22 to 25, wherein each of the combustion chambers has at least two corresponding exhaust valves and the at least two in the combustion chamber associated with the combustion cycle. By independently controlling the intake valves so that the at least two intake valves follow different valve lift curves and / or have different opening / closing time points , the ignition control unit can output high or low torque. An ignition controller that adjusts the air supply in each combustion cycle to generate the combustion.   27. The ignition controller according to any one of claims 22 to 26, wherein each of the combustion chambers has at least two corresponding exhaust valves.   28. The ignition controller according to any one of claims 22 to 27, wherein the ignition level determination module uses a sigma delta converter to determine a combustion cycle to be ignited at a high output and a combustion cycle to be ignited at a low output. Features an ignition controller. 29. The ignition controller of claim 28, wherein the sigma delta converter is
Analog parts,
Digital parts,
With programmable logic,
An ignition controller, implemented using at least one of program instructions executed on a processor. 30. The ignition controller of any one of claims 22 to 29, wherein the ignition controller is in each low torque combustion cycle.
Supplying fuel to the corresponding combustion chamber and burning in the corresponding combustion chamber such that a net torque is delivered by the low torque combustion cycle so that the corresponding combustion cycle does not stop during the low torque combustion cycle. The ignition controller further comprising:

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